FR2731754A1 - Car friction clutch with pressure disc - Google Patents

Abstract

The car friction clutch has a pressure plate fixed but able to be axially displaced relative a housing. At least one actuating spring is located between housing and plate, which operates the pressure plate in the direction of a clutch plate. A wear compensating device ensures a practically constant force on the pressure plate via the spring. An operator is provided to engage and disengage the clutch, which during disengagement provides a progressive reduction in the transferred moment force.

Description

 The present invention relates to a friction clutch, in particular for motor vehicles, comprising a pressure plate which is connected in a non-rotating manner but nevertheless with the possibility of limited axial translation with a casing, between which and the pressure plate acts at least one spring thrust, which urges the pressure plate towards a clutch disc that can be clamped between this plate and a back pressure plate, like a flywheel.

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

 The object of the present invention is to improve friction clutches of this kind with regard to operation and service life. In particular, thanks to the invention, the forces necessary for the actuation of such friction clutches must be reduced and it must be ensured to obtain, during the service life of the clutches, a declutching force the appearance of which is practically constant. In addition, the friction clutches according to the invention must be able to be produced in a particularly simple and economical manner.

 According to the invention, this result is obtained by the fact that there is provided a corrective means compensating for the wear of the friction linings of the clutch disc and which produces a practically constant stress on the pressure plate by the push spring. , the friction clutch comprising actuation means for producing a clutch and a disengagement as well as a system which ensures, during the disengagement process, in a partial zone of the actuation stroke of the actuation means, a gradual reduction of the torque that can be transmitted by the friction clutch or by the clutch disc. Using a system of this kind, it is also possible to obtain that, during the clutch engagement process friction and at the start of tightening the friction linings between the pressure plate and the counterpressure plate, there is a gradual or progressive increase in the torque that can be transmitted by the friction clutch ion.

 With the arrangement according to the invention of a friction clutch, it is ensured that the annular tightening spring has almost always, during the service life of the friction clutch, the same preload when the friction clutch is engaged, thus making it possible to obtain a practically constant stress on the pressure plate. In addition, using the additional system provided, which ensures a gradual decrease in the torque that can be transmitted by the friction clutch during a clutch release process, it is possible to obtain a reduction, or a minimum change, in the release force or maximum release force required. This is due to the fact that the aforementioned system produces an actuation assistance of the friction clutch, in particular during the disengagement process. To this end, the aforementioned system can comprise means having an axial elasticity and which exerts on the means actuation and / or on the clamping spring and / or on the pressure plate and / or on the counterpressure plate, a reaction force which is oriented in the opposite direction to the force exerted by the clamping spring on the pressure plate and which intervenes accordingly.

 It may be particularly advantageous for the aforementioned system to be associated with the friction clutch so that, during the disengagement process, it produces in a partial region of the axial offset stroke parts of the pressure plate which are stressed by the clamping spring, a gradual reduction in torque that can be transmitted by the friction clutch or by the clutch disc.

 For many applications, the aforementioned system can advantageously be arranged in the force transmission line between the pivot bearings of the actuating means or else between the clamping spring and the fixing zones, such as bolts, of the casing on the back pressure plate.

 For other applications, it may however also be advantageous if the aforementioned system is arranged in the force transmission line between the pivot bearings of the actuating means or else between the clamping spring and the friction surface of the plate. pressure. An arrangement of this kind has been proposed for example in the German patent application DE-OS 37 42 354 and in the German patent application DE-OS 1 450 201.

 For other applications, it may be particularly advantageous for the aforementioned system to be disposed axially between two bearing parts of the friction linings provided on the clutch disc, that is to say being constituted by what is called a "pad support", for example by means of elastic segments disposed between the pads. Arrangements of this kind are known, for example, from German patent application DE-OS 36 31 863.

 Another possibility for obtaining a progressive increase or reduction of the torque has been proposed in the German patent application DE-OS 21 64 297, where the flywheel is divided into two parts and where the component forming the backpressure plate is supported axially and elastically with respect to the component connected to the output shaft of the internal combustion engine.

 For the operation and the arrangement of a friction clutch in accordance with the invention, it can be particularly advantageous that the aforementioned system allows a flexible axial adaptation between components of the clutch, this system being arranged and arranged in such a way that, when the clutch is open, the force acting on the aforementioned system is minimal and that, in the process of closing the clutch, that is to say in the clutch stroke of the clutch, the force acting on the above system gradually increases to the maximum, this rise advantageously occurring only in a partial area of the closing stroke, or clutch stroke, of the actuating means or of the pressure plate. It may be particularly advantageous for the above system to be arranged in such a way that the gradual increase or decrease in the torque which can be transmitted by the friction clutch takes place in at least approximately 40 to 70% of the stroke. actuation of the actuation means and / or of the maximum axial stroke of the pressure plate. The residual area of the corresponding stroke is necessary to cut the force transmission line correctly and to compensate for any deformations that may occur in the clutch components, such as the clutch disc, the pressure plate and the pressure plate. against pressure.

 In order to minimize the forces necessary for actuating the friction clutch according to the invention, it may be particularly advantageous for the clamping spring to include, at least in part of the clutch release stroke of the clutch. friction, a decreasing profile of the force-stroke curve, i.e. the clamping spring has, at least in a partial region of its compression or deformation stroke, a decreasing pace of force generation. can be obtained by ensuring that, in the disengagement process of the friction clutch, the force exerted by the spring of the aforementioned system opposes the force of the clamping spring so that, in a partial region of the stroke release, the tension or deformation of the clamping spring is assisted by the spring force of the aforementioned system, in which case simultaneously, due to the decreasing or decreasing profile of the spring-force curve manifested in the area of declutching, the force exerted by this spring on the pressure plate or on the friction linings decreases. The force effectively necessary for a declutching of the friction clutch will be determined, insofar as it does not there are no spring actions intervening additionally and overlapping, from the difference between the force exerted by the aforementioned system and the force exerted by the clamping spring.

When the pressure plate is moved away from the friction linings or when the clutch disc is released by the pressure plate, the clutch force must still be exerted or the necessary clutch force will be essentially determined by the clamping spring. The force-stroke characteristic curve of the aforementioned system and the force-stroke characteristic curve of the clamping spring can be mutually adapted as regards their gears so that, when the clutch disc is released by the pressure plate , the force required for actuation of the clamping spring has a comparatively low level. Thus, by causing the spring characteristic or the force characteristic of the aforementioned system to approach or even become identical to the characteristic of the clamping spring until the clutch plate is released by the pressure plate, it will be obtained that for the actuation of the clamping spring, only a very small force is necessary, and even practically no force in an extreme case.

 As a clamping spring, it is judicious and particularly advantageous to use an annular spring which can, on the one hand, pivot around an annular pivot bearing which is carried by the casing and which, on the other hand, urges the pressure. In this regard, the annular spring may include an annular body from which project radially inwardly of the tabs which constitute the actuating means. The actuating means can however also be constituted by levers, which are pivotally mounted for example on the casing.

The clamping force of the pressure plate can however also be exerted by other types of springs, such as for example helical springs, which are arranged in the friction clutch so that the axial force exerted by them on the plate pressure is greatest in the clutched state of the friction clutch and this force decreases during the disengagement process. This result can be obtained for example by giving the helical springs inclined positions relative to the axis of rotation of the friction clutch.

 I1 can be particularly advantageous that the annular spring is held in the casing in a pivoting manner between two supports so as to create what is called a pressure clutch. In such clutches, the actuating means involved in disengaging the friction clutch are usually biased towards the pressure plate.

The invention is however not limited to clutches of the type operating under pressure, but it also applies to clutches of the type operating under traction, in which the actuation means involved in the disengagement and the friction clutch are usually used away from the pressure plate.

 In a particularly advantageous manner, the friction clutch according to the invention can comprise an annular spring which is arranged so that it comprises a force-stroke curve of sinusoidal profile and which is mounted so that, in the 'clutched state of the friction clutch, its operating point is located on the decreasing part of the characteristic curve, following the first maximum force. In this regard, it may be particularly advantageous for the annular spring to have a force ratio of between 1: 0.4 and 1: 0.7 between the first maximum force and the minimum force located afterwards.

 It can also be particularly advantageous for the friction clutch to be actuated by means of a declutching device acting on the actuating means, such as for example on the ends of the legs of the annular spring, in which case the device clutch may include a clutch pedal, arranged in a similar manner to an accelerator pedal and arranged in the passenger compartment of the motor vehicle. Such an arrangement of the clutch pedal can be particularly advantageous because, with the arrangement conform according to the invention, the force, or the pace of force, necessary for disengaging the friction clutch can be maintained at a very small level, which makes it possible to obtain, by means of a clutch pedal arranged in a similar manner to an accelerator pedal, better ability to dose the actuating force.

 Thanks to the arrangement in accordance with the invention of a friction clutch and the corresponding possibility of reducing the maximum forces to be exerted on the clamping spring during the service life of this clutch, the components can be reduced correspondingly by as regards their dimensions or their strengths, which makes it possible to obtain a significant reduction in manufacturing costs. In addition, due to the reduction of the clutch forces, the friction and elasticity losses in the clutch and in the clutch system are reduced and thus the efficiency of the assembly constituted by the friction clutch and the device. clutch release has improved considerably. It is thus possible to arrange the assembly in an optimal manner and to considerably improve the comfort obtained with the clutch.

The arrangement according to the invention is applicable and in general to friction clutches and in particular to clutches such as those which have been proposed for example in German patents
DE-29 16 755, DE-29 20 932, in German patent applications DE-OS 35 18 781, DE-OS 40 92 382, in French patent applications FR-OS 2 605 692, FR-OS 2 606 477, FR-OS 2,599,444, FR-OS 2,599,446, in British patent GB-1,567,019, in the US Patents
United States US-4,924,991, US 4,191,285, US-4,057,131, in Japanese utility models JP-3-25026, JP3-123, JP-2-124326, JP-1-163218, JP- 3-19131, JP-3-53628 and in Japanese patent application JP-OS 51-126452.

 The present invention further relates to the older German patent applications DE-P 42 07 528.9 and DE-P 42 06 904.1, the contents of which relate expressly to the content published for the present invention.

 The use of a friction clutch provided with an integrated or automatic compensation system at least for the wear of the linings - so that one is guaranteed to obtain a tightening force, approximately constant at least for the duration of the friction clutch, of the clutch disc - is advantageous in particular in relation to clutch groups in which the friction clutch, the clutch disc and the back-pressure plate, such as for example a steering wheel, form a mounting unit or a module. In such an assembly unit, it is advantageous, for cost reasons, for the clutch housing to be connected to the back-pressure plate via a non-separable connection, such as for example a welded connection or a bonding by forming, for example by plastic deformation of material. With a bonding of this kind, it is possible to remove the fixing members commonly used, such as bolts. In such mounting units, it is practically impossible to replace the clutch disc or the clutch linings, due to exceeding the wear limits, without destroying the components, such as for example the clutch housing.

When using a clutch with a wear compensation system, the mounting unit can be arranged in such a way that it ensures correct operation during the entire service life of the vehicle.

Thus, thanks to the arrangement according to the invention, the wear reserve of the clutch disc and the wear correction reserve of the friction clutch and of the clutch module can be dimensioned sufficiently large so that the service life of the clutch, and therefore also the service life of the mounting unit, safely corresponds at least to that of the vehicle.

 According to another characteristic of the invention, it can be particularly advantageous that a friction clutch comprising a wear correction means is combined with a so-called dual-mass flywheel, in which case the friction clutch can be mounted, with the interposition of a clutch disc, on one of the inertia masses which can be connected to a transmission while the second inertia mass can be coupled to the output shaft of an internal combustion engine. Flywheels with two inertia masses, with which the friction clutch according to the invention can be used, are known for example from German patent applications DE-OS 37 21 712, 37 21 711, 41 17 571, 41 17 582 and 41 17 579. The entire content of said patent applications also corresponds to the published content of the present invention so that the features described in the aforementioned patent applications must be appropriately combined with the peculiarities described in the present patent application. In particular the clutch housing or the clutch cover can be connected to the inertia mass supporting them by means of a link which cannot be separated without destruction, as has been described and represented by example for different embodiments in German patent application DE-OS 41 17 579.

 Thanks to the combination of a friction clutch with a means making it possible to at least compensate for the wear of the friction linings, it is also possible to optimize the design of the friction clutch, in particular of the accumulator. energy exerting the clamping force of the clutch disc. This energy accumulator can then be arranged in such a way that it exerts practically only the clamping force of the clutch disc which is necessary for the transmission of the desired torque. The energy accumulator can be constituted by at least one annular spring or by several helical springs. In addition, it is advantageous to use a friction clutch with automatic wear compensation in combination with flywheels with two masses d inertia, in which elastic torsional dampers which are arranged between the two inertia masses are located radially outside the clutch disc or else the outer friction circle of the friction surface which can be coupled to the transmission . In flywheels with two flywheels of this type, the diameter of the friction circle of the clutch disc must be smaller than in conventional clutches so that the clamping force must be increased in correspondence with the ratio between the medium friction spokes in order to be able to transmit a motor torque of a defined value. When using a conventional clutch, this would lead to an increase in the release force. Thanks to the use of a clutch with automatic wear compensation and with progressive reduction of the torque that can be transmitted by the declutching disc during the declutching stroke, as is provided in accordance with the invention, it is possible to avoid an increase in the release force, unlike what occurs in a conventional clutch.

 It can thus be ensured, thanks to the arrangement according to the invention of a friction clutch, that despite a reduction in the outside diameter of the friction linings, and also in the clamping force necessary for this purpose, the force of clutch can be kept low. Thanks to this reduction in the declutching force, also the stress on the bearing by means of which the two masses of inertia are mounted so as to be able to rotate relative to one another, is reduced. In addition, thanks to the wear compensation system, the service life of the clutch is increased, so that there is no longer any need to replace the components, in particular the clutch disc, during the service life. of the motor vehicle. Thus the clutch cover can be rigidly connected to the inertia mass which can be coupled to the transmission, for example by means of rivets or welds. This is particularly advantageous when there is a limited installation volume or else limited contours of the clutch housing, which no longer allow a connection of the clutch cover to the flywheel in a conventional manner. located on the transmission side.

 In a friction clutch provided with an integrated means of correcting the wear of the friction linings, in the case of a conventional fixing of the clutch group consisting of the friction clutch and the flywheel on the shaft output of an internal combustion engine, axial, angular and three-dimensional oscillations are transmitted to the clutch group, being produced by the output shaft of the internal combustion engine, such as in particular a crankshaft.

In order that the clutch unit or the wear-correcting means are not disturbed in their function by such oscillations and so that in particular an undesirable corrective action of the wear-correcting means is prevented, it is necessary to take into account, when of the design of the corrective means of wear, of the inertia forces of the components which act on this corrective means. To avoid the undesirable side effects caused in particular by axial and three-dimensional vibrations, and also the corresponding higher costs of designing a corrective means serving to compensate for the wear of the friction linings, it is proposed, in accordance with another characteristic of the invention, to isolate to a large extent the clutch group comprising the corrective means for wear with respect to the oscillations axial and bending which are generated by the output shaft of the internal combustion engine. This can be achieved by ensuring that the clutch unit can be coupled to the output shaft of the internal combustion engine via a component which is axially elastic or which has axial flexibility. of this component is then defined so that the axial and three-dimensional or bending oscillations produced by the output shaft of the internal combustion engine on the clutch unit are counterbalanced by this elastic component, or at least damped to a degree such that it is ensured that the friction clutch operates correctly, in particular its wear correcting means. Such elastic components are known, for example, from European patent applications EP-OS 385 752 and 0 464 997 as well as from the document "SAE Technical Paper 9 003 91".

The content of these publications must also correspond to the published content of the present invention. Thanks to the use of an elastic component, it is possible to eliminate an undesirable wear compensation, caused by axial oscillations of the pressure plate relative to the clutch cover - especially when the friction clutch is disengaged - by oscillations of the flywheel and / or oscillations of the annular spring. Such oscillations can cause, in clutch groups or even clutch units not comprising a means eliminating at least essentially said oscillations, like in particular an axially flexible disc, a modification of the setting independently of the state of wear of the clutch disc, so that the annular spring of the friction clutch could undergo an adjustment producing a reduction in the tightening force , in such a way that the transmission of the desired torque would no longer be ensured.

 According to another feature of the invention, a friction clutch comprising an integrated or automatic wear compensation system, which can in particular be arranged in accordance with the present invention, can be used in a transmission group, in particular for motor vehicles, which consists of an automatic or semi-automatic transmission and a friction clutch, arranged between a drive motor, such as an internal combustion engine, and a transmission and which can be actuated by being controlled or adjusted to the less depending on the actuation of the transmission. The friction clutch can preferably be actuated completely automatically.

An automated or fully automatic actuation of a friction clutch has been described for example in German patent application DE-OS 40 11 850.9, so that reference will be made to this document with regard to the mode of action and the necessary means.

