RU2128792C1 - Friction clutch; automobile with friction clutch; clutch (versions); clutch for internal combustion engine; unit of drive - Google Patents

Abstract

FIELD: mechanical engineering. SUBSTANCE: friction clutch has housing, friction, pressure and cover disks with pressure spring mounted between these disks and housing. Clutch is provided with engaging and release devices, adjusting device for compensation for wear of friction disk and for constant (as far as possible) pressing of cover disk by pressure spring. Automobile equipped with such clutch is provided with pedal-operated clutch release system. According to first version, clutch unit includes friction clutch and spring device for gradually decreasing the torque transmitted by clutch at one section of clutch release stroke. According to second version, clutch unit includes torsional damper; pressure disk is just one part of fly-wheel; fly-wheel consists of two parts which are turnable relative to each other overcoming action of torsional damper. Clutch unit for internal combustion engine includes friction clutch connected with output shaft of internal combustion engine through axially flexible member. Unit of drive includes gear-box and friction clutch mounted between drive engine and gear-box. EFFECT: reduction of force required for operation of clutch; constant force profile in clutch release force distribution; simplified construction; low cost. 24 cl, 39 dwg

Description

 The invention relates to a friction clutch, in particular for automobiles, with a pressure disk (pressure plate), which is without rotation, but with the possibility of limited axial displacement is connected to the body, and at least one compressible pressure spring acts between the body and the pressure plate, which loads the pressure spring the plate in the direction clamped between the last and the pressure plate (back pressure plate), like a flywheel, the clutch disc.

 Such couplings have become known, for example, from German Patent Application Laid-Open No. 2460963, German Patent No. 2441141 and 898531, as well as German Published Application No. 1267916.

 The basis of this invention is the task of improving such friction clutches with respect to functioning and service life. In particular, the forces necessary for the operation of this kind of friction clutches must be lowered and a virtually unchanged diagram of the shutdown forces throughout the entire service life must be guaranteed. The task is also to simplify the manufacture and reduce its cost.

 According to the invention, this is achieved by the fact that there is an automatically adjusting device that compensates for the wear of the friction plates of the clutch disc, which produces practically unchanged force loading of the pressure plate by means of the pressure spring, and the friction clutch has actuating means for coupling and disengaging, as well as a safety device that is at least on the part of the working path of the actuating means and / or the distribution path of the pressure plate, a gradual decrease in the moment and transmitted by a friction clutch respectively. clutch disc. By means of such safety devices, it is equally achieved that during the start-up of the friction clutch and at the beginning of the clamping of the friction plates between the pressure plate and the back pressure plate, a gradual acc. progressive growth of the moment transmitted by the friction clutch.

 By means of the friction clutch defined by the invention, it is ensured that the pressure cup spring, taking into account the friction clutch service life, almost always has the same pre-tension when the friction clutch is engaged, and thus a practically unchanged force load of the pressure plate is set. In addition, by means of an additional safety device that gradually reduces the moment transmitted by the friction clutch during the tripping process, a decrease or minimization of the course of the tripping force curve or the maximum necessary tripping force can be achieved. This should be explained by the fact that the safety device supports the work, in particular the process of disengagement of the friction clutch. For this, the safety device may have axially spring-loaded elastic means that exert a reaction force on the actuating means and / or on the pressure plate and / or on the back pressure plate, which is directed opposite to the pressure spring exerted on the pressure plate and is connected in series.

 Particularly preferred is the arrangement of the friction clutch safety device in such a way that during the disengagement process on a partial portion of the axial displacement path of the pressure plate-loaded areas of the pressure plate, the friction clutch transmitted gradually increases, respectively. clutch disc moment.

 For some applications, a safety device may advantageously be provided in the power flow between the rotary support of the actuating means or between the hold-down spring and attachment points, such as threaded joints, housings on the back pressure plate.

 However, for other applications, it may be preferable that the safety device is provided in the power flow between the rotary support of the working means or between the clamp spring and the friction surface of the pressure plate. This kind of arrangement was proposed, for example, laid out by the application of Germany N 3742354 and laid out by the application of Germany N 1450201.

 For other applications, it may be particularly preferable when the safety device is provided axially between the two rear sides of the clutch disc friction plates located on the rear sides, and is therefore formed by the so-called “clutch spring”, for example by means of the clutch spring segments provided between the claddings. Such safety devices have become known, for example, by way of the laid-out application of Germany N 3631863.

 A further opportunity to achieve progressive growth or to reduce the moment was proposed by the German application N 2164297, in which the flywheel is made up of two parts, and the axial spring-loaded part forming the back pressure plate is supported with respect to the structural part connected to the output shaft of the internal combustion engine.

 For the functioning and construction of the friction clutch defined by the invention, it may be especially appropriate when the adjusting safety device allows axial, springy elasticity between the clutch parts, the safety device being located and designed so that when the clutch is open, the force acting on the safety device is very small , and in the process of closing the coupling, therefore, along the clutch path (inclusion) of the coupling, the force acting on the safety device gradually rises to a maximum, and this rise is advisable only part way circuit or path of the actuating means or the clutch pressure plate. It is particularly preferred that the safety device is designed in such a way that a gradual decrease (resp. Gradual increase) of the moment transmitted by the friction clutch occurs at least about 40-70% of the working path of the actuating means and / or the maximum axial path of the pressure plate. The residual portion of the corresponding path is necessary for perfect separation of the power flow and for equalizing, possibly, existing deformations of the coupling parts, in particular, such as the clutch disc, pressure plate, and also the back pressure plate.

 In order to minimize the forces required for the operation of the friction clutch of the invention, it is particularly preferred that the compression spring, at least on one part of the friction clutch disengagement path, possesses a load-path digressive characteristic, which therefore means that the spring the clamp, at least on one part of its compression or deformation stroke, has a falling force characteristic. Thus, it is achieved that during the process of disengaging the friction clutch, the force of the spring of the safety device counteracts the force of the clamping spring, so that compression or deformation of the clamping spring is supported by the springing force of the safety device on the part of the trip, and at the same time, due to the digressive or falling characteristics of the dependence force - the path of the clamping spring; the force exerted by the latter on the clamping plate or on the friction section decreases adku. The force curve that is effectively necessary for disengaging the friction clutch is obtained, since there are no additional overlapping spring actions from the difference between the force characteristic applied by the safety device and the force characteristic of the pressure spring. When lifting the pressure plate from the friction plates or releasing the friction disk by means of the pressure plate, the necessary remaining characteristic of the disengagement force or the necessary disengagement force is determined mainly by the pressure spring. The dependence of the force - the path of the safety device and the dependence of the force - the path of the clamping spring can thus be adjusted to each other, so that when the clutch plate is opened by the clamping plate, the force required to operate the clamping spring is at a relatively low level. Thus, by approximating or equating the characteristics of the spring or the characteristics of the force of the safety device to the characteristics of the clamping spring, only a very small amount of force may be necessary before the clutch plate is disconnected by the clamping plate, and in the extreme case, the response force of the clamping spring is practically not needed.

 A disk spring is particularly suitable as a compression spring, which, on the one hand, can be swiveled around the ring-shaped rotary support around the weight housing and, on the other hand, loads the pressure plate. In this case, the disk spring may have an annular housing, from which radially inwardly directed reeds form the actuating means. Actuating means can also be formed by levers, which, for example, are pivotally mounted on the housing. The pressure force of the pressure plate can, however, be applied by other types of springs, such as coil springs, which are thus located in the friction clutch, so that the axial force exerted by them on the pressure plate in the locked state is greatest, and this force decreases during the disengagement process . This can be done, for example, by tilting the position of the coil springs with respect to the axis of rotation of the friction clutch.

 It is particularly preferred that the disc spring is deflectably supported on the housing between two linings to form a so-called clutch of a compressed structure. With such couplings, the actuating means for turning off the friction clutch are usually loaded in the direction of the pressure plate. However, the invention is not limited to clutches of a compressed design, but also encompasses couplings of a stretched structure in which the controls are loaded to turn off the friction clutch, usually away from the pressure plate.

 Particularly preferred for the friction clutch defined by the invention is a disk spring, which is designed in such a way that it has a sinusoidal force-path dependence and which is set in such a way that in the compressed state of the friction clutch its operating point is provided on the digressive section of the characteristic following the first power maximum . Moreover, it is preferable when the Belleville spring has a force ratio of 1: 0.4 - 1: 0.7 between the first power maximum and the subsequent minimum.

 In addition, it is preferable when the friction clutch is driven through a shut-off system acting on actuating means, such as, for example, on the tops of a disc spring, and the shut-off system may have a clutch pedal made similar to an accelerator pedal and located in a car chamber. Such a clutch pedal is particularly preferred since, due to the embodiment defined by the invention, the force necessary for turning off the friction clutch or the power characteristic can reach a very low level, so that better adjustment of the work force is possible through a clutch pedal similar to the accelerator pedal.

 By the implementation of the friction clutch defined by the invention and by the associated possibility of reducing the pressure spring forces that appear as much as possible during the life of the friction clutch, the parts can correspondingly decrease or decrease in their strength, as a result of which a significant reduction in manufacturing cost can occur. In addition, by reducing the disengagement forces, friction and elasticity losses in the clutch and the shutdown system are reduced, and thereby the efficiency of the friction clutch / shutdown system is significantly improved. At the same time, the entire system can be optimally designed and thereby significantly improve grip comfort.

 The claimed device is applicable generally to friction clutches and, in particular, to those proposed, for example, by the German Federal Patent No. 2916755, the German Federal Patent No. 2920932, the German Federal Application N 3518781, the German Federal Application N 4092382, the French Patent Application N 2605692, the international patent application France N 2606477, laid out by the application of France N 2599444, laid out by the application of France N 1599446, UK patent N 1567019, US patent N 4924991, US patent N 4191285, US patent N 4057131, IP-CM 3-123, IP-CM 2-124326, IP-CM 1-163218, IP-03 51-126452, IP-CM 3-19131, IP-CM 3-53628.

 The invention also refers to earlier applications of Germany P4207528.9 and Germany P4206904.1, the contents of which are explicitly related to the content of the disclosure of this invention.

 The use of a friction clutch with automatic or automatic alignment of at least the wear of the plates - thereby guaranteeing an almost unchanged at least for the life of the friction clutch clamping force of the clutch disc - is particularly preferable in connection with coupling units in which the friction clutch, clutch disc and a back pressure plate, such as in a flywheel, forms a mounting unit or module. In this kind of mounting unit, for economic reasons, it is preferable when the coupling body is connected to the back pressure plate through an integral connection, such as, for example, a welded connection or a geometrical connection, for example, by plastic deformation of the material. With this type of connection, commonly used fasteners, such as bolts, can be removed. In such assemblies, it is practically impossible to replace the clutch disc or clutch plates due to exceeding the wear limits without destroying the structural parts, such as in the clutch housing. By using a wear adjusting clutch, the mounting unit can be designed in such a way that it guarantees smooth operation over the entire life of the vehicle. Thus, based on the embodiment defined by the invention, the amount of the reserve for wear of the clutch disc and the reserve for adjusting the friction clutch or the coupling module can be selected such that the service life of the clutch and thereby the service life of the mounting unit reliably achieve at least the service life of the vehicle.

 According to a further development of the invention, it may be particularly preferable when the friction clutch having a wear-adjusting safety device is combined with the so-called two-mass flywheel, the friction clutch being mounted on one flywheel connected to the transmission, while the second flywheel is connected to the output shaft internal combustion engine. Two-mass flywheels in which the friction clutch defined by the invention can find application have become known, for example, from the laid-out applications of Germany 3721712, 3721711, 4117571, 4117582 and 4117579. The entire contents of these applications relate to the contents of the disclosure of this invention, so that described in these applications features in any way can be combined with features described in this invention. In particular, the coupling body or the coupling cover through a connection that is not detachable without destruction can be connected to the flywheel supporting it, as described and shown, for example, for various embodiments in German application laid out 4117579.

 By using a friction clutch with a safety device that compensates for at least wear of the plates, furthermore, optimization of the parameters of the friction clutch, in particular creating a clamping force for the clutch disc of the force accumulator, can be carried out. This power accumulator can therefore be carried out in such a way that it creates practically only the clamping force of the clutch disc, necessary for transmitting the desired torque. The power accumulator may be formed by at least one disk spring or a plurality of coil springs. In addition, it is preferable to use one self-regulating friction clutch in conjunction with two-mass flywheel wheels, in which a torsionally elastic damper located between both flywheel masses is provided radially outside the clutch disc or the outer diameter of the friction surface of the flywheel connected to the gear. With this kind of dual-mass flywheel, the rubbing diameter of the clutch disc should be less than with conventional couplings, so that the clamping force should increase in accordance with the ratio of the average friction radii in order to transmit a certain engine torque. When using a conditional clutch, this would lead to an increase in disengagement force. However, by using a wear adjusting clutch with a progressive reduction in the torque transmitted by the clutch disc according to paragraph 1 of the claims, lowering of the tripping force can be achieved, as a result of which the increase of the tripping force can be eliminated, or by decreasing the friction clutch, a reduction of the tripping force with respect to the conditional can be achieved coupling.

 Thus, by the friction clutch as defined by the invention, it can be ensured that, despite the reduced outer diameter of the friction plates and, as a consequence, the necessary high clamping force, the disengaging force can remain low. Due to the lower decoupling force, the load on the rolling bearing is also reduced, through which both flywheel masses are rotated relative to each other. In addition, by controlling wear, the service life is increased, so that replacement of parts, in particular the clutch disc, is no longer necessary during the life of the vehicle. Thus, the coupling cover can be rigidly connected to the flywheel connected to the transmission, for example, by riveting or welding. This is particularly advantageous when there is limited installation space or limited contours of the coupling cap that no longer allow coupling of the coupling cover to the flywheel located on the gearbox side of the gearbox in the usual manner by screwing.

 With a friction clutch with an integrated adjusting safety device for wear of the plates when the coupling unit consisting of a friction clutch and a flywheel is conditionally mounted, axial, rotational and oscillatory vibrations that are excited by the output shaft of the internal combustion engine are transmitted to the coupling unit as, in particular, the main shaft. In order to prevent the clutch unit or the adjusting safety device from operating such vibrations and, in particular, to suppress the undesirable regulation of the wear-compensating safety device, the inertia forces of those structural parts that act on this safety device must be taken into account when performing the adjusting safety device. In order not to cause these unwanted side effects caused in particular by axial and oscillating vibrations or the higher costs associated with them for calculating the parameters of the adjusting safety device to compensate for wear of the plates, according to another idea of the invention, the clutch assembly containing the adjusting safety device is largely disconnected relative to axial and bending vibrations excited by the output shaft of the internal combustion engine. As a result of this, it can happen that the coupling unit is connected to the output shaft of the internal combustion engine through an axially elastic or spring-pliable structural part. The rigidity of this structural part is selected in such a way that the axial and oscillating or bending vibrations generated by the driven shaft of the internal combustion engine on the clutch assembly are quenched or suppressed at least to such an extent that the friction clutch is guaranteed to function perfectly, in particular its adjusting safety device. Such resilient structural parts have become known, for example, from European Patent Laid-open No. 0385752 and 0464997, as well as from CAE Technical Paper 900391. The content of these publications should also relate to the content of the disclosure of this invention. By using one resilient structural part, it is possible to eliminate unwanted wear control caused by axial vibrations of the pressure plate relative to the clutch cover - in particular when the friction clutch is off - by means of flywheel vibrations and / or Belleville springs. Such oscillations in coupling units or coupling assemblies without suppressing these oscillations, at least mainly by a safety device, in particular, such as an axial-elastic disk, can lead to variable adjustment regardless of the state of wear of the clutch plate, and the friction disk spring the clutch could be adjusted in the force of pressure down to the power minimum, as a result of which the transmission of the desired moment would no longer be guaranteed.

