RU2156896C2 - Coupling unit and its friction clutch - Google Patents

Abstract

FIELD: devices ensuring normal operation of machines. SUBSTANCE: coupling unit has friction clutch and control facilities. Friction clutch has case 2 accommodating hold-down strip 3, plate spring 4, uncoupling member 22, and friction straps 7 whose position, considering their wear, is adjusted by means of springs; the latter turn adjusting rings 17 with wedge-shaped slopes 19. Coupling force is applied to turn center 12 of plate spring 4. EFFECT: improved reliability and reduced manufacturing cost. 117 cl, 51 dwg

Description

 The invention relates to a clutch assembly with a friction clutch provided with a pressure plate that can be connected without rotation, but with axial limited displacement with a support disk, the pressure plate being pressed with at least one pressure spring in the direction of the clutch plate clamped between the pressure plate and support disk, and there is at least one adjustment device that compensates for the wear of the friction linings of the clutch disc, which causes the application of almost constant force I to the pressure plate with the help of a compression spring, in addition, the friction clutch has controls that are controlled by an axially disengaging release element, for example, a disengaging fork mounted for rotation in the transmission housing.

 A clutch assembly of a similar design and similar control is described in French Patent Laid-Open No. 2582368. Controls of such a clutch assembly may be loaded with a disengaging system or disengaging means, as proposed, for example, in US Pat.

 In clutch assemblies or friction clutches with a built-in adjusting device that compensates for at least wear on the friction linings of the clutch disc, a problem arises, especially in connection with the so-called mechanical disengagement systems, in which the clutch pedal is transmitted through the traction and lever system and / or flexible traction with the intermediate inclusion of at least one trip support on the control means of the friction clutch, which consists in the fact that due to the tolerances available throughout the kinematic chain does not provide a constant axial position of disengagement member acting on the control means, the control means relatively of loaded zones may be created whereby a relatively large uncertainty disengaging clutch stroke control or stroke transmitted to the control means. Due to this uncertainty, the adjustment function may at least deteriorate, and in extreme cases the regulation function of this device may not be preserved. In addition, there may be cases when the controls cover an unacceptably large distance, as a result of which an unintended adjustment can occur, due to which either the clutch does not open flawlessly, or the preliminary compression or position of the clamping spring changes, as a result of which the force created by it becomes insufficient in order to ensure reliable transmission of torque.

 The basis of the invention is the task to eliminate these disadvantages and create a coupling node of the above type, which provides a reliable function of the adjusting device that compensates for the wear of the friction linings. In addition, the unit can be manufactured in a particularly simple manner and without high costs.

 According to the invention, this is achieved due to the fact that a device is provided that compensates for the axial displacement of the position of the control means or the axial displacement of the parts of the controls that are affected by the trip element relative to the trip element or trip means. Such a device is particularly preferred for coupling units, in which, according to a further development of the invention, depending on at least the wear of the friction linings, the control means are displaced in the axial direction of the disengagement movement, since due to this, a virtually clearance-free transmission of forces between the trip element or the trip means and controls. This also ensures that the controls can always have the same offset. Thus, when transferring forces between the trip element and / or the trip means and the control means, there are practically no gaps.

 It is particularly preferable to provide or provide an equalizer operation along the axis between the trip element and the controls. The equalization device may, however, be located in another place, for example, in accordance with the action between the trip element and the trip means. In combination with the invention under consideration, it is preferable if the trip element is mounted on an axial guide, preferably located on the drive side, for example, on a guide tube surrounding the input shaft of the transmission.

 It may be deemed advisable, especially in the case of clutch assemblies with a friction clutch having a housing mounted on a support disk, for example in the form of a cover made of sheet material with a bottom facing the trip element, to position and ensure the operation of the leveling device along the axis between the controls and the bottom . In addition, it may be preferable if the clamping spring is made in the form of a disk spring located between the clutch housing and the pressure plate having a spring-shaped annular main body and radially extending from it and forming tongue control means.

 To ensure reliable adjustment by means of a leveling device, it can be especially advantageous if it, when the clutch unit or friction clutch is in the engaged state, provides the desired adjustment automatically or independently.

 The equalization device may have an annular element, which, even in the on state of the friction clutch, is adjacent axially to the controls. Using this ring-shaped structural element, it is possible to equalize the varying distance between the loaded areas of the control means and the trip element. For the functioning of the leveling device, it is preferable if it has an adjustment protruding upward in the axial direction or a ramp which can be made on an annular structural element.

 Incident slopes for tuning can interact with cylindrical rolling bodies or rolling bodies, which have a shape similar to spherical. Particularly preferred, however, is the interaction of the ramp with the ramp, since due to the appropriate choice of the angle of this ramp, self-braking can be carried out with axial jamming of the ramp. Oncoming counterclones can also be located on an annular structural element.

 In order to reduce the manufacturing cost of the friction clutch, it may be preferable if at least one part of the equalization device is made of plastic. Plastic parts of this type can be injection molded. Thermoplastics such as polyamide are particularly suitable as plastics. Structural elements having adjusting slopes can be particularly preferably axially displaced when controlling the clutch assembly or friction clutch. In addition, it may be especially advisable if the structural elements bearing the oncoming slope and the oncoming counterclone have the possibility of rotation relative to each other, and one of these structural elements can be fixed without the possibility of rotation relative to the friction clutch, in particular relative to the clutch housing.

According to another idea of the invention, the equalization device can be designed so that it, when viewed in the direction of disengagement of the clutch assembly, acts or adjusts freely, but it is self-braking in the opposite direction to the disengagement. To this end, the ramp and / or ramp counterclone are designed so that they have an axial angle of elevation lying in the range from 5 to 20 ^o , preferably in the range from 7 to 11 ^o . It is preferable to perform an adjustment slope so as to provide self-braking due to friction. In any case, it should be ensured that the control slope has the effect of self-braking and that additional means are not required to prevent an undesirable return stroke, however, if necessary, such means can be provided.

 To ensure reliable operation of the leveling device, it may be advisable if at least one structural element bearing a ramp and / or ramp is made spring loaded in the direction of regulation. In this case, the loading of the spring preferably takes place in such a way that practically no effect is exerted on the function of the remaining springs, in particular clamping springs or plate springs and springs, which load the plates that are axially bent. A particularly preferred embodiment can be achieved due to the fact that parts having a ramp and a ramp are loaded or tightened in the direction of regulation with at least one energy accumulator located between them, for example coil springs. Using such a tightening, these parts move apart when viewed in the axial direction, in opposite directions, i.e. They are moved along the axis from each other with the help of an energy accumulator and an adjustment slope. Due to this, when the clutch is switched off, the equalization device can be clamped unobtrusively between the loading zones of the controls and the clutch cover and / or the trip element.

 According to another particularly preferred embodiment of the invention, the clutch assembly may have a device for restricting the disengagement of at least controls. For this, a restrictive stop may be provided, limiting the stroke of the trip element and / or trip means in the direction of trip. The limit stop can preferably be made in such a way that the part forming the equalization device abuts the clutch cover after a certain trip. However, the restriction can be carried out due to the fact that the restrictive element has zones that, after a certain trip, abut against a part that is rigidly fixed along the axis. In addition, it may be preferable if the trip element has a restriction in the direction of inclusion, which can also be made in the form of a stop. In a preferred embodiment, the equalization device can be designed in such a way that with its help the trip element has axial support in the engaged state of the clutch assembly. A constant control stroke of the clutch assembly can be ensured due to the fact that the part forming the equalizing device has, in the direction of disengagement and clutch, zones of effective travel restriction interacting with the stop zones. This part, in a preferred embodiment, can be made in the form of a part of an equalization device loaded with a trip element, the limit stops being mounted on the coupling body or can be formed by this housing. Limitation of the control stroke of the clutch assembly can, however, be ensured due to the fact that corresponding stops are provided on the part guiding the trip element in the axial direction.

 In addition, these stops interact with the part, which is connected to a support ring, made non-continuous around the perimeter. The restriction of the trip course in at least one axial direction can, however, also be carried out between the circumferential support ring of the trip element and the rotating part, for example, the clutch housing.

 According to a preferred embodiment of the invention, it may be preferable, in particular, to reduce to a minimum level of the tripping force or the maximum necessary tripping force, if there are means which, during the tripping process, cause, at least part of the control path of the controls, a gradual decrease in the moment transmitted from the friction clutch and clutch disc. These means can be performed, for example, due to the so-called spring-loaded pads, which is between the friction pads of the clutch disc, sandwiched between the pressure plate and the support disk. A particularly preferred embodiment of the friction clutch according to the invention can be obtained by virtue of the fact that the clamping spring, which can preferably be in the form of a disk spring, is rotatably mounted on the housing. Between two linings, one of which is facing the pressure plate, is spring-loaded in the direction of the pressure disk spring, and the maximum disengagement force transmitted by the pressure spring when the friction clutch disengages to the spring-loaded liner increases with wear of the liner and becomes greater than the reaction force or supporting force acting on a spring-loaded pad. When using leaf spring elements designed to transmit torque between the pressure plate and the clutch housing and / or the so-called spring load of the plate, as is known, for example, from the laid-out application of Germany 3631863, the forces exerted by these springs on the pressure plate when determining the force which acts on the spring-loaded pad, especially because these forces are superimposed on each other. This means that the fast-acting disengaging force that increases when there is sufficient wear on the linings must be greater than the resulting force obtained from the aforementioned forces and diverted from the changing diameter of the cup spring to allow adjustment. Particularly appropriate may be the possibility of axial displacement of the spring-loaded plates. In a preferred embodiment, the pressure plate spring may have such a characteristic that, based on its structurally determined location in the friction clutch during expansion, due to the wear of the friction linings, the force exerted by it and, at the same time, the level of passage of the release force increase, and in the position changed or deformed Compared with the originally mounted position, they decrease during the trip from the set maximum value. By using this arrangement and laying down the pressure plate springs, it is possible to ensure that with the wear of the linings occurring, it is possible to adjust the balance between at least the maximum disengaging force of the friction clutch and the force that counteracts the spring plate or the resulting force exerted in the area of the running diameter on the pressure plate springs.

 The clutch assembly or friction clutch can be made in a preferred way so that the axially displaced spring loaded plate can move with the pressure plate beyond the wear reserves of the friction clutch. During the friction clutch during the service life of the adjusting device, continuously or in small steps, the adjustment spring may slightly shift in the direction of the pressure plate. Due to this displacement, it can be ensured that the disk spring resting on the pressure plate is deformed, so that the force exerted by it decreases, balancing the force that counteracts the spring-loaded pad or the already mentioned resulting counter force with the decoupling force. Thus, when the spring-loaded pad is displaced, the maximum decoupling force of the clutch or the pressure plate spring decreases again.

 It may be particularly preferable if the compression disk spring in the friction clutch is mounted in such a way that there is a falling force-time characteristic on at least part of the disengagement zone, preferably practically on the entire disengagement zone of the clutch. The location of the clamping spring may be such that in the disengaged state of the friction clutch, the clamping spring reaches almost the lowest point or minimum of its sinusoidal force-displacement characteristic.

 The force counteracting the spring-loaded pad can be created in a preferred manner by an energy accumulator, which generates a substantially constant force in at least a predetermined control range. Especially suitable for this is a disk spring which is suitably made and inserted in a predetermined position into the friction clutch.

 The invention is not limited to the friction clutches described above, but can be widely used for friction clutches or clutch assemblies with an adjustment device that compensates for the wear of the friction linings of the clutch discs.

 In addition, the invention relates to a friction clutch, in particular for cars, containing a pressure plate that is connected without the possibility of rotation, but with the possibility of limited axial displacement with the housing, and between the housing and the pressure plate is axially clamped by a clutch plate spring, which, on the one hand , can rotate around a pivot bearing installed in the housing, and on the other hand, the pressure plate is loaded in the direction of the clutch plate sandwiched between it and the support disk, similarly to ovomu wheel, which has a control device that compensates for wear on the clutch disc lining.

 Automatic adjusting devices, which should create an almost constant application of the load on the pressure plate by means of a pressure plate spring, are known, for example, from the laid-out application 2916755 or 3518781. At the same time, adjusting devices having the possibility of displacement depending on at least one sensor are located or act between the pressure plate and the pressure plate spring. Due to the deflection of the pressure plate to the housing through the tangentially arranged leaf springs, the force of which, if it is directed against the pressure force of the disk spring, can be relatively small, it is possible to axially reject the relatively small mass of the pressure plate with the friction clutch disengaged, i.e. in this case, to lift from the cup spring, as a result of which the clutch functioning is not only impaired, but there is even a risk for safe operation, because, in particular, the adjustment device is adjusted in the open state until the pressure plate adheres to the clutch plate, as a result of which the clutch does not can carry out the separation. For this reason, such adjustment devices are not used in practice.

 The basis of the invention is the task of eliminating these shortcomings and creating adjusting devices of the type indicated above, which in practice can be applied on a wide basis under different operating conditions, which have a simple design and high reliability with a long service life, which require a small area and which are cheap in manufacture. In addition, the necessary decoupling forces must be small, remain insignificant during the service life, and, in addition, the service life of the friction clutches should be significantly increased.

 According to the invention, this problem is solved due to the fact that in the friction clutch with a pressure plate loaded by a Belleville spring, in which the clamping force is created by a Belleville spring, which, on the one hand, rests on one structural element, such as a housing, and on the other hand, has the ability turning around a rotary support located in a circle around the housing, between the cover and the disk spring there is an independently functioning adjusting device that moves depending on wear an overlay on the side of the housing away from the housing, which has the possibility of further transportation from the feeder, and the cup spring is under the action of a supporting force in the direction of the rotary support. This supporting force is constantly available, due to which the Belleville spring has a support relative to the disengagement force only with a force closure, but not using means with a geometric closure. In this case, the spring in its working area has a degressive characteristic, in particular, such that the supporting force and the force of the Belleville spring are coordinated with each other so that the supporting force in the initial position of the Belleville spring and without changing taper and after the process of breaking the Belleville spring more than the force counteracting the supporting force created by the cup spring; when the taper of the cup spring changes due to wear, the support force on the part of the trip of the cup spring is less than the force created by the cup spring acting against the support force. In this case, the supporting force can be created by one single spring element or at least basically one single spring element or a system of spring elements. The words "support force" means the sum of the forces of all springs acting against the force of action of a Belleville spring, if they occur, i.e. in the same way, or just the forces created by leaf springs (transmitting torque or retraction), (residual) spring loaded plates or their "replacement".

 As an energy accumulator, which generates at least mainly a supporting force, a spring can be used in an expedient manner, which, after adjustment, changes its shape, for example a disk spring. The energy accumulator creating the supporting force can also be made in the form of a flat spring.

 A Belleville spring creating a support force can be applied directly to the Belleville spring, for example, at a radially axially displaceable height located on the side of the bearing cover.

 Particularly preferred is the location of the adjusting device between the cup spring and the cover. The adjusting device may have, in a particularly preferred embodiment, on-going surfaces, for example slopes.

 The invention ensures that the disk spring is maintained for almost the entire service life of the friction clutch of constant taper or tension in the engaged state of the friction clutch and the load is practically constantly applied to the pressure plate and thereby to the clutch disc, regardless of the wear of the friction linings, pressure plate or other elements, for example, cover supports either from the side of the cover or from the side of the pressure plate, plate springs or friction surfaces of the disk flywheel.

 Using the features according to the invention, it is furthermore ensured that the weight of the pressure plate does not increase due to the weight of the adjusting device. In addition, it is located in the zone in which it is protected from the effects of abrasion of the plate and where it is further removed from the heat source created by friction.

 A particularly preferred embodiment of the friction clutch according to the invention can be achieved due to the fact that the pressure disk spring according to the invention is rotatably mounted in the housing between two plates, one of which, facing the pressure plate, is spring-loaded in the direction of the pressure disk spring, and the force acting from when the clutch disengages the spring plate, when the plate is worn, it increases and becomes greater than the force that counteracts the spring on the pad or supporting force. In this case, the pressure plate spring has such a characteristic that, based on its location in the friction clutch in the unloading direction determined by the wear of the linings, the force created by it and, at the same time, the necessary decoupling force increases first, but at the position changed or deformed in comparison with the initial position during installation, the force created by it during tripping is reduced. Due to this design and arrangement of the pressure plate spring, it is ensured that when the plates are worn, the balance between the force generated by the pressure plate spring on the plate during disengagement and the force opposing the spring-loaded plate can be adjusted, because when the supporting force is increased by the force of a disk spring transferred by a disk spring to the plate, the spring of the sensing element is displaced from the plate located on the side of the cover, and the adjustment The driven device can rotate further under the influence of the feeder. Due to this, the plate is displaced in the axial direction until the force created by the sensor stops further rotation and further axial displacement of the plate.

 As already mentioned, it may be particularly preferable if the compression disk spring is integrated in the friction clutch in such a way that it will have, at least part of the disengagement zone, preferably practically over the entire disengagement zone of the friction clutch, a falling force characteristic. In this case, the mounting position of the pressure plate spring can be such that, in the disengaged state of the friction clutch, the pressure disk spring reaches almost the minimum or the lowest point of its sinusoidal force-stroke characteristic or exceeds it.

 A counter-force applied to the spring-loaded pad is preferably created using an energy accumulator having a substantially constant force in at least a set sub-setting zone. For this purpose, a disk spring, suitably made and installed with a predetermined tension, is installed in the friction clutch.

 The adjusting device according to the invention can be particularly advantageously used in friction clutches equipped with a pressure plate spring, loading the pressure plate with radial upper zones and mounted in the housing radially located inside the zones between the two swivel plates. In such a design, a disk spring can act as a two-arm lever.

 However, the invention is not limited to friction clutch with cup springs, which simultaneously have a tongue-shaped release lever on a cup spring, but also extend to other clutch designs in which, for example, a cup spring is controlled via an additional lever.

 In order to ensure perfect wear control or optimum clamping force at the friction clutch, it can be especially preferable if the counter support provided on the side of the pressure disk spring opposite the spring-loaded plate is made so that it can automatically or independently displace axially in the direction of the pressure plate, but while this is independently or automatically blocked in the opposite direction using a special device. Counter support regulation, i.e. the support located on the side of the lid can be carried out with the help of an energy accumulator, which loads this counter support in the direction of the pressure plate or against the action of the pressure disk spring. Thus, the counter support can be automatically adjusted in accordance with the wear of the plates due to the displacement of the spring-loaded plates, and a clearance-free swivel bearing of the pressure plate spring can be provided.

 The counter-support may be axially displaceable by means of an adjustment device located between the pressure plate spring and the cover. In this case, the adjusting device may have an annular part, which, at least in the engaged state of the friction clutch, is axially loaded by a clamping cup spring.

 By turning the ring-shaped part during wear and during the trip, the pivot bearing can be adjusted in accordance with the wear of the pads. To this end, the adjusting device may have, in accordance with a particularly preferred embodiment (or the annular part of this adjusting device) axially rising ramps. In addition, it may be preferable if the annular part carries a counter-support, the latter being in the form of a wire ring. This wire ring can be inserted into the continuous annular groove of the part and connected to it by means of a geometric closure. In this case, the geometric closure can be made in the form of a latch.

 Oncoming slopes for adjustment can interact with cylindrical or having the shape similar to spherical rolling bodies. However, it may be especially preferable if the incline slopes function in conjunction with the incident counterclones corresponding to them, since in this case, self-braking can be carried out by axially clamping the incline due to the appropriate choice of the angle of these incline slopes. The oncoming contraclones may be located on an annular part, which may be located between the part bearing the oncoming slopes and the cover. However, a particularly simple construction can be obtained by arranging the oncoming counterclones in the housing. The latter can be particularly easy to implement in a housing made of sheet material, since in this case the oncoming counterclones can be stamped. In this case, stamping can be carried out in the radially passing zone of the body.

 In order to provide economical manufacture of the friction clutch, it may also be preferable if at least part of the adjusting device is made of plastic. Such plastic parts can be injection molded. As an artificial material, thermoplastics, for example polyamide, are particularly preferred. The use of artificial materials is possible because the adjusting device is located in an area that is only exposed to minor heat. In addition, due to the low mass, an insignificant moment of inertia is also created.

According to another inventive idea of the invention, the adjusting device can be made in such a way that, when viewed in the direction of disengagement of the friction clutch, it acts freely, and in the direction opposite to the direction of disengagement, it is self-braking. To do this, ramps and / or ramps are designed so that they have an axial angle of elevation lying in the range from 4 to 20 ^o , preferably from 5 to 12 ^o . It is preferable to perform ramps and / or ramps in such a way that they provide self-braking due to friction. Self-braking can also be achieved due to a geometric circuit, in which, for example, one of the slopes is smooth and the other is profiled or when both slopes are made profiled. Due to this, there is no need to use additional means to prevent unwanted reverse bias.

 The adjusting device can be made in a particularly preferred and simple way if the feed device acting in the circumferential direction is a prestressed spring that loads at least one part bearing oncoming slopes and / or one part bearing oncoming counterclones or counter-padding zones by springing in the direction of adjustment. The springing can preferably be carried out in such a way that there is practically no effect on the functioning of the remaining springs, in particular the control disk spring, and the springs loading the axially deflecting plate.

 In many applications, it may be preferable if the adjusting device has several offset adjusting elements, such as, for example, adjusting wedges or rolling elements, offset in the radial and / or circumferential direction. In addition, it may be preferable if the adjusting device depends on the speed. So, for example, centrifugal force acting on individual elements of the adjusting device can be taken into account to control and / or lock the adjusting device under certain operating conditions of the internal combustion engine. In particular, the adjusting device can be blocked at a certain number of revolutions by means depending on centrifugal force, which can occur, for example, when the number of revolutions is at least close to the number of revolutions of idle speed or the number of revolutions is lower than the number of revolutions of idle speed, as a result of which the adjustment associated with wear, occurs only with a small number of revolutions. The advantage of this is that there is no arbitrary adjustment that may occur due to vibration at high speeds.

 A particularly simple and reliable design of the adjusting device can be obtained when parts that can be displaced relative to the housing, equipped with ramps and / or ramps, are spring loaded. If there is only one corresponding part with corresponding incident slopes or zones, which has the possibility of displacement relative to the housing, it is made loaded. It may be particularly preferable if the spring creates a force acting in the circumferential direction.