In the drive units known up to now and comprising automatic or semi-automatic transmissions and conventional friction clutches, there have hitherto been significant problems with regard to the actuation of the clutch and the design of the actuation means necessary for this purpose, such as for example piston / cylinder mechanisms and / or electric motors. Because of the relatively large declutching forces which are necessary in conventional clutches, it is necessary to provide actuators which are very powerful or which have large dimensions. This results in large dimensions, heavy weights and significant costs. Also such large actuating members have relatively long reaction times because of their weight inertia. When using actuating cylinders,
it is also necessary to involve a large flow of pressurized fluid so that also the feed pump must have relatively large dimensions, to safely obtain the desired actuation time of the corresponding friction clutch . To partially eliminate the aforementioned drawbacks, it has been proposed for example in the German patent application DE-OS 33 09 427 to reduce the actuating force necessary for disengaging the clutch by means of corresponding compensating springs, in order to be able to thus use small actuators. However, the declutching force to be exerted in conventional clutches, however, varies very greatly during service life, i.e. the declutching force is relatively small. in new condition and then increases during service life as the wear on the linings increases, a compensating spring only reduces part of the normally required declutching force. Taking into account all the tolerances, and despite the use of compensating springs, it would be necessary for the actuating members to have a declutching capacity which is greater than that necessary for a new conventional clutch. of a friction clutch according to the invention, comprising a system for compensating the wear of linings, in combination with a propulsion unit consisting of an engine and an automatic or semi-automatic transmission, the force of declutching can be considerably reduced compared to the prior art, and in particular directly in the clutch, in which case the value or shape of the declutching force obtained in the new clutch is practically maintained invariant throughout the service life of this Lastly, we obtain important advantages with regard to the design of the actuating members because their power of drive and their capacity actuation can be kept in correspondence at a reduced value, so that also the forces or pressures exerted in the entire clutch system are reduced in correspondence. As a result, losses occurring in the clutch system due to friction or elasticity of the components will be eliminated or reduced to a minimum.

Other characteristics and advantages of the invention will be highlighted in the following description, given by way of nonlimiting example, with reference to the accompanying drawings in which Figure 1 shows in elevation a friction clutch according to to the invention, Figure 2 is a sectional view taken along line II-II of Figure 1, Figure 3 shows an actuating ring used in the friction clutch of Figures 1 and 2, Figure 4 is a sectional view taken along the line
IV-IV of Figure 3, Figure 5 shows a support ring used in the friction clutch shown in Figures 1 and 2, Figure 6 is a sectional view taken along the line
VI-VI of Figure 5, Figures 7 and 7a show a spring which exerts a torsional force on the actuating ring, Figures 8 to 11 show diagrams with different characteristic curves from which we can deduce the cooperation between the different elastic and corrective elements involved in the friction clutch according to the invention, Figures 12 and 13 show another possibility of arrangement of a friction clutch according to the invention, Figure 13 being a view in section taken along line XIII of Figure 12, Figure 14 is an elevational view of the actuating ring used in the friction clutch of Figures 12 and 13, Figures 15 to 17 show details of another friction clutch comprising a compensating means, FIGS. 18 and 19 represent diagrams with different characteristic curves from which the cooperation of the annu spring can be deduced clamping area and friction lining supports as well as the resulting effect on the variation of the declutching force of the friction clutch, FIG. 20 is a partial elevation view of another friction clutch conforming to FIG. Figure 20a is a partial elevation view taken in the direction of arrow A in Figure 20, Figure 21 is a sectional view taken along the line
XXI of Figure 20, Figure 22 is a partial elevational view of an actuating ring usable in a friction clutch such as that of Figures 20 and 21, Figures 23 and 24 show other alternative embodiments of 'friction clutches according to the invention, Figure 25 is an elevational view of an actuating ring which would be usable in a friction clutch according to Figures 12 and 13 or 20 and 21, Figures 26 and 27 show other alternative embodiments of friction clutches, Figure 28 is an elevational view of a friction clutch according to the invention, Figure 29 is a sectional view taken along line II-II of Figure 28, Figure 30 is a sectional view taken along the line
III-III of Figure 28, Figure 31 is a partial sectional view taken along line IV-IV of Figure 28, Figure 32 is an elevational view of an actuating ring used in the clutch Figures 28 and 29, Figures 33 and 34 show details of other embodiments of friction clutches according to the invention, Figure 35 shows a flywheel divided into two masses of inertia, with a shock absorber torsional oscillations, as well as a friction clutch according to the present invention, and Figures 36 and 37 show a torque transmission device comprising a friction clutch according to the invention.

The friction clutch 1 shown in the
Figures 1 and 2 comprises a housing 2 and a pressure disc 3, connected in a non-rotating manner to the housing but which can, however, perform a limited axial translation. I1 is provided axially between the pressure disc 3 and the cover 2 an annular tightening spring 4, which can pivot around a pivoting bearing 5 of annular shape carried by the casing 2 and which biases the pressure disc 3 in the direction a back pressure plate 6, rigidly connected to the casing 2 by means of bolts 6a, this back pressure plate can for example be a flywheel, so that the friction linings 7 of the clutch disc 8 will be tightened between the friction surfaces of the pressure disc 3 and the back pressure plate 6.

 The pressure disc 3 is connected in a non-rotating manner to the casing 2 by means of elastic blades 9, oriented in a circumferential direction or tangentially. In the exemplary embodiment shown, the clutch disc 8 comprises what are called elastic lining segments 10 which, as is known, ensure a progressive increase in the torque when the friction clutch is clutched. 1, by the fact that they allow, over a limited axial offset of the two friction linings 7 towards one another, a gradual increase in the axial forces acting on the friction linings 7. And it is however possible to also use a clutch disc in which the friction linings 7 would be mounted axially practically rigidly on a carrier disc.

 In the embodiment shown, the annular spring 4 comprises a base body 4a of annular shape, receiving the clamping force and from which radially inwardly extend the actuating tabs 4b. The annular spring 4 is then mounted in such a way that it urges the pressure disc 3 through zones located radially more to the outside and that it can tilt around the pivot bearing 5 by zones situated radially more to the interior.

 The pivoting bearing 5 comprises two pivoting supports 11, 12, which in this case consist of rings of metal wire and between which the annular spring 4 is held or clamped axially.

The pivoting support 11 provided on the side of the annular spring 4 which is directed towards the pressure disc 3 is stressed axially by a force in the direction of the casing 2 by means of an energy accumulator 13.

The energy accumulator 13 is constituted by an annular spring or by a component 13 in the form of an annular spring, which bears by its outer marginal zone 13a against the casing 2 and which urges axially, by parts located radially more towards inside, the pivoting support 11 in the direction of the annular actuating spring 4 and consequently also in the direction of the casing 2. The annular spring 13, provided between the pressure disc 3 and the annular actuating spring 4, comprises an outer marginal zone 13b of annular shape, from the inner edge of which projecting radially inwardly of the tabs 13c which are applied against the pivoting support 11.

 For the support of the component 13 in the form of an annular spring, provision is made in the embodiment shown on the casing 2 for additional means 14 which constitute a pivoting support for the component 13 in the form of an annular spring. These additional means can be constituted by independent parts 14 in the form of segments which are attached or riveted and which can be distributed uniformly around the periphery. The means 14 may however be constituted by an intrinsically closed part and in the form of a circular ring. In addition, the support means 14 can be created directly by forming from the casing 2, for example by imprints created in an axial zone of the casing 2 or else by parts in the form of legs which, after the positioning and the clamping of the component 13 in the form of an annular spring, are forced back below the outer edge of this component 13 by deformation of material. In addition, there may exist between the support means 14 and the component 13 in the form of an annular spring a connection or locking of the bar-type so that the component 13 in the form of an annular spring is initially prestressed and its radially outer zones can be applied axially on the support means 14.Then, by producing a corresponding rotation of the component 13 in the form of an annular spring with respect to the casing 2, it is possible to apply the support zones of the component 13 against the means of support 14. The support zones of the component in the form of an annular spring 13 can be formed in this respect by tabs projecting radially outward from the base body 13b of annular shape.

 For anti-rotation security of the annular actuating spring 4 and, where appropriate, of the component 13 in the form of an annular spring as well as for centering of the metal wire rings 11, 12, centering means are provided, s extending axially and in the form of rivets 15.

The rivets 15 each comprise an axially extending rod 15a which passes axially through a recess provided between adjacent tabs 4b of the annular spring and which can be held in part by zones 13d formed on the associated tab 13c of the annular spring 13 .

 The annular spring component or the annular spring 13 is arranged as a detector spring, which produces over a predetermined working distance at least approximately substantially constant force. This spring 13 makes it possible to absorb the declutching force acting on the ends 4c of the abovementioned tabs, so that an at least approximate balance is constantly established between the force produced during the declutching on the pivoting support 11 and the the opposing force exerted by the annular spring 13 on this pivoting support 11. By the expression "disengaging force" is meant the maximum force which is exerted, during the actuation of the friction clutch 1, on the ends of legs 4c or else on the parts forming disengaging levers) of the legs of the annular spring.

 The pivoting support 12 located on the housing side is mounted by means of a corrective or adaptation means 16 on the housing 2. This corrective or adaptation means 16 ensures that, in the event of an offset axial of the pivoting supports 11 and 12 in the direction of the pressure disc 3 or in the direction of the back-pressure plate 6, there can be no unintended play between the pivoting support 12 and the casing 2 or between the pivoting support 12 and the annular spring 4. It is thus ensured that an idle stroke does not occur unintentionally when the friction clutch 1 is actuated, so that one is assured of optimum performance and therefore of correct actuation of the friction clutch 1.The axial offset of the pivot bearings 11 and 12 occurs during axial wear on the friction surfaces of the pressure disc 3 and the back pressure plate 6 and also during wear of the linings ures de friction 7. The mode of action of the automatic correction system of the pivot bearing 5 will be further explained in detail in relation to the diagrams shown in Figures 8 to 11.

 The corrective means 16 comprises a corrective element biased by a spring and being in the form of an annular component 17, which is shown in FIGS. 3 and 4. This annular component 17 comprises contact ramps 18, oriented in a circumferential direction , inclined axially upwards and which are distributed around the periphery of this component 17. This corrective component 17 is mounted in the clutch 1 so that the contact ramps 18 are directed towards the bottom 2a of the casing. On the side of the corrective element 17 which is opposite to the contact rails 18, the pivoting support 12, constituted by a wire ring, is positioned centrally in a groove-shaped housing 19 (Figure 2) .

The housing 19 can be arranged in this respect so that the pivoting support 12 is also fixed in an axial direction on the corrective element 17.

This result can for example be obtained by ensuring that at least in places, the zones of the corrective element 17 which are adjacent to the housing 19 securely hold the pivoting support 12 or else form a nesting connection for the pivot support 12. When using different materials for the pivot support 12 and the corrective element 17, it may be advisable, to compensate for the differences in expansion occurring for large variations in temperature, that the 'pivoting support 12 arranged in the form of a metal wire ring is open, that is to say that it is divided at least in one place around its periphery, in order to allow movement of the ring 12 relative to the housing 19 in a circumferential direction and thus allow adaptation of the wire ring 12 to the diameter of the housing 19.

 In the embodiment shown, the corrective element 17 is made of plastic, such as for example a heat-resistant thermoplastic which can additionally be further reinforced by fibers. Consequently, the corrective element 17 can be manufactured in a simple manner in the form of an injection molded part. The corrective element 17 can however also be manufactured in the form of a profiled piece of sheet metal or else by sintering. In addition, during a corresponding selection of materials, the pivoting support 12 can form a single piece with the corrective element 17.

The pivoting support 11 may be formed directly by the detector spring 13. For this purpose, the ends of the legs 13c may include imprints or corresponding deformations, such as for example moldings.

 The corrective ring 17 is centered by the axially oriented zones 15a of the rivets 15 distributed uniformly around the periphery. In this regard, the corrective ring 17 has centering contours 20 which are constituted by recesses extending in a circumferential direction and which are located radially inside the pivoting support 11. To create recesses 21 , the corrective ring 17 comprises, in an inner marginal zone, protuberances 22 extending radially inwards and which radially delimit the internal contours of the recesses 21.

 As shown in Figure 3, there is provided, considering a circumferential direction, respectively five contact ramps 18 between the recesses 21 distributed uniformly. The recesses 21 are profiled in a circumferential direction so that they allow, between the corrective ring 17 and the casing 2, at least a relative angle of rotation which guarantees, throughout the service life of the friction clutch 1, a corrective compensation for the wear and tear occurring on the friction surfaces of the pressure disc 3 and of the backpressure plate 6 and also on the friction linings 7. This angle of corrective compensation can, depending on the arrangement of the contact rails, have an order of magnitude between 8 and 60 degrees, preferably an order of magnitude between 10 and 30 degrees. In the embodiment shown, this relative angle of rotation is of the order of 12 degrees, in which case the angle of application 23 of the contact ramps 18 is also of the order of 12 degrees. This angle 23 is chosen so that the friction occurring during mutual application of the contact ramps 18 of the corrective ring 17 and the opposing ramps 24 of the support ring 25 shown in Figures 5 and 6 prevents a slip between the contact ramps 18 and 24. Depending on the pair of materials involved in the area of the contact ramps 18 and the opposing contact ramps 24, the angle 23 can be between 5 and 20 degrees.

 The corrective ring 17 is biased by spring in a circumferential direction, and in particular in the direction of rotation for corrective compensation, that is to say in the direction which produces, by coming into contact of the ramps 18 with the counter-ramps 24 of the support ring 25, an axial offset of the corrective ring in the direction of the pressure disc 3, that is to say which produces in an axial direction a spacing with respect to the radial part 2a of the casing. 'embodiment shown in Figures 1 and 2, the elastic stress of the corrective ring 17 is provided by at least one branch spring 26 of annular shape, which may for example have two turns and which is provided with one of its ends of a branch 27 oriented radially, which is connected in a non-rotating manner to the corrective ring 17, while the spring comprises at the other end a branch 28 oriented axially, which is hooked in a non-rotatable manner urnant on the casing 2. The spring 27 is mounted with elastic preload.

 The support ring 25 shown in FIGS. 5 and 6 is also constituted by an annular piece, which includes contact counter-ramps 24 which create surfaces complementary to the surfaces delimited by the contact ramps 18, in which case the surfaces delimited by the contact ramps 18 and the contact counter-ramps 24 can also be congruent. The angle of inclination 29 of the contact counter-ramps 24 corresponds to the angle of inclination 23 of the contact ramps 18. As shown in a comparison of FIGS. 3 and 5, the contact ramps 18 and the contact counter-ramps 24 are similarly distributed in a circumferential direction.

The support ring 25 is connected in a non-rotating manner to the casing 2. For this purpose, the support ring 25 has on its periphery recesses 30 in which the ends of the rivets 15 engage.

 In Figure 2, there is shown by broken lines another branch spring 26a of annular shape, which can be bent in its end portions, like the branch spring 26, to ensure a non-rotating connection with d on the one hand the casing 2 and on the other hand the corrective element 17.

This spring 26a is also mounted with elastic preload so that a torsional force is exerted on the corrective element 17. The use of two branch springs 26, 26a can be advantageous in many applications because, during a rotation of the friction clutch 1 under the effect of centrifugal forces acting on the springs 26 or 26a, there is an amplification of the force of the springs.

Thanks to the use of two branch springs, it is possible to compensate for the amplification of force exerted for example on the spring 26 by means of the force exerted by the branch spring 26a. To this end, the branch springs 26 and 26a are wound in such a way that they produce on the corrective element 17, at least under the effect of centrifugal forces, forces which act in mutual opposition in the circumferential direction. two branch springs 26, 26a may have one or more turns and in addition these branch springs 26, 26a may have different turns diameters, as shown in Figure 2, in which case the centrifugal forces normally acting on the springs 26, 26a and which would produce circumferential forces of different sizes on the corrective element 17, can be compensated, at least approximately, by a corresponding arrangement of the thicknesses of the metal wires and / or by a corresponding choice of the number of turns of the different springs 26, 26a. In FIG. 2, the spring 26 is arranged radially inside the corrective element 17 and the spring 26a is arranged radially outside the corrective element 17.

The two springs could however, in a corresponding arrangement, also be arranged radially inside or radially outside the corrective element 17.

 In Figure 7, the branch spring 26 is shown in plan view. In the relaxed state of the branch spring 26, the branches 27, 28 are offset by an angle 31 which can have an order of magnitude between 40 and 120 degrees. In the embodiment shown, this angle 31 has an order of magnitude of 85 degrees. The relative position of the branch 27 relative to the branch 28 has been designated by 32, which the latter occupies when new friction linings 7 are mounted in the friction clutch 1. The position of the branch 27 which corresponds to the maximum admissible wear on the friction linings 7. The correction angle 34 has an order of magnitude of 12 degrees in the embodiment shown. The spring 26 is arranged so that, in the relaxed state of this spring 26, there is only one turn of metal wire 35 between the two branches 27, 28.