 According to another embodiment of the invention, a friction clutch with automatic or automatic compensation, which (the clutch) in particular can be made in accordance with this invention, can advantageously find application in the drive unit, which consists of an automatic or semi-automatic transmission and located between the drive motor type of internal combustion engine, and a transmission driven with control or regulation at least depending on the operation of the transmission of the friction gear you. The friction clutch is driven mainly fully automatically. Automated or fully automated control of the friction clutch was proposed, for example, laid out by the application of Germany 4011850.9, so that in relation to the principle of action and the necessary funds refer to this printed material.

 With the previously known drive units with automatic or semi-automatic transmission and conditional friction clutch, until now, there were significant problems in actuating the clutch and in choosing the parameters of the necessary functional elements, such as, for example, cylinder-piston units and / or electric motors. Due to the relatively high shutoff forces required for conditional couplings, actuators (functional elements) with very strong or large parameters are required. This means large structural volume, high weight and high costs. Also, the actuators thus calculated due to their mass inertia 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, laid out by the application of Germany 3309427, for example, it is proposed to reduce the control force to turn off the clutch by means of corresponding compensation springs, so that actuators with smaller dimensions can be used. Since the disengagement force with conditional couplings still fluctuates very strongly during the service life, i.e. the decoupling force in the new state is relatively insignificant and increases over the service life with increasing wear of the plates, through one compensation spring only one part of the usually required decoupling force can be reduced. Taking into account all the tolerances, despite the use of compensation springs, the breaking power of the actuators will be needed, which is more than the breaking power of the new conditional clutch. Through the use of one inventive friction clutch with compensation for wear of the plates in connection with one drive unit consisting of one engine and one automatic or semi-automatic transmission, the breaking force can be significantly reduced in relation to the previously mentioned prior art, namely, in the clutch, and this the magnitude of the breaking force or the characteristic of the breaking force of the new coupling remains virtually unchanged over the entire service life of the latter. As a result of this, significant advantages are obtained for calculating the parameters of the actuators, 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, losses occurring in the switching system due to friction or elasticity of structural parts are eliminated or reduced to a minimum.

 In addition, from the patent DE 2460963A1, a friction clutch is known comprising a pressure disk, a friction disk, a pressure disk, a pressure spring, a housing, wherein the pressure disk is not rotatable, with the possibility of limited axial displacement, connected to the housing, and at least between the housing and the pressure disk at least one compression spring that presses the pressure plate in the direction of the friction disk pressed between it and the pressure disk, and the pressure disk is a flywheel.

 A coupling assembly is known from the same document, comprising a housing, a pressure disk, a friction disk, a pressure disk, a pressure spring, and the pressure disk is inverted, with the possibility of limited axial displacement, connected to the pressure disk, at least one pressure spring mounted on the housing and presses the pressure plate against the direction of the friction disk pressed between it and the pressure disk.

 Compared to this source, the friction clutch according to the invention is characterized in that it is equipped with an adjustment device to compensate for the wear of the friction linings of the friction disk and to ensure at least approximately constant pressure of the pressure disk by the spring, while the friction clutch is equipped with actuating means for engaging and disengaging the clutch, and also a safety device, which is designed to provide, during the release of the coupling, on one of the sections of the actuator’s working stroke means of gradual reduction of the moment transmitted by the friction clutch.

 In this case, the safety device is designed to ensure during the process of disengagement of the clutch on one of the axial displacement sections of the pressure disk-loaded sections of the pressure plate to gradually reduce the moment transmitted by the friction clutch.

 The safety device (10) may be located between the actuating means (46) or the clamping spring (4) and the sections for mounting the housing on the clamping disk.

 The safety device (10) can also be located between the actuating means or the pressure spring and the friction surface of the pressure plate, as well as between the friction plates of the coupling disk in the axial direction.

 In this case, the safety device provides the possibility of axial displacement of the mentioned structural elements of the coupling, the safety device being located in such a way that when the coupling is disconnected, the force acting on the safety device is the smallest, and when the coupling is turned on, the force acting on the safety device gradually rises to maximum, and this rise occurs in at least one section of the working stroke of the inclusion of the coupling.

 The safety device has the ability to gradually decrease or gradually increase the moment transmitted by the friction clutch (1) by at least 40-70% of the working stroke of the actuating means.

 According to a preferred embodiment, the hold-down spring on at least one part of the friction clutch disengaging stroke has a regressive elastic characteristic.

 In this case, the compression spring is a disk spring, which is mounted with the possibility of rotation around an annular rotary support mounted on the housing, and is supported on one side by a pressure disk.

 According to a suitable embodiment, the disk spring has an annular part, from which radially inwardly directed protrusions form the actuating means, the disk spring being mounted rotatably between two supports on the housing.

 According to another embodiment, the Belleville spring has a sinusoidal elastic characteristic and when the friction clutch is engaged, the elastic characteristic of the Belleville spring has a working point at the next maximum regressive section of the elastic characteristic, the ratio between the first maximum and the minimum following it is 1: 04- 1: 0.7.

 A friction clutch for a car, including a clutch disengagement system designed to control a clutch, comprising a clutch pedal similar to an accelerator pedal, is characterized in that one of the elements of the disengagement system engages with actuating means of the friction clutch.

 The coupling assembly according to the invention differs from that mentioned in DE 2460963A1 in that the coupling assembly is equipped with an adjusting safety device to compensate for the wear of the friction plates of the friction disk and to ensure at least approximately constant pressing of the pressure plate by the pressure spring, the coupling unit is provided with actuating means for coupling and disengaging the coupling unit as well as a safety device with at least one spring medium located in series with the compression spring tion to effect during the disengagement process of the coupling unit on one of the sections of the working stroke actuating means constant decrease the torque transmitted through the friction disc.

 According to a second embodiment of a coupling assembly comprising a housing, a pressure disk, a friction disk, a pressure disk, a pressure spring, a torsional vibration damper, wherein the pressure disk is inversely connected to the pressure disk with a possibility of axial displacement, the pressure spring presses the pressure disk in the direction to the friction disk clamped between it and the pressure disk, and also equipped with an adjusting safety device to compensate for the wear of the friction plates of the friction disk and providing at least approximately constant pressing of the pressure plate by the pressure spring, characterized in that the coupling unit is equipped with actuating means for coupling and uncoupling the coupling unit and a safety device that is able to gradually realize during coupling disengagement of the coupling unit reduction of the moment transmitted through the friction disk, in addition, the clamping disk is the first part of the flywheel, two parts of the cat cerned are rotatable relative to each other against the action of the vibration damper mounted between two parts of the flywheel, the second portion of the flywheel is coupled to the output shaft of the internal combustion engine.

 It is advisable that the housing is connected to the part of the flywheel formed by the pressure disk through an indestructible detachable connection, and the torsional vibration damper is located radially outside the outer friction diameter of the clutch friction disk.

 According to yet another embodiment, a coupling assembly for an internal combustion engine comprising a friction clutch including a pressure plate that is inversely rotatable with limited axial displacement and is connected to a pressure disk and at least one pressure spring that presses the pressure disk against the clampable between it and the pressure disk, which serves as a flywheel, the friction disk, characterized in that the friction clutch is equipped as an adjustment device to compensate for wear of the friction the clutch friction disk plates and providing at least approximately constant pressure of the pressure plate by means of the pressure spring, the coupling is also equipped with actuating means for engaging and disengaging the clutch and a safety device that is able to disengage the clutch in one of the sections of the actuator working stroke carry out a gradual decrease in the moment transmitted through the friction clutch, the friction clutch being connected through an axially elastic element to the outlet the internal combustion engine shaft, while the rigidity of the part is selected so that the axial, oscillatory or bending vibrations excited by the output shaft of the engine on the friction clutch are damped or canceled by the said elastic element to a level that guarantees the best functioning of the friction clutch, in particular its adjusting device .

In this case, the stiffness of the axial-elastic element is chosen such that the control forces necessary for disengaging the coupling without significant axial displacement of the coupling unit are transmitted by the elastic element;
the safety device is provided with spring means, which is arranged in series with the pressure spring;
and the pressure disk is connected to the output shaft of the internal combustion engine through a torsional vibration damper.

 In addition, from DE 3828128 a drive unit is known comprising a gearbox and a friction clutch located between the drive motor and the gearbox including a pressure plate which is non-rotatable, with the possibility of limited axial displacement, connected to a pressure disk connected to the output shaft of the drive the engine, at least one pressure spring, which presses the pressure disk against the friction disk of the clutch clamped between it and the pressure disk, which performs the function of a flywheel.

 The gear assembly according to the invention is characterized in that the gearbox is an automatic or semi-automatic gearbox, and the friction clutch is provided as an adjusting device to compensate for wear of the friction plates and to ensure at least approximately constant force loading of the pressure plate by means of the pressure spring, and in addition, the clutch is equipped with actuating means for coupling and disengaging and a safety device that is able during the process a disengaging portion at one working stroke of the actuating means to carry out a gradual decrease in torque transmitted to the friction clutch.

According to a suitable embodiment, the hold-down spring in at least one portion of the disengagement stroke has a regressive elastic characteristic;
the safety device has the ability to gradually decrease or gradually increase the moment transmitted by the friction clutch by at least 40-70% of the working stroke of the actuating means and / or pressure disk.

The invention is explained in more detail using figures 1-37. On them are presented:
FIG. 1 - a friction clutch defined by the invention,
FIG. 2 is a section along the line II-II of figure 1,
FIG. 3 is a permutation ring used in the friction clutch of FIG. 1 and 2,
FIG. 4 is a section along the line IV-IV of figure 3,
FIG. 5 is a support ring used with the friction clutch of FIG. 1 and 2,
FIG. 6 is a section along the line VI-VI of figure 5,
FIG. 7 and 7a - spring, exerting a torque on the permutation ring,
FIG. 8-11 are diagrams with various characteristics, from which it should be concluded that the individual spring and adjusting elements interact with the friction clutch defined by the invention,
FIG. 12 to 13 show another possibility of performing the friction clutch defined by the invention, wherein FIG. 13 is a section along line XIII-XIII of FIG. 12,
FIG. 14 is a general view of the permutation ring used in the friction clutch of FIG. 12 and 13
FIG. 15-17 - details of the friction clutch with compensating safety device,
FIG. 18 and 19 are diagrams with graphical characteristics, according to which it can be concluded about the interaction of the pressure plate spring and the lining spring, as well as the resulting effect on the friction clutch disengagement characteristic,
FIG. 20 illustrates embodiments of a friction clutch as defined by the invention,
FIG. 20a is a view in the direction of arrow A of figure 20,
FIG. 21 is a section along the line XXI of figure 20,
FIG. 22 is a view of an adjusting ring used in the friction clutch of FIG. 20-21,
FIG. 23 and 24 are other embodiments of the friction clutches defined by the invention,
FIG. 25 is a perspective view of a permutation ring that would be used in the friction clutch of FIG. 12 and 13 or 20-21,
FIG. 26 and 27 are additional embodiments of friction clutches,
FIG. 28 is a perspective view of a friction clutch as defined by the invention,
FIG. 29 is a section along the line II-II of figure 28,
FIG. 30 is a section along the line III-III of figure 28,
FIG. 31 is a section along the line IV-IV of figure 28,
FIG. 32 is a perspective view of the friction clutch used in FIG. 28 and 29 permutation ring,
FIG. 33 and 34 are details of other embodiments of the friction couplings defined by the invention,
FIG. 35 is a flywheel divided into two masses with a torsional vibration damper and with a friction clutch according to this invention,
FIG. 36 and 37 show a torque transmission device of the invention with a friction clutch.

 The friction clutch 1 shown in figures 1 and 2 has a housing 2 and a pressure disk 3 connected with it without the possibility of rotation, but with the possibility of limited axial displacement ring-shaped rotary bearings 5 and loads the pressure plate 3 in the direction of rigidly connected through screws 6a with the housing 2 of the pressure plate 6 (back pressure plate), as, for example, in the flywheel, and as a result the friction plates 7 di ka 8 clutch friction surfaces are clamped between the pressure plate 3 and counterpressure plate 6.

 The clamping disk 3 is rotatably connected to the housing 2 through circumferentially oriented or tangentially leaf springs 9. In the illustrated embodiment, the coupling disk 8 has so-called lining elastic segments 10, which, as it is known, provide a progressive increase in torque when the friction clutch 1 is turned on, while they, through the limited axial displacement of both friction plates 7 towards each other, make possible a progressive increase in axial forces, acting x on the friction linings 7. However, one could also apply the clutch plate, wherein the friction linings 7 would be substantially axially fixedly arranged on the carrier disc.

 In the illustrated exemplary embodiment, the disk spring 4 has a ring-shaped main part 4a that generates a clamping 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 locates the pressure plate 3 located radially further outward and the inclined radially further inward sections tilt around the rotary support 5.

 The swivel bearing includes two swivel plates 11, 12, which are formed here by wire rings and between which the disk spring 4 is axially fixed or clamped. The rotary support 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 by means of a power accumulator 13. The force accumulator 13 is formed by a single disk spring or a similar disk spring by a structural part 13, which supports its outer edge portion 13a axially loads the rotary support 11 towards the control disk spring 4 and thereby towards the housing 2. Provided between the pressure disk 3 and the control disk spring 4, the disk spring 13 has an outer annular edge portion 13b, from which extend radially inward tongues 13c, which are supported on the rotary support 11.

 To maintain the structural part 13, similar to a Belleville spring, in the presented embodiment, additional means 14 are fixed on the housing 2, which form a swinging support for the Belleville-like spring of the structural part 13. These additional means can be formed by welded or riveted segmented parts 14; which can be evenly distributed around the circumference. Means 14 can be formed, however, also an annular, closed part. In addition, the support means 14 can be formed directly from the housing 2, for example, by means of indentations made in the axial region of the housing 2 or by reed cuts, which, after insertion and clamping of a structural part 13 like a disk spring, are sealed under the outer edge portion of this part 13 by deformation of the material. In addition, a bayonet connection or a fixation can take place between the support elements 14 and the disk-like spring 13, so that the disk-like spring 13 is first pre-tightened and its radially outer portions can be axially fed through the bearings 14. After that by appropriate rotation of the disk spring part 13 with respect to the housing 2, the supporting sections of the structural part 13 can be brought into contact with the supporting medium 14. In this case, the supporting sections of the structural part 13 in the form of a disk spring can be formed by a cantilever projecting radially outwardly on the annular main part 13b.

 To fix the rotation of the control disk spring 4 and, if necessary, a structural part 13 similar to a disk spring, as well as to center the wire rings 11, 12, axially passing centering means in the form of rivet elements 15 are fixed on the housing 2. Rivet elements 15 have at least one axially extending shaft 15a, which extends axially through a notch provided between adjacent tongues 4b of the Belleville spring, and which can partially be wrapped around portions 13c, is formed bubbled for mating with the notch 13c of the tongue plate spring 13.

 A Belleville spring-like structural part or Belleville spring 13 is configured as a sensor spring, which produces at least a substantially approximately constant force on a given working path. Through this sensing spring 13, the clutch disengaging force acting on the tongues of the tongues 4c is perceived, and there is always at least an approximate balance between the force exerted on the pivot bearing by means of a switching force and the opposing force exerted by the disk disk spring 13 on this swing bearing 11. By turning off the force should be understood the maximum force that is exerted during the control of the friction clutch 1 on the tip 4c of the reeds or on the release arms of the reeds of the Belleville spring.

 An oscillating support 12 located on the housing side through the adjustment compensation device 16 is supported on the housing 2. This adjustment device 16 ensures that, with the axial displacement of the swinging supports 11 and 12 in the direction of the pressure plate 3 or in the direction of the back pressure plate 6, an undesirable gap can occur between the swinging support 12 and the housing 2 or between the swinging support 12 and the disk spring 4. This ensures that there are no unwanted dead or idle paths when bringing the action of the friction clutch 1, as a result of which there is an optimal efficiency and thereby perfect control of the friction clutch 1. The axial displacement of the swing bearings 11 and 12 is carried out with axial wear on the friction surfaces of the pressure plate 3 and the back pressure plate 6, as well as the friction plates 7. The principle of operation of the automatic control of the rotary support 5 is described in more detail in connection with the diagrams according to figures 8-11.

