SE522574C2 - Coupling unit with friction clutch coupling - Google Patents

Abstract

The pressure plate (3) is movably connected to a counter pressure plate (6). At least one plate spring presses the pressure plate towards a coupling disc (8) clamped between the pressure plate and the counter pressure plate. Controls for engaging and disengaging the friction clutch coupling are operated by an axially displaceable shifter (22) between which and the controls is a compensating control. The friction clutch coupling has a housing fixed to the counter pressure plate and which has a base facing the shifter.

Description

20 25 30 u u | n n nu 522 574 2 trouble-free operation of the wear of the friction pads compensating the adjusting device is given. Furthermore, the unit must be manufacturable in a particularly simple and cost-effective way.

According to the invention, this is achieved in that an axial spread of the position of the actuating means or the axial spread of the position of the sections of the actuating means affected by the disconnector in relation to the disconnecting or disengaging means equalizing device is provided. Such a device is particularly advantageous in clutch assemblies in which, according to a design of the invention, the actuators are displaced in the axial direction of the disengagement movement due to the wear of at least the friction pads, as a practically game-free power transmission can thereby be ensured between the disengagement and the disengagement means. This also guarantees that the operating means can always be moved with the same amount. Due to the force between the disconnector and / or the disconnection means and the operating means, practically no play can occur.

It can be particularly advantageous if the leveling device is arranged or acts axially between the switch-off and the operating means. However, the equalizing device can also be arranged in another place, e.g. effective between disconnectors and disconnectors. In connection with the present invention, it is advantageous if the disconnector is accommodated on a preferably alternately arranged axial control, such as e.g. a guide tube surrounding a shift input shaft.

Especially in the case of coupling assemblies with a friction coupling which has a lid which can be fastened to the back pressure plate, such as e.g. sheet metal lid, with a bottom facing the disconnector, it may be suitable if the disconnecting device is arranged or acts axially between the operating means and the bottom. Furthermore, it can be advantageous if the pressing spring is formed by a plate spring axially prestressable between a coupling lid and the pressure plate, which has a resilient annular basic body and tongues forming from this radially inwardly extending actuator.

In order to guarantee a trouble-free readjustment through the Equalizer, it can be particularly advantageous if this in the engaged state of the clutch unit or of the friction clutch automatically or automatically guarantees the desired readjustment, while the operation of the friction clutch is blocked automatically or automatically.

The equalizing device can have an annular structural part, as also in P32724SE01 / LK 10 15 20 25 30 522 574 3. u a n a ø ø no engaged condition for the friction clutch axially abuts the actuators. Due to this annular construction part, the possibly changing distance between the actuating portions of the actuating means and the disconnector can be equalized. For the function of the leveling device, it can be advantageous if they have axially adjusting post-adjustment ramps or ramp ramps, whereby these can be arranged on the annular construction part.

The ramp ramps can cooperate with cylindrical or wheel-like unrolling bodies for post-adjustment. It can be particularly advantageous, however, if the ramp-up ramps cooperate with counter-ramp ramps, then by corresponding selection of the ramp-up angle for these ramps a self-inhibiting at axial prestressing of the ramps can take place. The counter-ramp ramps can also be supported by an annular structural part.

In order to guarantee an inexpensive production of the friction coupling, it can furthermore be advantageous if at least a part of the leveling device is made of plastic material.

Such plastic material parts can be produced by injection molding. As plastic material, thermoplastics, such as e.g. polyamide.

In a particularly advantageous manner, the structural parts having the post-adjustment ramps can be displaceable in the axial direction when operating the coupling unit or the friction coupling. Furthermore, it may be suitable if the construction ramps and counter-ramp ramps carrying the structural parts are rotatable relative to one another, wherein one of these construction parts can be rotationally fixed in relation to the friction coupling, in particular in relation to the coupling cover.

According to a further idea, the leveling device can be designed in such a way that - viewed in the direction of disengagement of the coupling unit - it appears freewheel-like or readjustable, but is still self-inhibiting in the direction of disengagement. For this purpose, the ramp ramps and / or counter-ramp rails can be designed in such a way that in the axial direction they have a pitch angle which is between 5 and 20 °, preferably in the order of magnitude of 7 to 11 ". Advantageously, the post-adjustment ramps are formed in such a way that a self-inhibition takes place through frictional intervention. Thus, in any case, it must be guaranteed that the readjustment ramps have a self-restraining action, so that no additional means are required, in order to avoid an undesired reset. If necessary, however, such means may be provided.

In order to guarantee a trouble-free function of the self-acting leveling device, it may be appropriate if at least one ramp ramp and / or a structural part carrying the counter-ramp ramps is spring-actuated in the post-adjustment direction. The spring action can then take place in an advantageous manner in such a way that the function of the other springs, such as especially the compression or disc spring and the spring acting on the axially resilient counter-bearing, is not affected at all or practically. A particularly advantageous design can be guaranteed in that the structural parts having the ramp ramps and the counter-ramp ramps are actuated or biased by at least one force magazine arranged between them, such as a helical spring, in post-adjustment direction. By means of such a clamping, these two structural parts, viewed in the axial direction, are pressed in opposite directions, i.e. by means of the power magazine and the post-adjustment ramps axially away from each other. When the coupling is engaged, the equalizing device can thereby be clamped without play between the actuating portions of the operating means and the coupling cover and / or the disconnector.