In the remaining area of the periphery, two turns are located axially one above the other. The spring 26a is arranged in a similar manner to the spring 26 but it however has a larger turn diameter and another direction of prestressing than that of the corrective element 17 of FIG. 2. The force exerted by the spring 26 on the corrective ring 17 is however larger than that exerted by the spring 26a.

 In the new state of the friction clutches 1, the contact ramps 18 and the contact counter-ramps 24 forming axial protrusions 18a, 24a engage axially as much as possible in each other, which means that the rings 17 and 25 located one on the other require for their mounting a minimum axial volume.

 In the embodiment of Figures 1 and 2, the contact counter-ramps 24, or the protuberances in the form of bosses 24a constituting them, are created by means of an independent part. The contact counteramps 24 can however be directly created by the casing 2, for example by imprinting protuberances which can penetrate into the volume delimited by the casing. The impression operation is particularly advantageous in the case of sheet metal sensors or covers which are made in one piece.

 To maintain the adjusting ring 17, before mounting the friction clutch 1, in its retracted position, this comprises, in the region of the protrusions 22, attachment regions 26 for a retaining means, which can also bear against the casing 2. Retaining means of this kind can be provided during the manufacture or assembly of the friction clutch 1 and be removed after mounting the friction clutch 1 on the steering wheel 6, so as to allow the activation of the corrective device 16. In the embodiment shown, there is provided for this purpose in the cover or the housing 2 of the recesses 37 elongated in a circumferential direction and there is provided in the corrective ring 17 a hollow or a recess 38. The elongated recesses 37 distributed in a circumferential direction must have in this regard at least one dimension such that the corrective ring 17 can be turned in retraction in cor responds to the largest possible wear correction angle. Also after assembly of the friction clutch 1, a tool can be engaged axially through the slots 37 of the cover and be introduced into the recesses 38 of the corrective ring 17. Then this ring 17 can be rotated in retraction. means of the tool so that it is offset in the direction of the radial zone 2a of the casing 2 and that it takes its minimum axial spacing with respect to this zone 2a.

The corrective ring 17 is then fixed in this position, for example by means of a fastener or a pin which is engaged in recesses located in coincidence in the cover and the corrective ring 17 and which prevents rotation relative of these two pieces.

This pin can be removed after mounting the friction clutch 1 on the flywheel 6 so that, as already stated, the corrective device 16 is released. The slots 37 provided in the casing 2 are arranged such that, during disassembly or after disassembly of the friction clutch 1 relative to the flywheel 6, the corrective ring 17 can be brought into its retracted position. To this end, the clutch 1 is initially disengaged so that the annular actuating spring 4 exerts no axial force on the pivoting support 12 and thus ensures correct rotation of the corrective ring 17 is ensured. .

 We will now describe in more detail, in relation to the characteristic curves indicated in the diagrams of Figures 8 to 11, the operating mode of the friction clutch 1 according to the invention.

 The curve 40 in FIG. 8 represents the axial force produced as a function of the variation in conicity of the annular spring 4, and in particular in the event of deformation of this annular spring 4 between two supports, the radial spacing of which corresponds to the radial spacing between the pivot bearing 5 and the radially outer bearing diameter 3a against the pressure disc 3. On the abscissa, the relative axial displacement between the two supports is plotted and on the ordinate the force produced by the annular spring. Point 41 represents the position occupied by the annular spring 4 when the clutch 1 is closed, i.e. the position in which the annular spring 4 exerts, in the corresponding position, the maximum clamping force on the pressure disc 3. Point 41 can be shifted up or down along curve 40 by modifying the conical mounting position of the annular spring 4.

 The curve 42 represents the axial separation force which is absorbed by the elastic segments 10 for supporting friction linings, this force acting between the two friction linings 7. This axial separation force acts in opposition to the axial force exerted by the annular spring 4 on the pressure disc 3. It is advantageous that the axial force which can be absorbed by elastic deformation of the elastic segments 10 corresponds at least to the force exerted by the annular spring 4 on the pressure disc 3, this force which can advantageously also be greater. During disengagement of the friction clutch 1, the elastic segments 10 relax, and in particular over the distance 43. Over this distance 43, corresponding to a corresponding axial offset of the pressure disc 3, the disengagement process of the clutch 1 is assisted, which thus means that a maximum release force must be exerted which is smaller than that which would correspond to the mounting point 41 in the event of the non-existence of the elastic segments 10 supporting the friction linings. When the point 44 is exceeded, the friction linings 7 are released and, in correspondence with the decreasing zone of the characteristic curve of the annular spring 4, the declutching force still to be exerted is considerably reduced compared to that which would correspond to the point 41. The declutching force of the clutch 1 decreases until the minimum or the low point 45 of the characteristic curve 40 of sinusoidal shape is reached.

When the minimum 45 is exceeded, the required declutching force increases again, the declutching stroke in the region of the leg ends 4c being chosen so that, even if the minimum 45 is exceeded, the declutching force does not exceed the maximum declutching force corresponding to point 44 but preferably remains below this force. So point 46 should not be exceeded.

 The spring 13 serving as a force detector has a stroke-force curve corresponding to the curve 47 in FIG. 9. This characteristic curve 47 corresponds to that which is produced when the taper of the component 13 in the form of an annular spring is modified from the relaxed position, and in particular between two pivot supports, which are radially spaced from each other by a distance which corresponds to the radial spacing between the pivot supports 11 and 14. As shown by the characteristic curve 47, the component 13 in the form of an annular spring has an elastic stroke 48, in which the axial force produced by it remains practically constant. The force produced in this zone 48 is then chosen so that it corresponds at least approximately to the clutch release force corresponding to point 44 in Figure 8. The support force to be exerted by the sensor spring 13 is reduced, compared to the force of the annular spring 4 corresponding to point 44, in correspondence with the lever arm or transmission ratio of this annular spring 4.

This transmission ratio is in most cases of an order of magnitude between 1: 3 and 1: 5 but however for many applications it can also be larger or smaller.

 The aforementioned transmission ratio concerning the annular spring corresponds to the ratio between the radial spacing of the pivoting support 5 with respect to the support 3a and the radial spacing of the pivoting support 5 with respect to the diameter of the circle d 'support 4c, for example for a declutching bearing.

 The mounting position of the element 13 in the form of an annular spring in the friction clutch 1 is chosen so that this element can perform, in the region of the pivoting support 5, an axial elastic stroke in the direction of the friction linings 7, this stroke corresponding on the one hand at least to the axial course of corrective adaptation of the pressure disc 3 in the direction of the back pressure plate 6, which results from the wear of the friction surfaces and of the linings friction, and also ensuring the generation of an axial support force at least approximately constant for the pivot support 5.

This means that the linear zone 48 of the characteristic curve 47 should have at least a length corresponding to the aforementioned wear distance, preferably being greater than this wear distance, because thus it would also be possible to compensate at least partially mounting tolerances.

 To obtain a substantially constant or defined release point 44 of the friction linings 7 when the friction clutch 1 is disengaged, it is possible to use a so-called double segment suspension between the friction linings 7, that is to say a suspension of linings in which different spring segments are provided in pairs back to back, the different pairs of segments being able to comprise a certain axial prestress with respect to each other so that the force axial exerted in total by the suspension of linings corresponds at least, when the clutch disc 8 is not tightened, to the disengagement force exerted, in correspondence at point 44, on the annular spring 4, being of preferably slightly higher. By means of prestressing the elastic means provided between the linings, it is possible to obtain that the losses occurring under the effect of penetration of the segments into the c at the rear removed from the linings during the service life are at least substantially offset or counterbalanced. By the expression "losses due to penetration" is meant losses resulting from a depression or penetration of the segments in the rear side of the linings. It is advisable that the prestress in the elastic suspension provided between the linings is of the order of magnitude from 0.3 mm to 0.8 mm, preferably of the order of magnitude of 0.5 mm. By a corresponding limitation of the axial elastic stroke between the two friction linings 7 and also by a defined prestress of the elastic suspension acting between the friction linings, it is also possible to obtain that at least when the friction clutch 1 is disengaged , the pressure plate 3 is retracted by a defined distance 43 by the elastic suspension provided between the friction linings. To obtain a defined distance 43, the axial travel between the friction linings can be limited by corresponding stops both in the direction of relaxation and in the direction of contraction of the elastic suspension 10 of the friction linings. , suspensions such as those known for example from German patent application P 42 06 880.0, which is expressly related to the subject of this application, can advantageously be used in connection with the present invention. patent.

 To be sure of an optimal functioning of the friction clutch 1 or of the corrective device ensuring an automatic compensation of the wear of the friction linings, it is obvious that, by considering the curve 49 of clutch release force of the Figure 10, the forces initially exerted by the elastic suspension 10 of the friction lining and the detector spring 13 on the annular spring 4 and thus adding up the force exerted on the annular spring 4 only by the detector spring 13 after the spacing of the pressure disc 3 with respect to the friction linings 7, are greater, but nevertheless at least equal, to the declutching force exerted in the zone 4c of the ends of the legs of the annular spring and modifying during the stroke release in correspondence to Figure 10.

 The considerations made up to now correspond to a well-determined mounting position of the annular spring 4 and we have not yet taken account of wear and tear occurring on the friction linings 7.

 In the event of axial wear, in particular of the friction linings 7, the position of the pressure disc 3 is shifted in the direction of the back pressure plate 6, so that there is a change in the taper and therefore also of the clamping force exerted by the annular spring in the engaged state of the friction clutch 1, and in particular in the direction of an increase. This modification makes point 41 move in the direction of point 41 'and point 44 moves in the direction of point 44'. Under the effect of this modification, the balance of forces which existed initially when the clutch 1 was disengaged is suppressed in the region of the pivoting support 11 existing between the annular actuating spring 4 and the detector spring. 13.

The increase, caused by wear of the linings, of the clamping force exerted by the annular spring on the pressure disc 3 also produces a change in the evolution of the declutching force in the direction of an increase. The evolution of the declutching force thus generated is shown in Figure 10 by the dashed line curve 50. Under the effect of the increase of this declutching force, it occurs, during the declutching process of the friction clutch 1, a reaction on the axial force exerted by the detector spring 13 on the annular spring 4 so that the detector spring 13 retracts, essentially in the region of the pivoting support 5, d 'An axial distance corresponding essentially to the wear of the friction linings 7.During this retraction phase of the detector spring 13, the annular spring 4 bears against the stressing zone 3a of the pressure disc 3 so that this annular spring 4 has its taper which is modified, as also the energy stored in this spring, or the torque stored in it, and therefore also the force exerted by the annular spring 4 on the pivoting support 11 or the detector spring 13 and on the pressure disc 3. This modification occurs as described in relation to FIG. 8 in the direction of a reduction in the forces exerted by the annular spring 4.This modification occurs until the axial force exerted by the annular spring 4, in the region of the pivoting support 11, on the detector spring 13 is in equilibrium with the opposing force produced by the detector spring 13. This means that, on the diagram in Figure 8, points 41 'and 44' will again move towards points 41 and 44. After this balance has been restored, the pressure disc 3 can again be removed from the friction linings 7. During this phase of corrective wear compensation during a disengagement process of the friction clutch 1, the corrective element 17 of the wear corrective device 16 is moved in rotation by the prestressed spring 26, so that also the pivoting support 12 is offset in correspondence with the wear of the friction linings and that one is thus guaranteed to obtain for the annular spring 4 a pivoting support 5 free of play. Following the wear compensation process, the evolution of the declutching force again corresponds to curve 49 in Figure 10. Curves 50 and 51 in Figure 10 represent the axial travel of the disc. pressure 3 in the case where the release force varies as a function of the stroke in correspondence with curves 49, 50.

 In the diagram in FIG. 11, the variation of the force exerted, in a declutching process, has been represented graphically on the casing 2 or else on the spring 13, the extreme values not being highlighted. From the engaged position corresponding to Figure 1, the housing 2 and therefore also the pressure disc 3 are initially subjected to a force which corresponds to the mounting point 41 (Figure 8) of the annular spring 4. During the process of declutching, the axial force exerted by the annular spring 4 on the casing 2 or on the pivoting support 12 decreases in correspondence with the curve 42 of FIG. 11, and in particular up to point 53. 53 in the declutching direction, one would obtain, in the case of a conventional clutch where the annular spring is pivotally mounted at an axially fixed point on the casing, that is to say where the pivoting support 11 is axially non-flexibly connected with the casing 2, an axial reversal of the direction of action of the force exerted by the annular spring 4 on the casing 2 at the radial level of the pivoting support 5. In the clutch according to the invention, in the area of the pivoting support 5, the axial reversal of the direction of the force produced by the annular spring 4 is absorbed by the detector spring 13.

When point 54 is reached, the annular spring 4 moves away from the stressing zone 3a of the pressure disc 3. Up to at least at this point 54, the disengagement process of the friction clutch 1 is assisted by the axial force exerted by the elastic suspension 10 of the friction linings. The force exerted by this elastic suspension 10 decreases as the declutching stroke increases in the area 4c of the tab ends or else as the axial disengaging stroke of the pressure disc 3 increases. The curve 52 thus represents a result, manifested in the disengagement process, between on the one hand the disengagement force acting in the zone 4c of the tab ends and on the other hand the axial force exerted in the radial zone 3a on the annular spring 4 by the elastic suspension 10 of the friction linings.

When the point 54 is exceeded in the declutching direction, the axial force exerted by the annular spring 4 on the pivoting support 11 is counterbalanced by the opposing force exerted by the annular spring 13 forming a detector, these two forces being in equilibrium at least after the discharge of the friction linings 7 by the pressure disc 3 and, during the continuation of the declutching process, the axial force exerted by the detector spring 13 in the region of the pivoting support 5 advantageously becomes slightly greater than the release force applied.

The partial area 55 of the characteristic curve 52 of the diagram in FIG. 11 shows that, when the declutching stroke increases, the declutching force or else the force exerted by the annular spring 4 on the pivoting support 11 becomes smaller than the declutching force corresponding to point 54. The dashed line curve 56 corresponds to a state of the friction clutch 1 in which wear has occurred in an area of the friction linings 7 but where, however, no corrective action was carried out in the area of the pivoting support 5. It can also be seen in this case that the variation in the mounting position of the annular spring 4 which is caused by wear generates an increase in the forces exerted on the casing 2 and on the pivoting support 11 or else on the detector spring 13. This has in particular the consequence that the point 54 is moved in the direction of the point 54 ′, which means that, when a new proce ssus disengaging the friction clutch 1, the axial force exerted by the annular spring 4 on the sensor spring 13 in the area of the pivoting support 11 will become greater than the opposing force exerted by this spring 13, so that the corrective compensation process described above is produced by an axial expansion of the detector spring 13. By this corrective compensation process, the point 54 'is again shifted towards the point 54, so that the state desired balance is restored in the region of the pivot bearing 5 between the annular spring 4 and the detector spring 13.

 In practice, the corrective action described above takes place continuously or else in very small steps so that the large shifts of points and characteristic curves represented on the diagrams with a view to a better understanding of the invention do not occur. normally not occur.

 During the service life of the friction clutch 1, some functional parameters or operating points may be changed.

Thus for example, during an incorrect operation of the friction clutch 1, there may be an overheating of the suspension of linings 10, which can result in compression, that is to say a decrease in the axial elasticity of the suspension of linings 10. By means of a corresponding profiling of the characteristic curve 40 of the annular spring 4 and of a corresponding adaptation of the elasticity curve 47 of the detector spring 13, it is however possible to '' obtain safe operation of the friction clutch. Axial compression of the lining suspension 10 would only result in the annular spring 4 assuming a lowered position compared to the position shown in the
Figure 1, in which case the clamping force exerted by the annular spring 4 on the pressure disc would be slightly lower, as shown in relation to the characteristic curve 40 of Figure 8. In addition, there would be a corresponding axial deformation of the detector spring 13 and consequently a corresponding axial offset of the pivot support 11.

 According to another feature of the invention, the resulting support force which acts on the annular actuating spring 4 can increase as the wear of the friction linings 7 increases. This increase can then be limited to a part of the maximum admissible wear stroke for the friction linings 7. The increase in the support force for the annular actuating spring 4 can then be produced by means of an arrangement corresponding to the detector spring 13. In Figure 9, there is shown in broken lines, and designated by the reference 47a, a corresponding profile of the characteristic curve in the area 48. By means of an increase in the support force for the actuating annular spring 4 as wear increases, it is possible to at least partially compensate for a reduction in the clamping force exerted by the actuating annular spring 4 on the pressure plate 3, caused by a reduction in the efficiency of the suspension of friction linings, for example due to a depression of the segments in the linings.