 The adjusting device 16 includes an elastically loaded adjusting element in the form of an annular structural part 17, which is shown in figures 3 and 4, the annular structural part 17 has circumferentially extending and axially increasing ramps 18 that are distributed around the circumference of the structural part 17. Adjusting element 17 is thus inserted into the sleeve 1 so that the ramps 18 face the housing base 2a. On the side of the adjusting element 17 facing away from the ramps 18, the swinging support 12 formed by the wire ring is positioned centrally in the grooved recess (base element) 19 (FIG. 2). In this case, the base element 19 can be made in such a way that the swinging support 12 on the adjusting element 17 is also fixed in the axial direction. This can be done, for example, by the fact that at least partially the sections of the adjusting element 17 adjacent to the base element 19 fix the swinging support 12 with a clip or form a latch for the swinging support 12. When using different materials for the swinging support 12 and the adjustment element 17 may be appropriate in order to compensate for differences of expansion arising at large temperature changes when the rotary support 12 made in the form of a wire ring is open, therefore, circumferentially separated at least in one place, whereby it becomes possible to move the wire ring 12 relative to the base element 19 in the circumferential direction and thereby the wire ring 12 can be matched with the diameter of the base element 19.

 In the presented exemplary embodiment, the adjusting element 17 is made of plastic, such as, for example, from heat-resistant thermoplastic, which can additionally be reinforced with fibers. As a result of this, the adjusting element 17 can be manufactured in a simple manner as an injection molded part. However, the adjusting element 17 can also be manufactured as a molded part of sheet metal or by sintering. In addition, with the appropriate choice of material, the pivot bearing 12 can be integral with the adjusting element 17. The swinging bearing 11 can be formed directly by the sensor spring 13. For this, the tongue of the tongues 13c can have a corresponding extruded recess or shape, such as a groove.

The adjusting element 17 is centered by axially extending portions 15a of rivets 15 evenly distributed around the perimeter. Moreover, the adjusting ring 17 has centering contours 20 which are formed in the circumferential direction of the recess 21, which are located radially inside the pivot plate (bearings) 11. To form the recesses 21 the adjusting ring 17 has on the inner edge portion radially inwardly extending cams 22 which define the radially inner contours of the recesses 21. As can be inferred from FIG. 3, discerned, when viewed in the circumferential direction, between evenly distributed recesses 21 as necessary 5 of the incident slopes 18. The recesses 21 are made in the circumferential direction so that they allow at least one torsion angle of the adjusting ring 17 with respect to the housing 2, which provides for the entire service life of the friction clutch 1 regulation appearing on the friction surfaces of the pressure plate 3 and the back pressure plate 6, as well as friction plates 7 wear. This angle of adjustment, in accordance with the calculation of the incline, can be about 8-60 degrees, preferably about 10-30 degrees. In the illustrated exemplary embodiment, this rotation angle is located in region 12, with the installation angle 23 of the incline slopes 18 also located in region 12 ^o . This angle 23 is chosen so that the friction that occurs when squeezing the ramps 18 of the adjusting ring 17 and the rams 24 shown in Figures 5 and 6 of the support ring 25 prevents slipping between the ramps 18 and 24. In accordance with the pairing of materials in the region of the ramps 18 and the counter-sloping slopes 14, the angle 23 may lie in the range between 5 and 20 degrees.

 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 counterclones 24 of the support ring 25, axially displaces the adjusting ring in the direction of the pressure plate 3, i.e., thus, in the axial direction from the radial body portion 2a. In the embodiment shown in FIGS. 1 and 2, spring loading of the adjusting ring 17 is provided by at least one ring-shaped twisted bending spring 26, which may have, for example, two turns and at one of its ends has a radially extending shoe 27, which prevents the adjusting ring from turning 17, at its other end, has an axially extending shelf 28, which is fixedly mounted, without the possibility of rotation, is hung on the housing 2. The spring 27 is inserted elastically clamped.

 The support ring 25 shown in FIGS. 5 and 6 is also formed by an annular structural part, which has oncoming counterclones 24, which form additional surfaces to surfaces bounded by incident slopes, and surfaces limited by incident slopes 18 and incident counterclones 24 can also be congruent. The installation angle 29 of the oncoming counterclones 24 corresponds to the angle of 23 oncoming slopes 18. As is evident from the comparison of Figures 3 and 5, the oncoming slopes 18 and oncoming counterclones 24 are likewise distributed in the circumferential direction. The support ring 25 is rigidly, without the possibility of rotation, is connected to the housing 2. In addition, the support ring 25 has recesses 30 distributed around the circumference through which rivet nozzles of the rivets 15 pass.

 Figure 2 shows a dashed line another ring-shaped twisted bending spring 26a, which, similarly to the twisted bending spring 26, can be bent at its end sections in order to provide a rigid, without the possibility of rotation, connection with the housing 2 on one side and with the adjusting element 17 with the other side. This spring 26a is also installed elastically clamped, so that it exerts a torque on the adjusting element 17. The use of two twisted bending springs 26, 26a may be preferable for other applications, since when the friction clutch 1 rotates as a result of acting on the springs 26 or 26a of centrifugal forces, spring force reinforcement appears. Through the use of two twisted bending springs, the force gain appearing, for example, on the spring 26 can be compensated by the force produced by the twisted bending spring 26a. For this, the coiled bending springs 26 and 26a are folded 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 one or more turns, in addition, these coil bending springs 26, 26a may have different coil diameters, as shown in figure 2, and usually associated with this and acting on the springs 26, 26a centrifugal the forces that would create different circumferential forces on the adjusting element 17 can at least approximately equalize by appropriately calculating the thickness of the wire and / or the number of turns of the individual springs 26, 26a. In figure 2, the spring 26 is located radially inside the adjusting element 17, and the spring 26a is radially outside the adjusting element 17. However, both springs could also be located radially inside or radially outside the adjusting element 17.

The figure 7 presents 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 of the order of 85 ^° . Position 32 represents the relative position of the bend 27 with respect to the bend 28, which it occupies with the new friction plates 7 in the friction clutch 1. Position 33 represents the position of the bend 27, which corresponds to the maximum allowable wear of the friction plates 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. In the remaining peripheral region, two turns of wire lie one above the other. The spring 26a is made similar to the spring 26, but has a larger coil diameter and a different clamping direction with respect to the adjusting element 17 according to figure 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, which form the ramps 18 and the ramps 24, are axially engaged to a large extent, which means that the rings 17 and 25 located on each other need the smallest axial design space.

 In the exemplary embodiment of FIGS. 1 and 2, the oncoming counterclones 24 or the cam-shaped protrusions 24a forming them are formed by a special structural 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, in particular, in sheet cases or covers, which are made as a whole.

 In order to hold the adjusting ring 17 before assembling the friction clutch 1 in its pulled position, it has in the area of the cams 22 engagement sections 36 for a rotary or stop means, which, on the other hand, can be supported on the housing 2. Such stop means can be provided with the manufacture or assembly of the friction clutch 1 and after mounting the friction clutch 1 to the flywheel 6 is removed, as a result of which the adjusting device 16 is activated. In addition, in the presented embodiment, in the roofs e or the housing 2 is provided in a circumferential direction with oblong elongated recesses 37, and in the adjusting ring 17 a recess or protrusion 38. In this case, the oblong recesses 37 lying in the circumferential direction must have at least one such passage that the adjusting ring 17 can rotate back according to the maximum possible angle of wear control. Also, after assembly of the friction clutch 1, the twisting tool can be axially guided through the grooves 37 of the cover and inserted 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 in relation to to section 2a at the smallest 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 6, so that, as already mentioned, the adjusting device 16 is released. The grooves 37 in the housing 2 are made in such a way that when removing or after dismantling the friction clutch 1 from the flywheel 6 the adjusting ring 17 may be brought into its retracted position. To this end, the clutch 1 is first disengaged, so that the control cup spring 4 does not exert axial force on the swinging support 12 and thereby guarantee a perfect rotation of the adjusting ring 17.

 In connection with the diagrams shown in FIG. 8-11, the characteristics should now explain in more detail the principle of operation of the previously described friction clutch.

 Line 40 in figure 8 denotes the axial force produced depending on the change in the taper of the disk spring 4, namely, when the disk spring 4 deforms between the supports, whose radial distance corresponds to the radial distance between the pivot bearing 5 and the radially outer support diameter 3a of the pressure disk 3. On the abscissa axis represents the relative axial path between the two supports, and the ordinate axis represents the force generated by the cup spring. Point 41 represents the inserted position of the disk spring 4 with the clutch 1 closed, thus the position where the disk spring 4 for the corresponding inserted position exerts maximum pressing force on the pressure plate 3. Point 41 by changing the conical inserted position of the disk spring 4 can move up and down along line 40.

 Line 42 represents the axial expanding force generated by the lining spring segments 10, which acts between the two friction plates 7. This axial expanding force counteracts the axial force exerted by the Belleville spring 4 on the pressing disk 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 disk 3, and it can advantageously also be greater. When the friction clutch 1 is disengaged, the spring segments 10 are unclenched, namely, the path 43. For this path 43, which also corresponds to one corresponding axial displacement of the pressure disk 3, the clutch 1 is switched off in this way, which means that a lower maximum tripping force must be created than one that would correspond to the insertion point 41 in the absence of cladding spring segments 10. When the point 44 is exceeded, the friction plates 7 are released, and on the basis of the degressive section In addition, the release force still exerted is significantly reduced with respect to that which would correspond to point 41. The release force of the coupling 1 is reduced until it reaches a minimum or a lower point 45 of the sinusoidal characteristic 40. When the minimum 45 is exceeded, the necessary release force again increases, and the path of disengagement in the region of the tip 4c of the protrusions is selected in such a way that automatically, when passing through a minimum of 45, the tripping force does not exceed the corresponding maximum point 44 The total release force, preferably remains below it. Thus, point 46 should not be exceeded.

 The spring 13, which serves as a force sensor, has a path-force relationship characteristic, respectively, of line 47 of figure 9. This characteristic 47 corresponds to that which is created when the disk-like structural part 13 changes in its taper from an unpressed position, namely, between two rotary supports that have a radial gap that corresponds to the radial distance between the swinging supports 11 and 14. As shown in graphical characteristic 47, a plate-like structural detail l 13 has one spring path, during which the axial force produced by it remains almost unchanged. In this case, the force created in this section 48 is selected so that it at least approximately corresponds to the clutch disengaging force corresponding to point 44 of figure 8. The support force exerted by the sensor spring 13 in relation to the corresponding point 44 is reduced corresponding to the gear ratio of the lever of this disk spring 4. This gear ratio in most cases is on 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 an out-of-line bearing.

 The insertion position of the plate-like spring element 13 in the friction clutch 1 is selected in such a way that it can make in the region of the swinging support 5 an axial spring stroke in the direction of the friction plates 7, which corresponds to at least the axial adjustment stroke of the pressure plate 3 in the direction of the back pressure plate 6 which occurs due to wear of the friction surfaces and friction plates and provides at least approximately constant axial support force for the pivot bearing 5. This this means that the linear portion 48 of the graphic characteristic 47 should have at least a length that corresponds to the aforementioned wear path, preferably more than this wear path, since as a result of this, the installation tolerances can also be aligned at least partially.

 In order to obtain an unchanged or a certain point of opening 44 of the friction plates 7 when the friction clutch 1 is turned off, the so-called double segmented lining spring between the friction plates 7 can be used, thus, a lining spring in which separate spring segments are provided in pairs back-to-back, moreover, individual pairs of segments can have a certain axial pre-tension (displacement) relative to each other, so that in the aggregate, provided by the lining spring The axial force corresponds to at least the disengaging force of the cup spring 4 corresponding to point 44, preferably lies slightly higher. By pretensioning the spring elements provided between the plates, it can be achieved that losses occurring during the service life when laying segments in the rear side of the plates are at least substantially aligned or compensated. By laying losses should be understood the losses that arise as a result of introducing segments into the rear side of the plates. It is advisable when the displacement provided between the plates of the spring is about 0.3 - 0.8 mm, preferably 0.5 mm By appropriately limiting the axial spring travel between the two friction plates 7, as well as by a certain displacement of the spring acting between the friction plates, it can also be achieved that at least when the friction clutch 1 is turned off, the pressure plate 3 is pressed out in a certain stroke 43 by means of linings of a spring. In order to obtain a certain stroke 43, the axial stroke between the friction plates can be limited by the respective stops both in the expanding direction and in the clamping direction of the cover spring 10. Preferably, as they are known, for example, they are used as cover springs, for example , from patent application P 4206880.0, which should be clearly accepted for the subject of this application.

 In order to ensure optimal functioning of the friction clutch 1 or an adjusting device that guarantees automatic compensation of the wear of the plates, it makes sense that when considering the characteristics of the tripping force according to FIG. 10, the forces exerted on the disk spring 4 primarily by the casing spring 10 and the sensor spring 13 and summing up, as well as the force after the pressure disk 3 is detached from the friction plates 7, then exerted only by the sensor spring 13 on the disk spring 4, are correspondingly larger , or at least equal to (a) the release force acting in the region 4c of the tip of the protrusions of the Belleville spring and changing over the course of the trip, respectively, figure 10.

 The previous consideration corresponds to a very specific mounting position of the disk spring 4, but the wear of the friction plates 7 has not yet been taken into account.

 With axial wear, in particular friction plates 7, the position of the clamping disk 3 is displaced in the direction of the backpressure plate 6, as a result of which the taper changes and thereby also exerts a cup spring when the friction clutch 1 of the clamping force is on, namely, in the sense of an increase. This change makes point 41 go in the direction of point 41 ', and point 44 in the direction of point 44'. By this change, the force balance initially existing when the clutch 1 was turned off in the region of the pivot bearing 11 between the control disk spring 4 and the sensor spring 13 is violated. An increase in the pressing force caused by the wear of the plates by the disk spring of the pressure disk 3 also displaces the travel of the switching force in the sense of increase. The resulting characteristic of the breaking force resulting from this is shown in FIG. 10 by the dashed line 50. By increasing the characteristic of the breaking force during the process of disengaging 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 is mainly in the region of the pivot bearing 5 weakens due to the axial stroke, which mainly corresponds to the wear of the friction plates 7. During this phase of deposition of the sensor spring 13, the disk spring 4 supports lives on the loading portion 3a of the pressure plate 3, so that this cup spring 4 changes its taper and, thus, the energy stored in it or the torque stored in it and, as a result, the force exerted by the cup spring 4 on the rotary support 11 or the sensor spring 13 and the pressure disk 3. This change occurs, as can be seen from figure 8, in the sense of reducing the forces generated by the spring disk 4. This change occurs 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 produced by the sensor spring 13. This means that in the diagram according to figure 8, the points 41 'and 44 'again go in the direction of points 41 and 44. After the equilibrium is again established, the pressure disk 3 can again come off from the friction plates 7. During this phase of the wear control during the release of the friction clutch 1, the adjusting element nt 17 of the adjusting device 16 is rotated by means of a pre-tensioned spring 26, as a result of which the rotary support 12 is additionally displaced according to the wear of the plates, and thus, the clearance-free rotary support 5 of the disk spring 4 is provided. 10. Lines 50 and 51 of figure 10 represent the axial stroke of the clamping disk 3 when the dependence of the clamping force - the path, respectively, lines 49, 50.