According to a particularly advantageous design of the invention, the coupling assembly can have a device, for limiting the disengagement movement of at least the operating means. For this, a restriction stop can be provided, which limits the distance of the disconnector and / or the disconnecting means in the disconnection direction. The limiting stop can advantageously be formed by a structural part forming a leveling device after a certain disengagement distance striking the coupling cover. However, the limitation can also take place in that the disconnector has portions which, after a certain disconnection distance, come into abutment against an axially fixed structural part. Furthermore, it can be advantageous if the disconnector also has a restriction in the connection direction, which can also be formed by a stop. Advantageously, the equalizing device is designed in such a way that by means of it the disconnector is axially supported in the engaged state of the coupling unit. The constant fl distance of the coupling unit can also be guaranteed by the fact that a structural part forming a leveling device in the direction of disengagement and the direction of connection has effective road limiting sections, which cooperate with abutment sections. Advantageously, this structural part can be formed by the structural part of the equalizing device influenced by the disconnector, wherein the limiting stops can be arranged on the coupling cover or be formed through this cover. However, the limitation of the coupling unit for the fl travel distance can also take place by arranging corresponding stops on the structural part which controls the disconnector in the axial direction. Preferably, these stops cooperate with a structural part which is connected to a non-rotating bearing ring of the disconnector. However, the limitation of the disengagement distance in at least one axial direction can also take place through the rotating bearing ring of the disengagement and a structural part rotating therewith, such as e.g. the coupling cover.

According to a further embodiment of the invention, in particular for minimizing the disengagement force course or the maximum required disengagement force, it can be particularly advantageous if means are provided which during the disengagement process at least over a part of the displacement distance of the operating means provide a gradual reduction the clutch disc transferable torque. These agents may, for example, be formed by a so-called Coating suspension, which are arranged between the friction linings of the coupling disc which can be clamped between the pressure plate and the back pressure plate.

A particularly advantageous design of a friction coupling according to the invention can be achieved in that the pressing spring, which can preferably be formed by a disc spring, is pivotally supported against the cover between two supports, of which the pressure plate facing is spring-loaded in the direction of the pressing spring. the disengagement of the friction clutch against the spring-loaded support exerted the maximum disengagement force during coating wear increases and becomes greater than the counter-force or support force acting on the spring-loaded support. When using leaf spring elements arranged for torque transmission between the pressure plate and the coupling cover and / or by a so-called Coating suspension, such as one already known, for example, from DE-OS 36 31 863, the forces exerted by these springs on the compression spring must be taken into account in determining the force acting on the spring-loaded support, namely because these forces are superimposed. This means that in the event of a sufficient coating wear in the short term, the increased disengagement force must be greater than the resulting force arising from the above-mentioned forces and attributable to the rotational diameter of the disc spring, in order to enable a readjustment. It may be particularly suitable if the spring-loaded support is axially displaceable. Advantageously, the pressing plate edema can have such a spring characteristic that, based on its constructively defined mounting position in the friction clutch, in a relief due to friction lining wear, the force exerted by them and thus also the level of disengagement force ratio increases in position the maximum force exerted by this decreases during a disconnection process. By such an arrangement and design of the pressing plate spring it can be guaranteed that in case of occurrence 1732724550 !! LK 10 15 20 25 30 522 574 6 Coating wear An equilibrium can always be set again, at least between the maximum disengagement force for the friction clutch and the counter-force acting on the capercaillie-supported support or the counter-force acting on the pressure plate in the area of the unrolling diameter.

The coupling assembly or friction coupling can advantageously be constructed in such a way that the axially displaceable, fi spring-loaded supports by means of the wear reserves of the friction coupling are jointly fl replaced with the pressure plate. During the post-adjustment of the post-adjustment device during the service life of the friction clutch, gradually or in small steps, the spring-loaded support can be slightly displaceable in the direction of the pressure plate. By this displacement it can be guaranteed that the disc spring then supporting the pressure plate undergoes a further deformation, so that the force exerted by it, as already mentioned, decreases, until the counter-force acting on the spring-loaded support or the resulting counter-force already mentioned is in equilibrium. with the disengagement force. Thus, when the spring-loaded support is displaced, the maximum disengagement force of the coupling or the pressing plate spring decreases again. It can be particularly advantageous if the pressing plate spring is built into the friction clutch in such a way that it has a decreasing force-path characteristic at least over a part of the disengagement area, preferably practically over the entire disengagement area of the clutch. The mounting position of the compression spring can then be such that in the disengaged state of the friction clutch the compression spring practically reaches the minimum or valley point for its sinusoidal force-path course.

The counter-force exerted on the spring-loaded support can advantageously be generated by a force magazine, which mainly applies a constant force, at least over the intended readjustment range. In a particularly advantageous manner, a disc spring built into the friction clutch in a correspondingly designed manner and in a prestressed state is suitable for this purpose.

The invention is not only limited to the friction couplings described above, but is generally useful in friction couplings or coupling assemblies with a post-setting device compensating at least for the wear of the friction pads of the coupling disc.