 In addition, for the design and operation of the friction clutch, it may be preferable if the sensor spring, made in the form of a disk spring, for example a disk spring, with its radial outer zone rests on a rigidly axially located part, such as a housing, and the following sections located radially inside acts on the run-in pads opposite to the cover. These run-in pads can also be made integrally with the sensor spring, as a result of which the sensor spring in the form of a disk spring also forms a plate. To hold the sensor spring in tension, support sections may be provided in the housing. These supporting sections can be made in the form of separate supporting elements located on the housing. However, it is preferable if the supporting sections are made in one piece with the housing, for example, stamped or cut and deformed sections can be on the housing, which form an axial support for the sensor spring from below.

 For the functioning of the friction clutch, in particular to reduce the disengagement force or the maximum necessary disengagement force, it is especially advisable if the friction disk sandwiched between the pressure plate and the support disk has friction linings between which a so-called spring is provided, as is known, for example, from the laid out German applications 3631863. Through the use of such clutch discs, a control effect is supported, in particular the process of disengaging the friction clutch. This is due to the fact that in the engaged state of the friction clutch, the spring-loaded lining when they are tensioned causes a reaction force on the pressure plate, which is directed opposite to the force exerted by the pressure plate spring or the set spring on this pressure plate. In the process of disengagement during the axial displacement of the pressure plate, it is first pushed back under the action of the spring of the plate, and at the same time, due to the relatively sharply falling portion of the pressure plate spring characteristic in the disengagement zone, the force exerted by it on the pressure plate decreases. With a decrease in the force generated by the pressure plate spring on the pressure plate, the force of the reverse action of the spring plate on this pressure plate also decreases. The force required to disengage the friction clutch is obtained due to the difference between the recoil force of the spring loaded plates and the clamping force of the pressure plate spring. After loosening the lining spring, that is, when lifting the pressure plate from the friction linings or when the clutch disc is released by the pressure plate, the necessary disengagement force is determined mainly by the pressure plate spring. The force-path characteristic of the plate spring loading and the force-path characteristic of the pressure plate spring can be particularly advantageously so coordinated with each other that when the clutch plate is released by the pressure plate, the force required to control the pressure disk disk is at a lower level. Thus, due to the achieved coordination or alignment of the spring loading characteristics of the lining with the characteristics of the pressure plate spring, until the clutch disc is released using the pressure plate, only very small, in extreme cases, practically absent, control force of the pressure plate disk is required to overcome the residual pressure. Furthermore, the characteristic of the hold-down disk spring can be calculated in such a way that after releasing the clutch disc, the force generated by the hold-down disk spring, or the force necessary to deflect the hold-down disk spring, is very low relative to the hold-down force exerted by this hold-down disk the spring in the engaged state of the friction clutch. Executions are also possible in which, when releasing the clutch disc with the pressure plate, a very small or practically no force is required in order to control the pressure disc disk to disengage the clutch. Such friction clutches can be calculated in such a way that the control forces are of the order of 0-200 N.

 According to an additional idea of the invention, the friction clutch can be calculated in such a way that approximately when the clutch plate is released using the pressure plate, the axial force exerted by the pressure disk spring is within zero, and when the release process continues, the force created by the pressure disk spring can become negative, that is, there is a change in the force of action of the pressure plate spring to the opposite. This means that with a fully disengaged friction clutch, it practically remains open by itself, and the clutch process can only be carried out by applying force from outside.

 In addition, the invention relates to a friction clutch, in particular for automobiles, equipped with a pressure plate, which is rotatably but axially displaceable to the housing, with at least one tightening spring that acts between the housing and the pressure plate to load a pressure plate in the direction of the clutch disc sandwiched between it and the support disc, for example, a flywheel.

 Such couplings are known, for example, from German Patent Laid-Open No. 2460963, German Patents 2441141 and 898531, and Accepted Application 1267916.

 In addition, the basis of the invention is the task to create a friction clutch of the specified type, the functioning and service life of which is significantly improved. In particular, using the invention, the forces necessary to control such friction clutches can be reduced, and, in addition, an almost constant force is required for disengagement during their entire service life. Furthermore, the friction clutches of the invention can be manufactured in a particularly simple and economical manner.

 According to the invention, this is achieved due to the fact that there is an adjustment device that automatically compensates for the wear of the friction linings of the clutch disc, which creates an almost constant application of force to the pressure plate by means of a compression spring, and the friction clutch has controls for engaging and disengaging, and also has a device which, during the disengagement process on at least part of the control path of the controls and / or the disengagement course of the pressure plate, causes a gradual decrease ix torque transmitted by a friction clutch or clutch disc. Using such a device, it is equally possible to ensure that during the process of disengaging the friction clutch and at the beginning of the clamping of the friction linings, a gradual or progressive decrease in the moment transmitted by the friction clutch occurs between the pressure plate and the support disk.

 By making the friction clutch according to the invention, it is ensured that the pressure cup disk has almost the same tension in the engaged state of the friction clutch during the entire service life of the friction clutch and thereby practically constant loading of the pressure plate. In addition, using an additional device that causes a gradual decrease in the moment transmitted by the equalization device, a reduction or minimization of the passage of the decoupling force or the maximum necessary decoupling force is achieved. This is due to the fact that the device supports control, in particular the process of disengaging the friction clutch. For this, the device may have an axially spring-loaded deflection means, which causes a reaction force on the controls and / or on the pressure plate, and / or on the support disk, which is directed opposite to the force exerted by the pressure spring on the pressure plate, and acts sequentially.

 It may be particularly preferable if the friction clutch device is positioned in such a way that during the disengagement process on a part of the axial movement of the portions of the pressure plate loaded by the pressure spring, the torque transmitted from the friction clutch or clutch disc is gradually reduced.

 In many applications, the device can be located on the power flow line between the rotary support of the controls or between the pressure spring and fasteners, for example threaded connections of the housing on the support disk.

 For other applications, however, it may be preferable if the device is on the power flow line between the rotary support of the controls or between the pressure spring and the friction surface of the pressure plate. Such a device was proposed in the laid out applications of Germany 3742354 and 1450201.

 For other applications, it may be particularly preferable if the device is provided axially between the friction linings of the clutch disc located back to back, that is, the so-called “spring-loaded linings” are formed, for example, by means of spring segments located between the linings. Such devices are known, for example, from German Patent Laid-Open No. 3631863.

 Another possibility of providing progressive growth or reducing the moment is proposed in the laid-out application of Germany 2164297, in which the flywheel is made of two parts, and the part forming the back pressure plate is axially spring-loaded on the part connected with the output shaft of the internal combustion engine.

 For the functioning and design of the friction clutch according to the invention, it may be particularly preferred if the device provides axial spring deflection between the parts of the clutch, the device being positioned and designed so that when the clutch is open, the force acting on the device is the smallest, and when closing the clutch, that is, during the course of the clutch coupling, the force acting on the device gradually increases to a maximum, and it is advisable to have this growth only on a part of the closing stroke or ode clutch control means or the pressure plate. It may be especially advantageous if the device is designed in such a way that a gradual decrease or gradual increase in the moment transmitted by the friction clutch by at least about 40 to 70% of the control stroke of the control means and / or maximum axial stroke of the pressure plate occurs. The remaining area of the corresponding path is necessary for reliable separation of the power flow and for alignment of possible deformations on the coupling parts, for example, the clutch disc, the pressure plate, and also the support disk.

 In order to minimize the forces necessary to control the friction clutch according to the invention, it may be particularly advantageous if the compression spring, at least in part of the compression or deformation stroke, has a decreasing characteristic. Due to this, it can be ensured that during the disengagement of the friction clutch, the force of the spring of the device counteracts the force of the pressure spring, as a result of which the action or deformation of the pressure squeeze under the action of the force of the spring of the device is retained on the part of the disengagement stroke, moreover, due to the degressive or the falling force-stroke characteristic of the compression spring, the force applied by the latter to the pressure plate or friction plates is reduced . The applied force necessary for the effective disengagement of the friction clutch is created if there is no additional overlapping action of the springs based on the difference between the power flow generated by the device and the power flow of the pressure spring. When the pressure plate is withdrawn from the friction linings or the clutch disc is released by the pressure plate, the necessary remaining passage of the decoupling force or the necessary decoupling force is determined mainly by the clamping spring. The force-path characteristic of the device and the force-path characteristic of the pressure spring can be matched to each other so that when releasing the clutch disc by the pressure plate, the force required to control the pressure spring is at a relatively low level. Thus, by approximating or even matching the characteristics of the spring or the characteristics of the force of the device with the characteristics of the pressure spring until the clutch plate is released by the pressure plate for the pressure spring, a very small, in the extreme case, almost absent, control force is required.

 A disk spring is particularly suitable as a compression spring, which, on the one hand, can rotate around an annular rotary support located in the housing, and on the other hand, it loads the pressure plate. In this case, the cup spring may have an annular housing, from which the tongues forming the control elements radially extend inward. However, the controls can be made in the form of levers mounted, for example, on the housing. The pressing force of the pressure plate can, however, be created by springs of a different type, for example, coil springs, which are located in the friction clutch in such a way that the axial force applied to the clutch plate by them is the greatest in the friction clutch engaged state, and during the disengagement process this force decreases . This can be done, for example, by installing helical springs at an angle relative to the axis of rotation of the friction clutch.

 It may be particularly advantageous if the cup spring is rotatably supported on the housing between two pads to create a so-called “compressible type clutch”. In such couplings, the controls for disengaging the friction clutch are typically loaded in the direction of the pressure plate. However, the invention is not limited to lifting type couplings, but also relates to pulling type couplings in which the controls for disengaging the friction clutch are loaded in the usual manner away from the pressure plate.

 In a particularly preferred embodiment, the friction clutch according to the invention may have a disk spring, designed in such a way that it has a sinusoidal force-path characteristic and is integrated in such a way that in the engaged state of the friction clutch its operating point is located in the area of the degressive characteristic following first maximum strength. It may be particularly preferable if the Belleville spring has a force ratio between 1: 0.4 and 1: 0.7 between the first maximum force and the next minimum force.

 In addition, it may be particularly preferable if the friction clutch is controlled through a disengagement system that interacts with controls, for example, with the tips of the disc spring tongues, the disengagement system may have a clutch pedal similar to the accelerator pedal and located in the passenger compartment. This embodiment of the clutch pedal may be particularly preferable, since due to the implementation according to the invention, the force required to disengage the friction clutch, or the nature of the applied force, can be kept at a very low level, due to which the clutch pedal, similar to the accelerator pedal, provides better metering power management. Due to the implementation of the friction clutch in accordance with the invention and due to the associated possibility of reducing the force of the compression spring that occurs during the life of the friction clutch, its parts can be correspondingly reduced in size or reduced in strength, due to which significant cost reduction can be achieved. manufacturing process. By reducing the disengagement forces, the friction losses and the loss of elasticity of the clutch in the disengagement system are also reduced, and due to this, the efficiency of the clutch / disengagement system is significantly increased. Thanks to this, the whole system can be optimally designed, and due to this, the coupling comfort is significantly improved.

 The implementation according to the invention can be widely used for friction clutches and in particular for those described in German patents 2916755, 2920932, German patent applications 3518781, 4092382, French patent applications 2605692, 2906477, 2699444, 2599446, UK patent 1567019, US patent 4057131, Japanese industrial designs 3-25026, 3-123, 2-124326, 1-163218, Japanese Patent Laid-open No. 51-126452, Japanese Industrial Design 3-19131 and 3-53628.

 The use of a friction clutch with independent or automatic compensation of at least wear on the linings — which ensures at least a constant clutch plate pressing force over the life of the friction clutch — is preferable, especially in combination with clutch units in which the friction clutch the clutch disc and the support disc, such as a flywheel, form a single mounting unit or module. In such an assembly, for reasons of economy, it is preferable if the coupling body is inseparably connected to the support disk, for example by means of a welded joint or by means of a geometric closure, for example, with plastic deformation. Thanks to this connection, the use of conventional fasteners, such as screws, may be unnecessary. For such assemblies, replacing the clutch disc or friction linings is practically impossible outside the wear line without destroying the part, such as the clutch housing. Thanks to the use of a clutch, which is adjustable due to wear, the mounting unit can be designed in such a way as to ensure smooth functioning throughout the entire service life of the vehicle. Thus, with the implementation according to the invention, the wear reserves of the clutch disc and the reserves of regulation of the clutch or clutch module can be so increased that the service life of the clutch and at the same time the service life of the mounting block are guaranteed to reach the service life of the entire vehicle.

 According to a further refinement of the invention, it may be particularly advantageous if the friction clutch having the adjusting device due to wear is combined with the so-called dual mass flywheel, the friction clutch being mounted with an intermediate clutch plate gasket on one of the fly masses connected to the gear, and the second fly mass has the ability to connect to the output shaft of the internal combustion engine. Double-mass flywheels in which the friction clutch according to the invention can be used are known, for example, from German Patent Laid-open Applications 3721712, 3721711, 4117571, 4117582, 4117579. The general content of these applications also relates to the content of the disclosure of this invention, therefore, the features described in of these applications may in any way be combined with the features described in this invention. In particular, the coupling body or the coupling cover can be connected to the flywheel supporting them by means of a connection that cannot be disconnected without breaking, as is shown, for example, in various forms of execution in the laid-open application of Germany 4117579.

 Due to the use of the friction clutch with a device that compensates for at least wear on the linings, optimization can also be carried out in the calculation of the friction clutch, and in particular the energy accumulator, which supplies the force to the clutch disc. This energy accumulator can be calculated in such a way that it brings the clamping force to the clutch disc, which is practically necessary only to transmit the desired torque. The energy accumulator can be made in the form of at least one disk spring or a plurality of coil springs. In addition, it is preferable to use a self-adjusting friction clutch in combination with dual-mass flywheels, in which an elastically rotating shock absorber located between both flywheels is radially outside the clutch disc or the outer diameter of the friction surface of the flywheel connected to the gear. In such dual mass flywheels, the friction diameter of the clutch disc should be smaller than that of conventional clutches, as a result of which the clamping force should increase in accordance with the ratio of the average friction radii to ensure accurate transmission of engine torque. With conventional couplings, this could lead to an increase in decoupling force. Due to the use of the clutch, which is adjustable due to wear, with a progressive decrease in the torque transmitted from the clutch disc during the disengagement, according to Clause 1, it is possible to reduce the disengagement force, thereby eliminating the increase in the disengaging force or due to the corresponding calculation of the friction clutch a significant reduction in decoupling force can be achieved compared to a conventional clutch.

 Thus, due to the implementation of the friction clutch according to the invention, in spite of the reduced outer diameter of the friction lining and the higher clamping force required thereby, it is possible to use a lower decoupling force. Due to the lower decoupling force, the load on the rolling support is also reduced, with the help of which both flywheel masses are rotated relative to each other. In addition, due to regulation, the clutch service life is increased due to wear, which does not require replacement of parts, in particular the clutch disc, during the life of the car. Thus, the coupling cover is rigidly connected to the fly mass associated with the transmission, for example, using rivets or welding. This is especially preferred when there is a limited mounting cavity or limited contours of the coupling cap that do not allow coupling of the coupling cover to the flywheel from the transmission side with conventional screws.

 In a friction clutch with an integrated device for adjusting due to wear of the linings during normal fastening of the clutch assembly consisting of a friction clutch and a flywheel on the output shaft of the internal combustion engine, axial vibrations during rotation and oscillatory vibrations excited by the output shaft of the internal motor are transmitted to the clutch assembly combustion, for example a crank shaft. In order to prevent such fluctuations from adversely affecting the coupling unit or adjusting device during operation, and in particular to suppress the need for undesirable adjustment of the device to compensate for wear, it is necessary to take into account the centrifugal forces of those parts that act on this device when calculating the adjusting device . In order to eliminate unwanted side effects caused by axial or oscillating vibrations, or the associated high costs when calculating an adjusting device to compensate for wear of the linings, according to another idea of the invention, the coupling assembly having the adjusting device is substantially separated from axial and bending vibrations generated by the output shaft of an internal combustion engine. This can be achieved due to the fact that the clutch assembly can be connected via an axially elastic or spring element to the output shaft of the internal combustion engine. In this case, the rigidity of this part is set so that the axial and oscillating or bending vibrations created by the output shaft of the internal combustion engine are damped or suppressed by this elastic element at least to such a size as to ensure the smooth functioning of the friction clutch, in particular its adjusting device. Such resilient details are known, for example, from European Laid-open Application No. 0385752 and 0464997, as well as SAE Technical Paper 900391. The content of these publications also refers to the contents of the disclosure of this application. Thanks to the use of the elastic part, it becomes possible to eliminate the need for adjustment due to wear caused by axial vibrations of the pressure plate relative to the clutch cover, especially when the friction clutch is disengaged, flywheel vibrations and / or Belleville springs. Such vibrations can result in clutch assemblies or clutch assemblies without a device that at least substantially suppresses these vibrations, for example, an axially deflecting disc, to a variable adjustment independent of the state of wear of the clutch disc, and the disk spring of the friction clutch could be regulated relative to the pressing force minimum force, in connection with which it would be impossible to ensure the transmission of the desired torque.

 According to another embodiment of the invention, a friction clutch with self or automatic compensation, which can be made in accordance with this invention, is preferably used in the drive unit, in particular in a car, which consists of an automatic or semi-automatic transmission and located between the drive motor of an engine type internal combustion and transmission driven with control or regulation, at least depending on the operation of the friction clutch cottages. The friction clutch is driven mainly fully automatically. Automated or fully automated control of the friction clutch was proposed, for example, in the laid-out application of Germany 4011850.9, therefore, with regard to the principle of operation and the necessary funds, a link to this publication is given.

 Significant problems have still existed in drive units with automatic or semi-automatic transmission and conventional friction clutch, which are still known, for actuating the clutch and selecting the parameters necessary for this functional elements, such as, for example, piston-cylinder assemblies and / or motors . Due to the relatively high decoupling forces required in conventional couplings, working (functional) elements having high strength and large dimensions are required. This leads to large structural dimensions, large weight and high costs. Also, thus calculated working parts due to their inertial masses are relatively slow in response time. When servocylinders are used, a greater volumetric flow of the working fluid is also required, so that the feed pump must also be relatively large in order to guarantee the desired response time of the corresponding friction clutch. In order to partially eliminate the previously mentioned drawbacks, German Patent Laid-Open No. 3309427 proposes, for example, to reduce the control force for disengaging the clutch by means of corresponding compensation springs, so that smaller working parts could be used. Since the decoupling force in the known couplings fluctuates very much during the service life, i.e. the decoupling force in the new clutch is relatively insignificant and increases during the service life with increasing wear on the linings, only one part of the usually required decoupling force can be reduced by means of a compensation spring. . Taking into account all the tolerances, despite the use of compensation springs, such a decoupling power of the working elements would be necessary that is greater than the decoupling power of a conventional coupling. When using the friction clutch according to the invention with compensation for wear of the linings in combination with a drive unit consisting of an engine and an automatic or semi-automatic transmission, the disengagement force can be slightly slightly reduced in relation to the previously mentioned prior art directly in the clutch, moreover, this value of the disengaging force or the same characteristic the decoupling force of the new coupling remains virtually unchanged throughout its entire service life. As a result of this, significant advantages are obtained in calculating the parameters of the working elements, since their drive power or control power can remain low, and accordingly lower forces and pressures appearing in the entire shutdown system. Due to this, friction losses or elasticity of structural parts that appear in the trip system are eliminated or reduced to a minimum.

The invention is explained in more detail using FIG. 1-48, where:
in FIG. 1 is a sectional view of a coupling assembly according to the invention;
FIG. 2 is a leveling device shown in cross section and on an enlarged scale;
FIG. 3 is a view in the direction of arrow III in FIG. 2;
FIG. 4 - an adjusting ring adjacent to the means for disengaging the friction clutch, view along arrow IV in FIG. 2;
FIG. 5 is a section along line VV of FIG. 4;
FIG. 6 - anti-adjusting ring used in the clutch nodes according to FIG. 1 in the form of arrow III in FIG. 2;
FIG. 7 is a section along line VII-VII in FIG. 6;
FIG. 8 is a detail of an embodiment of a surge device shown in FIG. 2;
FIG. 9 is a sectional detail of a clutch assembly according to the invention;
FIG. 10 and 11 show a wear adjustable ring that can be used in the clutch assembly according to the invention, for example according to FIG. nine;
FIG. 12 is a sectional view of the clutch assembly according to the invention, FIG. 12a is a circular sector of the sensor spring applied in FIG. 12;
FIG. 13 is a partial view in the direction of arrow XIII in FIG. 12;
FIG. 14 is yet another embodiment of a friction clutch according to the invention;
FIG. 15 is a schematic diagram illustrating a release system of a clutch assembly according to the invention;
FIG. 16 is yet another embodiment of a friction clutch according to the invention having a brake for an adjusting ring;
FIG. 17 - friction clutch according to the invention;
FIG. 18 is a section along line II-II in FIG. 17;
FIG. 19 is a mounting ring used in the friction clutch of FIG. 17 and 18;
FIG. 20 is a section along line IV-IV in FIG. 19;
FIG. 21 is a support ring used in the friction clutch of FIG. 17 and 18;
FIG. 22 is a section along line VI-VI of FIG. 21;
FIG. 23 and 23a - a spring exerting a cranking force on the installation ring;
FIG. 24-27 is a diagram with various characteristics showing the interaction of the individual spring and adjusting elements of the friction clutch according to the invention;
FIG. 28 and 29 are yet another embodiment of a friction clutch according to the invention, wherein FIG. 29 is a section along line XIII of FIG. 28;
FIG. 30 is a view of a mounting ring used in the friction clutch of FIG. 28 and 29;
FIG. 31-33 - details of another friction clutch with equalization device;
FIG. 34 and 35 are diagrams with various characteristics, from which the interaction of the clamping spring and the lining spring is visible, as well as the influence exerted on the decoupling force of the friction clutch;
FIG. 36 is another embodiment of a friction clutch, a partial view, FIG. 36a is a partial view along arrow A of FIG. 36;
FIG. 37 is a section along line XXI of FIG. 36;
FIG. 38 is a partial view of the alignment ring used in the friction clutch of FIG. 36 and 37;
FIG. 39 and 40 are other embodiments of friction clutches according to the invention;
FIG. 41 is a mounting ring used in the friction clutch of FIG. 28 and 29 or 36 and 37;
FIG. 42-45 - additional embodiments of the friction clutches;
FIG. 46-48 are details of other embodiments of the friction clutch, wherein FIG. 47 is a partial view along arrow A of FIG. 46, and FIG. 48 is a section along arrows BB of FIG. 47.