 The diagram according to figure 11 shows a graphical characteristic of the force exerted during the shutdown process on the housing 2 or on the disk spring 13, and the extrema were cut off. On the basis of the switched on position according to FIG. 1, the force 2 acts primarily on the pressure disk 3, which corresponds to the point 41 (FIG. 8) of the installation of the disk spring 4. 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 the line 52 of figure 11, namely, to point 53. If the point 53 is exceeded in the disengagement direction with a conditional clutch, in which the cup spring is axially stationary, but rotatably mounted on to the housing, therefore, the rotary support 11 without the possibility of axial bending would be connected to the housing 2, there would be an axial reversal of the action of the forces by means of a cup spring 4 on the housing 2 along the radial height of the rotary support 5. With the clutch defined in the invention, in the region of the swinging support 5, the force generated by axial reversal produced by the disk spring 4 in the region of the pivot bearing 5, it is sensed by the sensor spring 13. When the point 54 is reached, the disk spring 4 is detached from the loading portion 3a I of the pressure plate 3. At least up to this point 54, the disengagement process of the friction clutch 1 is supported by the axial force created by the cushioning spring 10. In this case, the force generated by the cushioning 10 decreases with increasing disengagement in the region 4c of the protrusion or with increasing axial disengaging pressure disk 3. Thus, the line 52 represents the resultant considered for the shutdown process, acting on the one hand in the region 4c of the tip of the protrusions of the shut-off force and, on the other hand The axial force exerted by the disk spring 4 on the swinging support 11 is perceived by the counteracting force generated by the sensor disk spring 13, and these both forces, at least after unloading the friction plates 7 by means of the clamping disk 3, are in equilibrium, and when the shutdown process continues, the axial force exerted by the sensor spring different 13 in the area of the rotary support 5, it becomes predominantly slightly larger than the corresponding breaking force. Section 55 of characteristic 52 of the diagrams according to FIG. 11 shows that with an increasing turn-off stroke, the breaking force or the force exerted by the disk spring 4 on the rotary support 11 will be less than the breaking force corresponding to point 54. The dashed line 56 corresponds to the state of the friction clutch, in which of the plates 7, wear has occurred, but regulation has not yet taken place in the area of the rotary support 6. It is also seen here that the change in the installation position of the disk spring 4 caused by wear is provided to increase the rotary support 11, or a sensor spring 13 to the housing 2 and effort. This, in particular, makes it possible that the point 54 leaves in the direction of the point 54 ', which makes it possible for the axial force exerted by the disk spring 4 on the sensor spring 13 in the region of the pivot bearing 11 to resume the disengagement of the friction clutch 1 more than the reaction force of the sensor spring 13, whereby the adjustment already described by expanding the sensor spring 13 occurs. Through this adjustment process, the point 54 'is again shifted in the direction of the point 54, whereby a state of equal And this in the swivel bearing 5 between the diaphragm spring 4 and the sensor spring 13.

 In practice, the described regulation takes place continuously or in very small steps, so that for the better understanding of the invention in diagrams, large displacements of points and displacements of characteristics do not occur in the usual way.

 During the operation of the friction clutch 1, some functional parameters or operating points may change. So, for example, due to the incorrect actuation of the friction clutch 1, overheating of the casing spring 10 may occur, which can result in upsetting, thereby reducing the axial springing of the casing springs or casing segments 10. However, by appropriately laying out the graphic characteristic 40 of the Belleville spring 4 and correspondingly fitting the characteristics 47 of the sensor spring 13, the operation of the friction clutch can be reliably operated. The axial upset of the casing spring 10 would only result in the disk spring 4 occupying a position pressed forward with respect to the position shown in FIG. 1, and the pressing force exerted by the disk spring 4 on the pressure disk would be slightly less, as is obvious in connection with characteristic 40 according to figure 8. 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 control disk spring 4 may increase with increasing wear. Moreover, the growth of part of the site in the aggregate of the maximum allowable wear rate of the friction plates 7 can be limited. In this case, the growth of the supporting force of the control disk spring 4 can occur by means of the corresponding calculation (calculation) of the sensor spring 13. Figure 9 shows the passage of the characteristic through section 48 by a dashed line and 47a. Due to the growth of the reference force of the control disk spring 4 with increasing wear at least partially, the decrease in the clamping force of the control disk spring 4 for the pressure plate 3 due to the lowering of the wrapping rod can be compensated reparation, for example, by nesting segments in plates. In this case, it may be particularly preferable when 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 casing. This means that with a decrease in the thickness of the disk in the region of the plates, therefore, a decrease in the distance between the friction surfaces of the plates due to the insertion of segments and / or one settlement of the spring elements and / or wear of the plates, the aforementioned supporting force should increase. It is particularly preferable when the increase in strength occurs in such a way that it is larger in the first section than in the second section adjacent to it, both sections being inside the region 48 according to FIG. 9. The last calculation is preferable because the largest part of said insertion between spring segments and linings occurs mainly during one period of time that is insignificant compared with the entire service life of the friction clutch and, accordingly, is practically stable uyutsya ratio between the spring segments and the friction facings. This means that with a certain insertion, there is no longer a change with respect to the insertion. An increase in the supporting force of the control disk spring can also occur for at least one part of the wear of the friction plates.

 In the preceding description of the adjustment process, to compensate for the wear of the friction plates, the axial forces accidentally exerted by the leaf spring 9 were not taken into account. When the leaf springs 9 are displaced in the sense of the pressure plate 3 is torn off from the corresponding friction plate 7, thus, in the sense of the pressure plate 3 is pressed against the disk spring 4 is the support of the tripping process. The axial force exerted 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. This is for the sake of a better understanding was not taken into account when describing the diagrams according to figures 8-11. The loading control disk spring 4 when the friction clutch 1 is turned off to the rolling support 12 located on the side of the cover, 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. The elements 9 of the leaf springs can be installed between the cover 2 and the pressure plate 3 in such a way that with increasing wear of the friction plates 7, the axial force exerted by the leaf springs 9 on the control disk spring 4 becomes greater. For example, in a stroke 48 according to the figure 9 and thus also due to the wear-compensating stroke of the adjusting device 16, the axial force created by the leaf springs 9 may have a characteristic according to line 47b. From figure 9 it can also be concluded that with increasing draft 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 travel according to characteristics 47b and the characteristic of the disk spring, a resultant force stroke is obtained which acts axially on the cup spring 4, namely, in the sense of pressing the cup spring 4 against the swinging support 12 located on the side of the lid. Ia according to one stroke line 47a is expedient computation touch plate spring so that it has a stroke characteristic curve according to the line 47c figure 9. Through the addition power characteristic according to the line 47c and a line 47b according to the characteristics obtained when the force characteristic curve according to the line 47a. Thus, by appropriately displacing the leaf springs 9, the support force exerted by the sensor spring can decrease (correspondingly, the characteristic of the support force). By appropriately performing and arranging the leaf spring elements 9, the upset of the cover springs and / or the insertion of the cover spring segments into the plates 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 working point or the same working section, so that the disk spring 4 exerts a substantially at least approximately constant clamping force on the clamping force during the life of the friction clutch plate 3. In addition, when laying out the friction clutch, in particular the sensor spring 13 and / or leaf springs 9, the adjusting springs 26 and / or 26a p acting on the adjusting element 17 must be taken into account the resulting axial force that counteracts the sensor spring 13 and / or leaf springs 9.

 When calculating the friction clutch 1 with the displaced leaf springs 9, it can also be taken into account that, by displacing the leaf springs 9, the axial force created by the pressure plate 3 on the friction plates 7 is affected. Thus, this means that when the leaf springs 9 are compressed the direction of the control disk spring 4, the clamping force exerted by the disk spring 4 is reduced by the force of preliminary compression (displacement) of the leaf springs 9. Thus, with this kind of friction clutch 1, the characteristic p the resulting clamping force of the pressure plate 3 or the friction plates 7, which is obtained by overlaying the graphical characteristics of the plate spring pressing force with the compression characteristic of leaf springs 9. On the proposal that if we consider the range of operation of the friction clutch 1, graphical characteristic 40 according to FIG. 8 represents the resulting force characteristic from 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 back pressure plate 6 as a result of wear of the plates, an offset of the resulting characteristic would be obtained in the sense of reduction. Figure 8 shows the line 40a in strokes, which corresponds, for example, to a total liner wear of 1.5 mm. As a result of this, the displacement of the line 40 in the direction of the line 40a during the service life of the friction clutch decreases the axial force exerted by the disk spring 4 on the sensor spring 13 when the friction clutch 1 is turned off, namely, as a result of the opposing moment exerted by the increasing spring wear of the leaf springs 9 on the disk spring 4. This opposing moment is due to the radial distance between the swing bearing 5 and the load diameter 3a between the control cup spring 4 and the clamp th plate 3. When laying out the friction clutch 1, it is especially important that the increase in compression of the leaf springs 9 due to wear of the plates is less than the increase in the shut-off force due to the same wear of the plates, which makes the sensor spring rotate 18 necessary to regulate. Otherwise the pressing force of the pressure plate 3 for the friction plates 7 would be reduced in the engaged state of the friction clutch, and regulation would not be possible at all.

 The friction clutch 101 shown in FIGS. 12 and 13 differs mainly with respect to the friction clutch 1 shown in FIGS. 1 and 2 in that the adjusting ring 117 is loaded with coil springs 126 in the circumferential direction. Regarding its function and principle of operation with respect to compensation of wear of the friction plates, the adjusting ring 117 corresponds to the adjusting ring 17 according to figures 2-4. In the shown embodiment, three view springs 126 are provided, which are evenly distributed around the circumference and compressed between the clutch housing 2 and the adjusting ring 117.

 As it follows, in particular, from figure 14, the adjusting ring 117 has radial protrusions or steps 127 on the inner circumference, at which arcuate coil springs 126 can be supported by one of their ends to load 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 as 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 figures 13 and 14, the ring 117 around the inner circumference can be made in such a way that at least mainly in the area of the extension of the springs 126 and preferably also through the angle of rotation of the ring 117 necessary for controlling wear or through spring expansion 126 there is a guide 129, which provides axial locking and radial support of the springs 126. The spring guides 129 in the presented embodiment are formed, when viewed in cross section, basically a semicircle bubbled recesses bounding surfaces of which are adjusted essentially to the cross section of the coil springs 126.

 This kind of embodiment has the advantage that, with the rotating friction clutch, the flawless guide of the springs 126 is given so that they cannot radially deflect. For additional locking of the coil springs 126, as shown in figure 13, the cover 2 may have axial formations 130 on its radially 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 adjusting springs 126 according to figures 12-14 has the advantage that, with the rotating 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 control forces applied by the springs 126 in the circumferential direction due to the springs formed between the turns of the springs and the adjusting ring 117 of the frictional resistance is reduced or completely destroyed. 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 at revolutions above 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 with an idle internal combustion engine, hence a non-rotating friction clutch 101, regulation can occur due to wear of the plates.

 Blocking the adjustment process during rotation of the friction clutch 1 or when exceeding a certain number of revolutions may be preferable also in the form of execution according to figures 1 and 2. In addition, on the housing 2 can be provided, for example, means that when exposed to centrifugal force on the adjusting element 17 they lock against rotation, namely, against the control force generated by the twisted bending spring 26 and / or 26a. In this case, the locking means can be formed by at least one weight, which is pressed radially outwardly, by means of centrifugal forces, which is supported, for example, on the inner flange of the ring 17 and can create friction there, which causes a blocking moment on the ring 17, which is greater than that exerted by the adjusting springs on the ring 17 torque.

 For radial support of at least one stretch of spring extension 126, support means carried by the housing 2 can also be provided. These support means, in the embodiment according to figures 12 and 13, can be integral with the stops 128. For this, the stops 128 can be made in the form of an angle, so that they have, as necessary, one section extending in the circumferential direction, which at least on one part of the passage section of one spring 126 enters into it. Due to this, at least one part of the coil of the spring can be held and supported in at least the radial direction.

 As can be inferred from FIG. 13, the wire ring 11 provided in FIG. 2 has fallen off and replaced by formations 111 located in the region of the tip of the protrusions of the sensor spring 113. Furthermore, the protrusions 113c in their tip region are barrel-shaped on their side .

 Figures 15-17 show another embodiment of a wear adjusting device defined by the invention, in which instead of an annular adjusting ring, separate adjusting elements 217 are used. These adjusting elements are evenly distributed around the circumference of the lid 202. The adjusting elements 217 are formed by button-shaped or disk-shaped structural parts, which have circumferentially extending and axially rising ramps 218. The ring-shaped adjusting elements 21 7 have one central recess or hole 219 through which axial rod-shaped nozzles 215a carried by the lid pass, so that the annular adjustment elements 218 are rotatably mounted on these nozzles 215a. Embossments 225 are provided on cover 202, which form oncoming counterclones 224 for slopes 218. Between one adjusting element 217 and cover 202 is spring-loaded element 226, which loads the adjusting element 217 in a direction of rotation that regulates. The spring element 226 may extend, as follows from figure 15, around one axial nozzle 215a, thus being made like a helical spring. At the end sections of one spring 226, formations such as bends or bent pods 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. With uniaxial displacement of the disk spring 204 or the sensor spring 213 in the area of the swinging support 205, the adjusting elements 218 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 the rotary or pivoting adjusting elements 217 shown in FIG. 14, wedge-shaped adjusting elements offset in the radial and / or circumferential direction to control wear could also be used. These wedge-shaped adjusting elements can have one oblong recess through which axial nozzles 215a can pass in order to guide the corresponding adjusting element. Wedge-shaped adjusting elements can act 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 the 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 adhesion between the ramps and the ramps cannot interfere with the regulation. To avoid such adhesion, at least one of these structural parts can be coated at least in the area of slopes or counterclones. By means of such coatings, in particular, corrosion can be eliminated by using two metal structural parts. Adhesion or gluing between structural parts forming adjusting slopes, in addition, can be eliminated by the fact that the structural parts resting on each other and forming slopes, as well as counterclones, are made of the same material with a different coefficient of expansion, so that as a result of temperature fluctuations appearing during the operating time of the friction clutch, in contact with the surface, which form the adjustment slopes, make a movement relative to each other. Thus, the structural parts forming oncoming slopes and oncoming counterclones are always kept movably relative to each other. Thus, adhesion or adhesion between these parts cannot occur, since due to various stretching, these parts are again constantly torn apart or disconnected. Separation of the adjustment slopes can also be achieved by the fact that, due to different strengths and / or performance of parts, centrifugal forces acting on these parts cause various elongations or displacements, which, in turn, eliminate the adhesion or adhesion of the parts.

 In order to avoid sticking between ramps and ramps, at least one adjusting device can be provided which, when the friction clutch is disengaged or when adjusting wear, exerts axial force on the adjusting element or adjusting elements. In addition, the adjusting element 17, 117 can axially connect to one structural part, which has sections that are axially mixed during wear. This connection can take place in particular in the area of the swinging support 5, namely, with a control cup spring 4 and / or sensor spring 13.

 The diagram according to FIG. 18 shows a characteristic 340 of a pressure cup disk, which has a lower point or a minimum of 345, in which the force exerted by the pressure cup disk is relatively small (about 450 Nm). The maximum of a Belleville spring with a characteristic of 340 path-force dependences is about 7600 Nm. The characteristic 340 is obtained due to the deformation of one disk spring between two supports radially distant from each other, namely, as described in connection with the characteristic 40 according to figure 8 and in connection with a disk spring 4.

 The characteristic 340 of the Belleville spring can be combined with the characteristic 342 of the coil spring. As can be inferred from Figure 18, the path-force relationship of the characteristic 342 of the casing spring segments approaches the characteristic 340 of the pressure plate spring, or both characteristics pass 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 of two vertically arranged one above the other points of the lines 340 and 342. This difference is indicated by 360. Actually, the necessary clamping force is reduced by the gear ratio of the control element levers, such as, for example, protrusions of a disk spring. This was also described in connection with the embodiment according to figures 1 and 2, as well as the diagrams according to figures 8-11.

 Figure 18 shows the strokes of 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, therefore, it 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 Belleville springs can be paired with the characteristic casing springs in accordance with line 442.