The invention further relates to a friction coupling, especially for motor vehicles, with a pressure plate, which is rotatably but axially limited slidably connected to a lid, wherein between the lid and the pressure plate a pressing plate spring is axially clamped, which is pivotable about a pivot bearing carried by the lid and affects the pressure plate in the direction of P32724SE01 / LK 10 15 20 25 30 522 574 7 uuoa of a clutch disc clampable between it and a flywheel, such as a flywheel, wherein a wear of the friction pads on the clutch disc compensating adjusting device is arranged.

Automatic post-setting devices, which are to provide a practically constant force effect of the pressure plate through the pressing plate spring, are known, for example, from DE-OS 29 16 755 and 35 18 781. Due to the attachment of the pressure plate to the lid by means of tangentially arranged leaf springs - whose force, since this is opposite to the pressing force of the disc spring only needs to be relatively low, it can pivot a relatively large mass containing the pressure plate axially when disengaged from friction, thus lifting it from disc spring, whereby the function of the coupling is not only affected, but the coupling becomes empty to a safety risk, since the post-adjustment device is set in the open state, until the pressure plate abuts against the coupling disc, and thus the coupling can no longer be separated. For this reason, it has not been possible to implement such readjustment devices in practice.

The object of the present invention was to obviate these disadvantages and to provide post-adjustment devices of the initially approached type, which are practical in practice on a broad basis and also in difficult operation, which have a simple construction and a lasting safe operation, and which further need a small built-in space and are affordable in manufacture. In addition, the required disengagement force must be low, maintained low during the service life and also increase the service life of friction clutches.

According to the invention, this object is solved by a friction coupling with a pressure plate loaded by a disc spring, in which the pressing force is generated by a disc spring, which is partly supported against a structural part, such as a lid and is pivotable about a pivot bearing arranged on the lid in a circular arrangement. between the lid and the plate teeth, an automatic, the side-sided support wear-dependent from the lid displacing adjusting device is effective, which is further transportable through a displacement device and the plate teeth stand in the direction of the pivot bearing under the action of a support force. This support force is suitably present permanently, so that the disc spring against the disengagement spring is only force-bonded and this can be supported by a capercaillie force and not by form-bonded means. The disc tether is then mounted over its working area with a degressive curve line, namely in such a way that the support force and disc P32724SE0l / LK 10 15 20 25 30 522 574 8 n u: | u ø ou fi the spring force is adapted to each other in such a way that the support force at the plate fittings' intended installation position and without wear-dependent conicity change and across the disc fi spring's disengagement path is greater disconnection path less than the force applied by the plate edema to the support force. The support force can then be applied through a single spring element or at least substantially through a single spring element or spring element system. By "support force" is meant also the sum of all capercaillie forces acting against the plate spring, to the extent that they appear noticeable, thus e.g. also or only the forces acting through the (torque transmission or lifting) leaf springs, the (residual) hos suspension of the coating fi suspension or their "replacement".

As the power magazine, which at least mainly applies the support force, a spring can suitably be used, which changes its shape through the readjustment, e.g. a plate tether. However, the support force on the power magazines can also be formed by leaf veins.

A support force on the disc spring can be stored directly against the disc spring, e.g. at the radial height of the axially displaceable, lid-sided support.

It is particularly advantageous if the post-adjustment device is arranged axially between the plate springs and the lid. The post-adjustment device may in a particularly suitable manner contain ramp surfaces, such as ramps.

The invention ensures that the disc spring, viewed over the life of the friction clutch, practically always has the same conicity or prestress when the friction clutch is engaged and a practically constant force effect of the pressure plate and thus the clutch plate is achieved, regardless of the wear of the friction linings, the pressure plate or other pressure plate sided supports, disc cleaner or the friction surfaces of the turntable. The measures according to the invention furthermore guarantee that the mass of the pressure plate is not increased by that of the post-setting device. It is further arranged in an area in which it is protected from the effects of the disc wear and in which it is at a longer distance from the source to the frictional heat.

A particularly advantageous design of a friction clutch according to the invention can be achieved in that the pressing plate springs are pivotally supported against the cover between two supports - of which the face facing the pressure plate is spring-loaded in the direction of the pressing plate spring - whereby in case of coating wear increases and then becomes greater than P32724SE0l / LK 10 15 20 25 30 522 574 9 the counter force or support force acting on the spring-loaded support. The compression plate spring thus has such a curve line course that, based on its constructively defined mounting position in the friction clutch, in the case of a clamping direction due to friction coating wear, the force applied thereto and thus the required disengagement force first increases and at a longer relative. position the force exerted by it decreases during the disconnection process. Such an arrangement and design of the compression disc spring ensures that in the event of wear of the coating, a balance can always be re-established between the force exerted by the compression disc on the supports during disengagement and the force acting on the spring-loaded support, since exceeding the support force the force plate dem edema displaces the sensor springs away from the cover-side support and the post-adjustment device can be rotated further by the force of the displacement device. Thus, the support is displaced axially, until the force exerted by the sensor prevents a widening and a further axial displacement of the support.

It can be particularly advantageous, as already mentioned, if the pressing plate spring is mounted in the friction clutch in such a way that it has a descending force curve line at least over a part of the disengagement area, preferably in practice over the entire disengagement area for the friction clutch. The mounting position of the pressing plate spring can then be such that in the disengaged state of the friction clutch the pressing plate spring reaches or exceeds in practice the minimum or valley point of its sinusoidal force-path course.