 Presented in FIG. 1, the clutch assembly has a friction clutch 1 with the housing 2 and a compression plate 3 connected with it without the possibility of rotation but with the possibility of limited axial displacement 2 of an annular swivel bearing 5 and a loading pressure plate 3 in the direction of a support disk rigidly connected to the housing 2, for example with a flywheel, as a result of which the friction pads 7 of the clutch disk 8 are clamped bend between the friction surfaces of the pressure plate 3 and the supporting disk 6.

 The pressure plate 3 is connected without the possibility of rotation with the housing 2 through oriented in the circumferential direction or tangentially leaf springs 9. In the illustrated embodiment, the clutch disc 8 has elastic segments 10, which, as is known, provide a progressive increase in torque when the friction couplings 1, while they, through the limited axial displacement of both friction linings 7 towards each other, make it possible to progressively raise the axial forces acting on the friction linings and 7. However, it can be applied as the clutch plate, wherein the friction linings 7 would be axially spaced substantially rigidly on the carrier disc.

 In the illustrated exemplary embodiment, the disk spring 4 has a ring-shaped main part 4a that creates a pressing force, from which the working tongues 4b extend radially inward. In this case, the cup spring 4 is installed in such a way that it loads the pressure plate 3 located radially from the outside and the sections located radially further inward and can rotate around the rotary support 5.

 The pivot bearing 5 includes two pivot bearings 11, 12, between which a disk spring 4 is axially fixed or clamped. The pivot bearing 11 provided on the side of the disk spring 4 facing the pressure disk 3 is axially loaded in the direction of the housing 2, for this the pivot bearing 11 is a part of a disk spring or similar to a disk spring of a part 13, which, with its outer edge portion 13a, is spring-loaded on the housing 2, the rotary bearings 11 being radially formed inwardly loaded axially against the force of action of the control disk spring 4 and thereby in the direction of the housing 2. Especially provided between the pressure plate 3 and the control disk spring 4, the disk spring 13 has an annular section 13b, from which the tongues 13c extend radially inward and form a pivot bearing eleven.

 In order to maintain the plate spring-shaped part 13 in the illustrated exemplary embodiment, bayonet joints or a lock are provided between the body 2 and the tongue-shaped protrusions 13a of the disk-shaped part 13 in the form of a disk spring 13.

 A Belleville spring-like structural part or Belleville spring 13 is made in the form of a sensor spring, which creates at least basically approximately constant force on a given working path. Through this sensor spring 13, at least a substantial part of the clutch disengagement forces acting on the tongues of the tongues 4c is perceived, and at least an approximate balance is constantly maintained between the force exerted on the pivot bearing 11 by the disengagement force and the reaction force exerted by the sensor disk spring 13 on this swinging support 11. By the decoupling force, we should understand the maximum force that is applied during the control of the friction clutch 1 to the tip 4c of the reeds or eplyayuschie levers diaphragm spring tongues.

 The swinging support 12 located on the side of the housing through the adjusting device 16 rests on the housing 2. This adjusting device 16 ensures that with an axial displacement of the swinging supports 11 and 12 in the direction of the pressure plate 3 or in the direction of the support disk 6, an undesirable gap between swinging bearing 12 and the housing 2, or between the swinging bearing 12 and a disk spring 4. This ensures that there are no unwanted "dead" or idle paths when actuating the friction clutch 1, consequence of which it holds an optimum efficiency, and thus excellent control of the friction clutch 1. An axial displacement of the rocking bearings 11 and 12 carried by axial wear on the friction surfaces of the pressure plate 3 and the support disc 6 and 7 of the friction linings.

 The adjusting device 16 includes an elastically loaded adjusting element in the form of an annular part 17, which has circumferentially extending and axially ascending ramps 18 that are distributed around the circumference of the band 17. The adjusting element 17 is integrated in the sleeve 1 so that the ramps 18 facing the base 2a of the housing.

 The adjusting ring 17 is elastically loaded in the circumferential direction, namely, in the direction of the adjusting rotation, therefore, in the direction which, by running the slopes 18 on the counterclaws 19, stamped on the bottom 2a of the cover, causes the axial displacement of the adjusting ring in the direction of the pressure plate 3, i.e. . in the axial direction from the radial section 2a of the housing.

 The operation of the automatic adjustment of the pivot bearing 5 or the adjusting device 16, as well as other possibilities for performing the adjusting device 16 are explained in more detail together in FIG. 17-18.

 The clutch includes a compensating device 20, ensuring that the release means of the friction clutch 1 formed by the tabs 4b of the disk spring are axially clearance-free and can be displaced by a constant amount of displacement 21. The balancing device 20 is located between the release element 22, covering the release bearing, and the tips of 4c reeds. The trip element 22 has the possibility of axial displacement on the schematically shown guide tube 23 for controlling the friction clutch 1. The guide tube 23 is mounted in a gear housing not shown in more detail and encompasses the input gear shaft, on which the clutch disc 8 is also rigidly mounted without turning. necessary for the axial displacement of the release element 22, is applied from the control means 24, which is made in the presented example of the execution schematically shown fork fork junction, which can also be located on the transmission side. However, other trip elements 22, hydraulically or pneumatically driven, i.e. trip elements having cylinder-piston assemblies driven by pressure means.

 The equalization device 20 is shown in FIG. 2 and 3 on an enlarged scale and includes an adjusting element in the form of an annular part 25 shown in FIG. 4 and 5. The ring-shaped adjusting element 25 in the illustrated embodiment has two sets of ramps 26, 27 offset radially, circumferentially and axially rising, which are distributed around the perimeter of part 25. As is especially clearly seen in FIG. 5, the ramps 26 located radially inside are offset in the circumferential direction relative to the ramps 27 located radially outside, namely approximately half the length of the ramp or pitch of the ramp. The adjusting member 25 is supported as shown in FIG. 1 and 2, with its end 25a directly on the tip 4c of the reeds. The ramps 26, 27 are facing in the opposite direction to the control means 4b. The adjusting element 25 is spring loaded in the axial direction, in particular in the direction of the adjusting rotation, i.e. in the direction that causes when inclinations 26, 27 run onto the counterclones 28, 29 of the support ring 30, not shown in FIG. 6 and 7 in more detail, the axial displacement of the adjusting element 25 in the form of a ring in the direction of the pressure plate 3, i.e., in the axial direction away from the trip element 22.

 As seen in FIG. 6 and 7, oncoming counterclones 28, 29 also form sets of oncoming slopes displaced from each other in both radial and circumferential directions. The slopes 26, 27 of the adjusting element 25 and 28, 29 of the support ring 30 are aligned with each other and interact axially with each other. With the circumferentially biased slopes, an exact central direction is ensured between the adjusting element 25 and the support ring 30. As is particularly clearly seen in FIG. 2, both parts 25 and 30 of equalization device 20 are axially arranged (see FIG. 2). The installation angle 31 of the oncoming contraclones 28, 29 (Fig. 7) of the support ring 30 corresponds to the angle 32 (Fig. 5) of the oncoming inclinations 26, 27 of the adjusting element 25. The support ring 30 can be rotatably connected to the housing 2, while being able to limited displacement in the axial direction on the path 21 of the direction of the coupling. Axial restriction is carried out through the radial zones 33 of the support ring 30, which in the engaged state of the friction clutch 1 is adjacent to the radial inner zones of the bottom 2a of the cover. This fit is carried out by springing the control means 4b. When the friction clutch 1 is disengaged, the travel path is limited by means of a shaped element 34 made of sheet material provided on the side of the support ring 30 facing the opposite direction of the control means 4b, to which the force from the releasing element 22 is applied in the diameter zone 35. This shaped element 34 of the sheet material also has radial zones 36, which, when disengaged, the friction clutch 1 can abut against the radially inner zones of the bottom 2a of the cover.

In the presented exemplary embodiment, the adjusting element 25 in the form of a ring, as well as the support ring can be made of heat-resistant artificial material, i.e. from thermoplastic, which can additionally be reinforced with fiber. Such parts are easy to manufacture by injection molding. The incline slopes 26, 27 and the incline counterclones 28, 29 are made in the circumferential direction so that they provide at least one angle of rotation between the two parts 25 and 30, which during the entire service life of the friction clutch 1 provides compensation for wear occurring on the surfaces the friction of the pressure plate 3 and the support disk 6, as well as on the friction linings 7. This adjustment angle may, depending on the calculation of the incline slopes, be about 30-90 ^o . In the exemplary embodiment shown in FIG. 3 it has a position of 37 and characterizes an angle of rotation of the order of 75 ^o . The angle 31 or 33 of the installation of slopes and counterclones can be in the range from 6 to 14 ^o , preferably 8 ^o , and the actual angle of 31 or 32 slopes and counterclones varies on the radial length of these slopes, since at a given angle of rotation it is necessary to cover the same difference in height . This means that ramps 31 or 32 decrease with increasing diameter.

 The circumferential application of force required to adjust element 25 is created using energy accumulators, which in the presented embodiment are made in the form of two tensioned coil springs 38, 39 located in an arc between the support ring 30 and the adjusting elements 25. These coil springs 38, 39 are supported onto the support ring 30, which is rotatably connected to the cover 2, and the adjusting ring 25 is rotated as soon as the controls or tabs 4b of the Belleville springs depart due to wear of the plates axially from the bottom 2a of the lid or trip element. As, in particular, seen in FIG. 3 and 6, the coil springs 38, 39 have circumferential rings 40, 41 suitable for the ring 30, in the form of a channel or a torus. As seen in FIG. 2, such a socket 40, fitted in its cross section to the turns of energy accumulators 38, 39, extends over more than half the circumference of the cross section of the springs 38 or 39, moreover, as can be seen in FIG. 3-6, there remains a corresponding spline-shaped hole 42, 43 on the side facing the controls 4b, and a corresponding spline-shaped hole 44, 45 on the side opposite to the controls 4b. The springs 38, 39 are fixed in the axial direction relative to the support ring 30 with surfaces defining the slots 40, 41. In order to insert the coil springs 38, 39, the sector sockets 40, 41 have a corresponding filling zone 46, 47 having such a width in the radial direction, which at least corresponds to the outer diameter of the coil springs 38, 39. Through these refueling zones 46, 47, the energy accumulators 38, 39 can be inclined inserted into the sector sockets 40, 41. After the unloaded coil springs 38, 39 are inserted into sector sockets 40, 41, the adjusting element 25 is connected to the support ring 30. For this purpose axial spouts 48, 49 made on the adjusting element 25 in the form of a ring, which simultaneously form force application zones or support zones for coil springs 38, 39, are inserted into zones of axial grooves 50, 51 adjacent in the circumferential direction to refueling zones 46, 47, due to which loading zones 48, 49 are applied to the end zone of the coil springs 38, 39 in an open state. The unloaded position of the energy accumulator 38 or 39 is seen in FIG. 3 and is indicated by 39a. The other end zone of the coil springs 38, 39 is supported by the sector nests 40, 41 provided in the circumferential direction of the bottom 53, 53a. Due to the rotation between the ring adjusting element 25 and the support ring 30, the springs 38, 39 can be compressed. After a certain angle of relative rotation, which is greater than the angular distance of the refueling zone 46, 47, the application zones of the loads 48, 49 of the ring-shaped adjusting element 26 are axially above the end zone of the slot-like openings 44, 45, due to which the adjusting element 25 is in the form of a ring and the support ring 30 can move towards each other until the ramps 26, 27 and the ramps 28, 29 come in contact with each other. The slot-like openings 44, 45 and the axial spouts 48, 49 are so aligned with each other that there is an axially snap joint between the two parts 25, 30. For this, the axial spouts 48, 49 have at their ends a hook-shaped portion 48a, which can abut against the radially located area of the support ring 30. Due to the additional relative rotation between the two elements 25, 30 in accordance with the angle 37 (Fig. 3) of the spring 38, 39 are installed in a stressed position along a length 54 corresponding to the friction clutch in the initial state. In this position, both elements 25, 30 can then be fixed using a device not shown in FIG. 3 means. This means can be a geometric closure acting between the two elements 25, 30, and after mounting the friction clutch 1 on the support disk 6, it can be removed, which drives the balancing device 20. The possible angle of adjustment to compensate, in particular, the wear of the linings corresponds the rotation angle indicated in FIG. 3 at 37. After this angle of rotation, the axial spouts 48, 49 abut against the end zones of the slot-like openings 44, 45 in the direction of the adjusting ring 25. One of these positions, corresponding to the position of the tense coil spring 38, 39, is indicated in FIG. 3 at 38a.

 In the new friction clutch 1, the axial spouts 26, 27 and 28, 29, forming oncoming slopes and oncoming counterclones, axially interact with each other. This means that the rings 25 and 30 located one above the other require minimal structural space.

 In the presented exemplary embodiment, the limitation of the control path in the direction of disengagement of the friction clutch 1 is provided by the friction element 34 of sheet material. According to not shown in FIG. 3 of the embodiment, the support zones necessary for this, which interact, for example, with the cover 2, are also provided on the trip element 22, namely, the support ring rotating with a friction clutch or a part connected to it. Axial restriction of the control path of the friction clutch 1 in one of at least one of the axial directions can also be achieved by using at least one axial stop for the trip element 22 provided on the guide tube 23.

 In addition, the trip element 22 can act directly on the control means 4b, and a corresponding compensating device is provided between the trip element 22 and the trip means 24.

 It is advisable if the trip element 22 applies a load in the direction of the control means 4b that does not affect the function of the friction clutch 1 and the leveling device 20.

 As seen in FIG. 2-4, the adjusting ring 25 has radially inside the cams 55, which form zones of influence for the means of rotation or restraint, which, if necessary, protection against rotation can be applied on the other hand to the housing 2 or to the support ring 30. Such restraining means may be provided when manufacturing or assembling the friction clutch 1 or leveling device 20 and are removed after mounting the friction clutch on the flywheel.

 The detail shown in FIG. 8 is an embodiment of the lower half of equalizer 20 of FIG. 1 and 2. In the embodiment of FIG. 8, an axial restriction is made between the leveling device 120 and the housing 102 in the engaged state of the friction clutch via a hook-shaped axial cantilever element 133 made integrally with the shaped element 134 of sheet material. The cantilever element 133 is molded on the outer edge of the shaped element 134 of sheet material serving as a spring element and interacting axially through the lid 102. At the free end of the cantilever element 133 facing the disk spring, there are radially outwardly extending zones 133a that radially trap the lid 102 its side facing the Belleville spring 104. This arrangement ensures that the axial forces exerted by the Belleville spring 104 on the equalization device 120 can support i spring member 134 made of sheet material, whereby the balancing device can absorb large axial forces from the equalization unit 20 of FIG. 2, in which the stops are made in the form of sections 33 of the support ring 30 made of artificial material. Such axial forces on leveling device 20 or 120 can occur, among other things, during transportation, i.e. when the friction clutch is not mounted, since in this state the main disk spring 4 or 104 is supported axially through the tabs of the spring on the support ring or compensating element 30, 130 made of plastic. The clamping element 134, made of sheet material, may have at least two, preferably three or more hook-shaped protrusions 133 distributed around the perimeter, preferably symmetrically or evenly. The thickness of the sheet for the clamping element 134 can be calculated in accordance with the axial forces that create the support. The ring 130 made of artificial material is connected to the clamping element 134 without the possibility of rotation. Similarly to FIG. 2, the clamping element or molded part 134 made of sheet material has zones 136 radially located outside, which, when viewed in the circumferential direction, pass between the hook-shaped protrusions 133 and serve to limit the disengagement path or to prevent an unacceptably large overrun with the stop in the housing 2 .

 The friction clutch part 201 shown in FIG. 9 has a substantially similar construction to the lower right portion of FIG. 1. In FIG. 9 shows a clutch housing 202, a pivot bearing 205 for a disk spring 204, an adjustment device 216 and an equalization device 220. With regard to the functioning of the adjustment device 216 and the equalization device 220, reference is made to the description of FIG. 1-8, as well as German patent applications P 4306505.8 and P 4239289.6, the contents of which should be considered as included in this application.

 In the embodiments of FIG. 9, a device 260 is provided for blocking the rotation of the adjusting element in the form of an adjusting ring 217.

 The device 260 for locking displacement or rotation provides, with the friction clutch 201 not yet mounted, the position of the adjusting element 217 relative to other parts, in particular the housing 202, with the locking device 260, in particular, it can be ensured that the adjusting element 217 in the new clutch 201 in its rear position, i.e., in practically zero position, in which regulation has not yet been carried out, it can be held in such a way that, at the same time, the adjusting ring 217 is in the swivel or the friction pad 212 is not loaded with a cup spring 204. The latter is caused by the fact that when the friction clutch 201 is not yet mounted or to install the friction clutch 201, the main cup spring 204 is axially supported through its tabs on the compensating device 220, as could also be seen in FIG. . 1 and 2. Due to this support, the main disk spring 204 presses the force sensor in the form of a disk spring 213 axially in the direction from the housing 202 or the adjusting ring 217, as a result of which the axial tension of the adjusting ring 217 is not provided in the direction of the housing 202. Without the ring rotation locking device 260 217, it would be impossible to shift the ring 217. When mounting the friction clutch 201 on the output shaft of the internal combustion engine, the ring 217 could therefore not have a retracted position for adjustment Due to wear and tear arising, in particular, on the friction linings of the clutch disc. In FIG. 9, the solid line shows the position of the parts corresponding to the friction clutch mounted on the flywheel. The dashed line shows the position of the disk spring 204 and the sensor spring 213, which corresponds to the position of a new unassembled friction clutch. As seen in FIG. 9, in an unassembled state of the friction clutch 201, there is an axial distance or clearance between the adjusting member 217 or the annular support 212 and the cup spring 204.

 The blocking of the adjustment process 260 for the adjusting device 216, provided, inter alia, for transporting the friction clutch, has at least a blocking element 261, which, at least in the un-mounted state of the friction clutch 201 and, if necessary, in the engaged state of the clutch 201 is held without the possibility of rotation relative to the housing 202 and to prevent rotation of the adjusting element 217 interacts with it. Lock member 261, as shown in FIG. 10 can have separate radially extending arms 262 that are radially outwardly rigidly connected to the adjusting ring 217, and can be clamped radially internally between an abutment 233 provided on a side of the housing 202 facing the disk spring 204 and the housing 202. Thereby a short circuit is formed between the adjusting element 217 and the housing 202. The arms 262 can be made similar to leaf springs and connected to each other radially inside using an annular zone 263. In the presented example, Ia shoulders 262 are connected by screws to the adjusting element 217, the shoulders 262 may, however, be connected to the adjusting member 217 via rivets or even have a zone for which the compound is non-rotatably with the adjusting element can be embedded in the synthetic material constituting the element.

 The equalization device 220 or its limiting stop 233 forms, in combination with the housing 220 and axially clamped zones of the locking element 261, a brake or clutch for the adjusting element 217, which operates when the friction clutch 201 is not turned on.

 When the clutch is released, the braking action or clamp of the locking element 261 or plate 262 is removed, so that the adjusting element or the adjusting ring 217 can be adjusted if necessary.

 It is preferable if the blocking element 261 or the plate formed like a leaf spring has an axial margin of spring or spring stiffness, and in the circumferential direction the element is made relatively rigid or without spring.

 In the embodiment shown in FIG. 11, the adjusting ring 317 made of artificial material has axially extending elastic plates 362 obtained directly on it by injection molding. The plates 362, as shown in combination with FIG. 10 can be connected radially inside using an annular portion.

 Shown in FIG. 12 and 13, the clutch assembly or friction clutch 401 has a housing made in the form of a cover 402 of sheet material, a pressure plate 403 connected to it, having the possibility of limited axial movement, but without the possibility of rotation, and also a pressure plate spring sandwiched between it and the cover 402 404. The pressure plate spring 404 is installed relative to the housing 402 as a two-arm lever, while it is mounted in the pivot bearing 405 with the possibility of tilt or rotation. Belleville spring 404, using zones located radially further outward from the annular swivel bearing 405, loads the pressure plate 403 in the direction of the friction linings 407 of the clutch disc 408 clamped between the pressure plate 403 and the flywheel. The transmission of torque between the pressure plate 403 and the cover 402 is carried out through leaf springs 409, which can be loaded in the direction of the removal of the pressure plate 403 from the friction linings 407.

 Belleville spring 404 has an annular main body 404a, as well as tongues 404b extending from it and directed radially inward.

 The pivot bearing 405 includes two pivot bearings 411 and 412, between which the Belleville spring 404 is axially held or clamped. The pivot bearings 411, 412 are arranged and configured similarly and have the same function as the pivot bearings 11 and 12 described in connection with FIG. 1. With regard to the parts acting on the pivot bearings 411 and 412, as well as the automatic adjustment function of the pivot bearing 405, reference is made to the description of FIG. 1 as well as FIG. 17-48.

 The release means of the friction clutch 401 made in the form of a disk spring tongues 404b are axially acted upon by the release device 422, thereby changing the taper of the disk spring 404. The release device 422, as described in conjunction with FIG. 1-7, may have a surge device 420. However, in disengagement systems of a self-adjusting trip support, such a surge device 420 is not required. In such a release system, the release device 422 may be coupled to an release support ring that rotates with the coupling 401 at least during the release process.

 In order to prevent an excessively large path of uncoupling of the clutch disengaging means formed by the disc spring tongues 404b, path limiting means 436 for the disc spring tongues 404b are arranged on the clutch 401 or the housing 402. The means for restricting the path 436 limit the path or angle of rotation of the Belleville spring 404 by axially supporting the tongues 404b of the Belleville spring and thereby axially receiving the disengaging force acting on the tripping device 422.

 In the illustrated exemplary embodiment, the means for restricting the path 436 are made in the form of an annular supporting portion 436 made radially by the inner portions of the cover 402, to which, after a certain axial stroke 421, the tongues 404c are adjacent. The annular supporting portion 436 is configured such that it is at least approximately located on the uncoupling diameter of the disc spring tongues, i.e. on the diameter at which the trip device 422 comes into contact with the disc spring tongues 404b. The support region 436 is axially between the tongues 404b of the spring or the tips 404c of the tongues of the spring and the clutch disc 408.

The annular region 436 is connected through radially extending ribs or jumpers 437 to the housing cover 402a. As can be seen from FIG. 13, in the illustrated embodiment, there are six such jumpers. However, in many applications, there can only be three such jumpers. In such friction clutches that require particularly high disengagement forces,
there may be more jumpers, for example nine.