 The figure 19 shows the progress of the associated graphical characteristics 340 and 342 or 440 and 442 downforce with the aim of disengaging the corresponding friction clutch exerted on the control levers, such as protrusions of a disk spring. As can be seen, the stroke 349 of the clamping force, coupled with the characteristics 340, 342, is constantly in the positive power zone, which means that to keep the clutch in the disengaged state, a force in the direction of disengagement is constantly necessary. The downforce stroke 449 associated with characteristics 440 and 442 has one 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 the friction clutch 501 shown in FIGS. 20, 20a and 21, the touch disk spring 513 is axially supported via 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 on the radial portions 502a, in the form of linings formed from the cover material. Cover plates 502a are formed from a substantially axially extending edge portion 502b of the cover, it being 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 you can conclude, in particular from figure 21, the pads 502a and the jaws or the protrusions 513d are thus set on top of 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 502a.

 In order to perfectly position the sensor spring 513 relative to the housing 502, while receiving the bayonet locking connection 514, at least three preferably uniformly distributed covers 502a around the circumference of the cover 502 are thus aligned with other portions of the cover that after one specific relative rotation between the sensor spring 513 and the lid 502, the corresponding protrusions 513d come into contact with the circumferential stop 502f, and thus, a further relative rotation between the sensor spring 513 and the cover 502. The stop 502f in the illustrated embodiment, as follows in particular from figure 20a, is formed by the axial step of the cover 502. From figure 20a, it also follows that the individual, 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 anti-unlocking between the sensor spring 513 and the cover 502, the stops 502g are formed by axial, 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 are axially inserted into the free cutouts 502c and 502d and extend to an axial arrangement on 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 to be located between the respective stops 502f and 502g, and all the protrusions 513d overlap the supports 502a located on the side of the cover. By performing bayonet locking as determined by the invention 514, it is ensured that when mounting the friction clutch 1, 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, for example 513b, which exerts the actual spring force of the sensor spring 513, 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 when 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 according to figures 1 and 2, this means that the axial clamping force 13 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 FIGS. 20-21, on the side of the housing, the oncoming counterclones 524 are formed by cam-shaped embossments located in the sheet housing 502. In addition, in this embodiment, clamping between the housing 502 and the adjusting ring 517, coil springs 526 are guided by guide mandrels 528, which made integrally with the adjusting ring 517 and extend in the circumferential direction. These guide mandrels 528 may, as follows in particular from figure 21, have an axial direction of one oblong cross section that is adapted to the inner diameter of the springs 526. The guides 528 extend in at least one partial section of the longitudinal extension of the springs 526 in 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 548, the installation of a friction clutch can be greatly facilitated.

 The figure 22 partially shows the adjusting ring 517. The adjusting ring 517 has radially extending formations 527 that bear rod-shaped, circumferentially extending guide sections 528 for coil springs 526. In the illustrated embodiment, the sections 528 for springs are integrally formed with plastic ring 517, manufactured as a part, molded under pressure. However, the spring direction portions or the spring locating portions 528 can also be formed by separate structural parts or all together by one single structural part, which can also be connected (respectively connected to the adjusting ring 517), for example, via latching. Thus, all the guide sections 528 can be formed by a single, if necessary, circumferentially open ring, which is connected to the adjusting ring 517 through at least three connecting places, preferably in the form of a latch.

 In the same way as described in connection with FIGS. 12 and 13, coil springs 526 can additionally, for example, be based on the action of centrifugal forces, be radially supported in correspondingly executed portions of the cover 502 and / or the adjusting ring 517.

 The supports for coil springs 526 located on the side of the cover are formed by plastic deformation from the cover material and axially extending wings or axial wall embossments 526. These support sections 526a of the springs 526 are expediently designed so that the respective ends of the springs are guided and thereby protected against unacceptable axial and / or radial misalignment.

 In the embodiment of the coupling 601 shown in FIG. 28, the sensor spring 613 is provided on the side of the housing 602 facing away from the pressure plate 603. Due to the location of the sensitive spring 613 outside the internal cavity of the housing, which receives the pressure plate 603, the thermal load of the sensor spring 613 can be reduced, thereby eliminating risk of settling of this spring 613 due to thermal overload. Also happening on the outside of the housing 602 is better cooling of the spring 613.

 The support provided on the flap side of the control disc spring 604 of the swing bearing 611 is provided through distance rivets 615, which extend axially through the corresponding recesses of the spring spring 604 and the housing 602 and are axially connected to the sensor spring 613. In the illustrated embodiment, the distance rivets 615 are connected to the 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. Thus, for example, the sensor spring 613 could have axially extending pads on the radially inner portion that support the swing bearing 611 with respective radial regions or directly form this swing bearing 611 with corresponding formations obtained by plastic deformation.

 In the exemplary embodiment of FIG. 24, the sensor spring 713 extends radially inside the pivot bearing 715 for the control disk spring 704. The sensor spring 713 is supported on the cover 702 in its radially inner portions, for which the cover 702 has axially extending through the corresponding grooves or recesses of the disk spring 704 pads 715 that axially support the sensor disk spring 713. According to another embodiment, also the sensor spring 713 in its inner edge region could have subcontroller which axially extend axially through corresponding holes 704 and the plate spring are supported by the cover.

 Shown in figure 25, the adjusting ring 817 can be used in one friction clutch according to figures 20-21. The adjusting ring 817 has radially inner formations 827 extending radially. The formations 827 obtained by plastic deformation have radial nozzles 827, which form supporting sections for circumferentially clamped coil springs 826 between the cap of the coupling and the adjusting ring 817. To carry out and facilitate the installation of the coil springs 826, a ring 528 is provided that is broken or open outer circumference. Ring 528 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 springs 826 located on the side of the cover are formed by axial overlays 826a of the clutch 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 when 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, the recesses in the formations 827a are open in one axial direction, and cutouts 826b - in another axial direction.

 In the embodiment of one friction clutch 901 shown in FIG. 26, the control disc spring 904 is supported in the disengaging direction in the middle portion of the disc spring 904 main part 904a. The main part 904a is radially outwardly supported on the pressure plate 903 and exits radially inward through the pivot bearing 905. This means that the rotation support 905 is relatively far removed from the inner edge of the main part 904a of the Belleville spring 905 or the groove ends that form the protrusions of the Belleville spring 904, compared to the previously known friction clutches. In the present exemplary embodiment, the ratio of the radial width of the portions of the main part provided radially inside the pivot support 905 to the portions of the main part radially outside the pivot support 905 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 disc spring main part 904a in the region of the rolling support 905 can be eliminated.

 In addition, the dashed line in FIG. 26 indicates an axial 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 the centering rivets or e fingers 915 may fall. Although not shown, the outer diametric alignment can also be carried out through the lining or embossment obtained from the material of the cover 902 by plastic deformation.

 With the friction clutch 901, the sensor spring is configured such that the force-producing main part 913a is provided radially inside the cams 903b. To maintain the control poppet spring 904, on the one hand, and to actually maintain on the cover 902, on the other hand, the sensor spring 913 has radial jaws or protrusions that extend, on the one hand, from the main part 913a radially inward and, on the other hand the sides, starting from the main part 913a, extend radially outward.

 In the embodiments of one friction clutch 1001 shown in FIG. 27, a force directed opposite to the disengaging force of the friction clutch or the rotational force of the control cup spring 1004 is applied by a sensor spring 1013, which is axially compressed between the housing 1002 and the pressure plate 1003. In one such embodiment, the control the disk spring 104 in the area of 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 plate located on the cover side or the support plate 1012 is guaranteed by the bias force of the sensor spring 1013. This sensor spring is laid out 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 or will be greater than the release force of the friction clutch 1001. It must be ensured that when there is no wear on the friction plates, the cup spring 1004 is constantly in contact with the support located on the side of the lid or rotary pads 1012. For this, in the same way as described in connection with the previous embodiments, coordination must be made between individual axially active forces and superimposed forces. These efforts, which are made by the sensor spring 1013, the casing spring provided between the spring plate 1003 and the housing 1002, the leaf spring elements, the control disc spring 1004, the tripping force of the friction clutch 1001 and the adjusting spring elements acting on the adjusting ring 1017, must be adjusted to each other accordingly .

 The friction clutch 1101 shown in FIGS. 28 and 29 has a housing 1102 and without rotation rotatably connected to it, but having limited axial displacement, a pressure plate 1103. A pressure plate spring 1104 is compressed axially between the pressure plate 1103 and the cover 110, which can deflect around carried by the housing 1102 of the annular support 1105 and loads the pressure plate 1103 in the direction of the back pressure plate 1106 rigidly connected to the housing 1102, such as, for example, in the flywheel, resulting in a friction the ionic plates 1107 of the clutch disc 1108 are clamped between the friction surfaces of the pressure disc 1103 and the back pressure plate 1106.

 The clamping disk 1103 is rotatably connected to the housing 1102 through the circumferential direction or tangentially oriented acting means in the form of leaf springs 1109. In the illustrated embodiment, the clutch disk 1108 has the so-called clutch spring segments 1110, which provide a progressive increase in torque during friction clutch couplings, while they, through the limited axial displacement of both friction plates 1107 towards each other, make possible a progressive increase in acting on friction plates 1107 axial forces. However, a clutch disk could also be used, in which friction plates 1107 would be axially practically rigidly applied to at least one bearing disk and provided for replacement in another place for the lining spring elements 1110, for example, between a disk spring 1104 and a pressure disk 1103.

 In the illustrated exemplary embodiment, the disk spring 1104 has a single pressure-generating annular main part 1104a from which radially inwardly extending control protrusions 1104b extend. In this case, the cup spring 1104 is inserted in such a way that it radially further externally located sections loads the clamping disk 1103 and tilts radially further inside the located sections around the rolling support 1105. This swing bearing 1105 includes two pivot pads 1111, 1112, which are here formed by wire rings between which the spring plate is axially fixed or trapped. For protection against rotation of the control disk spring 1104 and for centering, as well as fixing the wire rings 1111, 1112 relative to the housing 1102, latches in the form of rivets 1115 are fixed to the lid, which, if necessary, pass through one axially extending rod 1115a through the disk plate provided between adjacent protrusions 1104b spring cutout.

 The clutch 1101 has an axial compensating axial wear of the friction surfaces of the pressure disk 1103 and the back pressure plate 1106, as well as the friction plates 1107, which consists of a wear-compensating device 1116 provided between the pressure disk spring and the pressure disk 1103, as well as a pressure disk 1103 that limits the release stroke restrictive means 1117, which are designed as a track sensor.

 The restrictive means 1117 acting as wear sensors, as needed, have one sleeve 1118, which is rotatably based in one hole 1120 of the pressure disk 1103. The sleeve 1118 forms a groove 1121 through which axially two leaf spring elements 1122 pass. The leaf spring elements 1122 are supported against each other to a friend, wherein at least one leaf spring element is curved, preferably both leaf spring elements are curved in the opposite direction. The leaf spring elements 1122 are based in the sleeve with a certain pre-tension and thus have the ability to bias against the specified friction resistance with respect to the sleeve 1118 in the axial direction of the must 1101. The axial length of the leaf spring elements 1122 is selected so that when the friction clutch 1101 is engaged, these leaf spring elements in relation to one axially fixed structural part of the coupling - in the presented example of execution in relation to the outer edge section 1123 cor mustache 1102 - have a certain clearance 1124 which corresponds to a predetermined tripping stroke of the pressure plate 1103. When the friction clutch is engaged, the leaf spring elements 1122 with their end 1122a facing away from the housing 1102 come into contact with the back pressure plate 1106, which ensures that when the friction linings are worn 1107, the pressing disk 1103 in accordance with this wear of the plates is axially displaced relative to the leaf spring elements 1122, namely, against the action of the frictional closure between the leaf and spring elements 1122 and a sleeve 1118, which consists mainly of plastic or friction material.

 In the illustrated exemplary embodiment, the hole 1120 into which the sleeve 1118 is mounted by pressing in both axial and circumferential directions is provided in one cam 1125 of the pressure plate, which extends radially outward and is pivotally connected to one, as necessary, spring leaf element 1109 through rivet connection 1109a. The offset of the sleeve 1118 in the direction of the backpressure plate 1106 can also be eliminated by the fact that the sleeve 1118 at its end facing the housing 1102 has one shoulder 1118a through which it can be supported on the pressure disk 1103. The offset of the sleeve 1118 from the hole 1118 in the direction of the housing or the cover 1102 couplings can be eliminated by the fact that, as indicated by the dashed line in figure 1101, leaf springs 1109 partially overlap the sleeve 1118 and, if necessary, additionally rigidly compress axially into the hole 1120. The torsion of the sleeve can be eliminated, in addition, the fact that the sleeve has a profiling, in particular, a shoulder, which receives sections 1119 of leaf springs 1109 overlapping the sleeve.

 The wear compensating device 1116 has a spring-loaded disc spring 1104 in the form of a U-shaped cross-section of a ring 1126 of sheet material, which is shown in horizontal projection in figure 1105. Compensating part 1126 has at least a bottom 1127 on the bottom side of the spring 1104 of the bottom 1127 one annular axial protrusion 1128 or several protrusions 1128 which are predominantly evenly distributed around the circumference and are formed by grooves curved in the sheet material. The segmented protrusions 1128 ensure that in the region between the circumferentially adjacent protrusions 1128, radial gaps are formed between the disk spring main part 1104a and the expansion ring 1126, making it possible for cooling air to pass through. As can be seen, in particular, from figure 29, the compensation ring 1126 is centered relative to the pressure disk 1103. For this, the pressure disk 1103 has at least one step 1129 which positions the radially inner, axially extending wall 1130 of the compensation ring 1126 centered on the pressure disk 1103. The step 1129 can be formed by a circumferential closed surface, or provided circumferentially at a distance from each other, segmented surfaces. The expansion ring 1126 also has a radially outer axially extending wall 1131, which, together with the inner wall 1130 and the bottom 1127, forms an annular free cavity 1126a. Outwardly, the compensation ring 1126 has radial protrusions or cams 1132, which form the stops interacting with the counterparts 1133 of the axially displaceable structural parts in the form of leaf spring elements 1122 of the wear sensors 1117. The counters 1133 are formed by cams formed on the leaf spring elements 1122, which are directed radially inward and overlap the protrusions 1132. Due to this, the axial displacement of the compensation ring 1126 in the direction from the pressure disk 1103, therefore, in the direction of the housing, is limited by the counterclones 1133.

 A balancing device 1134 is provided between the compensation ring 1126 and the pressure disk 1103, which, when the friction clutch 1101 is disengaged and the plates are worn, makes it possible to automatically shift the compensation ring 1126, and when the clutch is engaged, it acts with self-braking, therefore, with blocking, which ensures that during the engagement phase of the friction clutch 1101, the combination ring 1126 maintains a certain axial position relative to the pressure disk 1103. This particular position m It can only change during the disengagement process and in accordance with the wear of the plates.

 The adjusting device 1134 includes several, preferably evenly distributed around the circumference pairs of wedges 1135, 1136, which are based in the annular free cavity 1126a of the sheet ring 1126. The wedges 1136 supported on the annular surface 1137 of the pressure disk 1103 are non-rotatable but axially displaceable connected to a ring 1126 of sheet material. For this, the leaf ring 1126 in the region of its axially extending walls 1130, 1131 has formations in the form of grooves 1138, 1139, which form protrusions in the region of the free cavity 1126a that are included in appropriately fitted recesses or grooves 1140, 1141 of the wedges 1136. Grooves 1140, 1141, or formations 1138 1139 extend in the axial direction of the sleeve 1101. As is apparent from Figure 30, the wedges 1135 are mainly axially based between the bottom 1127 of the sheet ring 1126 and the wedges 1136. The wedges 1135 and 1136 form circumferentially extending and axially rising oncoming slopes 1142, 1143, through which wedges 1135, 1136 mated with one pair are mutually supported. Wedges 1135, on the other hand, are supported at the bottom 1127 of ring 1126 and are offset with respect to this ring 1126 in the circumferential direction. The incline slopes 1142, 1143 are mutually compressed. For this, power accumulators in the form of coil springs 1144 are based in the annular cavity 1126a and are supported at one end on one wedge 1136 that is not rotatable with the ring 1126, and on the wedge 1135 displaced in the circumferential direction. For fixing the power accumulators 1144, the wedges 1135 , 1136 have protrusions 1145, 1146 at their ends facing the respective power accumulators, which enter the spring coils and thereby fix the ends of the spring. The springs 1144, in addition, are guided by sections 1130, 1131 of the walls and the bottom 1127 of the ring 1126.