The counter-force exerted on the spring-loaded support can advantageously be generated by a power magazine, which mainly applies a constant force at least over the intended readjustment area. In a particularly advantageous manner, a correspondingly designed disc spring mounted in the friction clutch is suitable for this purpose.

The post-adjustment device according to the invention can be used in a particularly advantageous manner in friction couplings with a pressing plate spring, which with radially outer portions acts on the pressure plate and is mounted over radially longer inwardly lying portions between two pivot supports on the cover. With this type of construction, the disc spring can act as a second lever.

However, the invention is not limited to friction couplings with disc springs, which at the same time have the disengagement lever designed in the form of disc spring tongues, but also extends to other types of coupling, in which e.g. the disc spring is actuated by P32724SEO1 / LK 10 15 20 25 30 522 574 10 additional levers.

In order to guarantee a trouble-free readjustment of the wear or an optimum pressing force for the friction coupling, it can furthermore be advantageous if the support supported on the side of the pressing plate spring facing the spring-loaded support is designed in such a way that it is automatic in the direction of the pressure plate. displaceable, and in the opposite direction through a device is automatically lockable. The post-adjustment of the counter-support, i.e. the lid-side support, can take place by means of a power magazine, which acts on this counter-support in the direction of the pressure plate or against the pressing spring. The counter-support can thus be automatically readjusted in accordance with the displacement of the fi-spring-supported support due to the wear of the coating, whereby a game-free swing storage of the pressing plate fi edges can be guaranteed.

The counter-support can be axially displaceable by means of a post-adjustment device arranged between the pressing spring and the lid. The post-adjusting device can then have an annular, thus inherently cohesive construction part, which at least in the engaged state of the friction clutch is axially actuated by the pressing plate spring. By turning the annular structural part in the event of wear and tear and during the disengagement process, the pivot bearing can be readjusted in accordance with the coating wear. For this purpose, in a particularly advantageous manner, the post-adjustment device or the annular structural part of this post-adjustment device can have axially adjusting post-adjustment grooves. Furthermore, it can be advantageous if the annular structural part carries the support, whereby the latter can be formed by a wire ring. This wire ring can be accommodated in a circumferential ring groove in the construction part and be connected to it by form bonding. The pre-connection can then be designed as a snap connection.

The ramp ramps can cooperate with cylindrical or ball-like roller bodies for post-adjustment. It can be particularly advantageous, however, if the ramp ramps cooperate with corresponding counter-ramp ramps, since thereby by self-adjustment of the ramp angle on these ramps a self-adjustment can take place at axial prestressing of the runner ropes. The counter-ramp ramps may be supported by a ring-like structural part, which may be arranged between the structural part carrying the ramp-side ramps and the cover. However, a particularly simple construction can be guaranteed by inserting the counter-ramp ramps into the lid. The latter can be done in a particularly simple manner at the plate lid, as the counter-ramp ramps can be pressed in. The pressing can then take place in radially extending portions of the lid.

P32724SEOl / LK 10 15 20 25 30 522 574 ll In order to guarantee a cost-effective production of the friction clutch, it can also be advantageous if at least a part of the post-adjustment device is made of plastic material. Such plastic material parts can be manufactured by injection molding. As plastic material, thermoplastics, such as e.g. polyamide. The use of plastic material becomes possible because the readjustment device is located in an area that is only slightly exposed to protective effects. In addition, due to the low weight, a small moment of inertia of mass is also obtained.

According to a further inventive concept, the post-setting device can be designed in such a way that - viewed in the disengagement direction of the friction clutch - it acts similarly to the freewheel, but is nevertheless self-inhibiting in the direction of disengagement. For this purpose, the ramp ramps and / or counter-ramp ramps can be designed in such a way that they have an inclination angle in the axial direction which is between 4 and 20 °, preferably in the order of magnitude of 5 to 12 ". Advantageously, the ramp ramps and / or the counter-ramp ramps are designed in such a way that a self-inhibition takes place by frictional intervention.

However, self-inhibition can also be achieved or supported by a form bonding, by e.g. one of the ramps is soft and the other is designed with a profiling, or in that both ramps have projections. These measures ensure that no additional funds are required, in order to avoid an unwanted recovery.

The post-adjusting device can be particularly advantageous and simple, if the displacement device acting in the circumferential direction is designed as prestressed mounted springs, which resiliently actuate at least one structural part carrying the ramp ramps and / or the counter-ramp running parts carrying the counter-ramp running parts in the direction of mounting. Spring action can then advantageously take place in such a way that the function of the other springs, such as in particular the actuating disc spring and the spring acting on the axially resilient support, is not affected at all or practically.

For many use cases it can be advantageous if the post-setting device has fl your displaceable post-setting elements, such as e.g. radially and / or circumferentially displaceable adjusting wedges or roller bodies. Furthermore, it can be advantageous if the post-setting device is speed-dependent. He can e.g. the centrifugal force acting on individual elements of the post-setting device is used for the operation and / or reading of the post-setting device at certain operating conditions of the internal combustion engine. In particular, the readjustment device can be blocked by centrifugal force. | u: dependent means from a certain speed, which e.g. can take place at at least approximately - idle speed or speed below the idle speed, so that the wear reset only takes place at low speeds. This has the advantage that no unwanted readjustments occur, which could occur due to oscillations at high speeds.