 The jumpers 437 extend from the bottom of the lid 402a or from the housing 402 radially inward or obliquely relative to the axis in the direction of the pressure plate 403 or the clutch disc 408. The support region 436 is displaced in the axial direction relative to the bottom 402b of the lid into the lid cavity. The tongues 404b of the spring enter the holes 438 made between the annular region 436, the radial housing of the cover 402a further located and the connecting jumpers 437. In the illustrated embodiment, for this purpose the tongues 404b of the Belleville spring are cut or axially radially inwardly opposed along their lengths the passage of jumpers 437. As can be seen from FIG. 13, the cup spring tongues 404b are grouped in three to form holes or spaces 438. Between the individual groups of three tongues, there are passages 439 for the jumpers 437. The passages 439 and the jumpers 437 are sized such that the cup spring 404 is flawlessly rotated.

 The reed spring 404b of the Belleville spring is inserted into the holes 438 during the installation of the friction clutch 401. For this, the Belleville spring 404 in the expanded state shown in FIG. 12 by a dash-dotted line, has tongues in the tip region 404c whose inner diameter 440 is larger than the outer diameter 441 of the annular support zone 436. Due to this, the cup spring 404 is axially inserted with its tongues 404b into the openings 438 of the cover 402 at least. mounting time of the friction clutch 401 or at the latest when mounting the friction clutch 401, for example on a flywheel, the disk spring 404 rotates, thereby reducing the inner diameter 440, limited by the tongues 404b poppet that spring. The friction clutch mounted on the flywheel has a Belleville spring 404 in the working position, in which the tongues 404c limit the inner diameter 442, which is smaller than the outer diameter 441 of the support zone 436. The Belleville spring 404 is rotatably mounted on the housing, and the tongues 404b are configured such that even after the turn 421, the inner diameter described by the tongues is smaller than the outer diameter of the support zone 436.

 The maximum possible control stroke 421, axially limited, is selected on the basis of this condition such that the coupling 401, after reaching the maximum allowable wear on the plates 407, has at least a still full set disengagement stroke necessary for perfect functioning, i.e. reliable separation of the clutch assembly 401. A clutch 401 or a sensor spring 413, which provides automatic compensation for wear of the plates in the clutch, and the adjusting device 416 are calculated so that in the new state of the friction The coupling 401 does not erroneously axially displace the pivot bearing 405 when the stroke 421 is fully engaged.

 In the example below, the numbers or operation or interaction between the support zone 436 and the disc spring tongues 404b are illustrated or demonstrated.

 The described release path of the friction clutch 401, taking into account the existing tolerances, is from 8.4 to 10 mm. The coupling 401 is calculated in such a way that in the new state an erroneous axial displacement of the pivot bearing 405 would be possible only with a trip of more than 14 mm. The stop 436 is located or is designed in such a way that, in the new state of the friction clutch, the zones that come into contact with the stop 436, in particular the tongue tip 404c, can pass along the axis at a distance 421 of 12.5 mm. When the tongues of the Belleville spring adhere to the stop 436 and the maximum disengaging force is applied, the lid can axially spring another 0.5 mm, which makes a total maximum axial stroke 421 of 13 mm possible.

 Assuming a maximum wear of the pads 407 may be 3 mm, the cup spring is biased towards the clutch disc during the life of the friction clutch 401 due to the axial displacement of its pivot bearing 405 by these 3 mm. The maximum possible disengagement stroke 421 is thus reduced from about 13 mm to about 10 mm, so that the coupling at the end of its service life still lies within the required disengagement tolerance of 8.4 to 10 mm.

 In the presented exemplary embodiment, the stop 436 is made integrally with the cover 402. However, this stop can be made as a separate part connected to the cover 402. The jumpers 437 can also be made as separate parts or integrally with the stop 436 filled in as a separate part.

 Clutch assembly 401 shown in FIG. 12 and 13, furthermore, has a device or means which, during operation of the clutch 401, at least in part of the speed range in which the clutch assembly 401 rotates during use, exerts an increase in the axial bearing force on the cup spring 404. Due to this increase in reference force can be prevented due to at least some negative factors in a certain range of speed during operation of the friction clutch 401 invalid adjustment based on unwanted axial deviation I or exposure of the sensor device as a sensor spring 413 that is not associated with the wear of the friction facings 407.

 In FIG. 12, means 450 are provided, depending on the number of revolutions or centrifugal forces, to increase the axial force acting on the run-in plate 411. The means 450, depending on the centrifugal force, are made in the form of a sensor disk spring 413 molded around the outer perimeter and axially rising in the direction of the cover 402 of the reeds 450. As can be seen from FIG. 12a, the sensor spring 413, made in the form of a disk spring, has tongues extending radially outwardly 413a, which, as can be seen from FIG. 12 and 13, axially rest on the cover 402. Between the protrusions 413a and the portions 451 of the cover 402, which provide axial support for them, there is a bayonet connection or clamp 452. The bayonet connection 452 is designed so that due to the axial joint control of the sensor spring 413 of the housing 402 and due to the subsequent relative rotation between the two parts, the protrusions 413a axially come into contact with the support sections 451 of the housing 402. When assembling the sensor spring 413 and the cover 402, before the rotation of both of these parts, the sleep sensor spring 413 ala resiliently axially tightened and released after the rotation, whereby the protrusions 413a are supported firmly on the cap 402. As seen from FIG. 12a, there is a tab 450 on either side of the radial protrusion 413a. When the clutch assembly 401 rotates, due to the centrifugal force acting on the tabs 450, a force is created which is superimposed on the force created by the sensor spring 413 due to its tightening, i.e. these forces add up, resulting in an increase in the force of the Belleville spring 404 in the area of the pivot bearing 411. This force, additionally generated by the tongues 450 on the pivot bearing 411, increases with increasing speed. However, this increase in force is limited due to the fact that, starting from a certain level of the number of revolutions, the tongues 450 are deformed or twisted under the influence of centrifugal force on them so that they rest on the housing 402 radially from the outside, as a result of which there is practically no further the growth of additional support efforts in the area of the rotary support 411 created by means 450, depending on the centrifugal force.

 When considering the ratios of axial forces or the balance of forces between the pivot bearing 411 and the Belleville spring 404, further consideration is given to the means 409 for transmitting torque in the form of a leaf spring. These means 409 for transmitting torque, made in the form of a leaf spring, can be tensioned between the housing 402 and the pressure plate 403 in such a way that during the entire service life of the friction clutch 401, the pressure plate 403 is axially pressed by the force transmitting means 409 to the cup spring 404. Due to this, the axial force generated by the torque transmission means 409 counteracts the force exerted by the cup spring 404 on the pressure plate 403 and thus folds the axial force exerted by the sensor spring 413 to the Belleville spring 404, both of which then axially counteract the disengagement force acting on the tongue tip 404c. The actual sensory force that counteracts the axial displacement of the Belleville spring 404 when the friction clutch assembly 401 does not rotate separately is the force created by the torque transmitting means 409 together with the sensor spring and acting on the Belleville spring 404. When the clutch assembly 401 rotates, this resulting force superimposed on a force depending on the number of revolutions or the centrifugal force created by the reeds 450.

 The clutch disc 408 has a device, for example in the form of a spring pad 453, which, when the friction clutch assembly 401 is operated on a part of the stroke of the pressure plate 403, provides a gradual decrease or gradual increase in the moment transmitted from the clutch disc 408, supports this device 450 until this friction clutch is released the pads 407 or the clutch disc 403 using the pressure plate 403, the axial support of the disk spring 404 relative to the adjusting element in the form of an adjusting ring 417. This ensures The fact is that at least until the friction liner 407 is released, the adjusting ring 417 remains axially clamped between the disk spring 404 and the housing or cover 402. If, when the clutch assembly 401 is disengaged, the pressure plate 403 is lifted from the linings 407 as an axial clamping force the main disk spring 404 is affected by the resulting force created by means 409 for transmitting torque in the form of a leaf spring, and a touch disk spring 413. This resultant touch I force counteracts the force of the trip supplied to the tip 404c tongues. In certain ranges of revolutions, in particular at high revolutions of the engine, vibrations can occur, generated, for example, by the engine, which cause axial vibration of the pressure plate 403. If the pressure plate 403 oscillates axially, then this pressure plate 403 can briefly rise from the main or of the cup spring 404, as a result of which the resulting force drops briefly, because in this case, the cup spring 404 is no longer affected by the axial force created by the means 409 for transmitting to utyaschego points having the form of the leaf spring. The consequence of this is that the balance of forces necessary for the purposeful adjustment of the device 416 is violated between the Belleville spring 404 or the disengaging forces acting on it and the resulting support force acting on this Belleville spring, and, in particular, under the operating conditions of the clutch 401 such the axial support force acting on the cup spring 404 is so small that the clutch adjusts prematurely or involuntarily, and the working point of the cup spring 404 is shifted in the direction imuma diaphragm spring. In addition, under certain conditions of the engine operating mode, especially at higher revolutions, there are especially large circumferential accelerations of the crank shaft, which create, on the basis of the inertia of the adjusting wheel 417, circumferential forces which, due to the adjusting incline 418 acting between the adjusting wheel 417 and the housing 402, , 419 can create an axial component on a cup spring 404, which is counter-directed to the resulting sensory force, as a result of which involuntary operation can also occur. ulirovka. Due to the occurring vibrations, the difference in frictional forces between the ramps 418, 419 may also decrease, resulting in an increase in the axial force acting on the cup spring 404 and generated by the adjusting spring 417a acting in the circumferential direction on the adjusting ring 417, as a result of which involuntary adjustment.

 To eliminate the above-described drawbacks in the coupling 401 without means 450 for transmitting a torque dependent on centrifugal forces, in the embodiment shown in FIG. 12-13, tongues 450 are provided, depending on centrifugal forces. These means 450, depending on the centrifugal forces, compensate for the interference associated with the centrifugal force, since they create a supporting force, depending on the number of revolutions or centrifugal forces, which runs parallel to the force created by the sensor spring 411.

 The tool, depending on the centrifugal force, can be made in such a way that the adjustment in the coupling 401, necessary in connection with the wear of the pads, is possible only when the coupling 401 is stopped or with a small number of revolutions. where critical vibrations occur, the adjusting device 416 can be practically blocked.

 In the embodiment of the friction clutch 501 shown in FIG. 14, the sensor spring 513 is located radially inside the disk spring support 505. The sensor spring 513 has an annular main body 513a from which tongues 513b radially extend inwardly. The sensor spring 513 is supported by these tongues 513b on an annular support zone 536, which is located and made similar to the annular support zone 436 according to FIG. 12 and 13. The sensory tabs 513b rests on the side of the support zone 536 facing the tips 504 from the disc spring tongues. The main body 513a also has tongues 513c radially outside, which, for the axial support of the cup spring 504, abut against the latter.

 The sensor spring 513 can be mounted on the cover 502 by conically deforming it in the compression direction at which the inner diameter 540 limited by the inner tongues 513b becomes larger than the outer diameter 541 of the support zone 536. In this case, the support tongues 513b can be inserted into the holes 538 of the cover 502 similar to that described in connection with reeds 404b and holes 438 in FIG. 12 and 13. After inserting the tongues 513b into the openings 533, the sensor spring 513 can be loosened, as a result of which the inner end zones of the tongues 513b are smaller in diameter and come into contact with the supporting surface 536.

 Another possibility is that for mounting the sensor spring 513 on the cover 502, at least a portion of the area of the inner tongues 513b is axially bent upward in the direction of the cover 502 so that it limits a larger inner diameter 540 than the outer diameter of the support zone 536. After sensor springs or tongues 513b are inserted into the openings 538 of the cover, the tongues 513b can be bent back so that they radially force their abutment against the support zone 536. By bending the tongues 513b from the position shown in FIG. 14 by a dashed line, they fall due to plastic deformation into the cup spring material, which is shown by a solid line. For plastic deformation of the sensor tongues 513b, they can axially rest on the tongues 504b or the tip 504c of the tongues of the spring springs 504. For the operation of bending the tongues 513b, a tool can be used that supports the tongues 504b of the spring spring 504c from above and acts from below on the tongues 513b of the sensor spring in particular about the diameter at which the tabs 513b are bent.

 The stops 436 and 536 to limit the disengagement path or the angle of rotation of the Belleville springs 404 and 504 have the advantage that they are integrated in the corresponding clutch 401 or 501 and work in the area of the reed springs 404b or 504b, which can ensure that that when the tongues of the Belleville springs 404b, 504b adhere to the stops 436, 536, the tongues of the Belleville springs cannot be axially deformed or are slightly deformed. In this way, it can also be ensured that the tongue springs 404b, 504b themselves, even in their position corresponding to the disengaged state of the friction clutch 401, 501, do not lie on the part of the clutch disc 408. FIG. 12 shows the position of the disk spring in a dashed line corresponding to the uncoupled clutch and indicated by 450. This can exclude the possibility that in the disengaged state of the friction clutch 401, the disk spring tongues 404b come into contact or can slip with the clutch disc 408 rotating relative to this clutch 401.

 In the exemplary embodiments shown in FIG. 12-14, in the region of the tongue points 404c, 504c of the cup spring tongues, there are stops 436, 536. However, these stops can be made differently and offset relative to the inner tongues 404c, 504c of the tongues radially outward. In this embodiment, however, it is advisable if the radial arm of the lever between the tongues of the tongues 404c, 504c and the radially farther outward stops is selected so that inadvertent deflection of the tongue springs 404b, 504b does not occur due to the disengaging force and the force of the stops .

 The said excessive or unacceptably large trip course can be caused by a trip system or a control system that acts on the clutch control, formed in the form of cup spring tongues, presented and described in the forms of execution. This control system typically comprises a trip support that acts on the friction clutch control, a control, such as a clutch pedal, and devices that transmit power between the trip support and the control. This section with devices for transmitting forces may have a receiving as well as a giving cylinder. In tripping systems with a receiving and returning cylinder, an unacceptable tripping path may be created that is outside the normal tripping path due to the fact that, due to the quick start and subsequent switching off of the friction clutch, the receiving cylinder may not return back enough, i.e. it will not reach the end position, as a result of which, when the re-uncoupling occurs very quickly after this re-uncoupling, although the path corresponding to the normal disengagement course passes, the common decoupling path for the clutch corresponds to the sum of the normal disengagement path and the residual return path that did not occur. Thus, the friction clutch may have a common control path, significantly exceeding the maximum permissible clutch disengagement path. This means that the reserve for the control path provided for the friction clutch may be exceeded.

 Using the features according to the invention or stops 36, 436, 536, it is possible to exclude an unacceptably large disengagement stroke or excessive travel when controlling friction clutches, while ensuring the necessary and predetermined normal disengagement stroke during the clutch service life.

 Thus, according to the invention, in general in couplings, and in particular in couplings with an adjusting device, at least one stop in the chain of devices designed to control the clutch is provided for compensating at least the wear of the friction linings, which prevents the clutch control means from running excessively work. Such an emphasis may, for example, limit the disengagement path of the disengagement support or the rotation path of the cup spring. However, such an emphasis can be installed elsewhere. In addition, the trip of the friction clutch can be limited by an updated constant value, when both in the direction of disengagement and in the direction of inclusion there is an exact limit in the form of a stop.

 This kind of restriction can be preferred in the area of the trip support, since in this zone the tolerance chain between the controls, for example the tongues of the disk springs of the friction clutch, and the part limiting to a certain stroke is small.

 If there is such a restriction or such kind of emphasis during disengagement, the process goes to an almost severe limitation, as a result of which overload of parts, in particular parts of the disengagement system, for example, in foot control systems, may occur, which may be undesirable for the driver. Therefore, according to a further refinement of the invention, a springy or resiliently deflecting means and / or means for restricting the pressure in the disengaging system is provided in the control line of the friction clutch, this means having an opening force that is at least slightly greater than the required maximum force or the required maximum pressure for clutch control. This ensures that under the action of the stop the clutch pedal can be pressed further or the control engine can move to the exact position. The creeping means provided in the control line of the friction clutch may be located between the clutch controls and the trip support, or between the latter and the trip control, as provided, for example, on the clutch pedals or trip motors.

 In FIG. 15, a trip system 601 is shown, and various arrangements are shown for limiting the maximum force exerted through the trip support 622 to the controls of the sleeve 604 and / or sleeve body 602.

 In FIG. 15 shows an axial stop 636 which, after a certain stroke of the release support 622, comes into contact with the housing 602 in the same way as described in connection with the stop 36 according to FIG. 1 and 2. However, the disengagement of the trip can occur otherwise, for example, as shown and described in FIG. 12-14. The trip system 601 has a delivery cylinder 650 and a receiving cylinder 651 connected through a conduit 652. A piston 653 of the receiving cylinder 651 has a trip support 622 and is axially mounted in the housing 654. The pressure chamber 655 is provided with hydraulic medium, for example, oil through a conduit 652. The cylinder assembly 650 has a housing 656, which, together with the piston 657 located therein, forms a variable pressure chamber 658. The pressure chamber 658 is connected through a pipe 652 to the pressure chamber 655. In the pressure chamber A return spring 659 is located for the chamber 658 for the piston 657. The piston 657 is axially displaceable by means of a clutch pedal or a control motor, for example an electric motor or a pump. The circuit of the working means of the trip system 601 is connected to the reservoir 660 of the working means. The delivery cylinder 650 is preferably connected via a conduit 661 directly to the reservoir of the working means 660.

 To limit the control force of the clutch acting on the release means 604 and / or the housing 602, in the embodiment according to FIG. 15, at least one means is provided in the working medium circuit of the tripping system 601, which limits to a certain value the pressure arising in the circulation circuit of the pressure means during the operation of the friction clutch. In the exemplary embodiment of FIG. 15 this means is at least in the form of a pressure limiting valve. In FIG. 15 shows various arrangements for such a pressure limiting valve. Such pressure limiting valve 662 may be located, for example, in piping system 652 and have a return pipe 663 leading to reservoir 660 with working means. Instead of a pressure limiting valve 662, a pressure limiting valve 664, which is located on or integrated in the housing, is connected to the pressure chamber 655 and is connected via the return line 665 to the working medium reservoir 660.

 In FIG. 15 provides yet another alternative arrangement for pressure limiting valve 666. The pressure limiting valve 666 is connected to the pressure chamber 653 of the delivery cylinder and can be located on or integrated into the housing. In addition, the pressure limiting valve 666 is diverted to the reservoir 660 with a working tool. For this, the pressure limiting valve 666 may have its own pipe or may be connected to a pipe 661.

 Another possibility for the arrangement of the pressure limiting valve 667 is that it can be integrated in the piston 657 of the delivery cylinder 650. On the discharge side, this valve 667 must also be connected to a reservoir 660 for a working medium or at least to an intermediate accumulator.

 Instead of a bypass valve, a hydraulic accumulator may be provided in the circulating circuit of the working medium, which limits the maximum pressure that occurs in the trip system, which relieves the system after the stops restricting the trip are activated by accumulating the working medium and acts almost like a buffer or battery.

 The clutch assembly 701 shown in FIG. 16 has, in a manner similar to that described in connection with the previous drawings, an adjusting device 716 for automatically compensating for wear occurring on the friction linings 707 of the clutch disc 708. In the illustrated embodiment, the principle of the construction and functioning of the adjusting device 716 corresponds to that described in connection with FIG. 12 and 13. The adjusting element or the adjusting ring 717 has contact or support zones 770 that can interact with the cup spring 704 during the disengagement process of the clutch assembly 701. The relative axial arrangement of the support zones 770 with respect to the interacting zones 771 of the disc spring 704 is thus determined that during the trip of the release zone 771 of the Belleville spring axially rely, at least indirectly, preferably directly on the contact zone 770 located on the adjusting ring 717. This mutual support is carried out, preferably, at least approximately, when reaching or slightly exceeding the predetermined disengagement stroke 772 or the corresponding angle of rotation of the Belleville spring 404 in the area of the tongue tip 704c. Such an excess of the predetermined tripping stroke 772 may occur due to an erroneously or incorrectly adjusted tripping system. Due to the axial support of the Belleville spring 704 on the supporting sections 770, the adjustment ring is protected from involuntary rotation. The Belleville spring 704 thus acts practically as a brake for the adjusting ring 717 when the predetermined trip 772 is exceeded.

 In the presented embodiment, the contact zone or support zones 770 are formed by an annular protrusion 773 formed on a ring 717 radially outside the pivot bearing 705. Instead of the annular radial protrusion, several radial protrusions 773 distributed around the perimeter can be used. The protrusion 773 extends to the outer edge of the disk-shaped protrusion springs 704. As soon as a certain trip path 772 is reached, the disk spring 704 is supported by its outer zones 771 on the support zones 770 of adjusting ring 717. When a certain tripping stroke 772 is exceeded, the rotation diameter of the cup spring 704 increases, since it shifts from the diameter of the rotary support 705 to the contact diameter between the areas 771 of the cup spring 704 and the support zones 770. Due to this displacement, the tripping force also decreases, necessary in the region of the ends 704c of the reeds, since the ratio of the lengths of the levers of the Belleville spring varies from i to i + 1, and in particular, because the Belleville spring, installed first before the trip 772 as a two-armed lever g, in excess stroke 772 rotates substantially as a single-armed lever. Due to this reduction in the decoupling force, it is also ensured that the disk spring 704 undergoes displacement under the action of the axial resultant supporting force generated, inter alia, by the sensor spring 713 and the leaf spring 709 or by the force applied in the direction of the housing 702 or the adjusting ring 717. Thus, the disk spring cannot be axially displaced axially away from the adjusting ring 717 or from the cover 702. When a certain cleavage stroke 772 is exceeded, the sensor spring 713 is axially deformed spring and, in particular, because the disk spring rises from the adjustment ring 717 in the area of the pivot bearing 705.

 The protrusion or cantilever portion 773 can preferably be made on the injection ring of the injection molding obtained by injection molding of artificial material 770. The maximum force acting axially on the protrusion 773 is obtained from the difference between the minimum disengagement force in the area of the tab spring tongues 704c and the axial sensory or reference force for the disk spring 704, applied by a sensor spring 713 and leaf spring elements 709. The protrusions 773 are designed so that they withstand this maximum force without significant deformations of.

 Another significant advantage is that the axial retraction of the pressure plate 703 remains almost constant, and thus, with an increase in stroke 772, the axial force exerted from the leaf springs 709 to the main disk spring 704 no longer decreases. Since the force generated by the leaf springs 709 is part of the resulting sensory force, based on the residual voltage of these leaf springs, the reliability of the operation when the friction clutch 701 is running excessively increases. Due to this, in the clutches, for example for cars, it is possible to provide tongues in the tip area 704c an excess stroke of approximately 0.5 to 2 mm, without impairing the functioning of the adjusting device 716.