 In the presented exemplary embodiment, the compensation ring 1126 is locked against rotation with respect to the pressure disk 1103. For this, as is clear from FIG. 31, axial protrusions in the form of pins 1147 are provided on the pressure disk 110, which extend axially through the recesses 1148 provided in the departure region 1132 By this anti-torsion stop, it is ensured that during operation of the friction clutch, the thrust areas of the linings 1132 are always positioned below the restrictive cams 1133 of the leaf spring elements 11 22.

In the presented exemplary embodiment, the wedges 1135, 1136 are made of heat-resistant plastic, such as, for example, thermosetting or thermoplastics, which can be further reinforced with fibers. Thus, the wedges 1135, 1136, acting as adjusting elements, can be manufactured in a simple manner as parts obtained by injection molding. However, it may also be preferable when at least one of the wedges 1135, 1136 of one pair is made of friction material, such as, for example, lining material. However, the wedges or adjusting elements 1135, 1136 can also be made as a molded part of a sheet material or as a sintered part. The angle of inclination, as well as the passage of the ramps 1142, 1143, are designed in such a way that for the entire service life of the friction clutch 1101, the regulation of the pressure disk 1103 and the back pressure plate 1106 and the friction wear plates 1107 on the friction surfaces is guaranteed. The wedge angle 1149 or the inclination angle 1149 of the ramps 1142, 1143 relative to the plane perpendicular to the axis of rotation of the friction clutch is selected so that friction that occurs when the ramps 1142, 1143 are pressed onto each other prevents slippage between these slopes. In accordance with the pairing of materials in the region of the incline 1142, 1143, the angle 1149 may lie in the range between 5 and 20 ^° , preferably it may be of the order of 10 ^° . The wedges 1135 displaced in the circumferential direction are arranged so that they point with their tip in the direction of rotation 1150.

 The clamp by means of power accumulators 1144 of the ramps 1142, 1143 and the angle of inclination 1149 are designed so that the resulting axial force acting on the adjusting ring 1126 is less than the necessary displacement force of the wear sensors 1122 of the restrictive means 1117.

 In addition, when calculating the Belleville spring 1104, it must be taken into account that the clamping force generated by it for the pressure disk 1103 must be increased to the necessary displacement force for the wear sensor 1122 and the clamping forces compressed between the cover 1102 and the pressure disk 1103 of the leaf springs 1109. In addition, Separate structural parts must be laid out so that the wear of the supports between the honey disc spring 1104 and the support ring 1126, as well as the wear of the stops between the wear sensors 1122 and the back pressure plate 1106 or between the sensors and 1102 and the case remains small compared to the wear plates in 1107.

 In order to avoid unwanted rearrangement between the ramps 1142, 1143 or the adjusting elements 1135, 1136 in FIG. 30, small protrusions can be provided in the region of at least one of the ramps 1142, 1143 that engage with another ramp. In this case, the protrusions can be laid out in such a way that adjustment is possible to compensate for wear, however, the sliding of the slopes relative to each other is prevented. It may also be especially advantageous when both on-going inclines 1142, 1143 have protrusions that fit into each other. These protrusions can be formed, for example, with a very small height sawtooth profiling, which makes possible the relative movement of the slopes 1142, 1143 only in the wear control ring. Such profiling is schematically shown in FIG. 3 along one section of the slopes 1142, 1143 and is indicated by 1143a. In applications where only one of the ramps 1142, 1143 has protrusions, the latter can be made in such a way that they have a higher hardness than the material forming the other ramp, so that at least slight penetration or trapping can take place protrusions on their supporting, incline slope.

 In order to prevent curved or wavy leaf springs 1122 from losing very high temperatures during the clutch of the pressure plate 1103, the clamping force is lost, the sleeves 1118 are made predominantly of the same material with low thermal conductivity and high friction coefficient. Wedges 1135, 1136 can be made of the same material.

 For better cooling, in particular of the pressure disk 1103, grooves radially extending and distributed around the circumference can be provided in the pressure disk 1103, one of which is represented by a dashed line in FIG. 29 and indicated by 1151. These radial grooves 1151, when viewed in the circumferential direction, are provided between two, as needed, adjacent pairs of wedges and extend between the ring 1126 and the clamping disk 1103. In the area of the springs 1144, the ring 1126 could have axial cuts extending from the bottom 1127, as a result of which between With a spring 1104 and a ring 1126, radial passages would be formed.

 In order to increase the wear resistance at different bearing points, the respective sections can be provided with a wear-resistant coating of the type, for example, hard chrome plating, molybdenum coating, or particularly wear-resistant structural parts can be provided in the area of the contact points. So, for example, plastic pads in the area of contact with the back pressure plate 1106 and the housing 1102 may be provided on contribution sensors 1122.

 The leaf springs 1109 transmitting torque to the pressing disk 1103 are clamped between the pressing disk 1103 and the housing 1102 in such a way that they disengage the pressing disk 1103 towards the housing 110 when the friction clutch 1 is released. however, until the restrictive means 1117 are in effect, the ring 1126 remains in contact with the cup spring 1104.

 The release path of the clutch in the region of the tip 1104 from the protrusions is selected mainly in such a way that when the clutch is disengaged, the outer edge of the disk spring 1104 is slightly raised from the ring 1126. This means, therefore, that when the friction clutch 1101 is disconnected from the disk spring in the diametrical loading region the pressure disk with a cup spring 1104 is larger than the stroke 1124 of the separation of the pressure disk 1103 set by the travel restrictive means 1122.

 Presented in FIG. 29, the relative position of the individual structural parts corresponds to the new state of the friction clutch. When axial wear, in particular friction plates 1107, the position of the pressure plate 1103 is shifted in the direction of the back pressure plate 1006, as a result of which there is primarily a change in the taper and thereby also the pressure exerted by the cup spring when the friction clutch 1101 is engaged, namely, preferably in the sense increase. This change makes it possible that the pressure disk 1103 changes its axial position with respect to the wear sensors 1122 axially supported on the back pressure plate 1106. Due to the force of the Belleville spring acting on the ring 1126, this ring 1126 follows the axial movement of the plates due to the axial movement of the back pressure plate 1103, as a result of which the thrust sections 1132 of the ring 1126 are axially detached from the employees as the counterpart of the sections in the form of cams 1133 of the wear sensors 1122, namely, by an amount that basically corresponds to wear of the plates. The expansion ring 1126 maintains its axial position during the engagement process with respect to the pressure disk 1103 because it is loaded with a cup spring 1104 in the direction of the pressure disk 1103, the wear compensating device 1134 is self-braking during the adhesion process, and therefore acts as an axial stop. When splitting the friction clutch 1101, the compression disk 1103 is loaded with leaf springs 1109 in the direction of the housing 1102 and is displaced until the feeler gauges 1122 come into contact with the housing 1102 along with the abutting portions 1123 of the housing. Prior to this release stroke, corresponding to the tear-off course of the pressure disk 1103, the axial position of the ring 1126 of the relative pressure disk 1103 is maintained. When the release process continues, the pressure disk 1103 axially stops, against which the ring 1126 follows the axial release trip of the disk spring in the region of the load diameter, and namely, until the abutting portions 1132 of the ring 1126 come into contact with the counter-abutting portions 1133 of the wear sensors 1122 again. The axial displacement of the compensation ring 1126 is carried out by wedges 1135, which are loaded with springs 1144. These wedges 1136 are displaced relative to the wedges 1136 in the circumferential direction until the ring 1126 is pressed against the counter supports 1133 of the wear sensors 1122. The release of the pressure disk 1103 in the illustrated embodiment is provided through leaf springs 1109, which are thus installed between the housing 1102 and the pressure disk 1103, so that they have an axial pre-tension that presses the pressure disk 1103 in the direction of the housing 1102. If the disk spring 1104, in addition , deviates in the direction of disengagement, then it is torn off by its radially outer section from the adjusting ring 1126, since the latter, as already described, is axially held by the wear sensors 1122 relative to eniyu the pressing disk 1103. Such at least a slight separation diaphragm spring 1104 relative to the adjusting ring 1126 during the disengagement process is particularly advantageous for the functioning of the regulating system 1117 + 1134.

 Defined by the invention, the adjustment system 1117 + 1134 ensures that the adjustment on the support ring 1126 by means of the adjustment wedges 1135, 1136 is always carried out in accordance with the amount of wear of the plates. This should be explained by the fact that the adjusting ring 1126 is axially clamped between the adjusting means in the form of wedges 1135, 1136 and the wear sensors 1122, as a result of which it is prevented that the compensating structural part in the form of the ring 1126 is rearranged by an amount greater than the corresponding wear of the plates. In addition, the calculation of the adjusting system defined by the invention ensures that also when shifting in the field of disengaging means, such as the tab spring protrusions 1104b, or during axial vibrations of the pressure plate, the adjustment system 1117 + 1134 cannot be shifted, since wear sensors 1121 also with a rigid support on the housing 1102 relative to the clamping disk 1103 are axially supported by a self-braking wear-compensating device 1134, namely, through the counters 1132. Thus, in a disengaged state clutch wear on the sensors 1122 may apply axial forces in the direction counter-pressure plate 1106, which are greater than the positive connection between the wear sensor 1122 and the pressure plate 1103 without that the wear sensors are displaced axially relative to this pressing disc 1103.

 The adjustment system defined by the invention ensures that, over the entire service life of the clutch, the Belleville spring operates in almost the same region of the characteristic and when the friction clutch 1101 is engaged, it has a practically constant clamping position and thereby also produces a practically constant clamping force on the clamping disk 1103. Thanks to it is possible to use a disk spring with a degressive graphical characteristic through the trip, namely, mainly in combination with one the clutch disk, the plates 1107 of which are spring-loaded to each other through the spring segments 1110, as a result of which the effective applied pressure can be brought to a relatively low level and during the service life of the clutch, since the characteristic of the plate springs changes slightly during the service life of the clutch constant. When one such clutch is disengaged, the Belleville spring 1104 deviates around its support 1105, and in one predetermined section of the axial stroke of the release of the pressure disk 1103, the spring segments 1110 are unclenched and thereby the axial force produced by the spring segments 1110 supports the process of disengaging the friction clutch 1101. Thus, this means that a lower maximum tripping force should be produced than that which, in the engaged position of the clutch 1101, theoretically follows from the mounting position of the cup spring 1104 and leaf springs 1109. As soon as the stretch of springing or the extraction of segments 1110 is exceeded, the friction plates 1107 are released, and on the basis of the degressive region of the characteristic in which the disk spring 1104 operates, then the applied decisive force is already significantly reduced with respect to that which would correspond the installation point or the installation position according to figure 1129. When the uncoupling process continues, the release force is lowered, namely, until a minimum or lower point predominantly sinusoidal graphical characteristics of the Belleville spring 1104.

 The adjustment system 1117 + 1134 shown in FIGS. 28 and 29 can advantageously be designed in such a way that when the friction clutch 1101 rotates, individual spring coils of the adjustment springs 1144 are supported on the outer wall 1131 of the adjustment ring 1126, and the regulating forces produced by the springs 1144 in the circumferential direction created between the threads of the springs and the adjusting ring 1126, the friction resistance is reduced or completely destroyed. Thus, the springs 1144 during rotation of the friction clutch 1101 due to supporting the action of the springs of the friction forces can be almost stationary. In addition, the adjusting wedges 1135 as a result of the centrifugal forces acting on them can also be radially supported on the wall 1131 of the adjusting ring 1126 and are protected against rotation by means of the frictional forces created between the wedges 1135 and the adjusting ring 1126. Due to this, it can be realized that, at least at speed ranges above the idle speed of the internal combustion engine, the wear-compensating device 1134 is not rotated by the springs 1144. Thus, the friction clutch 1101 can be laid out in such a way that the wear of the friction linings is compensated only when triggered the friction clutch 1101 at idle speed or at least approximately at idle speed. The locking of the adjusting ring 1126 can be carried out by means of the corresponding calculation of the wear-compensating device 1134, however, also in such a way that only when the internal combustion engine stops, therefore, when the friction clutch 1101 is not rotating or at very low speeds, adjustments to the wear of the plates can occur. The pair of materials between the slope-forming parts 1135, 1136, as well as the material of the structural parts interacting with them, are preferably selected so that during the service life of the friction clutch, the clutch preventing adjustment between the clones and the interacting remove structural parts cannot occur.

 To avoid such adhesion, at least one of these parts may be coated at least in the area of slopes or abutment surfaces. In order to avoid adhesion between ramps and ramps, the motet also has at least one safety device which, when the friction clutch is disengaged or when adjusting wear, exerts axial force on one or several adjusting elements that divide or break the ramps.

 In the new state of the friction clutch 1101, therefore, in the state that the clutch has, before laying the disc 1108, it is fixed on the back pressure plate 1106, the wedges 1135 are in relation to the position shown in FIG. 30 in a position farther back relative to the wedges 1136, so that the adjusting ring 1126 has its maximally extended position in the direction of the pressure disk 1103 and, thus, the pressure disk / adjustment ring assembly 1126 needs the smallest axial construction space. In order to hold the wedges 1135 before mounting the friction clutch 1101 in their drawn position, the wedges 1135 have working sections in the form of recesses 1152 for pivoting or supporting means. Such kind of support means can be provided during assembly of the friction clutch 1101 and removed after mounting the friction clutch 1101 on the flywheel 1106, as a result of which the adjusting device 1134 is activated. In the presented embodiment, provided, as follows from figures 30 and 32, in the adjusting ring 1126 in circumferential direction passing elongated grooves 1153, through which the working sections of the supporting means or the same torsion working body to enter the recesses 1152 can be passed. in the circumferential direction, the elongated recesses 1153 must have at least one extent that corresponds to a rotation in accordance with the greatest possible angle of wear control of the wedges 1135 in the circumferential direction. The wedges 1135 held in the new state of the friction clutch in the circumferential direction in their drawn position can be held in this position by wear sensors 1122 that support the adjusting ring 1126 in its drawn position. Self-regulating connections between the wear sensors 1122 and the pressure plate 1103 should be made in such a way that the necessary displacement force to bias the wear sensors 1122 relative to the pressure plate 1103 is greater than the resulting force exerted on the ring 1126, which is produced by the springs 1144 loading wedges 1135.

 In the exemplary embodiment of FIG. 28 and 29, the slopes 1143 can also be formed directly by the ring 1126, for example, by embossing the oblique surfaces 1143, and the ring 1126 should then be rotated relative to the pressure disk 1103 by the springs 1144. The wedges 1136 in this embodiment cannot rotate with the pressure disk 1103 or obtained directly on it by plastic deformation. In addition, with this kind of embodiment, the stops made in the form of overlays 1132 should extend in the circumferential direction according to the necessary adjusting torsion angle of the ring 1126 to ensure that there is a restriction between the wear sensors 1122 and the ring 1126 during the life of the friction clutch. With the last embodiment described, the adjusting ring 1126 and with the friction clutch 1101 mounted can be rotated in a simple manner radially from the outside, namely, in particular through the circumferential thrust plates 1132, which are accessible through the radial passes provided on the outer side surface of the clutch housing 1102. These radial passages can take, in particular, also torque-transmitting cams 1125 of the pressure plate 1103 as well as leaf springs 1109. The adjustment device defined by the invention also has the advantage that its principle can also be applied to so-called extended friction clutches in which the cup spring is deflectively supported by one outer edge portion and supported on the lid, and the pressure disk is loaded radially further inside the edge portion. One such clutch is shown in FIG. 33. A wear-compensating device 1234 is provided between the cup spring 1204 and the pressure disk 1203, which can be made in the same way as described in connection with FIGS. 28 and 29. The adjusting ring 1226, in turn, via the sensor rings elements 1217 interacts with wear sensors 1222. Rearrangement of the wear sensors 1222 relative to the pressure plate 1203 is carried out by abutment of the abutment sections 1222a to the housing or the cover 1202. The wear sensors 1222, in turn, carry stops 1233, which limit the axial stroke of the pressure plate 1203 during the disengagement process. For the smooth functioning of the adjusting device according to FIG. 33, ring 1226 is capable of at least one axial movement with respect to wear sensor 1222. This can be done due to the fact that there is a corresponding abutment connection 1233a with a gap between the wear sensors 1222 and the ring 1226 or also due to the fact that the ring 1226 has radial sections 1226a that are elastically deformable in the axial direction, therefore, have elastic flexibility.