A particularly simple and reliable construction of the post-adjustment device can be guaranteed in that the relatively displaceable parts, which have ramp-up ramps and / or counter-ramp ramps or counter-ramp sections, are loaded with capercaillie. If only a corresponding structural part is provided with corresponding ramps or sections, and which is displaceable in relation to the lid, this is affected. It can be particularly advantageous in this case if the spring load generates a force in the circumferential direction.

For the construction and function of the friction coupling, it may further be advantageous if the sensor spring designed as disc springs, such as disc spring, with its radially outer portion abuts against an axially fixed structural part, such as the lid, and with radially further inwardly facing portions affects the unrolling support. This unwinding support can also be designed in one piece with the sensor spring, so that the sensor disc spring also forms the support. For holding the sensor tether in the clamped position, the lid can carry support portions.

These support portions can be formed by individual support elements attached to the lid. However, it can also be advantageous if the support portions are formed in one piece with the lid, e.g. pressures or foamed and deformed portions can be arranged on the lid, which the sensor cap engages for axial support.

For the function of the friction clutch, especially for minimizing the disengagement force course or the maximum required disengagement force, it can be particularly advantageous if the clutch plate clampable between pressure plate and counterpressure plate has friction pads, between which a so-called Coating suspension, as has become known, for example, from DE-OS 36 31 863, is provided. The use of such a clutch disc facilitates the operation, in particular the disengagement process of the friction clutch. This is because in the engaged state of the friction clutch, the clamped lining exerts on the pressure plate a reaction force which is opposite to the force exerted by the pressing plate spring or the operating plate spring on this pressure plate. During the disengagement process, during the axial displacement of the pressure plate, it is first pressed back through the resiliently clamped lining suspension, whereby at the same time as a result of the relatively steeply falling curve line section of the pressing plate springs occurring in the disengagement portion, the force exerted on the pressure plate decreases. With the exception of that of the compression plate spring P32724SEßl / LK 10 15 20 25 30 522 574 13 n v | o c of the force exerted on the pressure plate also decreases the restoring force exerted by the coating tether on this pressure plate. The effective force required for disengaging the friction clutch is obtained from the difference between the restoring force of the coating and the compression force of the pressing disc spring. After relieving the coating suspension, i.e. when lifting the pressure plate from the friction pads or releasing the clutch disc through the pressure plate, the required disengagement force is mainly determined by the pressing plate guide. The force-path characteristics of the coating spring and the force-path characteristics of the compression disc spring can be so advantageously adapted to each other that when the clutch disc is released through the pressure plate the force required for the operation of the compression disc spring is at a low level. By intended adaptation or t.o.m. smoothing of the coating fi vein characteristic to the pressing plate spring characteristic until the release of the clutch disc can thus through the pressure plate only a very low, in the extreme case practically none, maneuvering force may be required for the pressing plate spring to overcome the remaining drive. Furthermore, the characteristic of the compression disc spring can be designed in such a way that after the clutch disc is released, the force then still directed by the compression disc spring to oscillate or the force required for the oscillation of the compression disc spring very low level. Designs are also possible in which, when releasing the coupling disc through the pressure plate, only a very low or practically no force at all is required, in order to operate the pressing plate spring for disengaging the coupling. Such friction clutches can be designed so that the actuating force is in the order of magnitude between 0 and 200 N.

According to a further invention, the friction clutch can be designed in such a way that at least when the clutch disc is released through the pressure plate, the axial force applied by the pressing plate spring is in the zero range, whereby when the disengagement process is continued the force action of pressure plate edema. This means that when the friction clutch is completely disengaged, it remains practically open by itself and only through external force action can the clutch process be initiated again.

The invention further relates to a friction coupling, especially for motor vehicles, with a pressure plate which is torsionally stable, but axially limitedly displaceable, connected to a cover, wherein at least one clampable compression spring is effective between the cover and the pressure plate, P32724SE01 / LK 10 15 20 25 30 522 574 ø u. o now 14 which acts on the pressure plate in the direction of a clutch disc which can be clamped between it and a counter-pressure plate, such as a flywheel.

Such couplings are known, for example, from DE-OS 24 60 963, DE-PS 24 41 141, DE-PS 898 531 and DE-AS 1 267 916.

The present invention was further based on the task of improving such friction couplings in terms of function and service life. In particular, the invention will reduce the forces required for the operation of such friction clutches and guarantee a virtually constant disengagement force course over their service life. Furthermore, the friction couplings according to the invention must be manufacturable in a particularly simple and economical manner. According to the invention this is achieved by arranging a wear of the friction lining of the clutch disc automatically compensating post-setting device, which provides a practically constant force effect of the pressure plate via the pressing spring, and the friction clutch has actuating means for the engagement and disengagement and has a device over a sub-portion of the actuating distance of the actuating means and or the disengagement path of the pressure plate, causes a gradual lowering of the torque transferable by the friction clutch or the clutch disc. By means of such a device it can also be achieved that during the engagement process of the friction clutch and at the beginning of the clamping of the friction linings between the pressure and counter-pressure plate there is a gradual or progressive construction of the torque transferable by the friction clutch.