 The restriction of the retraction of the pressure plate 703 can also be achieved by axially supporting the pressure plate 703 on the sensor spring 713 if a certain trip is exceeded. For this, corresponding elements, for example, cams, protrusions, etc., can be formed on the sensor spring 713 and / or on the pressure plate 703.

 The friction clutch 1 shown in FIG. 17 and 18, has a housing 2 and a pressure plate 3 connected to it without the possibility of rotation, but with the possibility of limited axial displacement. Axially between the pressure plate 3 and the cover 2, a clamping disk spring 4 is clamped, which can be rotated around an annular swivel support 5 mounted in the housing, and the compression plate 3 is loaded in the direction of the support disk 6, rigidly connected to the housing 2, for example, with a flywheel, the friction pads 7 of the clutch 8 are sandwiched between the friction surfaces of the pressure plate 3 and the supporting disk 6.

 The pressure plate 3 is connected to the housing 2 in the circumferential direction or tangentially directed leaf springs without the possibility of rotation. And the illustrated embodiment, the clutch disc 8 has the so-called spring segments 10 of the linings, which provide a progressive increase in torque when the friction clutch engages when it, due to the limited axial movement of both friction linings 7 towards each other, provides a progressive increase in axial forces acting on friction linings 7. However, it is also possible to use a clutch disc in which the friction linings 7 are located substantially rigidly on the carrier disk. In one of these cases, a “set of spacer springs” could be used, that is, springs in sequence with cup springs, for example, springs between a cap and a flywheel, between a cap and a pad on the side of the cap, and also between a cup spring and a pressure plate or due to the elasticity of the cover.

 In the illustrated exemplary embodiment, the disk spring 4 has an annular main body 4a that generates a pressing force, from which the working tongues 4b radially extend inward. In this case, the cup spring 4 is designed in such a way that it acts on the pressure plate 3 located radially outside the zones and has the ability to rotate the zones located radially inside the zones around the rotary support 5.

 The pivot bearing 5 includes two pivot bearings 11, 12, which are made in the form of wire rings and between which an axially held or pinched disk spring 4 is located. On the side of the disk spring 4 facing the pressure plate 3 there is a pivot bearing 11, which is loaded axially into the direction of the housing 2 by the force generated by the force accumulator 13. The force accumulator 13 is formed by a Belleville spring or a spring-shaped part 13, which rests on its outer edge zones 13a against pus 2, and with its radially inwardly located sections it loads the pivot bearing 11 against the force of action of the control disk spring 4 and thereby towards the housing 2. The disk spring 13 provided between the pressure plate 3 and the disk spring 4 has an outer annular edge zone 13b from The inner edge of which the tongues 13c radially inward extend, resting on the deflectable plate 11.

 To support the part 13, having the shape of a Belleville spring, in the presented embodiment, additional means 14 are fixed on the housing 2, which form a deflectable plate for the part 13, having the shape of a Belleville spring. These additional means can be made in the form of segmented elements 14 fastened together or riveted, which can be evenly distributed around the circumference. However, the means 14 can also be made in the form of an annular closed part. In addition, the support means can be molded directly on the housing 2, for example in the form of sections stamped in the axial direction of the housing 2, or in the form of sections having the shape of tongues, which, after installation in the clamp of the part 13, having the shape of a Belleville spring, are squeezed out by deformation of the material along the outer edge zones of this part 13. In addition, between the support means 14 and the disk-shaped spring 13, there may be a bayonet connection or blocking, so that the disk-shaped spring l 13 is first loaded, and its radially outer zones can axially wind up behind the support means 14. After this, due to the corresponding rotation of the part 13, having the shape of a disk spring, relative to the housing 2, the support zones of the part 13 come into contact with the support means 14. In this case, the support the zones of the disk spring-shaped part 13 may be formed by protrusions formed and directed radially outwardly on the annular main body 13b.

 To prevent the rotation of the disk spring 4 and, if necessary, the shape of the disk spring of the part 13, as well as to center the supports 11 and 12 in the form of wire rings, axially passing centering means in the form of rivet elements 15 are fixed on the body 2. Rivet elements 15 have a corresponding axially passing a shank 15a, which is axially passed through a section between two adjacent plate spring tongues 4b, which may partially be covered by zones 13 molded on the plate tongue 13c second spring.

 The part having the shape of a disk spring or disk spring 13 is made in the form of a sensor spring, creating at least a substantially constant force at a given working displacement. Using this sensor spring 13, a clutch disengaging force is applied to the tongues 4c of the tongues, and there is always at least an approximate balance between the force generated by the disengaging force on the pivot bearing 11 and the counter force exerted by the touch disk spring 13 on this pivot bearing 11. Under disengaging force means the force that acts on the tongues of the tongues 4c or on the lever for disengaging the tongues of the Belleville spring during operation of the friction clutch 1 and thereby counteracts the sensory ruzhin 13.

 The pivot bearing 12 is supported on the housing 2 through an adjustment device 16 located in the axial space between the cup spring 4 and the housing 2. This adjustment device 16 ensures that when the pivot bearings 11 and 12 are axially displaced in the direction of the pressure plate 3 or in the direction of the support disk 6, an undesirable gap cannot occur between the pivot bearing 12 and the housing 2 or between the pivot bearing 12 and the cup spring 4. This ensures that there is no "dead" or idle operation the friction clutch 1. The axial displacement of the rotary bearings 11 and 12 is carried out with axial wear on the friction surfaces of the pressure plate 3 and the support disk 6, as well as on the friction linings 7. The adjustment is carried out in devices according to the invention and when the rotary bearings 11, 12 are axially opposed to the disk spring, and when the disk spring is worn in the area of the cams of the lining on the pressure plate Well (3a) or opposite areas of the Belleville spring. The automatic adjustment of the pivot bearing is explained in more detail in connection with the diagrams of FIG. 24-27.

 The adjusting device 16 includes a spring-loaded adjusting element in the form of an annular element 17, which is shown in FIG. 19 and 20. The annular element 17 has circumferentially extending axially upward ramps 18 which are distributed along the perimeter of element 17. The adjusting element 17 in the coupling 1 is mounted so that the ramps 18 are facing the bottom 2a of the housing. On the side of the adjusting element 17 opposite to the ramps 18, in the recess 19 having the shape of a groove (Fig. 18), a rotary support 12 made in the form of a wire ring is centrally positioned. In this case, the recess 19 can be made in such a way that the pivot bearing 12 is fixed on the adjusting element 17 also in the axial direction. This can be done, for example, due to the fact that at least in certain sections of the zone of the adjusting ring 17 adjacent to the recess 19, the swivel support 12 is clamped in brackets or a snap connection is formed for the swivel support 12. When using different materials for the swivel support the supports 12 and the adjusting ring 17 may be appropriate to compensate for the difference in elongations arising from large changes in temperature, if the rotary support 12, made in the form of a wire ring, is open, i.e. divided at least in one place around the perimeter, due to which the wire ring is moved relative to the recess 19 in the circumferential direction, and due to this, the support 12 in the form of a wire ring can adapt to changes in the diameter of the recess 19.

 In the presented exemplary embodiment, the adjusting element 17 is made of plastic, for example a heat-resistant thermoplastic, which can additionally be fiber reinforced. Therefore, the adjusting element 17 is especially well made by injection molding. The adjustment element made of plastic with a low specific gravity, as already mentioned, has a low inertia, which also reduces the sensitivity to pressure fluctuations. The swivel plate can also be made in the form of a ring of artificial material. However, the adjusting element can also be made in the form of a shaped part made of sheet material or obtained by sintering. In addition, the pivot plate with the appropriate choice of material can be integral with the adjusting element 17. The pivot bearing 11 can be formed directly by the sensor spring. 13. For this point, the tongues 13c may have corresponding shapes, for example grooves.

 The adjusting ring 17 is centered in the axially extending zone 15a of the rivets 15, evenly distributed around the perimeter. For this, the adjusting ring 17 has centering loops 20 formed in the form of recesses 21 extending in the circumferential direction, which are located radially inside the pivot bearing 11. For the formation of the recesses 21, the adjusting ring 17 has cams 22 extending radially inward in the inner edge zone, limiting the radially inner the contours of the recesses 21.

As can be seen in FIG. 19, when viewed in the circumferential direction, between the equally distributed recesses 21 there are corresponding five incident slopes 18. The recesses 21 are made in the circumferential direction so that they allow the adjusting ring 17 to rotate at least a certain angle relative to the housing 2, which during the entire service life of the friction clutch 1 allows you to adjust due to wear that occurs on the friction surfaces of the pressure plate 3, as well as on the support disk 6, as well as friction linings 7, and Also, if it is necessary to wear the coupling itself, i.e., for example, supports 11, 12, the areas of the Belleville spring located between them, the cams of the pressure plate (3a) or the opposite zones of the Belleville spring 4. This adjustment angle depends on the calculation of the incline may range from 8 to 60 ^° , preferably may be in the order of 10 to 30 ^° . In the presented exemplary embodiment, this rotation angle is about 12 ^° , and the angle 23 of the installation of ramps 18 is also within 12 ^° . This angle 23 is selected so that the friction that occurs when the ramps 18 of the adjusting ring 17 and the ramps 24 of the support ring 25, shown in FIG. 21 and 22, causes a slip between the ramps 18 and 24. Depending on the material of the interacting pair in the region of the ramps 18 and the ramps 24, the angle 23 can be in the range from 4 to 20 ^o .

 The adjusting ring 17 is spring-loaded in the circumferential direction, in particular in the direction of the adjusting rotation, i.e. in the direction that causes when the incline 18 is incident on the counterclone 24 of the support ring 25, the axial displacement of the adjustment ring 17 and the direction of the pressure plate 3, i.e. in the axial direction from the radial section 2a of the housing. In the exemplary embodiment shown in FIG. 17 and 18, springing of the adjusting ring 17 is provided by at least one annular bending spring 26, which may have, for example, two turns and at one of its ends has a radially extending limb 27, which is rigidly connected to the adjusting ring 17, and at the other end has an axially extending limb 28, which is rigidly connected without rotation with the housing 2. The spring 26 is integrated with spring tightening.

 The support ring 25 shown in FIG. 21 and 22, is also made in the form of an annular part, which has the oncoming structures 24 forming surfaces additional to the surfaces bounding the oncoming slopes 18, and the surfaces bounding the oncoming slopes 18 and oncoming counterclones 24 can also be congruent. The installation angle 29 of the oncoming counterclone 24 corresponds to the angle 23 of the oncoming slope 18. As can be seen when comparing FIG. 19 and 21, oncoming slopes 18 and oncoming counterclones 24 in the circumferential direction are distributed similarly. The support ring 25 is connected to the housing 2 without the possibility of rotation. For this, the support ring 25 has recesses 30 distributed along the perimeter, through which rivets 15 pass.

 In FIG. 18, another dashed bending spring 26a is shown by a dashed line, which can be bent at the ends in the same way as bending spring 26 to enable rotation-free connection, on the one hand, with the housing 2, and on the other hand, with the adjusting element 17. This spring 26a also is integrated with spring tension, as a result of which it exerts a twisting force on the adjusting element 17. The use of two bending springs 26, 26a may be preferable in many applications, since when the friction clutch 1 rotates under the action of centrifugal forces applied to the springs 26 or 26a, the spring force is amplified. By applying two bending springs, it is possible to compensate for the increase in force occurring, for example, on the spring 26 due to the force generated by the bending spring 26a. For this, the bending springs 26 and 26a are bent so that they create, at least when centrifugal force is applied, forces on the adjusting element 17, which act oppositely in the circumferential direction. Both coil bending springs 26 and 26a may have different coil diameters, in addition, these coil bending springs may have different winding diameters, as shown in FIG. 18, the centrifugal forces normally associated with this and acting on the springs 26 and 26a, which would generate different circumferential forces on the adjusting element 17, can at least approximately equalize by appropriately calculating the wire thickness and / or the number of turns of the individual springs 26, 26a. In FIG. 18, the spring 28 is located radially inside the adjusting element 17, and the spring 26a is radially outside the adjusting element 17. However, both springs, if properly calculated, could also be located radially inside or radially outside the adjusting element 17.

In FIG. 23 shows a twisted bending spring 26 in horizontal projection. In the expanded state of the twisted bending spring 26, the bends 27, 28 are offset by an angle 31, which can be of the order of 40-120 ^o . In the illustrated exemplary embodiment, this angle 31 is about 85 ^° . 32 indicates the relative position of the bend 27 relative to the bend 28, which it occupies with the new friction linings 7 in the friction clutch 1. The position 33 denotes the position of the bend 27, which corresponds to the maximum allowable wear of the friction linings 7. The adjustment angle 34 in the presented embodiment is about 12 ^o . The spring 26 is designed in such a way that only one wire coil 35 passes between the two bends 27, 28 in the expanded state of the spring. Two turns of wire are axially arranged one above the other in the remaining circumferential region. The spring 26a is made similar to the spring 26, but has a larger diameter of turns and a different direction of clamping with respect to the adjusting element 17 according to FIG. 2. However, the force exerted by the spring 26 on the adjusting ring 17 is greater than the force of the spring 26a.

 In the new condition of the friction clutch 1, the axial toes 18a, 24a forming the ramps 18 and the rams 24 are substantially axially engaged, which means that the rings 17 and 25 located on each other clamp a very small axial structural space.

 In the exemplary embodiment of FIG. 17 and 18, the oncoming counterclones 24 or the cam-shaped protrusions 24a forming them comprise a single part. However, the oncoming counterclones 24 can be formed directly by the housing 2, for example by extruding cam-shaped protrusions that can extend into the cavity of the housing. Extrusion is preferably carried out, in particular, on cases made of sheet material or covers that are made in one piece.

 In order to hold the adjusting ring 17 before assembling the friction clutch in its pulled position, it has engagement portions 36 in the area of the cams 22 for pivoting or stop means, which, on the other hand, can be supported on the housing. Such a persistent means may be provided during the manufacture or assembly of the friction clutch 1 and after the friction clutch 1 is mounted on the support element 6, as a result of which the adjusting device 16 is activated. For this, in the presented embodiment, on the lid or in the housing 2, there are provided in the circumferential direction elongated recesses 37, and in the adjusting ring 17, a recess or ledge 38 is made. In this case, the elongated recesses 37 laid in the circumferential direction must have such a passage that full-time ring 17 could be rotated back in accordance with the maximum possible angle of wear control. After assembling the friction clutch 1, the turning tool can also be axially passed through the grooves 37 of the cover and inserted (into) or into the recesses 38 of the adjusting ring 17. After that, the ring 17 can be turned back with the tool so that it is shifted towards the radial section 2a of the housing 2 and is located relative to this section 2a at the minimum axial distance. Then, in this position, the adjusting ring 17 is fixed, for example, by means of a bracket or a pin, which enters the coaxial recess of the cover and the adjusting ring 17 and prevents the rotation of these two structural parts. This pin can be removed after mounting the friction clutch 1 on the flywheel so that, as already mentioned, the adjusting device 16 is released. The grooves 37 in the housing 2 are made so that when mounting or after dismounting the friction clutch 1 from the flywheel, the adjusting ring 17 can be installed in its extended position. For this, the clutch 1 is first disengaged, so that the control cup spring 4 does not exert axial force on the rotary support 12, thereby guaranteeing a perfect rotation of the adjusting ring 17.

 Another possibility to install the details of the friction clutch 1, already fixed in the internal combustion engine, is to rotate the adjusting element or the adjusting ring after mounting in the internal combustion engine or on the flywheel in the opposite direction or returning it back. To do this, with the help of an auxiliary tool, it is possible to act on the friction clutch, and the practically released ring 17 can be set in a position drawn relative to the pressure plate. After this, the coupling 1 is again connected, due to which the ring 17 initially retains this position pulled back.

 The ring-shaped adjusting element 17 or the support ring 25 may have two sets of axially displaced and axially rising ramps that are distributed around the perimeter of these parts. In this case, the radially inner ramps can be offset relative to the radially outside of the ramps that are in the circumferential direction, and in particular approximately half the length of the ramp or pitch of the ramp. Due to the bias in the circumferential direction, an impeccable central direction is ensured between the adjusting element 17 and the support ring 25.

 Using the diagrams shown in FIG. 24-27 characteristics in more detail explains the principle of operation of the above friction clutch 1.

 Line 40 in FIG. 24 denotes the axial force created depending on the change in the taper of the cup spring 4, namely, when the cup spring is formed between two supports, the radial distance between which corresponds to the radial distance between the rotary support 5 and the radially outer support diameter 3a on the pressure plate 3. The abscissa axis the relative axial path between the two pads is plotted, and in ordinate, the force generated by the cup spring. Point 41 represents the position of the disk spring 4 in plan, which is the mounting position of the disk spring 4 in the closed sleeve 1, i.e. the position at which the disk spring 4 in the corresponding mounting position exerts maximum clamping force on the pressure plate 3. Point 41 by changing the conical position of the disk spring 4 can move up or down along line 40.

 Line 42 is the axial expanding force generated by the lining spring segments 10, which acts between the two friction linings 7. This axial expanding force counteracts the axial force exerted by the Belleville spring on the pressure plate 3. Preferably, the axial force produced by elastic deformation of the spring segments 10 corresponds to at least the force exerted by the cup spring 4 on the pressure plate 3. When the friction clutch 1 is disengaged, the spring segments 10 open I, namely, on the path 43. For this path 48 corresponding to the axial displacement of the pressure disk, the process of switching off the coupling 1 is supported, and this means that a lower maximum tripping force must be created than that which would correspond to the mounting point 41 in the absence of cladding spring elements 10 (in the absence of springing of the plates). When crossing point 44, the friction plates are released, and on the basis of the degressive section of the characteristics of the Belleville spring, the still exerted decisive force is significantly reduced compared to that which would correspond to point 41. The decoupling force of clutch 1 is reduced until it reaches a minimum or lower points 45 of the sinusoidal characteristic 40. When the minimum 45 is exceeded, the necessary decoupling force increases again, and the decoupling path in the region of the tip 4c of the reeds is chosen so that When passing through a minimum of 45, the tripping force does not exceed the corresponding point 44 of the maximum tripping force, preferably remains below it. Thus, point 46 should not be exceeded.

 The spring 13 serving as a force transducer has a path-force relationship characteristic, respectively, of line 47 in FIG. 25. This characteristic 47 corresponds to that which is created when the Belleville spring-like structural part 13 changes in its taper from the unclamped position, namely between two pivot bearings that have a radial gap corresponding to the radial distance between the pivot bearings 11 and 14. As shown line 47 on the graph, the part 13, having the shape of a Belleville spring, has a path 48 of movement, on which the axial force generated by it remains almost constant. In this case, the force generated in this section 48 is selected so that it at least approximately corresponds to the disengagement force of the couplings according to point 44 in FIG. 24. The support force exerted by the sensor spring 13 with respect to the corresponding point 44 characterizing the force of the Belleville spring 4 is reduced in accordance with the gear ratio of the lever to the Belleville spring. This gear ratio in most cases is of the order of 1: 3-1: 5, but for some applications it can also be more or less.

 Said gear spring ratio corresponds to the ratio between the radial distance of the pivot bearing 5 to the bearing 3a and the radial distance of the bearing 5 to the contact diameter 4c, for example for a trip support.

 The mounting position of the element 13, made in the form of a disk spring, in the friction clutch 1 is selected in such a way that it can make an axial spring stroke in the area of the rotary support 5 in the direction of the friction linings 7, which corresponds to at least the axial adjustment stroke of the pressure plate 3 in the direction of the support disk 6, which occurs due to wear of the friction surface and friction plates and provides at least approximately constant axial supporting force on the rotary support 5. This means t, that the linear portion 48 characteristic curve 47 would have to have at least such a length that corresponds to said move when worn, preferably greater wear of stroke, since due to this can be at least partially compensated by mounting tolerances.

 In order to obtain a practically constant or specific opening point 44 of the friction plates 7 when the friction clutch 1 is disengaged, the so-called double segmented spring loading of the plates between the friction plates 7, i.e. spring, in which individual spring segments are arranged in pairs wall to wall, and the individual pairs of segments can have a certain axial pre-tension relative to each other. By tensioning the spring means present between the plates, it is possible to at least basically level or compensate for losses occurring during the service life by laying segments on the rear side of the plates. By laying losses should be understood the losses that occur due to the introduction of segments on the back side of the linings. Due to the corresponding limitation of the axial stroke of the spring between the two friction linings 7, as well as due to the exact spring between the friction linings, it is also possible to achieve that when the friction clutch 1 is released, the pressure plate 3 is squeezed out due to the spring between the plates . In order to obtain a certain axial stroke 43, the axial stroke between the friction linings can be limited by appropriate stops both in the expanding direction and in the clamping direction of the springs 10. As described in patent application P, for example, those used can be used as springs 4206880.8, the contents of which should be included in the subject matter of the application.

 In FIG. 26, line 49 shows the force required to disengage the clutch using a trip element interacting in the disk spring portion 4c to move the pressure plate from point 41 to point 44 (FIG. 24). In addition, line 49 shows the course of the tip of the tongues of the Belleville springs in section 4c.

 In order to ensure the optimal functioning of the friction clutch 1 or an adjusting device that automatically compensates for the wear of the plates, it is advisable that, when considering the actual course of application of the release force 49 of FIG. 26, primarily by means of the pad springs 10 and the sensor spring 13, the forces acting on the cup spring 4, which are larger than the force exerted by the cup spring 4 on the support 11, are created and stacked with each other and after the pressure plate is withdrawn from the friction linings 7, the force still generated by the sensor spring 13 on the disk spring 4 is greater than or at least equal to the decoupling force acting in the region of the tip 4c of the tongues of the disk spring and changing during the trip according to FIG. 26. In this case, the force created by the sensor disk spring 13 on the support 11 must be selected so that the rotation of the ring 17 under the action of the spring 26 is prevented, and thereby the axial displacement of the disk spring is prevented at least approximately in the mounting position of the disk spring point 41 of the ascending branch of characteristic 40 was exceeded.

 Until now, information has been provided that corresponds to the well-defined mounting position of the Belleville spring, in which there is still no wear on the friction linings 7.