 It can then be particularly advantageous for the support force of the annular actuating spring 4 to increase in proportion to the compression of the suspension of friction linings or alternatively in proportion to the depression of the segments in the friction linings. This means that when the thickness of the disc is reduced in the area of the linings, that is to say the reduction in the spacing between the friction surfaces of the linings due to the depression of the segments and / or compression of the lining suspension and / or wear of the linings, the abovementioned support force must increase.

In this regard, it is particularly advantageous that the increase in force is carried out in such a way that it is greater in a first partial zone than in a second partial zone following the first, the two aforementioned partial zones being located at the interior of the area designated by 48 in Figure 9. This latter arrangement is advantageous since the greater part of the abovementioned depression occurring between the elastic segments and the friction linings occurs mainly for a short period of time. compared to the total service life of the friction clutch and that thereafter the relations between the elastic segments and the friction linings practically stabilize. This means that, from a specific depression, there will no longer be any significant variation with regard to this depression.

An increase in the support force for the annular actuating spring can also occur for at least part of the wear of the friction linings.

 During the previous description of the correction process used to compensate for the wear of the friction linings, the axial forces possibly exerted by the elastic blades 9 were not taken into consideration. In the case of prestressing of the elastic blades 9 in the direction of a spacing of the pressure disc 3 relative to the corresponding friction linings 7, that is to say in the direction of compression of the pressure disc 3 against the annular spring 4, assistance occurs in the disengagement process. The axial force exerted by the elastic blades 9 is added to the forces exerted by the detector spring 13 and the annular spring 4 as well as to the declutching force. This was not taken into account when describing the diagrams in Figures 8 to 11 for better understanding.

The total force urging the annular actuating spring 4, in the disengaged state of the friction clutch 1, to apply it against the rolling element support 12 situated on the side of the cover is produced by adding the forces which will be exerted on the annular actuating spring 4 by the elastic blades 9, by the detector spring 13 and by the existing declutching force. The elastic blades 9 can then be placed between the cover 2 and the pressure plate 3 so that, when the wear of the friction linings 7 increases, the axial force exerted by the elastic blades 9 on the annular actuating spring 4 becomes larger. Thus, for example on the curve portion 48 of FIG. 9, and consequently also on the wear compensation stroke of the corrective means 16, the axial force exerted by the elastic blades 9 can vary in correspondence with curve 47b. Figure 9 also shows that, as the elasticity stroke of the detector spring 13 increases, there is an increase in the restoring force exerted by the elastic blades 9 on the pressure plate 3, this force also acting on the annular actuating spring 4. By a combination of the force variations corresponding to the characteristic curves 47b and to the characteristic curve of the annular spring, the resulting variation of the force acting axially on the annular spring 4 is obtained, and in particular in the direction of an application of the annular spring 4 against the pivoting support 12 located on the side of the cover. To obtain a variation corresponding to curve 47a, it is advisable to arrange the annular spring forming a detector so that it has a characteristic curve of a profile corresponding to that of curve 47c of the
Figure 9. By combining the force variation profile corresponding to curve 47c and the force variation profile corresponding to curve 47b, we then obtain the force variation profile corresponding to curve 47a. Thus, by means of a corresponding prestress of the elastic blades 9, it is possible to reduce the support force to be exerted by the detector-forming spring or else the profile of the support force. By a corresponding arrangement and arrangement of the elastic blades 9, it is also possible to at least partially compensate for a reduction in elasticity of the suspension of friction linings and / or a sinking of the elastic segments of friction linings in the linings. We can thus ensure that the annular spring 4 retains essentially the same operating point or the same operating range so that the annular spring 4 exerts, during the service life of the friction clutch, essentially an at least approximately constant clamping force on the pressure plate 3. Furthermore, when designing the friction clutch, in particular the detector spring 13 and / or the elastic blades 9, the force must be taken into account resulting axial produced by the compensating springs 26 and / or 26a acting on the corrective element 17 and which acts in the opposite direction to the forces produced by the spring 13 and / or the elastic blades 9.

 During an arrangement of the friction clutch 1 with elastic prestressed blades 9, it must also be taken into account that the axial force exerted by the pressure plate 3 on the friction linings 7 is influenced by this prestressing of the blades elastic 9. This means that, when the elastic blades 9 are preloaded towards the annular actuating spring 4, the clamping force exerted by the annular spring 4 will be reduced by a value corresponding to the preload force of the blades elastic bands 9. I1 thus establishes, in a friction clutch 1 of this kind, a profile resulting from variation of the clamping force of the pressure plate 3 or friction linings 7, which is obtained by combining the variation profile of the clamping force of the annular spring 4 with the tension variation profile of the elastic blades 9. Assuming that - considering the operating range of the friction clutch on 1 - the characteristic curve 40 of FIG. 8 represents the profile resulting from variation of forces of the annular actuating spring 4 and of the pre-stressed elastic blades 9 when the friction clutch 1 is new, one would obtain, during a decrease in the spacing between the pressure plate 3 and the back pressure plate 6 as a result of the wear of the friction linings, a shift in the variation profile resulting in the direction of a decrease. In Figure 8, there is shown in broken lines a curve 40a which corresponds for example to a total wear of 1.5 mm of the friction linings. Under the effect of this offset of the curve 40 in the direction of the curve 40a during the service life of the friction clutch, there is a reduction in the axial force exerted when disengaging the friction clutch 1 , by the annular spring 4 on the detector spring 13, and in particular because of the opposing torque exerted, during an increase in wear, by the elastic blades 9 on the annular spring 4. This opposing torque is produced as a function the radial spacing between the pivoting support 5 and the diameter 3a of the biasing circle between the annular actuating spring 4 and the pressure plate 3. When designing the friction clutch 1, it is particularly it is important to ensure that the increase in the tension of the elastic blades 9 resulting from the wear of the friction linings is smaller than the increase in the declutching force caused by the same wear of the friction linings and which produces, for the detector spring 13, the pivoting necessary for corrective wear compensation. Otherwise, the clamping force exerted by the pressure plate 3 on the friction linings 7 would decrease in the engaged state of the friction clutch and there could practically no longer be any corrective wear compensation.

The friction clutch 101 shown in the
Figures 12 and 13 differs in essence from the friction clutch 1 shown in Figures 1 and 2 in that the corrective ring 117 is biased in a circumferential direction by helical springs 126. With regard to its function and its mode of action to compensate for the wear of the friction linings, the corrective ring 117 corresponds to the corrective ring 17 of FIGS. 2 to 4. In the embodiment shown, three helical springs 126 are provided. which are uniformly distributed around the periphery and which are prestressed between the clutch housing 2 and the corrective ring 117.

 As can be seen in particular from FIG. 14, the corrective ring 117 has on its inner periphery radial projections or protrusions 127, against which the helical springs 126, which are arranged with an arc shape, can bear, by one of their ends with a view to biasing the corrective ring 117. The other end parts of the springs 126 bear against the stops 128 carried by the clutch housing 2. In the embodiment shown, these stops 128 are constituted by connecting elements in the form of bolts, which are connected to the cover 2. These stops 128 can however also be constituted by axial bosses which form a single piece with the clutch housing 2. Thus for example the stops 128 can be formed by indentations or bosses created by axial discharge of material in a sheet metal casing 2. As shown in particular in Figures 13 and 14, the ring 117 may be arranged on its inner periphery such that at least in the main in the area of extension of the springs 126 and preferably also in the torsion angle of the ring 117 necessary for compensation for wear , or else on the spring expansion stroke 126, a guide 129 is provided which provides axial support and radial support for the springs 126. The spring guides 129 are constituted, in the embodiment shown, by recesses the cross-section of which has essentially a semi-circular shape and the delimitation surfaces of which are essentially adapted to the cross-section of the coil springs 126.

 An arrangement of this kind has the advantage that, in the case of a friction clutch in rotation, a correct guiding of the springs 126 is obtained so that these cannot deviate axially.

For additional fixing of the helical springs 126, it is possible, as shown in FIG. 13, to provide the cover 2, in its radially inner marginal zone, with axial deformations 130 which cover the springs 126 in an axial direction. Instead of several deformations 130, the cover 2 can also include an inner edge 130 oriented axially and continuously around the periphery. The inner edge 130 can be used to limit the stress relief of the annular spring 4.

A guide of the corrective springs 126 of
Figures 12 to 14 has the advantage that, when the friction clutch 1 is in rotation, the individual turns of the springs 126 can bear radially, under the effect of centrifugal forces, against the corrective ring 117, in which case the actuating forces exerted by the springs 126 in a circumferential direction will be reduced or even completely eliminated because of the resistances exerted by friction between the turns of the springs and the corrective ring 117. Thus during a rotation of the friction clutch 101, the springs 126 can behave in a practically rigid manner (as a result of the friction forces assisting the action of the springs) .We can thus obtain at least for rotational speeds greater than the speed of idling of the internal combustion engine, the corrective ring 117 can no longer be turned by the springs 126.

 It is thus possible to obtain compensation for the wear of the friction linings only when the friction clutch 101 is actuated for the idle speed of the engine or at least approximately for the speed of slow motion.

The wear correction ring 117 can, however, also be locked in such a way that the wear on the friction linings is corrected only when the engine is stopped, that is to say when the friction clutch 101 does not rotate.

 A blockage of the wear correction process during a rotation of the friction clutch 1 or when exceeding a determined rotation speed can also be advantageous in an embodiment according to Figures 1 and 2 To this end, one can for example provide on the casing 2 means which exert an anti-rotation blocking under the action of a centrifugal force applied to the corrective element 17, and in particular in opposition to the actuating force. produced by the branch springs 26 and / or 26a. The locking means can then be constituted by at least one counterweight, which can be driven radially outwards under the effect of a centrifugal force, this counterweight being supported by example against the inner edge of the ring 17 and capable of producing in this zone a friction which exerts on the ring 17 a holding moment which is greater than the torsional moment exerted by the actuating springs on the ring 17 .

 For radial support of at least a partial area of the extent of the springs 126, it is also possible to provide support means which are supported by the casing 2. These support means can, in the embodiment of FIGS. 12 and 13, form a single piece with the stops 128. For this purpose, the stops 128 can be produced with an angular shape so that they each comprise an area extending in a circumferential direction and which covers at least one part of the extent of a spring 126. Thus at least part of the spring turns are guided and are supported at least in a radial direction.

 As shown in Figure 13, the wire ring 11 provided in Figure 2 is deleted and has been replaced by profiled parts 111 located in the area of the tab ends of the spring forming detector 113. For this purpose, the tabs 113c have been provided with a spherical profile in a region of their ends on the side directed towards the annular actuating spring 4.

 In Figures 15 to 17 is shown another alternative embodiment of a wear correction system according to the invention, where there is provided in place of a corrective ring of annular shape several corrective elements 217. These corrective elements are distributed uniformly around the periphery of the cover 202.

The corrective elements 217 consist of elements in the form of buttons or washers, which comprise a contact ramp 218 extending in a circumferential direction and inclined axially upwards. The corrective elements 217 of annular shape have a recess or central hole 219 which is crossed by the axial appendages 215a in the form of pins supported by the cover, so that the corrective elements 218 of annular shape are mounted in a rotary manner on these appendices 215a. On the cover 202, recesses 225 are provided which form contact counter-ramps 224 for the ramps 218. Between a corrective element 217 and the cover 202, there is provided an elastic element 226 prestressed so that the element corrective 217 is biased in the direction of rotation producing a wear correction. The elastic element 226 can, as shown in Figure 15, be arranged around an axial appendage 215a, that is to say, be arranged as a helical spring. In the end parts of a spring 226, profiled zones are provided, such as, for example, legs or branches serving to support one end of the spring against the casing 202 and the other end of the spring against the corresponding corrective element 217. During an axial offset of the annular spring 204 or of the detector spring 213 in an area of the pivoting support 205, the corrective elements 218 will be subjected to a torsion and the offset will be compensated by a sliding of the ramps 218 contact against the ramps 224.

 The axial support of the annular detector spring 213 against the casing 202 is ensured by means of bosses 214, which have been formed from the axially oriented zone of the casing 202 and which have been forced radially inwards below the external zones. - laughter of the spring forming detector 213.

 The corrective elements 218 of annular shape have the advantage of being able to produce a wear correction action practically independently of the centrifugal forces.

 Instead of the corrective elements 217, which can be turned or subjected to a torsion and represented in FIG. 14, one could also use individual corrective elements in the form of wedges, which could be offset in a radial direction and / or in a direction circumferential to the wear-correcting system. These wedge-shaped corrective elements may include an elongated recess through which an axial appendage 215a can be engaged in order to guide the corresponding corrective element. The corrective wedge-shaped elements can exert a corrective action under the effect of the centrifugal force acting on them. However, energy accumulators can also be provided which urge the wedge-shaped corrective elements in a direction of wear correction. For correct guiding of the wedge-shaped corrective elements, the casing 2 may include profiled parts. The corner surfaces of the corrective elements, oriented with a determined angle of attack with respect to a plane perpendicular to the axis of rotation of the friction clutch, can be provided on the side of the casing and / or on the side of the annular actuating spring. When using such individual wedge-shaped elements, it is advisable to make them from a light material so as to minimize the centrifugal forces acting on them.

 The matching of materials between the parts creating the wear compensation ramps is advantageously chosen so that, during the service life of the friction tie-back, it cannot occur between the contact ramps and the counteramps contact adhesion preventing wear compensation. To avoid such adhesion, at least one of the aforementioned parts may be provided with a coating at least in the area of the ramps or counter-ramps. By using such coatings, corrosion can be avoided in particular when two metal parts are used; adhesion or sticking between the parts forming the wear compensation ramps can also be avoided by making the parts '' applying one against the other and forming the ramps and counter-ramps in materials having different coefficients of expansion so that, under the effect of temperature variations occurring during the operation of the clutch at friction, the surfaces located in contact and constituting the wear compensation ramps can perform a movement relative to each other. Thus, the parts forming the contact ramps and the contact counter-ramps can always be kept mobile relative to each other. Thus, adhesion or bonding cannot occur between these parts because, due to the different expansions, these parts remain constantly free or separated from each other.

A separation of the wear correction ramps can also be obtained by the fact that, because of the different resistances and / or the different arrangements of said parts, the centrifugal forces acting on these parts produce different expansions or movements, which avoid in addition to adhesion or bonding of the parts.

 To avoid an adhesion connection between the contact ramps and the contact counter-ramps, it is also possible to provide at least one means which, when disengaging the friction clutch or else during a wear correction, exerts an axial force on the corrective element or elements. To this end, the corrective element 17, 117 can be axially coupled to a component which has zones which shift axially when wear occurs. This coupling can be established in particular in the region of the pivoting support 5, and in particular using the annular actuating spring 4 and / or the detector spring 13.

 In the diagram of Figure 18 is shown a characteristic curve 340 of the annular tightening spring, which has a low point or minimum 345, where the force exerted by the annular tightening spring is relatively low (about 450 Nm). The maximum force of the annular spring having the characteristic stroke-force curve 340 is of the order of magnitude of 7,600 Nm. The characteristic curve 340 is produced by deformation of an annular spring between two radially spaced supports, and in particular, like this has been described for the characteristic curve 40 of FIG. 8 and in relation to the annular spring 4.

 The characteristic curve 340 of annular spring can be combined with a characteristic curve 342 of friction lining spring. As shown in Figure 18, the profile of the characteristic curve force-stroke 342 of the springs or elastic segments for friction linings approximates the profile of the characteristic curve 340 of the annular tightening spring or the two characteristic curves are located only at a short distance from each other so that the corresponding friction clutch can be actuated with very little force. In the area of action of the friction lining springs, the theoretical clutch release force is deduced from the difference between two points, located vertically one above the other, from curves 340 and 342. Such a difference has been designated by 360. The effective disengaging force is reduced by the lever arm or corresponding transmission ratio of the actuating elements, such as, for example, the legs of the annular spring. This has also been described in relation to the embodiment of Figures 1 and 2 and also in relation to the diagrams shown in Figures 8 to 11.

 In Figure 18, there is shown in broken lines another characteristic curve 440 of annular spring and actuation, which has a minimum or a low point 445, where the force exerted by the annular spring is negative, that is to say say that it does not act in the clutch direction of the corresponding friction clutch but in the disengagement direction. This means that, if point 461 is exceeded during the declutching phase, the friction clutch remains automatically open. The characteristic curve 440 of annular spring can be associated with a characteristic curve of friction lining springs corresponding to curve 442.

 In Figure 19, the associated characteristic curves 340 and 342 or 440 and 442 show the variation profile of the disengaging force to be exerted, for disengaging the corresponding friction clutch, on the actuating levers like the ring spring lugs. As can be seen, the declutching force variation profile 349, which is associated with the characteristic curves 340, 342, is constantly in a range of positive force values, which means that, to keep the clutch in the declutched state, a force must be exerted constantly in the declutching direction. The declutching force variation profile 449, which is associated with the characteristic curves 440 and 442, comprises a partial zone 449a, in which the declutching force initially decreases and then passes from the positive to the negative domain so that the corresponding friction clutch requires no holding force in the disengaged state.