 In the embodiment shown in FIG. 34, the sensing elements 1317 are based directly on the main body of the pressure disk 1303. The wear sensors 1322 have thrust portions 1322a that cooperate with the cap portions 1323 forming the counters. The sections 1323 are made integrally with the fastening means 1302a through which the Belleville spring 1304 is rotatably mounted on the cover 1302. In the illustrated embodiment, the fastening or fixing means 1302a are formed as a whole from the cover material of the covers that extend axially through the Belleville spring 1304. A wear compensating device 1317 is provided radially outside of the sensing elements 1317 provided in the radial region of the disk spring main part 1304a.

 By the implementation of a single friction clutch as defined by the invention, not only can the clutch service life be increased due to the use of thicker friction clutch plates, therefore, the axial wear volume of the clutch plates can be increased, but also, in particular, the reduction of the tripping forces, namely, by using a single power accumulator with degressive during the course of the friction clutch disengagement, the force – path relationship in combination with at least one opposing power accumulator Yator acting on the pressure plate, a spring element, which, when coupling and uncoupling the clutch produces gradual rise or drop in the friction clutch torque transmitted via at least one portion of the working stroke clutch disc or the pressing stroke. This spring element is specially connected in series with a compression spring, such as, for example, a disk spring, of a friction clutch. Thus, by the implementation of one friction clutch as defined by the invention, a rather significant reduction in the breaking force can be achieved and this reduction is maintained or does not change during the service life of the friction clutch, which means in a relatively narrow tolerance field. In addition, in the friction clutch according to the invention, a tolerance field with a relatively steep force-path characteristic in the working area can be found. Such disk springs with conditional couplings would lead to a very strong rise in the tripping force during wear of the plates.

 In couplings without regulation defined by the invention, with increasing wear of the plates, the point 41 (FIG. 8) corresponding to the adhered state of the friction clutch first moves along line 40 in the direction of maximum 41a. To this point 41a, during one disengagement process, a drop in the release force occurs, however, collectively, the passage of the release force increases with respect to the passage of the release force in the new state of the friction clutch. Thus, the portion 43 moves to the left until the point 41 coincides with the maximum 41a. Point 44 shifts accordingly along characteristic 40. With further wear of the plates, the insertion point of the control disk spring corresponding to the friction clutch engagement is gradually displaced from the maximum 41a in the direction of point 41b, so that the compression force exerted by the disk spring gradually decreases. The clamping force exerted at point 41b by the actuating cup spring corresponds to the clamping force exerted at point 41 in the new state of the friction clutch. As soon as the maximum 41a is exceeded, during the disengagement process in at least one portion of the working path of the friction clutch, the pressing force increases. When the maximum allowable wear path or wear point 41b is reached, there is a rise in clamping force throughout the trip. This increase in the release force is also maintained when, as shown in FIG. 8, there is a spring of the plates or a substitute for the spring of the plates.

 When calculating the friction clutch and, in particular, its adjusting device, it should be taken into account that the main shaft of the internal combustion engine excites axial and oscillating vibrations on the flywheel, which are also transmitted to the friction clutch fixed to the flywheel. So that the friction clutch or the adjusting device can function flawlessly, this means that there is no undesirable rearrangement (regulation) due to such fluctuations, in the embodiment according to figures 1-27, therefore, in general, in embodiments with one adjusting device with a force sensor, the regulating force of this power sensor must be greater than the inertia forces that can act on this power sensor. These forces arise, in particular, due to the masses of the main disk spring, the adjusting ring or adjusting elements, one corresponding mass fraction of the power sensor and, if necessary, due to the mass of other structural parts with multiplication with the maximum possible axial acceleration of these parts or elements, which stems from axial and bending vibrations of the flywheel. So, for example, in the embodiment of FIG. 27, in which the sensor disk spring 1013 is supported on the clutch pressure plate 1003, the inertia of this clutch pressure plate 1003 must be taken into account. Therefore, it must always be ensured that the force exerted by the sensor spring is greater than the forces acting on it, which are formed by the mass of structural parts multiplied with the maximum possible acceleration, which, due to their inertia, act on the sensor spring. These inertia forces can adversely affect, in particular, during actuation of the friction clutch, in particular when the friction clutch is disengaged.

 In the embodiments according to figures 29-34, when laying out the wear sensors and wear-compensating devices, the forces arising due to the inertia of the individual structural parts and the axial and torsional vibrations acting on them should be equally taken into account.

 Thus, in general, when calculating the friction clutch with integrated wear-compensating means, the masses of elements to which axial and rotational vibrations can be transmitted and which act on the compensating means should be taken into account as necessary. In the examples of execution according to figures 28-34, in particular, details that affect the functioning of the tilting mechanism must be taken into account.

 Figure 35 shows in detail a flywheel 1401 that has one first or primary flywheel 1402 secured to an internal combustion engine not shown on the main shaft of the internal combustion engine, as well as a second or secondary flywheel mass 1403. One friction clutch 1404 is fixed to the second flywheel 1403 when one disc is inserted 1405, through which a gear not shown can also be engaged and disengaged. The flywheel masses 1402 and 1403 are mounted rotatably to each other through one support 1406, which is located radially inside the holes 1407 for holding the fixing screws 1408 for mounting the first flywheel mass 1402 on the output shaft of the internal combustion engine. Between the two fly masses 1402 and 1403 there is a damping device 1409, which has compression coil springs 1410 located in an annular cavity 1411 that forms one toroidal section 1412. The annular cavity 1411 is at least partially filled with a viscous medium, such as, for example, oil or grease and is located radially outside the extreme friction diameter of the coupling disk 1405, while having a larger diameter than the diameter of the coupling disk 1405.

 The primary fly mass 1402 is formed predominantly by the part 1413, which was made of sheet material. Structural part 1413 has one, basically radially extending, flange portion 1414 that carries radially inside one axial nozzle 1415 that is formed integrally, which is surrounded by drilled or punched holes 1407. The single-row rolling bearing 1406a of the rolling bearing 1406 with its inner ring 1416 is radially outside based on the end portion of the axial nozzle 1415. The outer ring 1417 of the rolling bearing 1406a carries a second flywheel made essentially as a flat disk-shaped body. For this, the fly mass 1403 has one central recess in which the bearing 1406a is based. Basically, the radially extending portion 1414 extends radially from the outside into a cup-shaped portion 1418, which at least partially wraps around the power batteries 1410 at least along their outer circumference and conducts or supports. The shell-shaped element 1419 fixed at section 1418 partially covers the circumference of the power accumulators 1410. The element 1419 is welded to the sheet element 1413 (at 1420). The toroidal section 1412, when viewed in the circumferential direction, is subdivided into separate bases in which power accumulators 1410 are provided. Separate bases, when viewed in the circumferential direction, are separated from each other by load sections for power accumulators 1410, which can be formed by pockets pressed into the sheet part 1413 and the cup-shaped element 1419. Combined with the second fly mass 1403, the load sections 1403 for the power accumulators 1410 are carried by the coupling cover 1422.

 The loading sections 1421 are formed by radial levers 1421, which are mounted on the axial section 1423 of the clutch cover 1422 in the illustrated embodiment and extend radially into the annular cavity 1412, namely, between the ends of the circumferentially adjacent power accumulators 1410. The axially extending section 1423 of the cover covers or covers the second fly mass 1403 with one segment 1423a and is rigidly connected to it, for example, through embossments introduced into the 1423a section, which enter the corresponding recesses of the fly mass 1403, or by means of another type of mount.

 The coupling cover 1422, centered on the outer contour of the fly mass 1403, at its end facing away from the loading sections 1421 has one mainly radially extending, ring-shaped section 1426, on which a Belleville spring 1427 acting as a two-arm lever is fixed, with a disc-shaped disk lying radially further from the outside spring 1427 loads the pressure plate 1428, as a result of which the friction plates 1429 of the coupling disk 1405 are axially clamped between the second fly mass 1403 and the pressure plate 1428. M forward friction plates 1429 provided the covering spring back in 1465.

As can be concluded from figure 35, the annular cavity 1311 or its toroidal section 1412 is located predominantly radially outside the extreme contours of the second flywheel mass 1403. Due to this, it serves for articulating the first flywheel mass 1402 with the output shaft of the internal combustion engine and bearing the toroidal section 1412 structural part 1413, which borders on the internal combustion engine, and the second flywheel mass 1403 radially inside the annular cavity 1411 at a relatively greater radial extent They can be located almost directly opposite each other, forming a gap or an air gap 1430, therefore, they can be practically adjacent with a small interval, as a result of which a very axially compact construction of an assembly consisting of a flywheel 1401, a clutch 1404 and a clutch plate 1405 becomes possible. . Sealing the annular cavity 1411 is provided by a seal 1431 active between the inner portions of the radial wall 1419 and the outer side surface of the cover 1422
Advantageously, this gap 1430 can serve to cool the flywheel 1401, namely while a stream of cold air is conducted through this gap. To obtain such cooling air circulation, the second flywheel mass 1403 has axial recesses 1433 radially inside the friction surface 1432, which extend in the direction of the structural part 1413 located on the engine side and flow into the intermediate cavity 1430. To improve cooling, the second flywheel mass 1403 may have other axial passages 1435, which lie radially further outside and on the side facing away from the friction surface 1432, are connected to the intermediate volume 1430, and on the side of the flywheel facing the coupling 1404, they exit 1433 radially outside the friction surface 1432. Radially inside the passages or recesses 1433, the flywheel 1403 has other passages 1434, which serve, in particular, for basing or holding the fixing screws 1408.

 To seal the annular chamber 1411, partially filled with a viscous medium, another seal 1436 is provided, which is formed by a part made in the form of a membrane or a disk spring, which extends radially in the intermediate space 1430.

 The cup-shaped housing 1419 carries a starting gear ring 1439, which is connected to it through a welded joint.

 Together with a coupling assembly consisting of a coupling 1404 and a disc 1405, shown in FIG. 35, the dual-mass flywheel 1402 + 1403 forms one assembly A, which as such is pre-assembled, so that it can be sent and stored and screwed onto the main shaft of the internal combustion engine in a simple and rational way. To assemble the assembly A, first of all, the coupling 1404 and the second flywheel 1403, when laying the coupling disk 1405, are connected to each other. After that, the subunit, consisting of the coupling 1404, flywheel 1403 and the disk 1405, is axially reduced to the structural part 1413, after which the cup-shaped housing 1419, which is based on the outer edge 1423 of the coupling cover 1422, can be brought into contact with the extreme regions of the part 1413 and welded with her (at 1420). Before the axial mixing of both structural parts 1413 and 1419, the springs 1410 were inserted into the toroidal region 1412. The structural unit A also has, with integration, a bearing 1406, which is inserted along the axial nozzle 1415. In addition, already in the holes 1407 of the flange section 1414 pre-mounted or still contained mounting screws 1408, namely in the form of screws 1408 with an internal hexagon. In this case, the screws 1408 are in the position, respectively, the lower half of the figure 35. The screws in this position are firmly held in the unit or node A.

 The coupling disk 1405 is clamped in a position pre-centered on the axis of rotation of the main shaft between the pressure plate 1428 and the friction surface 1432 of the second flywheel mass 1403 and, moreover, in such a position that the holes 1443 provided in the coupling disk are in such a position that during installation unit A on the output shaft of the internal combustion engine, a screw tool can move through them. Obviously, the holes 1443 are smaller than the heads 1440 of the screws 1408, so that an excellent and reliable fixation of the screws 1408 is ensured in unit A.

 Also in the disk spring 1427, namely, in the region of its protrusions 1427a, holes or cutouts 1444 are provided for the passage of the screw tool. The cutouts 1444 may be provided so that they form extensions or extensions of the grooves between the protrusions 1427a. The holes 1444 in the disk spring 1427, 1443 in the disk 1405 and 1434 in the fly mass 1403 overlap in the axial direction, namely, in such a way that with the required positioning of the assembly of the assembly A on the main shaft, the asymmetrical arrangement of the holes 1407 mounting a tool, such as, for example, a key with an internal hexagon, can perfectly pass through holes 1444, 1427 and 1443 and engage into recesses of heads 1440 of screws 1408.

 This type of combination unit A facilitates the installation of the flywheel, since various technological operations disappear, such as otherwise necessary centering of the coupling disk, insertion of the coupling disk, fitting of the coupling, insertion of a centering mandrel, centering of the disk itself, insertion of screws, as well as screwing couplings and centering mandrel removal.

 The friction clutch 1404 has an adjustment device 1445, which, similar to that described in connection with figures 1-27, provides wear compensation using the sensor spring 1446 and the adjusting ring 1447.

 The torque transmission device 1501 shown in FIG. 36 has a counter-pressure plate 1503 that is not rotatably connected to the main shaft K of the internal combustion engine and is fitted with a friction clutch 1504 when laying the disk 1505. The disk 1505 is based on the input shaft of a transmission not shown in more detail.

 The coupling cover 1522 has an axially extending portion 1523 which radially axially extends axially from the pressure plate 1528 and the friction plates 1529 of the coupling disk 1505. The end portion 1523a of the sleeve-shaped or pipe-shaped region 1523 of the lid covers or covers the backpressure plate 1503 and is rigidly connected to it through the embossments 1524 introduced into the embossing section 1524a, which enter the recesses of the backpressure plate 1504 provided on the outer circumference. The cover 1522 and the backpressure plate 1503 can, however, also be connected in another way, for example, via welded joints, by means of joints with screws or pins, which are also inserted predominantly in the radial direction.

 The coupling cover 1522, centered on the outer contour by the backpressure plate 1503, has one annular section 1526 that extends substantially radially inward, on which a disk spring 1527 acting as a two-arm lever is mounted. the plates 1529 are axially clamped between the back pressure plate 1503 and the pressure plate 1528. The cup spring 1527 has radial protrusions 1527a for engaging coupling 1504 through a trip system. To transmit torque between the pressure plate 1528 and the cover 1522, torque transmission means are provided, preferably in the form of leaf springs 1521, which are rigidly connected at one end to the cover 1522, preferably through rivet joints 1521a, and at their other end also through the rivet connection, to the pressure plate 1528. Preferably, the connection between the pressure plate 1528 and the leaf spring elements 1521 is through the so-called blind rivet connection, as it appears area claimed in the radial passage in the upper half of figure 35.

 The friction clutch 1504 or the torque transmission device 1501 has an adjustment device 1545, which, similarly to that described in connection with figures 1-27, by means of one sensitive spring 1546 and one adjustment ring 1547 provides wear compensation.

 In FIG. 35 and FIG. 36, the oncoming counterclones pressed into the cover material are made in such a way that they form, in the direction of rotation of the friction clutch, a single hole (1547a in FIG. 36) for air passage. Through one such embodiment, when the corresponding friction clutch is rotated, better cooling of the latter is achieved due to forced air circulation. In particular, thanks to this, the adjustment ring 1447 or 1547 made of plastic is also cooled, as a result of which the thermal load of this ring can be significantly reduced.