 When axial wear, such as friction linings 7, the position of the pressure plate 3 is shifted in the direction of the support disk 6, as a result of which there is a change in the taper of the disk spring (tongues 4c of the tongues are shifted to the right when looking at the plane of the sheet) and thereby there is a change in the pressing force exerted by the disk the spring in the engaged state of the friction clutch 1 in the direction of its increase.

 This change causes point 41 to move in the direction of point 41 ', and point 44 to move in direction 44'. This change leads to the fact that when the clutch 1 is released, the initially existing balance of forces is violated in the area of the rotary support 11 between the control disk spring 4 and the sensor spring 13. The increase in the pressing force of the disk spring to the pressure plate 3 caused by the wear of the plates causes a displacement of the application of the disengaging force in the direction its increase. The resulting characteristic of the disengagement force is shown by the dashed line 50 in FIG. 26. By increasing the release force characteristic during the release process of the friction clutch 1, the axial force exerted by the sensor spring 13 on the disk spring 4 is overcome so that the sensor spring 13 mainly weakens in the region of the pivot bearing 5 for the axial stroke, which mainly corresponds to the wear of the friction pads 7. During this deposition phase of the sensor spring 13, the disk spring 4 is supported in the loading portion 3a of the pressure plate 3, so that the disk spring changes its taper and thus the energy accumulated in it or the torque accumulated in it, and as a result also the force exerted by the disk spring 4 on the pivot bearing 11 or on the sensor spring 13 and on the pressure plate 3. This change occurs, as can be seen in FIG. . 24, in the direction of decreasing the forces generated by the spring disk 4. This change continues until the axial force exerted by the cup spring 4 in the area of the pivot bearing 11 on the sensor spring 13 is in equilibrium with the opposing force created by the sensor spring 13. This means that in the diagram according to FIG. 24 points 41 'and 44' again go in the direction of points 41 and 44. After equilibrium is again established, the pressure plate 3 can again move away from the friction linings 7. During this phase of wear adjustment, i.e. during the disengagement process of the friction clutch 1, the adjusting element 17 of the adjusting device 16 is rotated by the tensioned spring 26, as a result of which the swivel bearing 12 is further shifted according to the wear of the plates, and thus, the clearance-free swivel bearing 5 of the disk spring 4 is provided. After the adjustment process, the characteristic the breaking force again corresponds to line 49 of FIG. 26. Lines 50 and 51 of FIG. 26 represent the axial stroke of the pressure plate 3 with the characteristic pressure-path relationship, respectively, lines 49, 50.

 In the diagram of FIG. 27 shows a graphical characteristic of the force exerted during shutdown on the housing 2 or on the spring disk 13, with extrema being cut off on it. Based on the on position according to FIG. 17, the force 2, in particular the force corresponding to the point 41 (Fig. 24) of the installation of the disk spring 4, acts primarily on the pressure plate 3. During the shutdown process, the axial force exerted by the disk spring 4 on the case 2 or the pivot bearing 12, decreases, respectively, of line 52 (Fig. 27), namely, to point 53. When the point 53 is increased in the direction of disengagement with a conditional clutch, in which the cup spring is axially stationary, but rotatably mounted on the housing, therefore, the rotary support 11 without axial bending axis would be connected to the housing 2, there would be an axial reversal of the action of forces by means of a cup spring 4 on the housing 2 along the radial height of the rotary support 5. In the coupling according to the invention in the area of the rotary support 5, the force generated by the axial reversal produced by the cup spring 4 in the area of the pivot bearing 5, it is perceived by the sensor spring 13. When the point 54 is reached, the disk spring 4 is detached from the loading portion 3a of the pressure plate 3. At least to this point 54 tsepleniya friction clutch 1 is supported by the axial force generated by spring 10 laths. In this case, the force generated by this spring 10 decreases with an increasing disengagement course in the protrusion tip region 4c or with an increasing axial disengagement of the pressure plate 3. Thus, line 52 represents the resultant of the tip considered during shutdown, acting, on the one hand, in the region 4c of the tip the protrusion and, on the other hand, the axial force exerted in the radial region 3a on the Belleville spring 4 by means of the spring 10. When passing through the point 54 in the disengagement direction, the axial force exerted on the Belleville spring 4 on the rotary lining 11 is perceived by the opposing force generated by the sensor disk spring, both of these forces, at least after unloading the friction linings 7 by means of the pressure plate 3, are in equilibrium, and when the uncoupling process continues, the axial force exerted by the sensor spring 13 the area of the pivot bearing 5, becomes predominantly slightly larger than the corresponding disengagement force. Section 55 of characteristic 52 of the diagram according to FIG. 27 shows that with an increasing disengagement course, the disengaging force or the force exerted by the cup spring on the pivot bearing 11 will be less than the disengaging force at point 54. The dashed line 56 corresponds to the condition of the friction clutch 1, in which wear occurs in the area of the friction linings, but regulation has not yet taken place in the area of the pivot bearing 5. It can also be seen here that a change in the mounting position of the disk spring 4 caused by wear causes an increase in the pressure on the pivot bearing 11 or the sensor spring 13 and pus-2 effort. This leads, in particular, to the fact that the point 54 moves in the direction of the point 54 ', as a result of which, with the ongoing process of disengaging the friction clutch 1, the axial force exerted by the disk spring 4 on the sensor spring 13 in the region of the pivot bearing 11 is greater than the opposing force of the sensor spring as a result of which the already described regulation occurs by axial expansion of the sensor spring 13. By means of this adjustment process carried out by the spring 26, i.e. by turning the ring 17 and the axial displacement of the lining 12, the point 54 'is again shifted in the direction of the point 54, thereby again establishing the desired equilibrium state in the region of the pivot bearing 5 between the disk spring 4 and the sensor spring 13.

 In practice, the described control process occurs continuously or in very small steps, as a result of which large point movements and characteristic movements shown for better understanding in the diagrams do not occur in practice.

 During the service life of the friction clutch 1, some functional parameters or operating points may change. So, for example, due to improper actuation of the friction clutch 1, overheating of the lining springs 10 can occur, which can result in upsetting and, as a result, a decrease in the axial springing of the lining springs 10. However, by appropriately calculating the graphical characteristic 40 of the Belleville spring and adjusting the characteristic 47 of the sensor spring 13 accordingly, the friction clutch can operate reliably in operation. The axial settlement of the lining spring 10 would only result in the belleville 4 occupying the pushed spring relative to that shown in FIG. 17 to the position, and the pressing force exerted by the cup spring 4 on the pressure plate would be slightly less, as can be seen in connection with the characteristic according to FIG. 24. In addition, there would be a corresponding axial deformation of the sensor spring 13, and thus a corresponding axial displacement of the pivot bearing 11.

 According to another idea of the invention, the resulting support force acting on the guide disc spring 4 may increase with increasing wear. At the same time, growth in part of the site, together with the maximum allowable wear rate of the friction linings 7, can be limited. In this case, the growth of the supporting force of the control disk spring 4 can occur by means of a corresponding calculation of the sensor spring 13. In FIG. 25, the dashed line and 47a show the passage of the characteristic through section 48. Due to the increase in the support force of the control disk spring 4 with increasing wear, the pressure drop of the control disk spring for the pressure plate 3 can be compensated at least partially due to lowering of the lining spring, for example by inserting segments in the lining. In this case, it may be particularly preferable that the supporting force of the control disk spring 4 increases in proportion to the upsetting of the casing springs or in proportion to the insertion of the segments into the pads. This means that with a decrease in the thickness of the disk in the region of the linings, therefore, a decrease in the distance between the friction surfaces of the linings due to the insertion of segments and / or one settlement of the spring elements of the linings and / or wear of the linings, the mentioned supporting force should increase. It is particularly preferable if the increase in force occurs in such a way that it is greater in the first region than in the adjacent second region, both regions being inside region 48 of FIG. 25. The latter is preferable if the majority of the said insertion between the spring segments and the pads occurs mainly during an insignificant time span compared to the entire service life, and accordingly, the ratio between the spring segments and the friction pads is practically stabilized. This means that with a certain investment there are no more changes regarding further investment. An increase in the supporting force of the control disk spring can also occur at least in terms of wear of the friction linings.

 In the above description of the control process to compensate for the wear of the friction linings, the axial forces that may turn out to be the leaf spring 9 were not taken into account. When tightening the leaf springs 9 in the form of a retraction of the pressure plate 3 from the corresponding friction lining 7, i.e. with respect to the pressing of the pressure plate 3 against the cup spring 4, the disengagement process is supported. The axial force generated by the leaf springs 9 is superimposed with the forces exerted by the sensor spring 13 and the Belleville spring 4, as well as with the tripping force. For illustrative purposes, this was not taken into account in the description of the diagrams of FIG. 24-27. The total loading control disk spring 4 with a disengaged friction clutch 1 against the rolling support 12 located on the side of the cover side, the total force is obtained by adding the forces that appear on the control disk spring 4 mainly by the leaf spring elements 9, the sensor spring 13 and the disengaging force. In this case, the leaf spring elements 9 can be installed between the cover 2 and the pressure plate 3 in such a way that with increasing wear of the friction linings 7, the axial force exerted by the leaf springs 9 on the control disk spring 4 becomes greater. Thus, for example, in a stroke 48 according to FIG. . 25 and thus also for the wear-compensating stroke of the adjusting device 16, the axial force generated by the leaf springs 9 may have a characteristic according to line 47b. From FIG. 25, it can be concluded that with increasing upsetting of the sensor spring 13, the return force exerted by the leaf springs 9 on the pressure plate 3 increases, which also acts on the control disk spring 4. By adding the force course according to characteristics 47b and the disk spring characteristic, a resultant power stroke is obtained that is axially acts on the disk spring 4, namely in the sense of pressing the disk spring 4 to the rotary support 12 located on the side of the lid. asno line 47a, and to the top permutation zone has an initial strength increase, then rolling in the constant region, it is expedient to calculate a touch plate spring so that it has a stroke characteristic curve according to the line 47c of FIG. 25. By adding the force characteristic according to line 47c and the characteristic according to line 47b, then a force graphic characteristic according to line 47a is obtained. Thus, by appropriately tightening the leaf springs 9, the support force exerted by the sensor spring corresponding to the support force characteristic can be reduced. By appropriately performing and arranging the leaf spring elements 9, the settlement of the overlay springs and / or the insertion of the overlay spring segments in the overlays can at least partially be compensated. Thus, as a result of this, it can be ensured that the disk spring 4 basically maintains the same operating point or the same working section, so that the disk spring 4 exerts a substantially at least approximately constant clamping force on the friction clutch pressure plate 3. In addition, when calculating the friction clutch, in particular the sensor spring 13 and / or leaf springs 9, the adjustment springs 26 and / or 26a ultiruyuschee axial force which counteracts the sensor spring 13 and / or the leaf spring 9.

 When calculating the friction clutch 1 with biased leaf springs 3, it should also be taken into account that by tightening the leaf springs 9, the axial force exerted by the pressure plate 3 on the friction linings 7 is affected. Thus, this means that when the leaf springs 9 are compressed in the direction of the control disk spring 4, the clamping force exerted by the disk spring 4 is reduced by the pre-compression force of the leaf springs 9. Thus, in this kind of friction clutch 1, a resultant characteristic is formed pressure force of the pressure plate 3 or friction linings 7, which is obtained by adding the graphical characteristics of the pressure of the Belleville spring with the compression characteristics of leaf springs 9. Assuming that, if we consider the range of operation of the friction clutch 1, graphical characteristic 40 according to FIG. 24 represents the resulting force characteristic of the control disk spring 4 and pre-compressed leaf springs 9 in the new state of the friction clutch 1, with a decrease in the distance between the pressure plate 3 and the support disk 6 as a result of wear on the pads, an offset of the resulting characteristic would be obtained in the sense of lowering.

 In FIG. 24 shows the line 40a, which corresponds, for example, to the overall wear of the linings of 1.5 mm. As a result of this, the displacement of the line 40 in the direction of the line 40a, which is allowed during the service life of the friction clutch, decreases the axial force exerted by the Belleville 4 on the sensor spring 13 when the friction clutch 1 is disengaged, namely, as a result of the opposing moment exerted by the increasing spring wear of the leaf springs 9 on the Belleville spring 4 This opposing moment results from the radial distance between the support 5 and the load diameter 3a between the control cup spring 4 and the pressure plate th 3.

 Presented in FIG. 28 and 29, the friction clutch 101 differs mainly from the friction clutch 1 shown in FIG. 17 to 18, in that the adjusting ring 117 is loaded with coil springs 126 in the circumferential direction. Regarding its function and principle of operation in relation to the compensation of wear of the friction plates, the adjusting ring 117 corresponds to the adjusting ring 17 according to FIG. 18-20. In the shown embodiment, three coil springs 126 are provided, which are evenly distributed around the circumference and are compressed by the clutch housing 2 and the adjusting ring 117.

 As follows, in particular, from FIG. 30, the adjusting ring 117 has radial protrusions or steps 127 on the inner circumference, on which one of its ends can be supported by arcuate coil springs 126 for loading the adjusting ring 117 in the circumferential direction. Other end sections of the springs 126 are supported on the supports 128 carried by the clutch body 2. In the illustrated embodiment, these stops 128 are formed by screw-like connecting elements that are connected to the cover 2. These stops 128 can be formed, however, also by axial formations, which are made in one whole with 2 clutch housing. Thus, for example, stops 128 can be formed by embossments or overlays axially extruded from the sheet body 2. As can be concluded, in particular, from FIG. 29 and 30, the ring 117 around the inner circumference can be made in such a way that at least mainly in the area of the location of the springs 126, and preferably also through the angle of rotation of the ring 117 necessary for controlling wear, or through the spring travel 126, there is a guide 129, which provides axial fixation and radial support of the springs 126. The guides 129 of the springs in the illustrated embodiment are formed when viewed in cross section, mainly by semicircular recesses, limiting over which are mainly adapted to the cross section of coil springs 126.

 This kind of implementation has the advantage that with the rotating friction clutch there is a perfect guide for the springs 126 so that they cannot radially deviate. To further lock the coil springs 126, as shown in FIG. 29, the cover 2 may have axial formations 130 in its radial inner portion that overlap the springs 126 in the axial direction. Instead of the individual formations 130, the cover 2 may also have one circumferential and axial inner flange 130. The inner flange 130 may serve to limit the expansion of the cup spring 4.

 The guide of the adjustment springs 126 of FIG. 28-30 has the advantage that, in the rotary clutch assembly 1, the individual turns of the springs 126 can be radially supported on the adjusting ring 117 under the influence of centrifugal force, and the regulating forces created by the springs 126 in the circumferential direction due to the frictional resistance between the coil springs and the adjusting ring 117 are reduced or completely disappear. Thus, the springs 126 during rotation of the friction clutch 101 can (due to the frictional forces suppressing the action of the springs) be practically stationary. Due to this, it can be achieved that, at least when the number of revolutions exceeds the idle speed of the internal combustion engine, the adjusting ring 117 cannot be turned by means of springs 126. Thus, it is achieved that the wear of the friction plates is compensated only when the friction clutch 101 is actuated with the number idle speed, or at least approximately at idle speed. However, the locking of the adjusting ring 117 can also be carried out in such a way that only when the internal combustion engine, therefore, the non-rotating friction clutch 101 is stopped, regulation can occur due to wear on the pads.

 Blocking the adjustment process during rotation of the friction clutch 1 or when exceeding a certain number of revolutions may also be preferable in the form of execution according to FIG. 1 and 2. To this end, for example, means can be provided on the housing 2, which, when subjected to centrifugal force on the adjusting element 17, lock against rotation, namely against the regulating force created by the twisted bending spring 26 and / or 26a. In this case, the blocking means can be formed by at least one radially pressed outward mass under the action of centrifugal forces, which rests, for example, on the inner flange of the ring 17 and can create friction there, which causes a blocking moment on the ring 17, the magnitude of which is greater than that exerted adjusting springs on the ring 17 torque.

 To radially support at least one portion along the length of the spring 126, support means located on the housing 2 may also be provided. These support means in the embodiment according to FIG. 28 and 29 can be made integrally with the stops 128. For this, the stops 128 can be made in the form of an angle, so that they have, as needed, one section extending in the circumferential direction, which is at least one part of the length of one spring 126 into her. Due to this, at least one part of the coil of the spring can be carried out and at least have support in the radial direction.

 As can be inferred from FIG. 29 provided in FIG. 18, the wire ring 11 is eliminated and replaced by elements 111 located in the region of the tip of the protrusions of the sensor spring 113. To this end, the protrusions 113c in the region of their tip on their side facing the control disk spring 4 are barrel-shaped.

 In FIG. 31-33, another embodiment of a wear adjusting device according to the invention is presented, in which separate adjusting elements 217 are used instead of a ring-shaped adjusting ring. These adjusting elements are evenly distributed around the circumference of the cover 202. The adjusting elements 217 are formed by push-button or disk-shaped structural parts that have circumferentially extending and axially rising ramps 218. The ring-shaped adjusting elements 217 have one central height a hole or hole 219 through which axial rod-shaped nozzles 215a mounted on the lid pass, so that the ring-shaped adjusting elements 217 are rotatably mounted on these nozzles 215a. Elements 225 are stamped onto the cover 202, which form oncoming counterclones 224 for slopes 218. Between the adjusting element 217 and the cover 202 a spring element 226 is clamped, which loads the adjusting element 217 in the direction of rotation that regulates, the spring element 226 can pass, as seen in FIG. 31 around one nozzle 215a, thus, be made similar to a helical spring. At the end sections of one spring 226, formations such as bends or bent shoulders are provided to support one end of the spring on the housing 202 and the other end of the spring on the corresponding adjusting element 217. When the disk spring 204 or sensor spring 213 is axially displaced in the region of the swing bearing 205 the adjusting elements 217 rotate and compensate for the displacement by running inclinations 218 on inclinations 224.

 The axial support of the sensor disk spring 213 on the housing 202 is carried out using pads 214, which were formed from an axially extending portion of the housing 202 and pressed radially inward under the outer regions of the sensor spring 213.

 The ring-shaped adjusting elements 218 have the advantage that they can be made substantially independent of centrifugal forces with respect to their regulatory action.

 Instead of those shown in FIG. 30 rotatable or rotatable adjusting elements 217 could also use wedge-shaped adjusting elements, offset in the radial and / or circumferential direction to regulate wear. These wedge-shaped adjusting elements can have one elongated recess through which the axial nozzle 215a can pass in order to guide the corresponding adjusting element. Wedge-shaped adjusting elements can operate with regulation based on the centrifugal force acting on them. However, power batteries may also be provided that load the wedge-shaped adjusting elements in the direction of regulation. For an impeccable direction of the wedge-shaped adjusting elements, the housing 202 may have oncoming formations. Wedge-shaped surfaces of adjusting elements extending at a certain incidence angle with respect to a plane extending perpendicular to the axis of rotation of the friction clutch can be provided on the side of the housing and / or on the side of the control disk spring. When using such wedge-shaped individual elements, it is advisable to manufacture them from a light material in order to minimize the centrifugal forces acting on them.

 The selection of material pairs between the structural parts forming the adjustment slopes is preferably carried out in such a way that, during the service life of the friction clutch, the control between the ramps and the ramps cannot interfere with the regulation. To avoid such a setting, at least one of these structural parts may be coated at least in the area of slopes and counterclones. By means of such coatings, in particular, corrosion can be eliminated by using two metal structural parts. Grasping or gluing between structural parts forming adjusting slopes, in addition, can be eliminated due to the fact that leaning against each other and forming slopes, as well as counterclones, structural parts are made of material with different expansion coefficients, so that as a result of temperature fluctuations, friction clutches that appear during operation, which are in contact with the surface, form adjusting slopes, move relative to each other. Thus, the structural parts forming oncoming slopes and oncoming counterclones are always held with the possibility of moving relative to each other. Thus, adhesion or seizure between these parts cannot occur, since due to various stretching, these parts are again again torn apart or disconnected. Separation of the adjustment slopes can also be achieved by the fact that, due to different strengths and / or parts, the centrifugal forces acting on these parts cause various elongations or displacements, which in turn eliminate the adhesion or setting of the parts.

 In order to avoid traction between ramps and ramp, at least one adjusting device can be provided that exerts an axial force on the adjusting element or on the adjusting elements when the friction clutch is released or when adjusting wear. In addition, the adjusting element 17, 117 can axially connect to one structural part having sections axially miscible during wear. This connection can occur, in particular, in the area of the swinging support 5, namely with the control cup spring 4 and / or the sensor spring 13.

 In the diagram of FIG. 34 shows a characteristic 340 of a pressure plate spring having a lower point or a minimum of 345, in which the force exerted by the pressure plate spring is relatively small (about 450 Nm). The maximum of a Belleville spring with a characteristic 340 of the "path-force" relationship is about 7600 Nm. The characteristic 340 is obtained due to the deformation of one cup spring between two supports radially distant from each other, namely as described in connection with the characteristic 40 according to FIG. 24 and in connection with a cup spring 4.

 The characteristic 340 of the Belleville spring can be combined with the characteristic 342 of the casing spring. As can be inferred from FIG. 34, the path-force relationship of the characteristic 342 of the spring segments of the pads is close to the characteristic 340 of the pressure plate spring, or both characteristics are only at a small distance from each other so that the corresponding friction clutch can be driven with very little force. In the working area of the cushioning spring, the theoretical tripping force is obtained from the difference between two vertically arranged, one above the other, points of the lines 340 and 342. This difference is indicated by 360. The actual necessary pressing force is reduced by the gear ratio of the control element levers, such as, for example, protrusions Belleville springs. This has also been described in connection with the embodiment of FIG. 17 and 18, as well as the diagrams of FIG. 24-27.

 In FIG. 34, the dashed line represents another characteristic 440 of the control disk spring having one minimum or one lower point 445, in which the force exerted by the disk spring is negative and, therefore, acts not in the clutch direction of the corresponding friction clutch, but in the direction of disengagement. This means that if point 461 is exceeded during the tripping phase, the friction clutch automatically opens. With the characteristic 440 of a disk spring, the characteristic of the casing springs can be mated in accordance with line 442, in order to obtain the minimum tripping force, one should strive to parallelly pass the characteristics of the casing springs 442 and the disk spring 440.

 In FIG. 35 shows graphical characteristics of 340 and 342, or 440 and 442 of a pressing force applied to disengage the corresponding friction clutch to control levers, for example, cup spring tongues. As can be seen, the stroke 349 of the clamping force of the presented characteristic 340 and 342 is constantly in the positive power zone, this means that to keep the clutch in the disengaged state, a force in the direction of disengagement is constantly necessary. The downforce stroke 449, represented by the characteristics 440 and 442, has a portion 449a in which the downforce first decreases and then passes from a positive to a negative force region so that the corresponding friction clutch in a disengaged state does not need holding force.