 In the embodiment of a friction clutch 501 which is shown in Figures 20, 20a and 21, the annular sensor spring 513 is pressed against the clutch cover 502 axially via a link 514 of the bar type. To this end, the detector spring 513 has tabs 513d, projecting radially from the outer periphery of the base body 513b of annular shape and which bear axially against radial zones 502a in the form of tabs created by forming towards outside of the cover material. The tabs 502a of the cover are formed towards the outside from the marginal zone 502b of the cover which is oriented substantially axially, and in this regard it is advantageous that the tabs 502a are initially created at least partly by a cutout 502c or 502d formed in the material of the cover. By at least partially trimming the legs 502a, they can be easily deformed in their set position. As shown in particular in FIG. 21, the legs 502a and the branches or tongues 513d are adapted relative to one another so that centering of the detector spring 513 can take place relative to the cover 502. In the exemplary embodiment shown, the legs 502a are provided for this purpose with a small axial offset 502e.

 To ensure correct positioning of the detector spring 513 relative to the housing 502 during the production of the locking link 514 of the bar-type, at least three tabs 502a are provided, advantageously distributed uniformly around the periphery of the cover 502 and which are adapted with respect to the other areas of the cover so that, after a defined relative rotation between the detector spring 513 and the cover 502, the corresponding tabs 513d are applied against a peripheral stop 502f, thus avoiding another relative rotation between the detector spring 513 and the cover 502. The stop 502f is constituted, in the embodiment shown, as shown in particular in Figure 20a, by an axial flange of the cover 502. The Figure further shows that at less certain, and preferably the three legs 502a, create another limitation of rotation 502g between the cover 502 and the lugs 513d of the spring forming detector 513. In the embodiment shown, the same lugs 502a constitute the anti-rotation safety devices 502f and 502g for the two directions of rotation. The stops 502g avoiding unlocking between the detector spring 513 and the cover 502 are constituted by axial chamfers, oriented in a radial direction, of the legs 502a. By means of the peripheral stops 502f and 502g, a positioning is defined, defined in a circumferential direction, of the detector spring 513 relative to the cover 502. For an embodiment of the locking link 514, the detector spring 513 is axially prestressed in direction of the cover 502 so that the tabs 513d penetrate axially into the cutouts 502c and 502d and are placed axially above the cover supports 502a. Then, the cover 502 and the detector spring 513 can be turned relative to each other until some of the tabs 513d come to bear against the rotation limiting stops 502f. In addition, a partial relaxation of the detector spring 513 occurs so that some of the legs 513d, considered in a circumferential direction, are placed between the corresponding stops 502f and 502g and that all the legs 513d are applied against the parts. support 502a located on the side of the cover. With the arrangement in accordance with the invention of the locking of the bar-type 514, it is ensured that, during the mounting of the friction clutch 1, the legs 513d do not come to be placed next to the supports 502a located on the side of the lid.

 In the embodiments which have been described so far, the basic body, for example 513b, in the form of a circular ring absorbing the elastic force proper of the detector spring 513, is located radially outside the loading area or the support area between the pressure plate and the annular actuating spring. For many applications, it may however also be advisable for the circular ring-shaped basic body of the detector annular spring to be arranged radially inside the stressing circle between the pressure plate and the actuating annular spring. This also means that, for an embodiment such as that of FIGS. 1 and 2, the base body 13d, absorbing the axial clamping force of the detector spring 13, is provided radially inside the stressing zone. 3a between the annular actuating spring 4 and the pressure plate 3.

 In the embodiment of FIGS. 20 and 21, the contact counter-ramps 524 situated on the side of the cover are formed by indentations formed in the form of bosses in the sheet metal casing 502. In addition, in this embodiment, the helical springs 526 compressed between the housing 502 and the corrective ring 517, are guided by pins 528, forming a single piece with the corrective ring 517 and extending in a circumferential direction. These guide pins 528 can, as shown in particular in FIG. 21, have in the axial direction an elongated section, which is adapted to the internal diameter of the springs 526. The guides 528 extend at least over a partial zone of the longitudinal extent springs 526 inside the latter. Consequently, at least part of the turns of the springs is guided and is supported at least in a radial direction. In addition, bending or ejection of the springs 526 in an axial direction can be avoided. Pins 548 make it much easier to mount the friction clutch.

 In Figure 22 is shown in part the wear compensation ring 517. This corrective ring 517 has protrusions 527 oriented radially inwardly and which carry the parts 528 in the form of pins, oriented in a circumferential direction and serving to guide the coil springs 526. In the embodiment shown, the parts 528 receiving springs form a single piece with the plastic ring 517 which is produced by injection molding.

The spring guide zones or the spring reception zones 528 may however also be constituted by independent parts or else they may all be joined in the form of a single part, the abovementioned part or parts being connected to the ring. fix 517, for example by means of an interlocking lock. Thus all the guide parts 528 can be constituted by a ring open if necessary on the periphery and which is coupled to the corrective ring 517 by means of at least three connection zones, preferably in the form of interlocking.

 In a manner analogous to that which has been described in relation to FIGS. 12 and 13, the helical springs 526 can furthermore bear additionally radially, for example under the effect of centrifugal forces, against zones of corresponding profile of the cover 502 and / or corrective ring 517.

 The supports provided on the side of the cover for the coil springs 526 are constituted by ribs formed by the material of the cover and extending in an axial direction or by indentations 526 forming axial walls. These bearing zones 526a provided for the springs 526 are in this respect advantageously arranged so that the corresponding ends of the springs are guided and are thus prevented from shifting excessively in an axial and / or radial direction.

 In the embodiment shown in Figure 23 for a clutch 601, the sensor spring 613 is located on the side of the housing 602 which is opposite to the pressure plate 603. The arrangement of the sensor spring 613 outside the volume inside the casing, which receives the pressure plate 603, makes it possible to reduce the thermal stress on the spring forming detector 613, so that the risk of blocking of this spring 613 under the effect of excessive thermal stress is avoided. Also there is a better cooling of the spring 613 on the outside of the casing 602.

 The support of the pivoting support 611, provided on the side of the annular actuating spring 604 which is opposite to the cover, is provided by means of spacing rivets 615, which extend axially through the corresponding recesses of the annular spring 604 and of the casing 602 and which are axially connected to the detector spring 613. In the embodiment shown, the spacing rivets 615 are riveted to the spring 613. Instead of spacing rivets 615, other means can also be used to establish a connection between the pivoting support 611 and the detector spring 613.

Thus, for example, the detector spring 613 could comprise, in a radially inner zone, legs extending axially and which would support the pivot support 611 by corresponding radial zones or even which would directly form this pivot support 611 by parts corresponding profiles.

 In the embodiment of Figure 24, the detector spring 713 extends radially inside the pivot support 615 provided for the annular actuating spring 704. The detector spring 713 is supported by its radially inner zones against the cover 702 which is provided for this purpose with lugs 715 extending through corresponding slots or recesses of the annular spring 704 and which axially support the annular spring 713 forming a detector. In accordance with another alternative embodiment, the spring 713 could also include in its inner marginal zone tabs which extend in an axial direction through corresponding openings of the annular spring 704 and which support it on the side of the cover.

 The wear corrective ring 817 shown in Figure 25 can be used in a friction clutch according to Figures 20 and 21. The corrective ring 817 has protrusions 827 projecting radially inward. The protrusions 827 include radial appendages 827a, which form support zones for the coil springs 826, compressed in a circumferential direction between the clutch cover and the wear corrective ring 817. To provide guidance and facilitate assembly helical springs 826, a ring 528 is provided which is interrupted or opened around its outer periphery. The ring 528 is connected to the radial protrusions 827a. To this end, the protuberances 827a may include recesses or grooves extending in a circumferential direction and which are arranged so that they form, in cooperation with the ring 528, a socket connection. The supports provided on the cover side for the wear correction springs 826 consist of axial lugs 826a of the clutch cover. The axial tabs 826a each have an axial notch 826b for receiving the ring 828. The notches 826b are arranged in this respect so that the ring 828 has, with respect to the legs 826a, a possibility of axial offset at least in correspondence to the wear stroke of the friction clutch. In this regard, it is particularly judicious that the recesses formed in the radial protrusions 827a to receive the ring 828, and the notches 826b are, considering them in an axial direction , oriented in opposite directions, or in other words, that the recesses formed in the protrusions 827a are open in one of the axial directions and that the notches 826b are open in the other axial direction.

 In the embodiment of a friction clutch 901 which is shown in Figure 26, the support of the annular actuating spring 904 in the disengagement direction is provided in a central area of the base body 904a of the annular spring 904. In the embodiment of a friction clutch 901 which is shown in Figure 26, the support of the annular actuating spring 904 in the disengagement direction is provided in a central area of the base body 904a of the annular spring 904. Radially outward, the base body 904a rests against the pressure plate 903 and it extends radially inward above the pivot support 905. This means that the pivot support 905 is relatively far from the inner edge of the base body 904a of the annular spring 905 or from the ends of the slots which form the legs of the annular spring 904, and this in comparison with the spring clutches ann ulcers known to date. In the embodiment shown, the ratio between the radial widths of the areas of the base body which are located radially inside the pivoting support 905 and of the areas of the base body which are located radially outside of the pivoting support 905 is of the order of magnitude of 1: 2. It is advisable that this ratio is between 1: 6 and 1: 2. With such support of the annular actuating spring 904, it is possible to avoid damage or excessive stress on the basic body 904a of the annular spring in the area of the pivoting support 905.

 In FIG. 26, an axial protuberance 903a which is provided on the pressure plate 903 has also been indicated in broken lines. With the aid of such protrusions 903a provided on the pressure plate 903, in particular in the region of the bosses of 903b support, it is possible to ensure centering of the annular actuating spring 904 relative to the clutch 901. I1 is also possible to obtain a maintenance of the annular actuating spring 904 relative to the cover 902 at by means of centering carried out from the outside in a radial direction, so that it is then possible to remove the rivets or centering pins 915 also shown in Figure 26. Although this is not indicated, the centering thus produced from the outside can also be carried out by means of tabs or indentations formed by pushing back the material of the cover 902.

 In the friction clutch 901, the sensor spring 913 is arranged such that the base body 913a producing the force is located radially inside the bosses 903b. To ensure on the one hand the support of the annular actuating spring 904 and to obtain on the other hand a support of the spring forming detector 913 on the cover 902, this spring has radial tabs or tongues which extend on one side radially inward from the base body 913a and on the other side radially outward from the base body 913a.

 In the alternative embodiment shown in FIG. 27 for a friction clutch 1001, the force oriented in the opposite direction to the disengaging force of the friction clutch or to the pivoting force of the annular actuating spring 1004 is generated by a detector spring 1013, which is compressed axially between the casing 1002 and the pressure plate 1003. In one embodiment of this kind, the annular actuating spring 1004 is not supported, in the pivoting or tilting 1005, by a pivoting support in the declutching direction. The application of the annular spring 1004 against the pivoting support or the support support 1012 provided on the side of the cover is ensured by the prestressing force of the detector spring 1013. This detector spring is arranged so that, during the disengagement process of the friction clutch 1001, the axial force exerted by this detector spring 1013 on the annular spring 1004 either becomes or becomes greater than the disengagement force of the friction clutch 1001. This must therefore be ensured that, when there is no wear on the friction linings, the annular spring 1004 remains constantly applied against the support located on the side of the cover or else against the pivoting supports 1012. In this regard, it is necessary to in a manner analogous to what has been described previously in relation to the other embodiments, of establishing a balance between the different forces acting and adding up in a axial direction.

These forces, which are produced by the detector spring 1013, by the suspension of friction linings, by the elastic blades optionally provided between the pressure plate 1003 and the casing 1002, by the annular actuating spring 1004, by the force clutch clutch 1001 and by the correcting springs acting on the wear corrective ring 1017, must in correspondence be balanced.

The friction clutch 1101 shown in the
Figures 28 and 29 comprises a housing 1102 and a disc or pressure plate 1103, connected in a non-rotating manner to the housing but which can however carry out a limited axial translation. An annular tightening spring 1104 is compressed axially between the pressure disc 1103 and the cover 1102 and it can move by pivoting around a pivoting support 1105 of annular shape which is supported by the casing 1102 so as to stress the disc pressure 1103 towards a back pressure plate 1106, connected to the casing 1102, such as for example a flywheel, so that the friction linings 1107 of the clutch disc 1108 are clamped between the friction surfaces of the disc 1103 and the backpressure plate 1106.

 The pressure disc 1103 is connected in a non-rotating manner to the housing 1102 by means of articulation means oriented in a circumferential or tangential direction and being in the form of elastic blades 1109. In the embodiment shown, the clutch disc 1108 comprises what are called elastic segments 1110 for friction linings, which ensure a progressive increase in the torque when the friction clutch is clutched, by the fact that they allow, during a limited axial offset of the two friction linings 1107 towards one another, a gradual increase in the axial forces acting on these friction linings 1107. However, a clutch disc could also be used in which the friction linings friction 1107 would be fixed axially rigidly on at least one bearing disc and where one could mount in another place the elastic segments es 1110 for friction linings, for example between the annular spring 1104 and the pressure disc 1103.

 In the embodiment shown, the annular spring 1104 comprises a base body 1104a of annular shape, which exerts the clamping force and from which radially inwardly extend the actuating tabs 1104b. The annular spring 1104 is then installed in such a way that it can urge the pressure disc 1103 through zones located radially more towards the outside and that it can tilt around the pivoting support 1105 through zones situated radially more towards the inside. The pivoting support 1105 has two pivoting support zones 1111, 1112, which in this case consist of metal wire rings and between which the annular spring is held or clamped. For an anti-rotation fixing of the annular spring actuator 1104 and for centering as well as for retaining the metal wire rings 1111, 1112 relative to the housing 1102, there are provided retaining means in the form of rivets 1115 which are fixed on the cover and which each extend by a rod 111Sa oriented axially through a cutout provided between adjacent tabs 1104b of the annular spring.

 The clutch 1101 has a wear correction means which has the function of compensating for axial wear on the friction surfaces of the pressure disc 1103 and of the backpressure plate 1106 and on the friction surfaces of the linings 1107, this corrective means consisting of a wear compensating device 1116, provided between the annular tightening spring 1104 and the pressure disc 1103, and also of limiting means 1117 having the function of limiting the declutching stroke of the pressure disc 1103 and which are arranged in the form of wear distance detectors.

 The limiting means 1117 acting as wear detectors each comprise a sleeve 1118 which is engaged in a non-rotating manner in a hole 1120 in the pressure disc 1103. The sleeve 1118 forms a slot 1121 through which two elastic blades extend axially. 1122. The elastic blades 1122 bear against each other and at least one elastic blade is curved; advantageously, the two elastic blades are curved in opposite directions from one another. The elastic blades 1122 are mounted in the sleeve with a defined preload and they can thus undergo an offset with respect to the sleeve 1118 in the axial direction of the clutch 1101 in opposition to a predetermined resistance exerted by friction. elastic 1122 is chosen such that, when the friction clutch 1101 is engaged, these elastic blades have, relative to an axially fixed component of the clutch - in the embodiment shown relative to the outer edge 1123 of the casing 1102 - a defined clearance 1124 which corresponds to the predetermined disengagement race of the pressure disc 1103.

When the friction clutch is engaged, the elastic blades 1122 can bear, by their end 1122a opposite the casing 1102, against the back-pressure plate 1106 so that it is ensured that, in the event of wear of the friction linings 1107, the pressure disc 1103 is axially offset relative to the elastic blades 1122 in correspondence with this wear of the friction linings, and in particular in opposition to the friction effect manifested between the elastic blades 1122 and the sleeve 1118 , which advantageously consists of a plastic or a material with a friction effect.

 In the embodiment shown, the hole 1120, in which the sleeve 1118 is fixed by fitting both in an axial direction and in a circumferential direction, is formed in a boss 1125 of the pressure plate, which extends radially towards the exterior and on which an elastic blade 1109 is respectively articulated by means of a riveted connection 1109a. A displacement of the sleeve 1118 in the direction of the backpressure plate 1106 can also be avoided because the sleeve 1118 comprises, at its end directed towards the housing 1102, a flange 1118a by means of which it can rest against the pressure disc 1103. An ejection of the sleeve 1118 from the hole 1120 in the direction of the casing or of the clutch cover 1102 can be avoided by the fact that, as indicated in phantom in Figure 28, the blades elastic 1109 are hooked partly radially on the bush 1118 and if necessary they are additionally clamped axially in the hole 1120. A rotation of the bush can also be prevented because this bush has a profiled part, in particular a recess which receives zones 1119 of the elastic blades 1109 hanging on the sleeve.