 The friction clutch 1504 or the counter-pressure plate 1503 is non-rotatable, but with a possibility of limited axial displacement, mounted on the output shaft K through an elastic or spring-flexible structural part 1550. In the presented embodiment, this structural part 1550 is made in the form of a cup, and its rigidity has such a value that axial and torsional or bending vibrations excited through the output shaft K on the friction clutch 1504 are damped by the elastic structural part 1550 or are sized to ensure that the friction clutch 1504 and, especially, its adjusting device function perfectly. Thus, through the axially compliant part 1550, the maximum possible separation of the coupling assembly 1504 with respect to the axial and bending vibrations of the output shaft, such as the main shaft, of the internal combustion engine, should occur. This can eliminate the fact that the coupling unit 1504 or its adjusting device 1545 is harmful to their functioning. Without the aforementioned disconnection of the coupling assembly 1504 relative to the main shaft K, an undesirable adjustment of the wear control device 1545 could occur, namely, due to the mass of structural parts and accelerations acting on the latter as a result of vibrations. Thus, without filtering vibrations of the structural part 1550, in particular when calculating the adjusting device 1545, the inertia forces produced by the components of this adjusting device would have to be specially taken into account, as a result of which expensive adjustment and / or additional means would be necessary to avoid what is not explained by wear plates adjusting the wear control device 1545.

 In the torque transmitting device 1501 according to FIG. 36, the wear adjusting device 1545 acts between the clutch cover 1522 and the cup spring 1527. However, the torque transmitting device 1501 could be provided with a friction clutch according to FIGS. 28-34, hence a friction clutch in which the wear adjusting the device acts between a cup spring and a pressure plate loaded by it.

 The backpressure plate 1503 is radially externally through screw connections 1551 rigidly connected to the axially elastic disk-shaped structural part 1550. Instead of threaded connections 1551 blind blind rivets could also be used, such as those shown in the upper half of Figure 35 in connection with the fastening of leaf springs on the pressure plate 1428. Radially inside the connection points 1551 between the disk-shaped part 1550 and the back pressure plate 1503 there is an axial clearance 1552 between these two parts 1550 and 1503, which is determined t the maximum amplitude of the axial vibrations between the two parts 1550 and 1503 in the same axial direction. By stopping the radially inner portions of the back pressure plate 1503 on the disk-shaped part 1550, the maximum axial displacement of the friction clutch 1504 in the direction of the output shaft K of the internal combustion engine can be limited. Under normal operating conditions, in particular with a perfectly functioning internal combustion engine, this kind of contact, however, does not occur. An annular backpressure plate 1503 covers one axial protrusion 1553, which is part of an annular or disk-shaped structural part 1554. This disk-shaped structural part 1554 can be rigidly connected to the radially inner portions of the elastic disk 1550. The elastic disk 1550 and the disk-shaped structural part 1554 are centered one at a time the annular protrusion 1555 of the shaft K and rigidly connected to it through threaded connections 1556. In this case, the radially inner portions of the disk-shaped part 1550 are axially clamped between one end 1557 of the shaft K and the annular structural part 1554.

 The axial nozzle 1553 of the part 1554 has at its end facing away from the elastic part 1550 radial sections 1558 that limit the axial displacement of the friction clutch 1504 or the back pressure plate 1503 in the other axial direction. Between sections 1558 and the back pressure plate 1503 with an unloaded elastic part 1550 there is an axial backlash 1559. This clearance 1559 in the axial direction is the same size as the backlash 1552. The back pressure plate 1503 over its entire inner lateral surface can be based practically without a gap on the axial nozzle or a protrusion 1553, so that the axial direction of the back pressure plate 1503 can thereby be ensured. However, it may be appropriate when there is at least a slight air gap between the inner side surface of the backpressure plate 1503 and the axial nozzle 1553, therefore, under normal operating conditions, there is no contact between these two structural parts.

 According to another embodiment, for damping vibrations still transmitted, despite the resilient structural member 1550, a safety device that destroys vibrational energy can be provided. Such a safety device can be formed by a friction joint, such as it is represented, for example, in FIG. 37. In the embodiment shown in FIG. 37, damping means 1560 are provided between the inner portions of the backpressure plate 1503 and the outer side surface of the nozzle 1553, which can be formed by a single circumferentially wavy ring, the waves of which extend radially. This ring 1560 is mounted with a clamp in the radial direction, as a result of which friction is created between this ring 1560 and the inner side surface of the back pressure plate 1503 in the presence of axial vibrations. Thus, the damped support of the back pressure plate 1503 on the nozzle 1553 occurs. The corrugated ring 1560 can be divided along its circumference, therefore, open.

 The disk-shaped resilient structural part 1550 is radially externally provided with a starting gear ring 1561.

 The disk-shaped part 1550, the back pressure plate 1503, the disk 1505 and the friction clutch 1504 form one assembly that is pre-assembled as such, so that it can be sent, stored and screwed onto the main shaft K of the internal combustion engine in a simple and rational way. The fixing screws 1556 in the form of hexagon screws are also pre-mounted, therefore, are contained in the assembly, namely, reliably.

 The friction clutch 1505 is clamped in a position pre-centered on the axis of rotation of the main shaft between the pressure plate 1528 and the back pressure plate 1503 and, moreover, in such a position that the holes 1562 provided therein, which are provided radially inside the spring damper of the clutch disk 1505, are in such the position that when mounting the unit on the shaft K the screw tool 1563 can pass. Also, the disk spring 1527, if necessary, has holes or cutouts 1564 for holding the screw tool ment 1563. The holes or cutouts 1564 of the Belleville spring 1527 are overlapped by the holes 1562 of the disk 1505, so that an assembly tool, such as, for example, an Allen key 1563, can perfectly fit into the profiling of the screw heads 1556.

 As already described in connection with other embodiments, by using a friction clutch 1504 with a safety device 1545 that compensates for at least wear of the plates, optimization can be made in the layout of the friction clutch, in particular the power accumulator exerting a clamping force on the clutch disc. This power accumulator can be laid out in such a way that it creates practically only the clamping force of the clutch disc necessary for transmitting the desired torque. The adjusting device 1545 ensures that the power accumulator 1527 in the engaged state of the coupling unit 1501 maintains practically the same installation position during the entire service life. In addition, there is also an adjustment device in the coupling assembly 1501 in the form of a cover spring 1565, which, when the assembly 1501 is disengaged or engaged in at least one portion of the working path of the pressure plate 1528, determines a gradual decrease or increase in the torque transmitted by the assembly, a reduction in the release forces can be achieved or release characteristics. Thus, by appropriate adjustment between the forces created by the safety device, such as the springing of the plates, and the force acting on the pressure plate or the force-force characteristics, the desired passage of the release force graph can be determined. Due to this, it is also possible to calculate the elastic part 1550 more optimally with respect to the desired damping of axial, bending or torsional vibrations, since the reduced decoupling forces acting on this elastic part have a secondary (subordinate) value. Thus, the working forces necessary for disengaging the coupling can be maintained without significant axial displacement of the coupling unit by means of the structural part.

 The invention is not limited to the presented and described examples of execution, but also covers options that can be formed by a combination of individual features described in connection with various forms of execution or elements. In addition, individual signs described in connection with the figures, or the principles of action, taken by themselves, can represent an independent invention.

Claims (24)

 1. A friction clutch (1) comprising a pressure plate (3), a friction disk (8), a pressure disk (6), a pressure spring (4), a housing (2), and the pressure disk (3) is not rotatable with the possibility of limited axial bias is connected to the housing (2), and between the housing and the pressure plate (3) is installed at least one compression spring (4), which presses the pressure plate (3) in the direction of the friction disk (8) clamped between it and the pressure disk (6) ), and the clamping disk (6) is a flywheel, characterized in that the coupling (1) is equipped with an adjustment device three (16) to compensate for the wear of the friction linings of the friction disk and provide at least approximately unchanged pressing of the pressure disk (3) by the spring (4), while the friction clutch (1) is equipped with actuating means (46) for coupling and disengaging the coupling, as well as safety device (10), which is designed to ensure during the release of the clutch (1) on one of the sections of the actuator actuator (46) the progressive reduction of the moment transmitted by the friction clutch (1).  2. The clutch according to claim 1, characterized in that the safety device (10) is designed to ensure during the disengagement of the clutch (1) in one of the axial displacement sections of the pressure plate (4) loaded by the compression spring (4) a gradual decrease in the transmitted friction coupling (1) of the moment.  3. The clutch according to claim 1 or 2, characterized in that the safety device (10) is located between the actuating means (46) or the pressure spring (4) and the sections (6a) for mounting the housing (2) on the pressure disk (6).  4. The clutch according to one of claims 1 to 3, characterized in that the safety device (10) is located between the actuating means (46) or the pressure spring (4) and the friction surface of the pressure plate (3).  5. The coupling according to one of claims 1 to 4, characterized in that the safety device (10) is located between the friction plates (7) of the disk (8) of the coupling in the axial direction.  6. The coupling according to one of claims 1 to 5, characterized in that the safety device (10) allows axial displacement of the aforementioned structural elements (6), (7) and (3) of the coupling (1), and the safety device ( 10) it is located so that when the clutch (1) is disconnected, the force acting on the safety device (10) is the smallest, and when the clutch (1) is turned on, the force acting on the safety device (10) gradually rises to the maximum, and this rise occurs in at least one area of the working ho a coupling of (1).  7. The coupling according to one of claims 1 to 4, characterized in that the safety device (10) has the ability to produce a gradual decrease or gradual increase in torque transmitted by the friction clutch (1) by at least 40 - 70% of the length of the actuator funds (46).  8. The clutch according to one of paragraphs.1 to 7, characterized in that the clamping spring (4) at least on one part of the working stroke disconnect the friction clutch (1) has a regressive elastic characteristic.  9. The clutch according to one of claims 1 to 8, characterized in that the compression spring (4) is a disk spring, which is mounted to rotate around an annular swivel bearing mounted on the housing (2), and rests on one side of the pressure disk (3).  10. The clutch according to claim 9, characterized in that the disk spring has an annular part (4a), from which radially inwardly directed protrusions (46) form the actuating means.  11. The clutch according to claim 9 or 10, characterized in that the disk spring (4) is mounted rotatably between two supports (11) and (12) on the housing (2).  12. The clutch according to one of claims 8 to 11, characterized in that the disk spring (4) has a sinusoidal elastic characteristic (Fig. 8) and when the friction clutch (1) is turned on on the elastic characteristic (Fig. 8) of a disk spring (4) ) there is a working point on the regression section of the elastic characteristic following the first maximum (41a), while the ratio between the first maximum (41a) and the minimum following it (45) is 1: 0.4 - 1: 0.7.  13. A car with a friction clutch, including a clutch disengagement system (1) designed to control the clutch (1), the disengagement system includes a clutch pedal made similar to the accelerator pedal, characterized in that the friction clutch is made according to one of paragraphs.1 - 12, and one of the elements of the trip system engages with actuating means (46) of the friction clutch.  14. A coupling assembly (1501) comprising a housing (1522), a pressure plate (1528), a friction disk (1505), a pressure plate (1503), a pressure spring (1527), while the pressure plate (1528) is not rotatable with the possibility of limited axial bias is connected to the pressure plate (1503), with at least one pressure spring (1527) mounted on the housing (1522) and presses the pressure plate (1528) in the direction of the friction disk (1505) clamped between it and the pressure disk (1503) characterized in that the coupling unit (1501) is equipped with an adjusting safety device (1545) for to compensate for the wear of the friction plates (1529) of the friction disk and to ensure at least approximately constant pressure of the pressure plate (1528) by the compression spring (1527), the coupling unit (1501) is equipped with actuating means (1527a) for coupling and uncoupling the coupling unit, as well as a safety device (1656) with at least one spring means arranged in series with the compression spring (1527) for carrying out during the uncoupling process of the coupling unit in one of the sections of the actuator operating stroke (1527a) gradually reducing the torque transmitted via the friction disk (1505).  15. A coupling assembly (1401) comprising a housing, a pressure disk (1428), a friction disk (1405), a pressure disk (1403), a pressure spring (1427), a torsional vibration damper, while the pressure disk (1428) is not rotatable with the possibility of limited axial displacement is connected to the pressure plate (1403), while the compression spring (1427) presses the pressure plate (1428) in the direction of the friction disk (1405) clamped between it and the pressure disk (1403), and it is also equipped with an adjusting safety device (1445) ) to compensate for the wear of the friction plates of the friction the disk (1405) and providing at least approximately constant pressure of the pressure plate (1428) with a compression spring (1427), characterized in that the coupling unit (1401) is equipped with actuating means (1427a) for coupling and uncoupling the coupling unit and a safety device (1465) which, during the release of the coupling unit on one of the sections of the actuator operating path (1427a), can gradually decrease the moment transmitted through the friction disk (1405), in addition, the clamping disk (1403) is the first part of the flywheel, the two parts of which are able to rotate relative to each other against the action of the torsional vibration damper installed between the two parts of the flywheel, the second part of the flywheel connected to the output shaft of the internal combustion engine.  16. The assembly according to claim 15, characterized in that the housing is connected to a part of the flywheel formed by the pressure disk (1403) through an indestructible detachable connection.  17. The assembly according to claim 15 or 16, characterized in that the torsional vibration damper (1410) is located radially outside the outer friction diameter of the friction disk (1405) of the clutch.  18. Coupling unit for an internal combustion engine comprising a friction clutch (1504) including a pressure plate (1528), which is inextricably rotated with limited axial displacement and is connected to a pressure plate (1503), and at least one pressure spring (1527) , which presses the pressure plate (1528) against the friction disk (1505) clamped between it and the pressure disk (1503), which functions as a flywheel, characterized in that the friction clutch (1504) is equipped as an adjustment device (1545) to compensate for the wear of the friction obk the clutch friction disk (1506) of the clutch (1504) and ensuring at least approximately constant pressure of the pressure disk by the clutch spring (1504), and actuating means (1527a) for coupling and disengaging the clutch and a safety device (1565), which during the disengagement of the clutch on one of the sections of the actuator actuator (1527a), gradually reduce the moment transmitted through the friction clutch (1504), and the friction clutch (1504) through the axially elastic element (1550) connected to the output shaft of the internal combustion engine, while the stiffness of the part (1550) is chosen so that axial, oscillating or bending vibrations excited by the output shaft of the engine on the friction clutch (1504) are damped or absorbed by the elastic element (1550) to the level that guarantees the best the functioning of the friction clutch (1504), in particular its adjusting device (1545).  19. The assembly according to claim 18, characterized in that the stiffness of the axial-elastic element (1550) is selected such that the control forces necessary for disengaging the coupling (1504) without significant axial displacement of the coupling unit are transmitted by the elastic element (1550).  20. The assembly according to p. 18 or 19, characterized in that the safety device (1565) is equipped with spring means, which is located in series with the pressure spring (1527).  21. The assembly according to one of claims 18 to 20, characterized in that the pressure disk (1503) is connected to the output shaft of the internal combustion engine through a torsional vibration damper.  22. A drive unit comprising a gearbox and a friction clutch (1) located between the drive motor and the gearbox including a pressure plate (3), which is inversely connected with a possibility of limited axial displacement to the pressure plate (6) connected to the output shaft the drive motor, at least one pressure spring (4), which presses the pressure disk (3) against the clutch disc (8) clamped between it and the pressure disk (6), which performs the function of a flywheel, characterized in that the gearbox is presented This is an automatic or semi-automatic gearbox, and the friction clutch (1) is equipped as an adjustment device (16) to compensate for the wear of the friction plates (7) and to ensure at least approximately constant force loading of the pressure plate (3) by means of the pressure spring (4), as well as actuating means (46) for coupling and uncoupling the friction clutch (1) and a safety device (10), which has the possibility during the disengagement process on one section of the working stroke of the actuating media tv (46) to implement a gradual decrease torque transmitted by the friction clutch (1).  23. The assembly according to claim 22, wherein the hold-down spring (4) in at least one portion of the disengagement stroke has a regressive elastic characteristic.  24. The assembly according to claim 22 or 23, characterized in that the safety device (10) has the ability to gradually decrease or gradually increase the moment transmitted by the friction clutch (1) by at least 40 - 70% of the working stroke of the actuating means (46 ) and / or pressure plate (3). Priorities for items:
11/26/91 according to claims 1, 2, 5, 6, 8 - 12 and 14;
04/18/92 according to claims 3, 4, 7 and 13.

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