The design of a friction clutch according to the invention ensures that the compression plate spring, viewed over the life of the friction clutch, practically always has the same preload when the friction clutch is engaged and thus gives a practically constant force effect of the pressure plate. Furthermore, by the further measure, which provides a gradual lowering of the torque transferable torque during a disengagement process, a reduction or minimization of the disengagement force course or the maximum required disengagement force can be achieved. This can be traced back to the device facilitating the operation, especially the disengagement process, of the friction clutch. For this, the device can have axially resilient resilient means which exert a reaction force on the actuating means and / or on the compression spring and / or on the pressure plate and / or on the counterpressure plate, which is opposite to the force exerted by the compression spring on the pressure plate and is connected in series.

It can be particularly advantageous if the device of the friction clutch is designed on P32724SE01 ILK 10 15 20 25 30 522 574 15 f a. | ø in such a way that during the disengagement process over a sub-section of the axial displacement distance of the pressure plate portion actuated by the pressing spring, a gradual lowering of the torque transferable by the friction clutch or clutch disc takes place.

For many use cases, the device can advantageously be arranged in the flow of force between the pivot bearing of the operating means or between the pressing spring and the fastening points, such as the screwing-on points, of the cover at the back pressure plate.

For other use cases, however, it can also be advantageous if the device is arranged in force fl between the pivot bearing of the actuating means or between the pressing spring and the friction surface of the pressure plate. Such an arrangement is e.g. proposed in DE-OS 37 32 354 and DE-OS 1 450 201.

For further use cases, it can be particularly advantageous if the device is arranged between two back-to-back friction linings of the coupling disc, i.e. is formed by a so-called "lining suspension", e.g. through the coil spring segments arranged between the pads. Such arrangements are known, for example, from DE-OS 36 31 863.

A further possibility, to achieve a progressive torque build-up or reduction, is proposed in DE-OS 21 64 297, in which the flywheel is formed in two parts and the structural part forming the counterpressure plate is supported axially resiliently against the structural part connected to the drive shaft of the internal combustion engine.

For the function and construction of a friction coupling according to the invention, it may be particularly suitable if the device enables an axial, resilient resilience between the coupling structural parts, the device in such a way that when the coupling is opened the force acting on the device is minimal and during the coupling closing process. the coupling path of the coupling, the force acting on the device gradually increases to a maximum, this increase compulsorily taking place only over a part of the closing path or the coupling path of the operating means or the pressure plate. It may be particularly advantageous if the device is designed in such a way that the gradual lowering or the gradual increase of the torque transferable by the friction clutch takes place over at least about 40 to 70% of the actuating distance of the actuator and / or the maximum axial distance of the pressure plate. The remaining portions of the corresponding section are needed for the trouble-free separation of the force fl and for the smoothing out of any deformations on the coupling structural parts, such as in particular the coupling disc, the pressure plate and the back pressure plate.

In order to minimize those for the operation of the friction clutch according to the invention P22724SEO1 / LK .Z .C 2.1 21 ".. n | a | oa: 16 required forces, it may be particularly advantageous if the pressing tether, at least over a part of the disengagement distance of the friction clutch, has a degressive force-path course, i.e. it means that the pressing tether , at least over a part of its compression or deformation distance, has a decreasing force course. at the same time, as a result of the degressive or decreasing force-path course of the pressing spring occurring in the disengagement process, the force exerted by the latter on the pressure plate or the friction pads decreases. storage fi spring effects occur, from the difference between the force applied by the device and the force of the pressing cap. During the lowering movement of the pressure plate from the friction pads or the release of the clutch disc through the pressure plate, the required residual disengagement force course or the required disengagement force is controlled mainly by pressing spring. The force-path characteristics of the device and the force-path characteristics of the pressing spring can be adapted to each other in such a way that when the coupling disc is released through the pressure plate the force required for the operation of the pressing spring is at a relatively low level. Thus, by approaching or even equalizing the spring characteristic or force characteristic of the device to the pressing spring characteristic for the release of the clutch disc through the pressure plate, only a very low, in extreme case practically no, actuating force for the pressing spring may be required.

A disc spring is particularly advantageously suitable as a pressing spring, which can be pivotable about an annular pivotal bearing supported by the lid and partly affects the pressure plate. In this case, the disc spring can have an annular body, from which radially inwardly directed tongues emanate, which form the operating means. However, the operating means can also be formed by the levers, such as e.g. are pivotally mounted at the lid. However, the pressing force of the pressure plate can also be applied through other types of springs, such as e.g. screw fi springs, which are arranged in the friction clutch in such a way that the axial force exerted by them on the pressure plate is greatest in the engaged state of the friction clutch and this force is reduced during the disengagement process. This can e.g. be done by tilting the coil springs in relation to the axis of rotation of the friction clutch.

P32724SE0l / LK 10 15 20 25 30 522 574 s a o n n 17 It may be particularly suitable if the disc spring is pivotally supported against the lid between two bearings, to form a so-called coupling of the pressed type. In such couplings, the actuator for disengaging the friction clutch is usually actuated in the direction of the pressure plate. However, the invention is not limited to couplings of the pressed type, but also includes couplings of the drawn type, in which the actuating means for disengaging the friction clutch is usually actuated in the direction away from the pressure plate.