 In the embodiment of FIG. 36, 36a, 37 of an embodiment of the friction clutch 501, the touch disc spring 513 is axially supported through a bayonet coupling on the clutch cover 502. To this end, the sensor spring 513 has linings 513a radially extending from the outer circumference of the annular main body 513b, which are axially supported in the radial portions 502a in the form of linings formed from the cover material. The cover plates 502a are formed from a substantially axially extending edge portion 502b of the cover, and it is advantageous when for this purpose the covers 502a are at least partially formed primarily through one free cut-out 502c or 502d of the cover material. By at least partially cutting the slats 502a, they can more easily deform to their predetermined position. As can be concluded, in particular, from FIG. 37, the pads 502a and the jibs or protrusions 513d correspond to each other so that the sensor spring 513 can be centered with respect to the cover 502. In the illustrated embodiment, the pads 502a also have one small axial stage 502e.

 In order to perfectly position the sensor spring 513 with respect to the housing 502 while receiving the bayonet locking connection 514, at least three in a preferred manner around the circumference of the cover 502, the uniformly distributed pads 502 are aligned with other portions of the cover so that after a certain relative rotation between the sensor spring 513 and the cover 502, the corresponding protrusions 513d come into contact with the circumferential stop 502f, and thus a further relative rotation between the cen ornoy spring 513 and the cover 502. The focus 502f illustrated in the exemplary embodiment as follows in particular from Fig. 36a is formed by the axial protrusion of the cover 502. From FIG. 36a, it also follows that the separate, preferably three, pads 502a form another rotation restriction 502g between the cover 502 and the protrusions 513d of the sensor spring 513. In the example shown, the same pads 502a form anti-rotation stoppers 502f and 502g for both directions of rotation. The stoppers 502g, which eliminate the unlocking between the sensor spring 513 and the cover 502, are formed by axially radially extending curved edges of the protrusions 502a. The circumferential stops 502f and 502g give a certain circumferential positioning of the sensor spring 513 with respect to the cover 502. To obtain a blocking connection 514, the sensor spring 513 is axially compressed in the direction of the cover 502 so that the protrusions 513d axially enter the free cutouts 502c and 502d and pass with an axial arrangement on the supports 502a in the cover. After that, the cover 502 and the sensor spring 513 can rotate relative to each other until some of the protrusions 513d come into contact with the rotation restrictions 502f. After this, the sensor spring 513 partially opens, so that some of the protrusions 513d, when viewed in the circumferential direction, come between the respective stops 502f and 502g, and all the protrusions 513d are superimposed on the supports 502a located on the side of the cover. By performing the bayonet lock 514 defined by the invention, it is ensured that when the friction clutch 501 is mounted, the protrusions 513d do not come to be located near the covers 502a from the cover side.

 In the previously presented exemplary embodiments, the annular main part exerting the actual spring force of the sensor spring 513, for example 513b, is provided radially outside the load section or the support section between the pressure plate and the control disk spring. However, for some applications, it may also be appropriate that the annular main part of the sensor disk spring is provided radially inside the loading diameter between the pressure plate and the control disk spring. Thus, for the embodiment of FIG. 17 and 18, this means that the main part 13b exerting an axial clamping force of the sensor spring 13 is provided radially inside the load section 3a between the working disk spring 4 and the pressure plate 3.

 In the embodiment of FIG. 36-37, the oncoming counterclones 524 located on the housing side are formed by cam-shaped elements housed in the sheet housing 502. In addition, in this embodiment, coil springs 526, sandwiched between the housing 502 and the adjusting ring 517, are guided by guide mandrels 528 made for integral with the adjusting ring 517 and passing in the circumferential direction. These guide mandrels 528 may, as follows, in particular, from FIG. 21, to have, in the axial direction, one elongated cross section adapted to the inner diameter of the springs 526. The guides 528 extend in at least one partial portion of the longitudinal extension of the springs 526 to the latter. Thus, at least one part of the coil of the spring can be guided and supported at least in the radial direction. In addition, longitudinal bending or popping of the springs 526 in the axial direction can be eliminated. By means of mandrels 528, the installation of a friction clutch can be greatly facilitated.

 In FIG. 38, the adjusting ring 517 is partially shown. The adjusting ring 517 has radial inward formations 527, bearing rod-shaped, circumferentially extending, guiding sections 528 for coil springs 526. In the illustrated embodiment, the spring-based sections 528 are integral with the plastic ring 517, manufactured as an injection molded part. However, the directions of the springs or the sections 528 based on the springs can also be formed by separate structural parts (or all together by a single structural part), connecting (respectively connecting) with the adjusting ring 517, for example through latching. Thus, all the guide sections 528 can be formed by a single, if necessary, circumferentially open ring connected to the adjusting ring 517 through three at least connecting places, preferably in the form of a snap fastening.

 In the same way as described in connection with FIG. 28 and 29, the coil springs 526 are additionally additionally, for example, based on the action of centrifugal forces, can be radially supported on the respective portions of the cover 502 and / or the adjusting ring 517.

 Supports for coil springs 526 located on the side of the cover are formed by flaps obtained by plastic deformation from the cover material and extending in the axial direction, or by embossments 526a forming the axial walls. These supporting sections 526a of the springs 526 are expediently designed in such a way that the respective ends of the springs are guided and thereby protected against unacceptable displacement in the axial and / or radial direction.

 In the embodiment of FIG. In embodiment 39 of the coupling 601, a sensor spring 613 is provided on the side of the housing 602 facing away from the pressure plate 603. Due to the position of the sensitive spring 613 outside the internal cavity of the housing that receives the pressure plate 603, the thermal load of the sensor spring 613 can be reduced, thereby eliminating the risk of settling springs 613 due to thermal overload. It is also better to cool the spring 613 on the outside of the housing 602.

 The swing support 611, provided on the side of the control disc spring 604 facing away from the lid, is supported through the distance rivets 615, passing axially through the corresponding recesses of the disk spring 604 and the housing 602, and axially connected to the sensor spring 613. In the illustrated embodiment, the distance rivets 615 are connected with a sensor spring 613. Instead of distance rivets 615, other means can also be used that form the connection between the rolling support 611 and the sensor spring 613. In the illustrated exemplary embodiment, the distance rivets 615 are connected to the sensor spring 613. Thus, for example, the sensor spring 613 could have axially extending pads on the radially inner portion supporting the rolling bearing 611 with respective radial regions or directly forming this rolling bearing 611 with corresponding formations obtained by plastic deformation.

 Instead of elements 615 riveted rigidly to the sensor spring, other arrangements, for example, elements pivotally attached to the sensor, can also be used.

 In the exemplary embodiment of FIG. 40, the sensor spring 713 extends radially inside the rotary support 715 for the control disk spring 704. The sensor spring 713 is supported on the cover 702 in its radially inner portions, for this the cover 702 has plates 715 axially passing through the corresponding grooves or recesses of the disk spring 704, axially supporting touch disc spring 713.

 Presented in FIG. 41, the adjusting ring 817 can be used in the friction clutch of FIG. 20-21. The adjusting ring 817 has radially inner formations 827 extending radially. The formations 827 obtained by plastic deformation have radial nozzles 827a forming support sections for circumferentially clamped coil springs 826 between the coupling cover and the adjusting ring 817. To conduct and facilitate the installation of coil springs 826, a ring 828 is provided that is broken or open on the outside circles. Ring 828 is connected to radial formations 827a obtained by plastic deformation. For this formation 827a can have in the circumferential direction passing recesses or grooves made in such a way that they form, together with the ring 828, a snap connection with a latch. The support spring 826 located on the side of the cover is formed by axial linings 826a of the coupling cover. The axial pads 826a have, as needed, one axial cutout 826b for the base of the ring 828. The cutouts 826b are designed so that the ring 828 with respect to the pads 826a has the possibility of axial displacement at least according to the wear path of the friction clutch. In addition, it is particularly expedient that the recesses introduced into the radial formations 827a for receiving the ring 828 and the recesses 826b, when viewed in the axial direction, are made in the opposite direction, or, in other words, that the recesses in the formations 827a are open in the same axial direction, and cutouts 826b in the other axial direction.

 In the embodiment of FIG. 42 of the embodiment of one friction clutch 901, the control of the poppet spring 904 in the disengaging direction is maintained in the middle portion of the main body 904a of the poppet spring 904. Radially outward of the base, the portion 904a is supported on the pressure plate 903 and exits radially inward through the pivot bearing 905. This means that the bearing 905 of the rotation is removed relatively far from the inner edge of the main part 904a of the Belleville spring 904 or of the groove ends forming the protrusions of the Belleville spring 904 compared to so far natural friction clutches. In the presented embodiment, the ratio of the radial width of the sections of the main part provided radially inside the support 905 of the rotation is about 1: 2. It is advisable when this ratio is 1: 6-1: 2. By maintaining the control disc spring 904 in this manner, damage or overloading of the main body 904 can be repaired, and damage or overload of the main part 904a of the disc spring 904 in the region of the bearing 905 can be repaired.

 In addition, in FIG. 42, a dashed line indicates a specific formation 903a provided on the pressure plate 903. Through such provided on the pressure plate 903, in particular in the region of the support cams 903b, of the formation 903a, the control disk spring 904 can be centered with respect to the clutch 901. Thus, the control disk spring 904 can be fixed by centering the outer diameter in the radial direction with respect to the cover 902 so that, as shown in FIG. 42 centering rivets or fingers 915 may fall off. Although not shown, the outer diametric alignment can also be carried out through pads or elements obtained by plastic deformation from the cover 902.

 In the friction clutch 901, the sensor spring 913 is configured in such a way that the main force portion 913a is provided radially inside the cams 903b.

 To maintain the control disk spring 904, on the one hand, and to actually maintain the cover 902, on the other hand, the sensor spring 913 has radial jaws or protrusions extending, on the one hand, from the main part 913a radially inward and, on the other hand starting from the main part 913a extends radially outward.

 As shown in FIG. In 43 embodiments of one friction clutch 1001, a force directed opposite to the disengaging force of the friction clutch or the rotational force of the control disk spring 1004 is applied by a sensor spring 1013 axially compressed between the housing 1002 and the pressure plate 1003. In one such embodiment, the control disk spring 1004 in the region rotation or rolling 1005 is not supported by the rotation support in the direction of disengagement. The adherence of the disk spring 1004 to the pivot bearing 1012 located on the side of the cover is guaranteed by the bias force of the sensor spring 1013. This sensor spring is designed so that during the disengagement of the friction clutch 1001, the axial force exerted by this sensor spring 1013 on the disk spring 1004 is equal to or greater than the release force of the friction clutch 1001. It must be ensured that when there is no wear on the friction linings, the cup spring 1004 is constantly in contact with the e lid support or pivoting supports 1012. In this same manner as was described in connection with the earlier embodiments, should be coordination between the individual active in the axial direction and overlapping one another effort. These forces created by the sensor spring 1013, the springs between the pressure plate 1003 and the housing 1002, the leaf spring elements, the disc spring 1004, the release force of the friction clutch 1001 and the adjusting spring elements acting on the adjusting ring 1017, must be consistent with each other.

 In the friction clutch 1101 of FIG. 44, the sensor spring 1113 is supported radially outside the annular support region 1112 from the cover side. In the present exemplary embodiment, there is mutual support between the control disk spring 1104 and the sensor spring 1113 on the pressure plate 1103 also radially outside the support diameter 1103a of the control disk spring 1104. To support the cover 1102, the sensor spring 1113 has a radially outwardly shaped protruding arms 1113b radially outwardly axially supported through the bayonet lock 1514 on the cover 1102 and secured against rotation in the same way as described in connection with FIG. 36-37. For mounting the sensor spring 1113, the cover 1102 has corresponding axial recesses 1502b into which the radially external support shoulders of the sensor spring 1113 can be axially inserted to create a bayonet lock 1514. The contact of the disk spring 1104 with the rotary plate or the support plate 1112 from the side of the cover is provided by the pressure force of the sensor spring 1113.

 Using FIG. 43, the operation of the coupling is explained in more detail. In this case, the force of the sensor spring is calculated in such a way that it corresponds to the disengagement force of the adjustment point. If the disengagement occurs after wear of the linings (or wear elsewhere) and the angle of the disk spring changed due to this, and as a result of a higher force on the disk spring, the latter first pivots around the support 1012 near the pivot point. Since at this point the decoupling force is equal to the force of the sensor together with the lining spring — the residual force — the disk spring rotates with further disengagement around the support on the pressure plate until an equilibrium is reached between the decoupling force and the sensor force. In this case, the cup spring is diverted from the support from the side of the cover and releases it for adjustment. With a further course of disengagement, the disengaging force continues to fall, the value of the sensor force rises and, through the pressure plate, presses the Belleville spring to the support 1012 from the side of the cover, around which a further rotation of the Belleville spring occurs. When the Belleville spring moves from the support from the cover side to the support from the side of the pressure plate, the Belleville spring changes the nature of its functioning as a two-shouldered lever. It rests on the pressure plate mainly with a decoupling force on the pressure plate and, due to this, is separated from the support from the side of the cover. With a further trip path, the force of the sensor spring exceeds the associated drop in force and again presses the cup spring to the support from the side of the cover, as a result of which the adjustment device is blocked and the regulation process ends. After that, the Belleville spring again acts as a two-arm lever while continuing the trip. The disk spring must be calculated taking into account the forces of all springs acting indirectly or directly on the disk spring. These include, in particular, the forces that are generated by the control disk spring and the parts of the corresponding equalizing or adjusting device axially displaced relative to the cover.

 The embodiment of FIG. 44 has the additional advantage that, in the engaged state of the friction clutch, the Belleville spring 1104 is practically loaded or acts as a two-armed lever, and thereby the Belleville spring 1104 is clamped between the support 1112 from the lid side and the support 1103a from the side of the pressure plate when the friction is disengaged the clutch 1101, the disk spring is supported, however, almost exclusively by the sensor spring 1113 and rotates around the support zone 1113a while axially displacing the support zone 1113a so that it then it acts as a one-armed lever.

 The sensor spring 1113 of FIG. 44 can, like the sensor springs in other drawings, be reckoned with any diameter of the control disk spring 1104, if appropriately calculated or selected. Thus, the support of the sensor spring 1113 can occur at a diameter of the disk spring 1104, which is between the turning area 1105 from the side cover and a supporting diameter 1103a from the side of the pressure plate. In addition, the support of the sensor spring 1113 can occur on a cup spring 1104 and radially inside the support diameter 1105 from the side of the cover. Moreover, as a rule, the axial support force generated by the sensor spring 1113, the greater, the smaller its reference diameter 1113a on the disk spring. In addition, the spring zone of the sensor spring 1113 with an almost constant force is greater, the farther the reference diameter 1113a between the springs 1104 and 1113 is located from the reference diameter from the side of the cover 1105 of the disk spring 1104.

 In the embodiment of FIG. 45 there is an adjustment device made in the same way as in the previous drawings, in particular with what is described in connection with FIG. 17-30. The control cup spring 1204 is rotatably disposed between the annular rolling bearings 1211 and 1212. The support 1211, which is adjacent to the pressure plate 1203, is loaded with a sensor spring 1213. The friction clutch 1201 has a device 1261 that ensures that the slopes of the adjusting ring 1217 with counterclones made from the side of the lid are not grasped during the life of the fractional clutch. In the illustrated exemplary embodiment, counterclones are provided on a support ring fixed without the possibility of rotation on the cover, similarly to the embodiment shown in FIG. 18. The adhesion between slopes and counterclones could lead to the fact that the desired regulation could not occur due to wear.

 The device 1261 is a tearing mechanism that, when the friction clutch 1201 is disengaged and if there is wear on the friction plates 1207, can exert axial force on the adjusting ring 1217, thereby preventing adhesion between slopes and counterclones. The mechanism 1261 includes an axially spring element 1262, which in the illustrated embodiment is axially connected to a disk spring 1204. The element 1262 has an annular membrane or disk spring spring main body 1262a, which is radially externally connected to the disk spring 1204. From the inner edge of the ring spring the main body 1262a pass along the perimeter of the axial plate 1263, passing through the axial recesses of the Belleville spring 1204. The plate 1263 have at their free end supporting elements in the form of bends 1264, which are included in the corresponding recesses 1265 of the adjusting ring 1217. The recesses 1265 are made in the form of a radial groove or a closed groove on the ring 1217. The distance between the bend 1264 and the recess 1265 in the cordless state of the friction clutch is selected so that at least for most of the tripping phase, there is no contact between the bends 1264 and the recesses 1265. Preferably, the bends 1264 only come into contact with the recesses 1265 when the friction clutch is fully disengaged, whereby the element 1262 can be elastically clamped between the adjusting ring 1217 and the cup spring 1204. This ensures that as soon as the axial displacement of the swivel bearing 1211 occurs, the adjusting ring 1217 is forced to rise from the incline slopes from the side of the cover. In addition, the mechanism 1261 must prevent the ring 1217 from being adjusted too much, for example due to erroneous adjustment of the tripping system, due to the fact that with a too large angle of rotation of the disk spring 1204 in the direction of disengagement, the spring element 1262 will press the adjusting ring 1217 to the cup spring 1204, as a result of which there will be a stop from turning the adjusting ring 1217 relative to the cup spring 1204. Thus, it must be ensured I mean that when point 46 is exceeded according to FIG. 24, in the disengagement direction, the adjusting ring 1217 will be prevented from turning relative to the cup spring 1204, since if the point 46 is exceeded, the return force of the sensor spring 1213 will be overcome, as a result of which, if the friction clutch discs are not worn out, regulation could occur. This would lead to a change in the operating point, i.e. changing the mounting position of the cup spring 1204, namely in the direction of decreasing the clamping force. This means that in FIG. 24, the operating point 41 would move along characteristic 40 in the direction of the minimum indicated by 45.

 In an exemplary friction clutch, the details of which are shown in FIG. 46-48, individual coil springs 1326 are placed on the plates 1328, made in one piece with the cover 1302 of the coupling. The plates 1328 are molded from the sheet material of the cover 1302 by forming, for example, stamping, U-shaped cutouts. The plates 1328 extend when viewed in a circumferential direction, arcuately or tangentially, preferably at least approximately at the same axial height as the adjacent portions of the cover. In FIG. 48, it can be seen that in the illustrated embodiment, the plate 1328 is offset approximately half the thickness of the material relative to the cover bottom region 1302b. The width of the plate 1328 is selected so that the coil spring provided on it can pass in both radial and axial directions.

 The adjusting ring 1317, loaded by the springs 1326 in the direction of adjustment, has on its inner diameter radially inwardly directed formations or protrusions 1327 extending between the cover 1302 and the cup spring 1304. The protrusions 1327 have an axially radially extending fork or U-shaped member 1327a, both teeth 1327b of which, extending in the axial direction, pass through a cutout 1302a in the lid. Adjusting springs 1326 rest on elements 1327a or their teeth 1327b.

 The adjusting ring 1317 is likewise supported by its ramps on the ramps 1324, formed on the cover 1302, as described in connection with the previously mentioned drawings. The formations made on the cover, forming counterclones 1324, are, however, made in such a way that they form ventilation holes 1324 in the direction of rotation of the clutch. Due to this design, when the corresponding clutch rotates, its better cooling due to forced air circulation. Particularly due to this, the adjustment ring 1317 made of artificial material is cooled, which significantly reduces the thermal load on this part.

 According to another embodiment, the sensory force that acts on the control disk spring of the friction clutch is applied, for example, by leaf springs located between the clutch housing and the pressure plate, and these leaf springs, the pressure plate and the body can be connected without rotation but with the possibility of limited axial displacement relative to each other. In this form of execution, it would not be necessary to have a special sensor spring, but it would be possible, for example, to perform leaf spring elements 9 of the friction clutch 1 according to FIG. 1 and 2 so that they additionally have the function of a touch disk spring 13.

 Due to this, there may be no need for a sensor spring 13 or a break-in ring 11. In this case, the leaf spring elements 9 would have to be designed in such a way that during operation of the friction clutch 1 and without wear of the plates, the control disk spring 4 remains adjacent to run-in support 12 from the side of the cover. However, as soon as the corresponding wear occurs on the friction plates 7, as a result of which the disengaging force of the disk spring 4 increases, the leaf spring elements 9 make it possible to adjust the disk spring 4 in accordance with the wear. The leaf spring elements integrated in the friction clutch have at least the most linear force-path characteristic for at least the most required adjustment of the friction clutch or pressure plate. Thus, this means that the leaf spring elements 9 must have the same as described in connection with FIG. 25 area of characteristic 28 according to characteristic 47 or 47a.

 The invention is not limited to the presented and described examples of execution, but also covers those options that can be formed by a combination of features or elements described in this invention. In addition, the individual features described in connection with the drawings or a description of the operation of the device may individually constitute independent inventions.

 Thus, the applicant intends to claim protection and others only mentioned in the description, especially in connection with the drawings, disclosed features of inventive value. Therefore, the foregoing claims represent only proposals for the protection of the invention, without limiting the possible wider scope of protection.