 The wear compensation device 1116 comprises a compensating component biased by the annular spring 1104 and being in the form of a sheet metal ring 1126 having a U-section and which is shown in plan view in Figure 32. The compensating component 1126 comprises, on the side of the bottom 1127 which is directed towards the annular spring 1104, at least one or more axial projections 1128 of annular shape, which are preferably distributed uniformly around the periphery and which are created by moldings formed by imprint in the sheet.

These segment-shaped protrusions 1128 ensure that, in zones situated between adjacent protrusions 1128 in a circumferential direction, radial passages are formed between the base body 1104a of the annular spring and the compensating ring 1126, these passages allowing a flow of cooling air. As shown in particular in Figure 29, the compensating ring 1126 is centered relative to the pressure disc 1103.

For this purpose, the pressure disc 1103 comprises at least one shoulder 1129, which ensures the centering with respect to the pressure disc 1103 of the radially inner wall 1130, extending axially, of the compensating ring 1126. The shoulder 1129 can be created by a closed surface extending around the perimeter or alternatively also by surfaces in the form of segments which are spaced apart from each other around the perimeter. The compensating ring 1126 further comprises a radially outer wall 1131, extending axially and which creates a free volume 1126a of annular shape in cooperation with the inner wall and the bottom 1127. Radially outside, the compensating ring 1126 has radial tabs or flanges 1132, which constitute stops cooperating with counter-stops 1133 of the components axially displaceable and being in the form of elastic strips 1122 of the wear detector 1117. The counter-stops 1133 are constituted by lugs formed on the elastic blades 1122, oriented radially inwardly and hooking onto the lugs 1132. Thus, the axial offset of the compensating ring 1126 in a direction away from the pressure disc 1103, that is to say -to say in the direction of the housing, is limited by the abutments 1133.

 Between the compensating ring 1126 and the pressure disc 1103, there is an adaptation device 1134 which, when the friction clutch 1101 is disengaged and when there is wear on the friction linings, allows automatic adjustment of the position of the compensating ring 1126 while, when the friction clutch is engaged, it acts irreversibly, that is to say with locking effect, so that is ensured that, during the clutch phase of the friction clutch 1101, the compensating ring 1126 maintains a defined axial position relative to the pressure disc 1103. This defined position can only be modified during a disengagement process and in correspondence to the wear which has occurred on the friction linings.

The wear correction device 1134 comprises several pairs of corners 1135, 1136, preferably distributed uniformly around the periphery and which are engaged in the free volume 1126a of annular shape of the sheet metal ring 1126. The corners 1136 resting against a surface 1137 of annular shape of the pressure disc 1103 are connected in a non-rotating manner with the sheet metal ring 1126 but they can however move in axial translation. To this end, the sheet metal ring 1126 comprises, in an area of its walls 1130, 1131 extending axially, profiled parts in the form of ribs 1138, 1139, which create in an area of the free volume 1126a projections s' engaging in recesses or grooves 1140, 1141, adapted in correspondence, corners 1136. The grooves 1140, 1141 or the projections 1138, 1139 are oriented in the axial direction of the clutch 1101. As also shown in FIG. 30, the corners 1135 are engaged substantially axially between the bottom 1127 of the sheet metal ring 1126 and the corners 1136. The corners 1135 and 1136 form contact or mutual application ramps 1142, 1143 extending in a circumferential direction, axially inclined upward and through which the corners 1135, 1136 corresponding to a pair are pressed against each other. The corners 1135 bear on the other hand against the bottom 1127 of the ring 1126 and they can move in a circumferential direction relative to this ring 1126. The contact ramps 1142,
1143 are applied one against the other. For this purpose, energy accumulators are provided in the form of helical springs 1144 in the annular volume 1126a, these springs pressing at one end against a wedge 1136 connected in a non-rotating manner to the ring 1126 and by their other end against a corner 1135 which can move in a circumferential direction. To hold the energy accumulators 1144, the corners 1135, 1136 comprise, at their ends directed towards the corresponding energy accumulators, projections 1145, 1146 which engage in the turns of the springs and which thus maintain their ends. The springs 1144 are further guided by the wall zones 1130, 1131 and the bottom 1127 of the ring 1126.

 In the illustrated embodiment, the compensating ring 1126 is prevented from rotating relative to the pressure disc 1103. For this purpose, there are provided, as shown in FIG. 31, on the pressure disc 1103 axial projections having in the form of pins 1147, which extend axially in recesses 1148 formed in the region of the lugs 1132. By means of this anti-rotation security, it is ensured that, during the operation of the friction clutch, the regions of stop of the tabs 1132 remain constantly positioned below the lugs 1133 for delimiting the elastic blades 1122.

 In the embodiment shown, the corners 1135, 1136 are made of a heat-resistant plastic material, such as for example a duroplastic or thermoplastic material, which can also be additionally reinforced with fibers. As a result, the corners 1135, 1136 acting as wear-correcting elements can be manufactured in a simple manner as injection molded parts. It may however also be advantageous that at least one of the corners 1135, 1136 of a pair is made of a friction-effect material, such as for example a material for friction linings. The corners or wear corrective elements 1135, 1136 can however also be produced in the form of sheet metal parts or else sintered parts. The angle of inclination as well as the extent of the contact ramps 1142, 1143 are chosen to so that, during the entire service life of the friction clutch 1101, a corrective compensation for the wear and tear occurring on the friction surfaces of the pressure disc 1103 and of the back pressure plate 1106 is ensured. and also on the friction linings 1107. The corner angle 1149 or else the inclination angle 1149 of the contact ramps 1142, 1143 relative to a plane perpendicular to the axis of rotation of the friction clutch is chosen so that, friction occurring during the application of the contact ramps 1142, 1143 against each other, slippage between said ramps is prevented. Depending on the pairing of materials in the area of conta ramps and 1142, 1143, the angle 1149 can have a value between 5 and 20 degrees, preferably of the order of magnitude of 10 degrees. The corners 1135 movable in a circumferential direction are arranged so that the vertices of said corners are oriented in the direction of rotation 1150.

 The mutual application of the contact ramps 1142, 1143 by the energy accumulators 1144 as well as the angle of inclination 1149 are chosen so that the resulting axial force acting on the wear corrective ring 1126 is smaller that the necessary translational force of the wear detector 1122 of the limiting means 1117.

 In addition, when designing the annular spring 1104, it must be taken into account that the clamping force to be exerted by this spring on the pressure disc 1103 must be increased by a value corresponding to the necessary force of translation of the detector. wear 1122 and a value corresponding to the clamping force of the elastic blades 1109 held between the cover 1102 and the pressure disc 1103.

In addition, the various components must be designed so that the contact wear between the wear detectors 1122 and the backpressure plate 1106 or else between the wear detectors and the housing 1103 remains low compared to wear occurring on friction linings 1107.

 To avoid an unintentional shift between the contact ramps 1142, 1143 or else the wear corrective elements 1135, 1136 in FIG. 30, it is possible to provide in one area at least one of the contact ramps 1142, 1143 small projections that hang on the other ramp. The protrusions can then be arranged in such a way that movement for wear correction is possible while, however, preventing the ramps from slipping relative to one another. It may also be particularly advisable for the two contact ramps 1142, 1143 to have projections which hook into one another. These projections can for example be created by profiled parts in the shape of saw teeth having a very low height and allowing a relative offset of the ramps 1142, 1143 only in the wear compensation ring. Such profiling has been shown schematically in Figure 30 in a partial area of the extent of the ramps 1142, 1143 and it has been designated by 1143a.

In applications where only one of the contact rails 1142, 1143 has projections, these can be arranged so that they have a greater hardness than the material of which the other contact rail is made, so that it can at least a slight sinking or catching of the projections on the contact rail against which they are supported.

 To prevent the elastic curved or wavy blades 1122 from losing their clamping force due to the very high temperatures occurring in the pressure disc 1103 during the clutching process, the sleeves 1118 are preferably made of a material having a small conductivity thermal and a large coefficient of friction. The corners 1135, 1136 can be made of the same material.

 To allow better cooling of the clutch, in particular of the pressure disc 1103, it is possible to provide in this pressure disc 1103 radially oriented grooves, distributed around the periphery and one of which is shown in broken lines in FIG. 29 and is designated by 1151. These radial grooves 1151 are arranged, considering a circumferential direction, between two pairs of adjacent corners respectively and they extend between the ring 1126 and the pressure disc 1103. Also in the area of the springs 1144, the ring 1126 could include axial cutouts starting from the bottom 1127 so as to create radial passages between the annular spring 1104 and the ring 1126.

 To increase the wear resistance of the different support zones, the corresponding zones can be provided with a wear-resistant layer, such as for example a hard chrome coating, a molybdenum coating, or it is also possible provide particularly wear-resistant components in the contact areas. Thus, for example, plastic wear pads can be provided on the wear detectors 1122 in the support zone against the backpressure plate 1106 and against the casing 1102.

 The elastic blades 1109 transmitting the torque to the pressure disc 1103 are clamped between this pressure disc 1103 and the casing 1102 so that, when the friction clutch 1101 is disengaged, the pressure disc 1103 is offset towards the casing 1102. It is thus ensured that practically during the entire disengagement phase or else until the entry into action of the limiting means 1117, the ring 1126 remains applied against the annular spring 1104.

 The clutch disengaging distance in the area of the leg ends 1104c is judiciously chosen so that, when the clutch is disengaged, the outer edge of the annular spring 1104 deviates a short distance from the ring 1126. This thus means that, when the friction clutch 1101 is disengaged, the displacement of the annular spring in the diametrical zone of stress of the pressure disc by the annular spring 1104 is greater than the distance of separation 1124 from the disc pressure 1103 which is defined by the stroke limiting means 1122.

 The relative position, shown in Figure 29, of the various components corresponds to the new state of the friction clutch. In the event of axial wear, in particular of the friction linings 1107, the position of the pressure disc 1103 is shifted in the direction of the backpressure plate 1106, so that initially there is a change in the taper, and consequently also of the clamping force exerted by the annular spring in the engaged state of the friction clutch 1101, and in particular advantageously in the direction of an increase. This modification causes the axial position of the pressure disc 1103 to change with respect to the wear detectors 1122 pressing axially against the back pressure plate 1106.

Under the effect of the force exerted by the annular spring on the ring 1126, this ring 1126 follows the axial movement of the back-pressure plate 1103 which is caused by the wear of the friction linings, so that the zones of stop 1132 of the ring 1126 deviate axially from the zones serving as counter-stops and being in the form of pins 1133 on the wear detectors 1122, this spacing having in particular a value which corresponds essentially to the wear of friction linings. The wear compensating ring 1126 maintains its axial position relative to the pressure disc 1103 during a clutching process because it will be biased by the annular spring 1104 in the direction of the pressure disc 1103 and the wear compensation device 1134. is irreversible during the clutching process, i.e. it acts as an axial lock. When disengaging the friction clutch 1101, the pressure disc is biased towards the housing 1102 by the blades elastic 1109 and it is moved until the wear detectors 1122 are applied against the casing 1102 or against the abutment zones 1123 of this casing. Until this disengaging distance is reached, which corresponds to the distance between the pressure disc 1103, the axial position of the ring 1126 is maintained with respect to the pressure disc 1103. declutching process, the pressure disc 1103 remains axially stationary whereas, on the other hand, the ring 1126 follows axially the declutching movement of the annular spring in the diametrical zone of stress, and in particular until the abutment zones 1122 of the ring 1126 again come into contact with the counter-stop zones 1133 of the wear detectors 1122. The axial offset of the wear compensating ring 1126 is produced by the corners 1135 which are biased by the spring 1144. These corners 1135 will be moved in one direction
circumferential with respect to the corners 1136 until the ring 1126 is tightened against the counter-stops 1133 of the wear detectors 1122. The movement of spacing of the pressure disc 1103 is ensured, in the embodiment shown , by means of the elastic blades 1109, which are arranged between the casing 1102 and the pressure disc 1103 so that they are subjected to an axial prestress which pushes the pressure disc 1103 in the direction of the casing 1102. When the annular spring 1104 continues to pivot in the declutching direction, its radially outer zone deviates from the wear corrector ring 1126 because the latter, as has already been described, is kept axially set back relative to the pressure disc 1103 by the wear detectors 1122. Such a separation, at least slight, of the annular spring 1104 with respect to the wear correcting ring 1126 during a declutching process is particular particularly advantageous for the functioning of the wear correction system 1117 + 1134.

 The wear correcting means 1117 + 1134 according to the invention ensures that the wear correction using the ring 1126 is always carried out, by means of the correcting corners 1135, 1136, in correspondence the wear value of the friction linings. This is due to the fact that the wear corrector ring 1126 is tightened axially between the correcting means in the form of the corners 1135, 1136 and the wear detectors 1122, which prevents the compensating component appearing under the shape of the ring 1126, is displaced by a greater distance than the corresponding wear of the friction linings.

In addition, thanks to the arrangement according to the invention of the wear-correcting means, it is ensured that also when the stroke is exceeded in the area of the declutching means, like the tabs 1104b of the annular spring, or although during axial oscillations of the pressure plate, there can be no offset of the wear correction means 1117 + 1134 because the wear detector 1121 is also supported axially, during a significant impact on the housing 1102 with respect to the pressure disc 1103 by means of the self-locking effect wear compensation device 1134, the aforementioned support being provided in particular by means of the abutments 1132. Thus, in the disengaged state of the friction clutch, axial forces can act on the wear detectors 1122 towards the back pressure plate 1106, these forces being greater than the forces establishing the connection between the wear detectors 1122 and the pressure disc 11 03, without the wear detectors being axially offset with respect to this pressure disc 1103.

 With the wear-correcting means according to the invention, it is ensured that, during the entire service life of the clutch, the annular spring operates in practically the same range of characteristic curves and comprises, in the engaged state of the friction clutch 1101, a clamping position remaining practically constant, which thus also allows the application of a practically constant clamping force on the pressure disc 1103.

It is therefore possible to use an annular spring having a decreasing characteristic curve of variation of the force as a function of the declutching distance, and in particular advantageously in combination with a clutch disc whose friction linings 1107 are resiliently suspended. relative to each other by means of elastic segments, so that the clutch force to be effectively exerted can be kept at a comparatively low level and can be kept practically constant during the service life of the clutch, provided that the characteristic curve of the friction lining springs does not vary appreciably during the clutch's service life. When disengaging such a friction clutch, the annular spring 1104 pivots around its support 1105 located on the cover, to which case, in a predetermined partial area of the axial disengagement race of the pressure disc 1 103, the elastic segments 1110 relax and thus the axial force exerted by these elastic segments 1110 assists the clutch release clutch 1101.

This thus means that a smaller maximum disengaging force must be exerted than that which theoretically results, in the clutched state of the clutch 1101, from the mounting position of the annular spring 1104 and of the elastic blades 1109. Provided that the elasticity or relaxation range of the elastic segments 1110 is exceeded, the friction linings 1107 will be released, in which case, due to the decrease in the corresponding zone of the characteristic curve in which the annular spring 1104 operates, the force of the clutch still to be exercised will already be considerably revised compared to that which would correspond to the installation point or the installation position in accordance with Figure 29. In the case of a continuation of the disengagement process, the disengagement force decreases further , and in particular at least until the minimum or the lowest point of the characteristic curve, preferably sinusoidal, of the spring 1104, is reached nt.

 The wear correction means 1117 + 1134 shown in FIGS. 28 and 29 can advantageously be arranged so that, during a rotation of the friction clutch 1101, the different turns of the wear corrective springs 1144 come bear against the outer wall 1131 of the corrective ring 1126 and the actuating forces exerted by the springs 1144 in a circumferential direction are reduced or even completely counterbalanced as a result of the resistances generated by friction between the turns of the springs and the wear correction ring 1126. The springs 1144 can thus, when the friction clutch 1101 rotates, behave in a practically rigid manner as a result of the friction forces which prevent the action of the springs. in addition, the corrective wear corners 1135 can, under the effect of the centrifugal forces acting on them, also bear radially against the wall 1131 of the corrective ring d wear 1126 and be prevented from rotating by the frictional forces generated between the corners 1135 and the wear corrective ring 1126. It is thus possible to obtain at least in ranges of rotational speeds greater than the idle speed of the internal combustion engine, the wear compensating device 1134 is not rotated by the springs 1144. Thus the friction clutch 1101 can be arranged in such a way as to compensate for the wear of the friction linings only when the friction clutch 1101 is actuated at idle speed or at least approximately at this idle speed. The blocking of the wear corrective ring 1126 can however be obtained by a corresponding arrangement of the compensating device 1134 so that a correction of the wear of the friction linings can occur only when the internal combustion engine is stopped, that is to say when the clutch with friction 1101 does not rotate, or else for very low rotational speeds.