In a particularly advantageous manner, the friction clutch according to the invention can have a plate spring, which is designed in such a way that it has a sinusoidal force-path course and is built in such a way that in the engaged state of the friction clutch its operating point is arranged thereon. first kraftrnaxirnum, degressive characteristic council. In this case, it can be particularly advantageous if the disc spring has a force ratio of 1: 0, 4 to 1: 0, 7 between the first force maximum and the subsequent minimum.

It can furthermore be particularly advantageous if the friction coupling is manoeuvrable by means of one on the actuating means, such as e.g. on the tip of the disc of the disc gear, engaging the disengagement system, the disengagement system being able to have a clutch pedal, which is designed in a manner similar to an accelerator pedal and is arranged in the passenger compartment of the motor vehicle. Such a design of the clutch pedal can be particularly advantageous, since by the design according to the invention the force required for disengagement of the friction clutch can be brought to a very low level so that by means of an accelerator pedal-like clutch pedal a better dosability of the actuating force is possible.

Due to the design according to the invention of the friction clutch and the associated possibility of reducing the maximum pressing forces occurring during the life of the friction clutch, the structural parts can be reduced correspondingly or reduced in their strength, whereby a considerable reduction in production can take place. By reducing the disengagement force series, friction and elasticity losses in the coupling and in the disengagement system are again reduced, and thus the efficiency of the friction clutch / disengagement system is significantly improved. As a result, the entire system can be designed optimally and thereby the coupling comfort is considerably improved.

The design according to the invention is generally useful for friction couplings and especially for those proposed by DE-PS 29 16 755, DE-PS 29 20 932, DE-OS 35 18 781, DE-OS 40 92 382, FR-OS 2 605 692 , FR-OS 2 606 477, FR-OS 2 599 444, FR-OS 2 599 446, GB-PS 1 567 019, US-PS 4 924 991, US-PS 4 191 285, US-PS 4 057 131, JP- P32724SFßl / LK 10 15 20 25 30 522 574 18 GM 3-25026, .TP-GM 3-123, JP-GM 2-124326, JP-GM 1-163218, JP-OS 51-126452, JP-GM 3-19131, and JP-GM 3-53628.

The use of a friction clutch with a self-acting or automatic smoothing of at least the coating wear - whereby a substantially constant clamping force of the clutch disc is guaranteed at least over the life of the friction clutch - is particularly advantageous in connection with clutch assemblies, in which the friction clutch, clutch a flywheel, forms a mounting unit or a module. In such a mounting unit, it is advantageous for cost reasons if the coupling cover with counter-pressure plate is connected by means of a non-releasable connection, such as e.g. welded joint or molded joint, e.g. by such a connection, the commonly used fastening means, such as screws, can be omitted. In such mounting units, a replacement of the coupling disc or of the coupling pads due to exceeding the wear limits is practically not possible without destructive structural parts, such as e.g. of the coupling cover. Through the use of a wear-adjustable coupling, the mounting unit can be designed in such a way that this over the entire vehicle life guarantees a trouble-free function. Thus, due to the design according to the invention, the wear reserves of the clutch disc and the post-adjustment reserves of the friction clutch or clutch module can be dimensioned so large that the clutch life and thus also the service life of the mounting unit certainly reaches that of the vehicle.

According to a further development of the invention, it can be advantageous if a friction clutch having a wear-resetting device is combined with a so-called two-mass flywheel, wherein the friction clutch during intermediate coupling of a clutch disc is mountable on the one swivel shaft connectable to a gearbox and the other swivel mass is connectable to the drive shaft of an internal combustion engine. Two-mass flywheels, in which the friction clutch according to the invention can be used, are e.g. known from DE-OS 37 21 712, 37 21 711, 41 17 571, 41 17 582 and 41 17 579. The entire contents of these applications belong to the background of the present invention, so that the features described in these applications can be arbitrarily combined with the the features described in the present invention. In particular, the coupling cap or the coupling cap can be connected to the pivot mass carrying it by means of a non-destructible connection, such as this e.g. shown and described for different embodiments in DE-OS 41 17 579.

Furthermore, by the use of a friction clutch with a device which at least equalizes the wear of the coating, an optimization can take place in the design of the friction clutch. now 19 lingen, especially by the biasing force for the clutch plate applied to the magazine.

This power magazine can thus be designed in such a way that it practically only applies the clamping force required for the transmission of the desired torque for the clutch disc. The power magazine can be formed by at least one disc spring or by a number of helical springs. Furthermore, the use of a self-adjusting friction clutch is advantageous in connection with two-mass flywheels, in which the torsionally elastic damper arranged between the two pivot masses is arranged radially outside the clutch disc or the outermost friction diameter on the friction surface of the pivot shaft connectable to the gear. In the case of such two-mass flywheels, the friction diameter of the clutch disc must be smaller than in conventional clutches, so that the compressive force must be increased in proportion to the ratio of the average friction radius, in order to be able to transmit a defined motor torque. When using a conventional clutch, this would lead to an increase in the disengagement force. However, by using a wear-resetting clutch with a progressive lowering of the torque transferable by the clutch disc according to claim 1, a decoupling force reduction can be achieved, whereby an increase of the disengagement force can be avoided or by corresponding design of frictional engagement and disengagement. conventional coupling is achieved.