Claims (117)

 1. The clutch assembly containing a friction clutch with a pressure plate connected to the support disk without the possibility of rotation, but with the possibility of limited axial displacement, and the pressure plate is loaded with at least one pressure spring in the direction of the clutch disc, sandwiched between the pressure plate and the support the disk, in addition, there is an adjustment device that compensates for at least the wear of the friction linings of the clutch disc, made with the ability to create an almost constant force on the pressure the plate by means of a clamping spring, and the friction clutch has control means for disengaging and engaging, characterized in that the control means are adapted to be actuated by means of a releasing means mounted with axial displacement, moreover, depending on at least wear friction linings, the control means are axially displaced in the direction of disengagement movement, and there is a device in the power flow between the release means and the control means necks least approximately compensating axial offset control means.  2. The clutch assembly according to claim 1, characterized in that the equalization device is located between the trip element and the controls.  3. The clutch assembly according to claim 1 or 2, characterized in that the friction clutch has a housing mounted on a support disk and made in the form of a cover made of sheet material having a bottom facing the trip element, and the equalization device is arranged to clamp between the controls and the bottom.  4. The clutch assembly according to any one of claims 1 to 3, characterized in that the pressure spring is made in the form of a disk spring axially clamped between the clutch housing and the pressure plate having a spring-shaped annular main body and tongues extending radially inward, forming control means.  5. The clutch assembly according to any one of claims 1 to 4, characterized in that the equalization device has adjusting slopes rising in the axial direction.  6. The clutch assembly according to claim 5, characterized in that the adjusting slopes are formed by at least incident slopes located on the annular part.  7. The clutch assembly according to any one of claims 5 or 6, characterized in that the adjustment slopes include oncoming counterclones interacting with oncoming slopes.  8. The clutch assembly according to claim 7, characterized in that the incident counterclones are likewise mounted on the annular part.  9. The clutch assembly according to any one of claims 5 to 8, characterized in that the parts having adjusting slopes are mounted with the possibility of displacement in the axial direction.  10. The clutch assembly according to any one of claims 6 to 9, characterized in that the parts carrying oncoming slopes and oncoming counterclones are mounted with the possibility of relative rotation relative to each other.  11. Clutch assembly according to any one of claims 6 to 10, characterized in that one of the parts having slopes and oncoming counterclones is mounted without the possibility of rotation relative to the friction clutch.  12. The clutch assembly according to any one of paragraphs.5 to 11, characterized in that the adjusting slopes have an upward angle, which causes self-braking under the action of friction on the adjusting slopes. 13. The clutch assembly according to any one of paragraphs.5 to 12, characterized in that the adjustment slopes have an ascending angle having values in the range from 5 to 20 ^o , preferably in the range from 7 to 11 ^o .  14. The clutch assembly according to any one of paragraphs.6 to 13, characterized in that at least one part carrying oncoming slopes and / or oncoming counterclones is spring loaded in the direction of adjustment.  15. The clutch assembly according to any one of claims 6 to 14, characterized in that the part having oncoming slope and / or part having oncoming counterclones is loaded or loaded in the direction of adjustment with at least one energy accumulator located between these parts in the form of a coil spring.  16. The clutch assembly according to any one of claims 1 to 15, characterized in that it is equipped with a device for restricting the movement of the trip.  17. The clutch assembly according to clause 16, wherein the device has at least one limit stop.  18. The clutch assembly according to item 16 or 17, characterized in that the restrictive device acts between the trip element and the clutch cover.  19. The clutch assembly according to any one of paragraphs.16 to 18, characterized in that the release element has a trip support, and the emphasis is mounted with the possibility of action between the support ring covering the friction clutch and the clutch cover.  20. The clutch assembly according to any one of paragraphs.16 to 19, characterized in that the restriction device is located between the guide tube for the trip element located on the transmission side and the trip element itself.  21. The clutch assembly according to any one of claims 1 to 20, characterized in that means are provided that cause, during the disengagement of the control means on the part of the trip, a gradual decrease in the moment transmitted by the friction clutch or clutch disc.  22. The clutch assembly according to any one of claims 1 to 21, characterized in that the clamping spring, made in the form of a Belleville spring, is mounted to rotate on the clutch housing, between two bearings, one of which, facing the pressure plate, is spring loaded in the direction pressure spring, and the force exerted by the cup spring when the friction clutch is disengaged to the spring support, increases when the pad is worn and exceeds the force acting on the spring support.  23. The clutch assembly according to item 22, wherein the spring-loaded bearing has the possibility of axial movement.  24. The clutch assembly according to claim 22 or 23, characterized in that with an increase in the disengaging force of the pressure plate spring, the spring-loaded support is displaced in the direction of the pressure plate.  25. The clutch assembly according to any one of paragraphs.22-24, characterized in that when the spring-loaded support is displaced, the decoupling force on the pressure plate spring decreases.  26. The clutch assembly according to any one of paragraphs.22 to 25, characterized in that the spring-loaded bearing is mounted with the ability to move so that a balance is established between the maximum disengaging force acting on the bearing and the counter-force exerted on this bearing.  27. The clutch assembly according to any one of paragraphs.22 to 26, characterized in that the pressure plate spring has a falling force characteristic on at least a portion of the trip stroke.  28. The clutch assembly according to any one of paragraphs.22 to 27, characterized in that the opposing force acting on the spring-loaded support is created by an energy accumulator having, basically, a constant force in the prescribed control range.  29. The clutch assembly according to any one of paragraphs.22 to 28, characterized in that the axially deflecting or shifting bearing is loaded with a disk spring, which serves as an energy accumulator.  30. A friction clutch, in particular for automobiles, comprising a pressure plate connected to the housing without the possibility of rotation, but with the possibility of limited axial displacement, and a Belleville spring is axially clamped between the housing and the pressure plate, creating a clamping force, mounted, on the one hand, with the possibility of rotation around the rotary support, and on the other hand with the possibility of loading in the direction of the clutch disc, sandwiched between it and the supporting disk, made in the form of a flywheel, characterized in that the pivot bearing is mounted on the housing with the possibility of axial displacement by means of an automatically acting adjusting device between the cover and the disk spring that compensates for at least the friction linings of the clutch disc and is moved further by the feed device, and the disk spring is under the action of the support force in the direction rotary support.  31. The friction clutch of claim 30, wherein the disk spring has a degressive characteristic in its working area.  32. The friction clutch according to claim 30 or 31, characterized in that the disk spring counteracts the disengagement force only by a force short circuit.  33. The friction clutch according to any one of paragraphs.30 to 32, characterized in that the disengagement force and the force of the Belleville spring are thus matched to each other, that the supporting force at a given mounting position of the Belleville spring and without changing the taper due to wear, and on the way the release of the cup spring is greater than the force that counteracts the support force generated by the cup spring when the taper of the cup spring due to wear changes, the support force on the part of the stroke of the cup spring is less, than the force exerted by the cup spring against the supporting force.  34. The friction clutch according to any one of paragraphs.30 to 33, characterized in that the supporting force is created by at least one energy accumulator in the form of a spring, which changes shape when the cup spring or support located on the side of the cover is worn due to wear.  35. The friction clutch according to any one of paragraphs.30 to 33, characterized in that the adjusting device is axially located between the cup spring and the cover.  36. The friction clutch according to any one of paragraphs.30 to 35, characterized in that the adjusting device has ramps in the form of slopes.  37. The friction clutch according to any one of paragraphs.30 to 36, characterized in that the supporting force is created by an element having the shape of a cup spring.  38. The friction clutch according to any one of paragraphs.30 to 37, characterized in that the disk spring, creating a supporting force, is located at the radial height of the axially displaced bearings on the disk spring.  39. The friction clutch according to any one of paragraphs.30 to 38, characterized in that the pressure plate spring is mounted to rotate between two bearings, one of which, facing the pressure spring, is spring-loaded in the direction of the pressure disk spring.  40. The friction clutch according to clause 39, wherein the bearing, spring-loaded with a supporting force, has the possibility of axial displacement.  41. The friction clutch according to any one of paragraphs.30 to 40, characterized in that when the spring-loaded support is displaced, the decoupling force of the pressure plate spring decreases.  42. The friction clutch according to any one of paragraphs.30 to 41, characterized in that the spring-loaded support is displaced so that an equilibrium is established between the disengaging force acting on the support and the back pressure on this support.  43. The friction clutch according to any one of paragraphs.30 to 42, characterized in that the counter-pressure force acting on the spring-loaded support is created by an energy accumulator having mainly a constant force in a predetermined control zone.  44. The friction clutch according to any one of paragraphs.30 to 43, characterized in that the energy accumulator, creating a supporting force, acts as a sensor.  45. The friction clutch according to any one of paragraphs.30 to 44, characterized in that the counter support provided on the side of the pressure disk spring opposite the spring support, can be displaced in the direction of the pressure plate, but can be fixed in the opposite direction.  46. The friction clutch according to any one of paragraphs.30 to 45, characterized in that a spring is provided as a feed device configured to move the adjusting device.  47. The friction clutch according to any one of paragraphs.30 to 46, characterized in that the adjusting device has an annular part axially loaded with a clamping cup spring in the engaged state of the friction clutch.  48. The friction clutch according to any one of paragraphs.30 to 47, characterized in that the adjusting device has adjusting slopes that rise in the axial direction.  49. The friction clutch according to claim 48, wherein the adjusting slopes are made on an annular part.  50. The friction clutch according to any one of paragraphs.47 to 49, characterized in that the annular part carries a counter-support.  51. The friction clutch according to any one of paragraphs.48-50, characterized in that the ramps interact with the corresponding or ramp counterclones.  52. The friction clutch according to Claim 51, wherein the incident counterclones are located on an annular part that is located between the component bearing the incident slopes and the cover.  53. The friction clutch according to paragraph 52, wherein the oncoming counterclones are located directly in the radially passing zone of the housing.  54. The friction clutch according to any one of paragraphs.30 to 53, characterized in that the adjusting device, when viewed in the direction of disengagement of the friction clutch, operates with free travel, and in the direction opposite to the direction of disengagement, is self-braking. 55. The friction clutch according to any one of paragraphs.48 to 54, characterized in that at least the incident slopes have an increasing angle having a value in the range from 4 to 20 ^o , preferably from 5 to 12 ^o .  56. The friction clutch according to any one of paragraphs.48 to 55, characterized in that the ramps have an angle of elevation that provides self-braking due to the friction force of the ramps with the zones of support of another part.  57. The friction clutch according to any one of claims 48 to 56, characterized in that at least one part bearing oncoming slopes and / or one part bearing oncoming counterclones or approach zones is spring loaded.  58. The friction clutch according to any one of paragraphs.30 to 57, characterized in that the adjusting device has several offset adjusting elements.  59. The friction clutch according to any one of paragraphs.30 to 58, characterized in that the adjusting device is dependent on the speed.  60. The friction clutch according to any one of paragraphs.30 to 59, characterized in that the adjusting device is blocked depending on the speed.  61. The friction clutch according to any one of paragraphs.30-60, characterized in that the adjusting device is blocked at a speed above a certain limit.  62. The friction clutch according to any one of paragraphs.30 to 61, characterized in that the adjusting device operates at idle speed or at a speed below idle speed.  63. The friction clutch according to any one of paragraphs.30 to 62, characterized in that the adjusting device is activated at a speed of almost zero.  64. The friction clutch according to any one of paragraphs.30 to 63, characterized in that the details of the adjusting device having oncoming slopes and / or oncoming counterclones, or areas, and displaced relative to the housing, are spring loaded.  65. The friction clutch according to item 64, wherein the spring is carried out by a force applied in the circumferential direction.  66. The friction clutch according to any one of paragraphs.30 to 65, characterized in that the sensor spring acting on the counter force relies on its housing with its radial outer region.  67. The friction clutch according to any one of paragraphs.30 to 66, characterized in that on the housing are made supporting sections for the sensor spring, creating a counter force.  68. The friction clutch according to any one of paragraphs.30 to 67, characterized in that between the friction linings of the friction clutch there is a spring or means for ensuring the spring of the lining.  69. The friction clutch according to any one of paragraphs.30 to 68, characterized in that the spring provided between the clutch plate linings has a path-force characteristic close to the path-force characteristic of the force exerted by the pressure plate in the course of the lining spring spring on the pressure plate.  70. The friction clutch according to any one of paragraphs.30 to 69, characterized in that in the disengaged state of the friction clutch, the force required to control the pressure plate disk or friction clutch is of the order of 150 to 150 Nm.  71. The friction clutch according to any one of paragraphs.30 to 70, characterized in that after releasing the clutch disc by means of the support disk, the pressure plate spring moves from a positive force-path characteristic to a negative force-path characteristic.  72. The friction clutch with a pressure plate connected to the housing without the possibility of rotation, however, with the possibility of limited movement in the axial direction, while at least one pressure disk spring in the mounted state of the friction clutch on the support disk loads the pressure plate in the direction of the clamped between it and the clutch disc support disk, and an adjustment device (16) is provided that automatically compensates for at least the friction linings of the clutch disc, which under holds almost constant the force exerted by the pressure plate spring on the pressure plate (3), and the friction clutch contains control means (4b) for disengaging and engaging, characterized in that the control means are driven by a releasing means (22), while means are provided ( 2a + 36,436) to limit the trip of the control means (4b, 404b).  73. Friction clutch with a pressure plate connected to the housing without rotation, but with limited movement in the axial direction, while at least one pressure disk spring in the mounted state of the friction clutch on the support disk loads the pressure plate in the direction of the clamped between supporting disk of the clutch disc, the disk spring (404) is mounted rotatably on the housing (402) and has an annular housing (404a), from which the control tabs (404b) radially extend inward, schayasya in that the friction clutch has a stop (436) limiting the stroke of the friction clutch and disengaged when operating region radial placement control latch (404b) or at least in the boundary region.  74. Friction clutch with a pressure plate, which can be connected to the support disk without the possibility of rotation, but with the possibility of limited movement in the axial direction, while at least one pressure disk spring loads the pressure plate in the direction of the clutch plate clamped between it and the support disk at the same time, an adjustment device (16) is provided that automatically compensates for at least the wear of the friction linings of the clutch disc, which maintains an almost constant force provided by the pressure plate spring onto the pressure plate (3), and the friction clutch comprises control means (4b) for disengaging and engaging, characterized in that the control means are driven by the release lever (22) controlled by the releasing means (24), depending on at least the wear of the friction linings of the control means (4b), they move axially in the direction of movement of the trip (IV), and the cup spring (4) has a spring characteristic (40 in FIG. 24) with a degressive force change cter with a minimum (45) of the region and with the region of increase in force adjacent to it, the pressure plate (3) when the clutch disengages inside this region of degressive force change in the minimum direction (45) releases the clutch disc and the adjusting device (16) contains sensitive springs (13) for axial emphasis and pivoting (5) of the disk spring (4) on the housing (2), characterized in that on the path of movement of the clutch release means, such as control means (4b), the release lever (22), medium VA (24) or the like, locking means (2a + 36,436) are provided that, after passing a minimum (45), prevent a subsequent increase in force beyond that value (46 in FIG. 24) that would cause the bending of the sensitive spring (13).  75. The friction clutch according to claim 74, wherein the restriction or abutment means prevents unacceptably high load and / or deformation of the control means (4b).  76. Friction clutch according to one of paragraphs. 74 and 75, characterized in that the means of restriction or emphasis (2A, 436) are provided at least on the housing (2, 402).  77. The friction clutch according to one of claims 74 to 76, characterized in that the restriction means have at least one limit stop fixed to the housing (436).  78. Friction clutch according to one of paragraphs. 74 - 77, characterized in that the means of restriction or stop (2A + 36) act between at least one part (34) of the switch lever and the clutch housing (2).  79. The friction clutch according to one of paragraphs. 74 - 78, characterized in that the displacement of the disengagement is limited by the stop of the control means (4b) in the limit stop fixed on the housing (436).  80. The friction clutch according to one of paragraphs. 74 - 79, characterized in that the pressure plate spring has an annular main part (4A), from which the tongues (4b) radially outward form the control means.  81. The friction clutch according to one of paragraphs. 74 - 80, characterized in that the restrictive means (436) limit the angular rotation of the pressure plate spring.  82. The friction clutch according to one of claims 74 to 81, characterized in that the restrictive means (436) are integrated into the friction clutch.  83. Friction clutch according to one of paragraphs. 74 - 82, characterized in that the restriction of the trip occurs by stopping the radially internal ends of the tongues of the Belleville spring (404c) in an emphasis fixed to the housing (436).  84. The friction clutch according to one of claims 74 to 83, characterized in that the limit stop (436) fixed to the housing is provided axially between the disc spring tongues and the support disk.  85. The friction clutch according to one of claims 74 to 84, characterized in that the restrictive stop fixed to the housing is formed by an annular support region (436) that extends through radially extending jumpers (437) and is connected to the housing.  86. The friction clutch as claimed in claim 85, characterized in that the jumpers (437) extend radially inward and axially inclined in the direction of the support disk (403) from the housing and pass through the slots (439) provided between the cup spring tongues.  87. The friction clutch according to one of paragraphs. 74 - 86, characterized in that it is installed with the ability to be actuated by a trip system acting on the control means, while means are provided (662, 664, 666, 667) to limit the effect on the control means (604) release forces.  88. The friction clutch according to one of paragraphs. 74 - 87, characterized in that between the friction linings of the clutch disc is provided spring-loaded linings (10).  89. The friction clutch according to one of paragraphs. 74 - 88, characterized in that it is made in the form of a pressure clutch.  90. Friction clutch with a pressure plate connected to the housing without the possibility of rotation, but limited to move axially, while at least one pressure spring in the mounted state of the friction clutch on the support disk loads the pressure plate in the direction of the disk clamped between it and the support disk clutch, while the friction clutch has control means (604) for disengagement and clutch, characterized in that the control means are actuated by the acting on them system Lenia, while in the disengagement system provides the means (662, 664, 666, 667) to limit the impact on the control means control force.  91. The friction clutch according to Claim 90, wherein the control force is limited by means provided on the transmission path of the decoupling system and retreating when a certain value of the control force is exceeded.  92. The friction clutch according to claim 91, characterized in that the retreating means is springy, respectively elastic.  93. The friction clutch according to one of claims 90 to 92, characterized in that the disengagement system has a working fluid circuit in which a valve limiting the maximum disengaging force is provided (662, 664, 666, 667).  94. The friction clutch according to one of claims 90 to 93, characterized in that the means limiting the maximum disengaging force is provided on the transmission path of the force between the clutch pedal or the control motor and the friction clutch controls, such as plate spring tabs (4b).  95. The friction clutch according to one of claims 90 to 94, characterized in that the trip system has a working fluid circuit in which a means limiting the maximum pressure is provided (662, 664).  96. The friction clutch according to one of claims 90 to 95, characterized in that the means provided for limiting the force acting on the control means (4b) of the friction clutch specify a maximum force that is at least somewhat greater than the maximum force required for clutch control.  97. The friction clutch according to one of claims 90 to 96, characterized in that the force limiting means are formed by a buffer, for example, a hydro- or spring battery.  98. A friction clutch comprising a control device with a drive motor, in particular for automobiles, while for engaging and disengaging the friction clutch between it and the drive motor, a control rod is provided, characterized in that a safety device is provided to the drive motor, by which, if exceeded, of a predetermined control force or a predetermined moment of the drive motor is limited to a predetermined value, the rotation moment of the drive motor tor, respectively, the driving force.  99. A friction clutch, in particular for cars with a pressure plate connected to the housing without turning, but limited to axially limited movement, a pressure cup spring is provided between the housing and the pressure plate, which on one side can rotate around a housing fixed on the body pivot bearings formed by pivot bearings located on both sides of the pressure disk spring and, on the other hand, in the mounted state of the friction clutch on the support disk loads the pressure plate in the direction of the clutch plate clamped between it and the support disk, such as a flywheel, characterized in that the pivot bearing located on the side of the housing (412) can axially move under the action of an automatic adjusting device (416) to compensate for at least wear the friction linings of the clutch disc, and the pressure cup disk (404) is under the action of a support force directed in the direction of the rotary support (412) located on the side of the housing, the magnitude of which changes depending on the speed.  100. Friction clutch according to claim 99, characterized in that the adjusting device (416) acts between the cover (402) and the cup spring (404).  101. Friction clutch according to one of paragraphs.99 and 100, characterized in that the supporting force loads the pivot bearing (411) provided on the side of the pressure disk spring, turned away from the housing.  102. The friction clutch according to one of claims 99 to 101, characterized in that the supporting force depends on the centrifugal force.  103. Friction clutch according to one of claims 99 to 102, characterized in that the pressure disk spring (404) is supported by a force only by a force closure.  104. The friction clutch according to one of claims 99 to 103, characterized in that the support force is generated by at least one disk-spring energy accumulator, which, when wear is caused by the movement of the pressure disk spring (404), respectively, of the support (412) with side of the lid changes shape.  105. The friction clutch according to one of paragraphs.99-104, characterized in that the supporting force increases with increasing speed.  106. The friction clutch according to one of claims 99 to 105, characterized in that there are means (450) depending on the centrifugal force, which create a supporting force that acts parallel to the main supporting force created by the disk-spring element (413).  107. The friction clutch as claimed in claim 106, characterized in that the centrifugal force-dependent means (450) make it possible to carry out wear-dependent control only when the friction clutch is stopped (401) or at low speeds.  108. The friction clutch according to one of claims 99 to 107, characterized in that the adjusting function of the adjusting device (416) is practically blocked when a certain speed is exceeded.  109. The friction clutch according to one of claims 99 to 108, characterized in that the means depending on the centrifugal force are formed by tongues (450), which are mounted on a disk-spring element that creates the main supporting force.  110. The friction clutch according to claim 109, characterized in that the tongues (450) form a single unit with a disk-spring (413) element.  111. A friction clutch with a pressure plate connected to the housing without the possibility of rotation, but limited in axial direction, with at least one pressure spring, for example a disk spring, acting between the housing and the pressure plate, and an adjustment device is provided (216 ), automatically compensating for at least the wear of the friction linings of the clutch disc, which in the unmounted condition of the friction clutch on the support disc is in the initial position and which further, during the life of the clutch plate, it provides an almost constant force on the pressure plate from the side of the pressure plate, and the friction clutch has controls (204) for disengagement and clutch, characterized in that the adjusting device (216) in the unmounted state of the friction clutch (201 ) is protected from unacceptable (unforeseen) displacement from its original position.  112. The friction clutch according to Claim 111, wherein the protection of the adjusting device (216) is removed at least after mounting the friction clutch (201) on the support disk.  113. The friction clutch according to one of paragraphs 111 or 112, characterized in that the adjusting device (216) has a rotary ring (217), which is protected by means of protection (261) from turning, at least in the unmounted condition of the friction clutch ( 201).  114. The friction clutch according to one of paragraphs 111 to 113, characterized in that an adjusting ring (717) is provided axially between the pressure plate spring (704) and the housing (702), as well as the radial slopes (773), in which the pressure plate rests spring (704) when exceeding a certain taper.  115. The friction clutch according to one of paragraphs 111 to 114, characterized in that the means (773) that prevent the adjusting device from moving in the unmounted condition of the friction clutch (701) in an unmounted friction clutch, also when exceeding a certain trip distance (772) control means (704c) block the adjusting device (716).  116. Friction clutch according to p. 115, characterized in that the locking of the adjusting device (716) occurs due to the fit provided on the pressure plate disk spring (704) to the areas (770) of the adjusting ring (717) of the adjusting device (716).  117. Friction clutch according to p. 116, characterized in that the outer lateral region of the pressure cup spring (704), when the pre-set taper is exceeded, rests on the radial slopes (773) of the adjusting ring (717).

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