 The pairing of materials between the components 1135, 1136 forming the wear correction ramps and also the material of the components cooperating with the aforementioned parts is advantageously chosen so that, during the service life of the friction clutch, it cannot create adhesion, preventing corrective wear action, between the ramps and the components cooperating with them. To avoid such adhesion, at least one of the aforementioned components may be provided with an appropriate coating at least in the area of the ramps or the bearing surfaces.

 To avoid an adhesion connection between the contact ramps and the contact counter-ramps, it is also possible to provide at least one means which, when disengaging the friction clutch or else during wear compensation friction linings, exerts an axial force on the wear corrective element (s), this force producing a separation or detachment of the ramps.

 In the new state of the friction clutch 1101, that is to say in the state which the clutch comprises before being fixed, with the interposition of the clutch disc 1108, on the counter plate. pressure 1106, the corners 1135 are located, with respect to the position indicated in FIG. 30, in a position still retracted with respect to the corners 1136, so that the wear corrective ring 1126 occupies its retracted position as far as possible in the direction of the pressure disc 1103 and that thus the assembly constituted by the pressure disc 1103 and the wear corrective ring 1126 requires the smallest axial installation volume. To maintain the corners 1135, before mounting the friction clutch 1101, in their retracted position, these corners 1135 have hooking zones, in the form of recesses 1152, for a retaining means. Such retaining means can be provided during the manufacture or assembly of the friction clutch 1101 and, after the mounting of the friction clutch 1101 on the flywheel 1106, they will be removed so that the corrective device 1134 wear can be activated.

In the exemplary embodiment shown, as shown in FIGS. 30 and 32, there are provided in the wear corrective ring 1126 elongated slots 1153, oriented in a circumferential direction and by means of which the attachment zones retaining means, or the torsion tool, can pass to engage in the recesses 1152. The elongated recesses 1153 oriented in a circumferential direction must then have at least an extent which corresponds to a rotation in the circumferential direction in relationship with the largest possible angle of wear correction by the corners 1135. In a new state of the friction clutch, the corners 1135, maintained in their retracted position in the circumferential direction, can be locked in this position by the wear detectors 1122, which fix the wear corrective ring 1126 in its retracted position. The connections which are automatically established between the wear detectors 1122 and the pressure disc 1103 must be designed so that the necessary translational force serving to offset the wear detectors 1122 relative to the pressure disc 1103 is more greater than the resulting force acting on the ring 1126 and which is produced by the springs 1144 biasing the corners 1135.

 In the embodiment of Figures 28 and 29, the ramps 1143 can also be formed directly by the ring 1126, for example by means of a recess of inclined surfaces 1143, in which case the ring 1126 must then be turned by the springs 1144 with respect to the pressure disc 1103. In such an embodiment, the corners 1136 are connected in a non-rotating manner to the pressure disc 1103 or else they are formed directly thereon. In addition, in such an embodiment, the stops arranged in the form of the tabs 1132 must be elongated in a circumferential direction corresponding to the necessary angle of rotation of the ring 1126 for wear correction, in order to ensure maintaining an axial delimitation between the wear detectors 1122 and the ring 1126 during the service life of the friction clutch. In the embodiment described last, the wear corrector ring 1126 can also , when the friction clutch 1101 is installed, be rotated in a simple manner from a position located radially on the outside, and in particular by means of the stop lugs 1132, extending in a circumferential direction and which are accessible via radial passages provided in the outer periphery of the clutch housing 1102. These radial passages can in particular also receive the bosses 1125 of transmiss torque ion of the pressure disc 1103 as well as the elastic blades 1109.

The corrective wear compensation means according to the invention also has the advantage that its principle can also be applied to so-called friction clutches operating in traction and in which the annular spring bears so pivoted by a radially outer marginal zone against a cover and urged by marginal zones located radially more inside the pressure disc. A clutch of this kind has been shown in Figure 33. Between the annular spring 1204 and the pressure disc 1203, there is provided a device 1234 for wear compensation which can be arranged in a similar manner to that which has been described in relation to Figures 28 and 29.

The corrective ring 1226 again cooperates with wear detectors 1222 by means of detector elements 1217. The adaptation of the wear detectors 1222 with respect to the pressure disc 1203 is ensured by applying the stop zones 1222a against the casing or the cover 1202. The wear detectors 1222 again carry stops 1233 which limit the axial travel of the pressure disc 1203 during a declutching process. To allow correct operation of the wear correction means of Figure 33, the ring 1226 has at least a slight possibility of axial movement relative to the wear detectors 1222. This result can be obtained by providing a corresponding stop link 1233a with play between the wear detectors 1222 and the ring 1226 or else however by ensuring that the ring 1226 has radial zones 1226a which are elastically deformable in an axial direction, that is to say which have elastic flexibility.

 In the embodiment shown in Figure 34, the wear detector elements 1317 are engaged directly in the base body of the pressure disc 1303. The wear detectors 1322 comprise abutment zones 1322a, which cooperate with zones 1323 of the cover which constitute counter-stops.

The areas 1323 of the cover form a single piece with a fixing means 1302a by means of which the annular spring 1304 is pivotally mounted on the cover 1302. In the embodiment shown, the fixing or fixing means holding 1302a are constituted by lugs integral with the cover material and which extend axially through the annular spring 1304. Radially outside the wear detectors 1317 provided in a radial zone of the base body 1304a of the annular spring, 1334 wear compensation device is provided.

 With the arrangement according to the invention of a friction clutch, it is not only possible to obtain an increase in the service life of the clutch by incorporating thicker friction linings, that is to say by increase in the axial wear volume of the linings, but also it is possible in particular to obtain a reduction in the declutching forces, and in particular by using an energy accumulator comprising a characteristic force-stroke curve which is decreasing over the declutching distance of the friction clutch, in combination with at least one elastic means acting in the opposite direction to the energy accumulator acting on the pressure plate and which ensures, during a clutch engagement and disengagement friction, a gradual increase or reduction of the torque that can be transmitted by the friction clutch over at least a partial area of the clutch actuation stroke or of the pr platen stroke ession. This elastic means is advantageously arranged in series with the clamping spring, such as for example an annular spring, of the friction clutch. I1 is thus possible to obtain, with the arrangement according to the invention of a friction clutch, a significant reduction in the release force; this reduction is retained or remains invariant during the service life of the friction clutch, which means that a comparatively narrow tolerance band is obtained.

Furthermore, in friction clutches according to the present invention, it is possible to use annular springs which have force-stroke characteristic curves having a relatively steep slope in the working range.

Annular springs of this kind would cause, in conventional clutches, a very large increase in the declutching force in the event of wear of the friction linings.

 In clutches which do not include the wear correction system according to the invention, point 41 (Figure 8) corresponding to the engaged state of the friction clutch moves, as the wear of the friction linings increases along the curve 40 towards the maximum 41a. Up to this point 41a, there occurs during a declutching process a decrease in the declutching force, but however in total the level of evolution of the declutching force increases relative to the evolution of the declutching force in l new condition of the friction clutch. This therefore means that the zone 43 progresses towards the left until the point 41 comes into coincidence with the maximum 41a. The point 44 moves in correspondence along the characteristic curve 40. In the event of increased wear of the friction linings, the point of intervention of the annular actuating spring, corresponding to the engaged state of the friction clutch, gradually moves from the maximum 41a towards point 41b, so that the clamping force exerted by the annular spring gradually decreases. The clamping force exerted by the annular actuating spring at point 41b corresponds to the clamping force exerted at point 41 in the new state of the friction clutch. As soon as the maximum 41a is exceeded, there occurs during the declutching process and at least in a partial area of the clutch actuating stroke, an increase in the declutching force. maximum allowable wear or wear point 41b, there is throughout the declutching stroke 43a an increase in the declutching force. This increase in the declutching force is also maintained when, as shown in FIG. 8, there is provided a suspension of friction linings or else a system 42a replacing the elastic suspension of friction linings.

 When designing a friction clutch and in particular its wear-correcting means, it must be taken into account that the crankshaft of the internal combustion engine produces in the flywheel axial and three-dimensional oscillations, which will also be transmitted to the friction clutch attached to the steering wheel. In order for the friction clutch or the wear-correcting means to function properly, that is to say so that an undesirable wear-out correction does not occur under the effect of such oscillations, it is necessary , in the embodiment of Figures 1 to 27, that is to say completely generally in embodiments comprising a wear-correcting means provided with a force detector, that the corrective force exerted by this wear detector is greater than the inertial forces which can act on the force detector. These forces result in particular from the masses of the main annular spring, the wear corrective ring or the wear corrective elements , of a corresponding proportion of mass of the force detector and, if applicable, the masses of other components, which are multiplied by the maximum possible axial accelerations of these components or elements, as a result of the axial and flex oscillations steering wheel ion. Thus for example in the case of an embodiment according to FIG. 27, where the annular spring forming a detector 1013 bears against the pressure plate 1003 of the clutch, one must also take into account the inertia of this plate of clutch pressure 1003. It must therefore be constantly ensured that the force exerted by the detector spring is greater than the forces acting on it and which are formed by the masses, multiplied by the maximum possible axial acceleration, of the components which influence by their inertia the spring forming the detector. These inertial forces can have a disturbing effect, in particular during the actuation of the friction clutch and in particular in the disengaged state of this friction clutch.

 In the embodiments of FIGS. 29 to 34, it is necessary, when designing the wear detectors and the wear compensating devices, to also take account of the forces generated because of the inertia of the various components as well as forces resulting from the axial and angular oscillations acting on them.

 Generally, when designing a friction clutch comprising integrated wear compensation means, it is thus necessary to take into account the masses of the elements to which the axial and / or angular oscillations can be transmitted and which act on the compensating means. In the embodiments of Figures 28 to 34, account must be taken in particular of the components which have an influence on the function of the ramp mechanism.

 In Figure 35 is shown a divided flywheel 1401, which has a first inertia mass, or primary mass 1402, which can be fixed on a crankshaft, not shown, of an internal combustion engine, as well as a second mass d inertia, or secondary mass 1403. On the second mass of inertia 1403 is fixed, with the interposition of a clutch disc 1405, a friction clutch 1404 by means of which a transmission, also not shown, can be coupled and uncoupled. The inertia masses 1402 and 1403 are mounted so as to be able to rotate relative to each other by means of a bearing 1406, which is arranged radially inwards relative to the holes 1407 for passage of the fixing bolts 1408 used to mount the first inertia mass 1402 on the crankshaft of the internal combustion engine. Between the two inertia masses 1402 and 1403 acts the damping device 1409 which includes the helical pressure springs 1410, which are housed in a volume 1411 of annular shape, constituting a zone 1412 of toroidal shape. The annular volume 1411 is filled at least in part with a viscous agent, such as, for example, oil or fat.

 The primary inertia mass 1402 consists essentially of a component 1413, which is made of sheet metal. Component 1413 comprises a flange-shaped zone 1414, oriented substantially radially and which carries radially inward an axial appendage 1415, forming a single piece with it and which is surrounded by holes or perforations 1407. The bearing 1406a has a row of bearing balls 1406 is mounted by its inner ring 1416 radially on the outside on the end part of the axial appendage 1415. The outer ring 1417 of the bearing 1406a carries the second mass of inertia 1403, arranged essentially like a flat disc shaped body.

To this end, the inertial mass 1403 comprises a central recess in which is housed the bearing 1406a.

The zone 1414 oriented essentially radially is extended radially outward by a zone 1418 arranged in the form of a shell, which at least partially surrounds the energy accumulators 1410, at least on their outer periphery, and which ensures their guiding or supporting them. The body 1419 arranged in the form of a shell and fixed on the zone 1418 partially surrounds the periphery of the energy accumulators 1410. The body 1419 is welded to the sheet metal body 1413 (at 1420).

The toroidal zone 1412 is divided, by considering it in a circumferential direction, into different housings in which the energy accumulators 1410 are arranged. By considering a circumferential direction, the different housings are separated from each other by stress areas of the energy accumulators 1410, these areas may be pockets formed by stamping in the sheet metal part 1413 and the shell-shaped body 1419. The stress areas 1421 of the energy accumulators 1410, which are connected to the second inertia mass 1403, are supported by the clutch cover 1422.

 The biasing zones 1421 are formed by radial arms 1421 which, in the example shown, are arranged in the axial zone 1423 of the clutch cover 1422 and which penetrates radially into the annular volume 1412, and in particular between the ends energy accumulators 1410 adjacent in a circumferential direction. The lid zone 1423 oriented axially envelops or surrounds by a part 1423a the second mass of inertia 1403 and it is firmly connected thereto, for example by means of bosses formed in the part 1423a and which hook in corresponding recesses of the inertial mass 1403, or else by another method of fixing.

 The clutch cover 1422, centered on the external contour of the inertia mass 1403, has at its end opposite to the stressing zones 1421 an annular zone 1426, oriented essentially radially inwards and on which is mounted pivotally an annular spring 1427 acting as a lever with two arms. This annular spring 1427 biases, by zones located radially more towards the outside, a pressure plate 1428, so that the friction linings 1429 of the clutch disc 1405 are clamped axially between the second inertia mass 1403 and the pressure plate 1428. An elastic suspension 1465 is provided between the friction linings 1429.

 As shown in Figure 35, the annular volume 1411, or its toroidal zone 1412, are arranged essentially radially outside the extreme outer contours of the second inertia mass 1403. Consequently, the component 1413, used for the articulation of the first inertia mass 1402 on the crankshaft of the internal combustion engine, carrying the toroidal zone 1412 and adjacent to the internal combustion engine, as well as the second inertia mass 1403, can be arranged practically directly opposite each other, over a relatively large radial distance and creating an interior volume or air gap 1430, that is to say be placed practically in positions spaced apart one from the other by a small distance, which makes it possible to obtain an axially very compact structure for the group consisting of the flywheel 1401, the clutch 1404 and the clutch disc 1405.

The sealing of the annular volume 1411 is ensured by a seal 1431, which acts between the interior areas of the radial wall 1419 and the exterior peripheral surface of the cover 1422.

 Advantageously, this intermediate volume 1430 can be used to cool the flywheel 1401, and in particular by passing a flow of cold air through this intermediate volume 1430. To produce such a circulation, on cooling air, the second mass of inertia 1403 comprises, radially inward of the friction surface 1432, axial recesses 1433, which extend in the direction of the component 1413 located on the engine side and which open into the intermediate volume 1430. To improve cooling, the second mass d inertia 1403 may include other axial passages 1435, which are located radially further outwards and which are in communication, on the side opposite to the friction surface 1432, with the intermediate volume 1430, opening radially to the outside the friction surface 1432 on the side of the inertial mass 1403 which is directed towards the clutch 1404. Radially inward of the passages or of course ts 1433, the inertia mass 1403 includes other passages 1434 which are used in particular for receiving or passing the fixing bolts 1408.

 To seal the annular chamber 1411, partly filled with a viscous agent, another seal 1436 is provided which is constituted by a part in the form of a membrane or annular spring, which extends radially in the intermediate volume 1430.

 The shell-shaped body 1419 carries a starter ring gear 1439, which is fixed to this body by welding.

 The clutch group, consisting of the clutch 1404 and the clutch disc 1405, forms with the flywheel with two inertia masses 1402 + 1403, shown in Figure 35, a unitary assembly A which is pre-assembled , shipped and stored in this form and which can be bolted to the crankshaft of an internal combustion engine in a particularly simple and rational way. For the assembly of the unitary assembly A, initially the clutch 1404 and the second mass of inertia 1403 are connected with each other with interposition of the clutch disc 1405.Then, the sub-assembly consisting the clutch 1404, the flywheel 1403 and the clutch disc 1405 is assembled axially with the component 1413, then the shell-shaped body 1419, which is received on the outer edge 1423 of the clutch cover 1422, is applied against the external zones of the component 1413 and can be welded (in 1420) with it. Before the axial assembly of the two components 1413 and 1419, the springs 1410 are put in place in the toroidal zone 1412. The unitary assembly A also comprises the bearing 1406, which has already been put in place on the appendix axial 1415. In the holes 1407 of the flange-shaped zone 1414, the fixing bolts 1408 have already been pre-assembled or put in place, which are in particular in the form of hexagon socket bolts 1408. At this Regarding, the bolts 1408 are located in a position corresponding to the lower half of Figure 35. The bolts are maintained in this position, without possibility of loss, in the group or in the unitary assembly A.

 The clutch disc 1405 is tightened, in a pre-centered position relative to the axis of rotation of the crankshaft, between the pressure plate 1428 and the friction surface 1432 of the second inertia mass 1403 and furthermore it is located in a position such that the openings 1443 provided in the clutch disc are placed so that, during assembly of group A on the crankshaft of the internal combustion engine, it is possible to pass a bolting tool. It is obvious that the openings 1443 are smaller than the heads 1440 of the bolts 1408, so that it is thus ensured that they are properly maintained, and without the possibility of loss, in assembly A.

 There are also provided in the annular spring 1427, and in particular in the region of its legs 1427a, openings or recesses 1444 allowing the passage of the bolting tool. The recesses 1444 can be provided in such a way that they constitute enlargements or enlargements of the slots provided between the lugs 1427a. The openings 1444 provided in the annular spring 1427, the openings 1443 provided in the clutch disc 1405 and the openings 1434 provided in the inertia mass 1403 are located mutually in coincidence in an axial direction, and in particular so that also in the case of an asymmetrical arrangement of the holes 1407 which is necessary for the following mounting of the assembly A on the crankshaft, it is possible to pass an assembly tool, such as for example a hexagon socket wrench, correctly through the openings 1444, 1427 and 1443 and to engage it in the recesses of the heads 1440 of the bolts 1408 .

 A complete assembly A of this kind considerably facilitates the mounting of the flywheel because it makes it possible to eliminate various operations, such as the operations previously necessary for centering the clutch disc, engaging the clutch disc, setting up the clutch, insertion of the centering pin, centering of the clutch disc itself, fitting of the bolts, bolting of the clutch and removal of the centering pin.

 The friction clutch 1404 includes a wear correction means 1445 which provides wear compensation, in a manner analogous to what has been described in relation to FIGS. 1 to 27, by means of a detector spring. 1446 and a wear correction ring 1447.

 The torque transmission device 1501, shown in FIG. 36, comprises a back-pressure plate 1503, which can be connected in a non-rotating manner to the crankshaft K of an internal combustion engine and on which a friction clutch 1504 is fixed with interposition of a clutch disc 1505.

The clutch disc 1505 can be mounted on the input shaft of a transmission, not shown in detail.

 The clutch cover 1522 has an axially oriented zone 1523 which maintains axially, radially outside, the pressure plate 1528 and the friction linings 1529 of the clutch disc 1505. The end part 1523a of the cover zone 1523, of tubular shape or in the shape of a sheath, envelops or surrounds the back-pressure plate 1503 and it is firmly connected thereto by means of bosses 1524 formed in the part 1523a and hooking in provided recesses on the outer periphery of the back pressure plate 1504. The cover 1522 and the back pressure plate 1503 can however also be connected in another way, for example by means of welds, by assemblies by means of bolts or pins, which are also arranged in a radial direction.

 The clutch cover 1522, centered on the outer contour of the back pressure plate 1503, has an annular zone 1526, oriented essentially radially inward and on which an annular spring 1527 is pivotally mounted. acting like a two-armed lever. By zones located radially on the outside, the annular spring 1527 urges the pressure plate 1528, so that the friction linings 1529 are clamped axially between the counterpressure plate 1503 and the pressure plate 1528. The annular spring 1527 includes radial lugs 1527a for actuating the clutch 1504 by means of a declutching system.For torque transmission between the pressure plate 1528 and the cover 1522, torque transmission means are provided. , preferably in the form of elastic blades 1521, which are rigidly connected at one end with the cover 1522, preferably by means of rivets 1521a, and the other ends of which are also connected by means of rivets with the pressure plate 1528. Preferably, the connection between the pressure plate 1528 and the elastic blades 1521 is ensured by the inter through what is called a blind rivet connection, as shown in an area of radial extent of the friction linings in the upper half of Figure 35.

 The friction clutch 1504, or the torque transmission device 1501, includes a wear correction means 1545 which provides wear compensation in a manner analogous to that which has been described in relation to FIGS. 1 to 27 , by means of a detector spring 1546 and a wear correction ring 1547.

 In FIGS. 35 and 36, the contact counter-ramps, formed by imprint directly in the material of the cover, are arranged so that they create, in the direction of rotation of the friction clutch, respectively an opening of air passage (1547a in Figure 36). By means of an arrangement of this kind, one obtains1 upon rotation of the corresponding friction clutch, better cooling of the latter by means of forced air circulation. In particular, this also provides cooling of the wear corrective ring 1447 or 1547, made of plastic, which allows to significantly reduce the thermal stress of this ring.

 The friction clutch 1504 or the back-pressure plate 1503 is fixed in a non-rotating manner, but nevertheless with a limited possibility of axial translation, on the output shaft K of the internal combustion engine by means of a component 1550 which is elastic or has elastic flexibility. In the embodiment shown, this component 1550 is arranged in the shape of a disc and its rigidity is calibrated so that the axial and three-dimensional or bending oscillations generated by means of the crankshaft K in the friction clutch 1504 are damped or counterbalanced by the elastic component 1550 to a degree such that it is ensured that the friction clutch 1504 operates correctly, and in particular its wear corrective means 1545. This must therefore be obtained, via of the axially elastic component 1550 insulation as large as possible of the clutch group 1504 with respect to the axial and bending oscillations of the output shaft of the internal combustion engine, such as the crankshaft. One can thus avoid a malfunction of the clutch group 1504 or of its wear correction means 1545. Without the aforementioned decoupling of the clutch group 1504 with respect to the crankshaft K, there could be an undesirable wear correction by the corrective means 1545, and in particular under the effect of the mass of the components and the accelerations acting on them under the effect of oscillations. In the absence of the component 1550 filtering the oscillations, we should therefore take particular account , in particular during the design of the wear correction means 1545, inertia forces produced by the components acting on this wear correction means, so that it would be necessary to carry out a complicated adaptation and / or to make intervene additional means to avoid actuation of the wear correction means which is not caused by wear of the friction linings.

 In the torque transmission device 1501 of FIG. 36, the wear compensating means 1545 acts between the clutch cover 1522 and the annular spring 1527. The torque transmission device 1501 could however also be fitted with a clutch friction according to Figures 28 to 34, that is to say a friction clutch in which the wear compensating means acts between the annular spring and the pressure plate biased by the latter.

The back pressure plate 1503 is rigidly connected, radially to the outside, by means of bolts 1551 with the axially elastic component 1550 in the form of a disc. Instead of bolts 1551, blind rivets could also be provided, such as those used for fixing elastic blades on the pressure plate 1428 and shown in the upper half of Figure 35. Radially inside the zones d hinge 1551 between the disc-shaped component 1550 and the back-pressure plate 1503, there is between these two components 1550 and 1503 an axial gap 1552 which determines in one of the axial directions the maximum amplitude of the axial oscillations between the two components 1550 and 1503. By abutment of the radially inner zones of the back-pressure plate 1503 against the disc-shaped component 1550, it is possible to limit the maximum axial offset of the friction clutch 1504 in the direction of the crankshaft
K of the internal combustion engine. However, under normal operating conditions, such as when the engine is operating properly, such contact does not occur. The annular back pressure plate 1503 surrounds an axial projection 1553, which is part of a component 1554 in the form of a ring or disc. This disc-shaped component 1554 can be securely connected with the radially inner regions of the elastic disc 1550. The elastic disc 1550 and the disc-shaped component 1554 are centered on an annular projection 1555 of the crankshaft K and are securely connected to this by means of bolts 1556. The radially inner zones of the disc-shaped component 1550 are thus clamped axially between a front surface 1557 of the crankshaft K and the component 1554 of annular shape.

 The axial appendage 1553 of the component 1554 comprises, at its end opposite to the elastic component 1550, radial zones 1558 which limit the axial offset of the friction clutch 1504, or of the back-pressure plate 1503, in the other direction axial.

Between the zones 1558 and the backpressure plate 1503, there is an axial gap 1559 when the elastic component 1550 is not stressed. This interval 1559 has in the axial direction a dimension similar to that of the interval 1552. The back-pressure plate 1503 can be mounted practically without play, by means of its inner peripheral surface, on the appendage or projection axial 1553 so that it can thus be ensured that the back pressure plate 1503 is axially guided. It may however be wise to provide between the inner peripheral surface of the back pressure plate 1503 and the axial appendage 1553 at least a small air gap so that, under normal operating conditions, there is no contact between the two aforementioned components.

 According to another variant, it is possible, to dampen the oscillations still transmitted despite the existence of the elastic component 1550, to obtain a means canceling the oscillation energy.

One such means can be constituted by a friction connection, such as that shown by way of example in Figure 37. In the alternative embodiment shown in Figure 37, it is provided between the interior areas of the counter plate pressure 1503 and the external peripheral surface of the appendix 1553 a damping means 1560, which can be constituted for example by a ring wavy in a circumferential direction and whose corrugations are oriented radially. This ring 1560 is mounted with clamping in a radial direction, so that friction is generated for example between this ring 1560 and the inner peripheral surface of the backpressure plate 1503 during the existence of axial oscillations. There is thus a damping effect in the bearing area of the back-pressure plate 1503 on the appendix 1553. The corrugated ring 1560 can be divided, that is to say open, around its periphery.

 Radially outside, the disc-shaped elastic component 1550 carries a starter ring gear 1561.

 The disc-shaped component 1550, the back-up plate 1503, the clutch disc 1505 and the friction clutch 1504 constitute a unitary assembly which is pre-assembled, shipped and stored in this form and which can be bolted on the crankshaft K of an internal combustion engine in a particularly simple and rational way. The fixing bolts 1556, in the form of hexagon socket bolts, are also already pre-installed, that is to say incorporated in the unitary assembly, and in particular in a manner preventing losses.

 The clutch disc 1505 is tightened, in a pre-centered position relative to the axis of rotation of the crankshaft, between the pressure plate 1528 and the counter-pressure plate 1503 and it is also maintained in a position such that the openings 1562 provided therein and located radially inside the elastic damper of the clutch disc 1505, are arranged in a position such that, during the assembly operation of the assembly on the crankshaft K, a 1563 bolting tool can pass through. Also the annular spring 1527 comprises, if necessary, openings or cutouts 1564 for the passage of the bolting tool 1563.

The openings or cutouts 1564 of the annular spring 1527 are coincident with the openings 1562 of the clutch disc 1505 so that the mounting tool, such as for example a hexagon wrench 1563, can be properly engaged in the 1556 bolt head bolt slots.

 As has already been described in relation to the other embodiments, it is possible, by using a friction clutch 1504 comprising a means 1545 which at least compensates for the wear of the friction linings, to optimize the arrangement of the friction clutch, in particular of the energy accumulator 1527 exerting the tightening force for the clutch disc. This energy accumulator can be arranged in such a way that it produces practically only the clamping force of the clutch disc which is necessary to obtain the desired torque. With the aid of the wear correction means 1545, it is ensured that the energy accumulator 1527 retains practically the same mounting position in the engaged state of the clutch assembly 1501 and throughout its service life. .In addition, using the means also provided in the clutch assembly 1501 and in the form of an elastic suspension 1565 of the friction linings, which ensures during the disengagement and the clutch of the group d clutch 1501, and over at least a partial area of the actuating stroke of the pressure plate 1528, a gradual decrease or increase in the torque that can be transmitted by the clutch unit, it is possible to obtain a reduction in forces release or the evolution of the release forces. Thus, by a corresponding adaptation between the forces exerted by the means such as an elastic suspension of friction linings and the energy accumulators acting on the pressure plate, or else by a corresponding adaptation of the force-stroke characteristic curves, of determine the desired evolution of the declutching forces. It is thus possible to design the elastic component 1550 in an optimal way to fulfill the desired function of damping of the axial, bending or three-dimensional oscillations because the reduced declutching forces which act on this elastic component are of secondary importance. Thus the actuation forces necessary for disengaging the clutch can be assisted by the aforementioned component without a substantial axial offset of the clutch group.

 The invention is not limited to the exemplary embodiments shown and described, but it also includes variants which can be obtained by combining different features or elements which have been described in relation to the various embodiments. In addition, certain features or modes of operation which have been described in relation to the Figures may represent, by considering them individually, a particular invention.

Claims (26)

 1 - Friction clutch, in particular for motor vehicles, comprising a pressure plate which is connected in a non-rotating manner, but however with the possibility of limited axial translation, with a casing, between which and the pressure plate acts at least one spring thrust, which urges the pressure plate towards a clutch disc which can be clamped between this plate and a back-pressure plate, such as a flywheel, and where a corrective means is provided compensating for the wear of the linings of friction of the clutch disc and which minimizes, during the service life of the friction clutch, the relaxation of the thrust spring normally occurring under the effect of the wear of the friction linings, clutch characterized in that the wear compensation function of the corrective means can be blocked in dependence on a state of actuation of the friction clutch and / or in dependence on an operating state of the em friction clutch.  2 - Friction clutch according to claim 1, characterized in that the blocking takes place as a function of the speed of rotation.  3 - Friction clutch according to one of claims 1 or 2, characterized in that the blocking is effected by means of elements dependent on the centrifugal force.  4 - Friction clutch according to one of claims 1 to 3, characterized in that the blocking of the corrective means is effected by means of a friction conjugation.  5 - Friction clutch according to one of claims 1 to 4, characterized in that the blocking of the corrective means is effected by means of a combination of shapes.  6 - Friction clutch according to one of claims 1 to 5, characterized in that the blocking of the corrective means is effected by means of weight members sensitive to the centrifugal force.  7 - Friction clutch according to one of claims 1 to 6, characterized in that the corrective means is blocked above a determined limit of the speed of rotation.  8 - Friction clutch according to one of claims 1 to 7, characterized in that the corrective means is blocked when the drive motor rotates at idle speed and at higher speeds.  9 - Friction clutch according to one of claims 1 to 8, characterized in that the corrective means can produce a corrective adjustment below the idle speed.  10 - Friction clutch according to one of claims 1 to 9, characterized in that the corrective means can produce a corrective adjustment practically only for a zero rotation speed.  11 - Friction clutch according to one of claims 1 to 10, characterized in that the corrective means can compensate for wear of the friction linings only when the friction clutch does not rotate.  12 - Friction clutch according to one of claims 1 to 11, characterized in that the corrective means is blocked when the friction clutch is at least approximately completely disengaged.  13 - Friction clutch according to one of claims 1 to 12, characterized in that the corrective means comprises a corrective adjustment ring whose corrective adjustment function is blocked when the friction clutch is at least approximately disengaged.  14 - Friction clutch according to one of Claims 1 to 13, characterized in that the corrective means comprises a corrective adjustment ring which is tightened against a clutch component, when the friction clutch is at least approximately disengaged, from so that its corrective adjustment function is blocked.  15 - Friction clutch according to one of Claims 1 to 14, characterized in that the corrective adjustment means produces a practically constant stress on the pressure plate by the push spring, the friction clutch comprises actuation means for producing a clutch and a disengagement as well as a system which ensures, during the disengagement process, in a partial area of the actuating stroke of the actuating means, a gradual reduction in the torque which can be transmitted by the friction clutch or by the clutch disc.  16 - Friction clutch according to one of claims 1 to 15, characterized in that this friction clutch comprises a system which ensures, during the disengagement process, in a partial zone of the axial stroke of offset of the zones of the pressure plate which are requested by the thrust spring, a gradual reduction in the torque that can be transmitted by the friction clutch.  17 - Friction clutch according to one of claims 15 or 16, characterized in that the aforementioned system is provided in the force transmission line between the actuating means or the thrust spring and the fixing zones of the housing on the plate back pressure.  18 - Friction clutch according to one of claims 15 or 16, characterized in that the above system is provided in the force transmission line between the actuating means or the thrust spring and the friction surface of the pressure plate.  19 - Friction clutch according to one of claims 16 to 18, characterized in that the aforementioned system is provided axially between the friction linings of the clutch disc.  20 - Friction clutch according to one of claims 16 to 19, characterized in that the aforementioned system creates a possibility of elastic axial adaptation between components of the clutch, said system being arranged so that, when the clutch is open, the force acting on the system is the smallest and that, when the clutch is closed, the force acting on the system gradually increases to the maximum, this increase occurring at least in a partial area of the stroke closing.  21 - Friction clutch according to one of claims 16 to 20, characterized in that the aforementioned system ensures the gradual reduction or the gradual increase of the torque that can be transmitted by the friction clutch in at least approximately 40 to 70% of the actuating stroke of the actuating means.  22 - Friction clutch according to one of claims 1 to 21, characterized in that the thrust spring comprises, at least in part of the clutch clutch disengaging stroke, a force-stroke curve which is decreasing.  23 - Friction clutch according to one of claims 1 to 22, characterized in that the thrust spring consists of an annular spring, which can pivot on one side around a pivoting zone of annular shape carried by the housing and which requests on the other side the pressure plate.  24 - Friction clutch according to claim 9, characterized in that the annular spring comprises an annular body, from which protrude tongues directed radially inwards and constituting the actuating means.  25 - Friction clutch according to one of claims 23 or 24, characterized in that the annular spring is pivotally held in the housing between two supports.  26 - Friction clutch according to one of claims 23 to 25, characterized in that the annular spring has a force-stroke characteristic curve of sinusoidal shape and, in the engaged state of the friction clutch, the operating point of the spring annular is located on the decreasing zone of the characteristic curve according to the first maximum of force and the annular spring has a ratio of forces ranging between approximately 1: 0.4 and 1: 0.7 between the first maximum of forces and the minimum of force located as a result.