Thus, the design according to the invention of a friction clutch can guarantee that despite the reduced friction coating outer diameter and the thereby higher pressing force, the disengagement force can be kept low. Due to the low disengagement force, the load on the roller bearing is also reduced, through which the two pivotal masses are mutually rotatable. Furthermore, the wear-and-match adjustment increases the service life of the clutch, so that a replacement of the parts, especially of the clutch disc during the service life of the motor vehicle, is no longer required. In this way, the coupling cover can be fixedly connected to the pivot mass connectable to the gearbox, e.g. by riveting or welding. This is especially advantageous if there is a limited installation space or limited contours for the clutch bell, which no longer enable a connection of the clutch cap to the alternating flywheel in the usual way by screwing.

In the case of friction coupling with integrated post-adjustment device for the coating wear, during conventional attachment of the friction clutch and flywheel coupling unit to the drive shaft of an internal combustion engine, axial, rotational and sliding oscillations are transmitted to the clutch unit, which is further through the combustion engine Such as e.g. crankshaft. Because the coupling unit or the post-adjustment device in its function is not affected by such oscillations and in particular an undesired post-adjustment of the wear compensation device is suppressed, the inertial forces of such structural parts which act on this device must be taken into account in the design of the post-adjustment device. In order to avoid these undesired side effects, in particular caused by axial and sliding oscillations, or the associated higher cost of designing the post-adjustment device for smoothing the coating wear, according to a further think tank, the post-adjustment device the bending oscillations. This can be done by the coupling unit being connected to the drive shaft of the internal combustion engine by means of an axially elastic or resiliently resilient structural part. The rigidity of this structural part is adapted in such a way that the axial and sliding or bending oscillations generated by the internal combustion engine drive shaft on the coupling unit are at least damped or suppressed to such an extent that a trouble-free function of the friction clutch, in particular its post-adjustment guaranteed. Such elastic structural members are known, for example, from EP-OS 0 385 752 and 0 464 997 as well as SAE Techical Paper 9 003 91. The contents of these publications are likewise part of the background of the present invention. By using an elastic structural part, it is possible to avoid an undesired wear adjustment, caused by axial oscillations of the pressure plate relative to the clutch cover - especially when the friction clutch is disengaged - by flywheel oscillations and / or oscillations of the disc spring. Such oscillations can in the case of coupling units or coupling units without a device at least substantially suppressing these oscillations, such as in particular an axially resilient disc, lead to a changed setting independent of the wear condition of the coupling disc, whereby the friction clutch disc spring could be reduced. of the desired torque could no longer be guaranteed.

According to a further embodiment of the invention, a friction clutch with an automatic or automatic equalization, which may be specially designed according to the present invention, can advantageously be used in a drive unit, especially for motor vehicles, which consists of an automatic or semi-automatic gearbox and an intermediate drive motor. . such as the internal combustion engine and the gearbox arranged, at least depending on the operation of the gearbox controlled or regulated manoeuvrable friction clutch. The friction clutch is P32724SE0l / LK 10 15 20 25 30 522 574 21 - | | n s. preferably fully automatically manoeuvrable. An automated or fully automatic operation of a friction clutch is proposed, for example, by DE-OS 40 11 850.9, which is why, with regard to the mode of operation and the required means, reference is made to this document.

In the hitherto known drive units with automatic or semi-automatic gearbox and conventional friction clutch, there have hitherto been considerable problems for the clutch operation and the design of the actuating means required for this, such as e.g. piston / cylinder units and / or electric motors. Due to the relatively high disengagement forces required in conventional couplings, very strong or large-dimensioned actuating means are required. This means large construction volumes, high weight and high costs. Likewise, due to their mass inertia, such large-shaped actuating means are relatively slow in their activation time. In addition, when using actuating cylinders, a larger volume flow of pressure unit is required, which is why the supply pump must also be dimensioned relatively large, in order to guarantee the desired operating time for the corresponding friction clutch. In order to partially obviate the above-mentioned disadvantages, it is proposed, for example, by DE-OS 33 09 427 to reduce the actuating force for the disengagement of the clutch by corresponding compensating springs, in order thereby to be able to use smaller dimensioned actuating means. However, since the disengagement force in conventional couplings fluctuates very sharply during the service life, ie. the disengagement force is in the new state relatively low and increases during the service life with increasing coating wear, by means of a compensating spring only a part of the normally required disengagement force can be reduced. Taking into account all tolerances, despite the use of compensating springs, a disconnection effect from the actuating means is required, which is greater than that of a new conventional coupling. By using a friction clutch according to the invention with coating wear compensation in connection with a drive unit, consisting of a motor and an automatic or semi-automatic gearbox, the disengagement force in relation to the above-mentioned technique can be reduced very considerably, namely directly in the clutch, this disengagement force - the course of force of the new coupling throughout its service life is maintained practically unchanged. In this way, significant advantages are obtained for the design of the actuating means, as their driving power or operating power can be kept correspondingly low. whereby also the forces or pressures occurring in the total disconnection system are correspondingly lower. Thereby, the losses occurring in the disengagement system due to friction or elasticity of the structural parts are eliminated and reduced to a minimum.

P32724SE0l / LK 10 15 20 25 30 522 574 22 o n | | n n With the aid of Figs. 1-48, the invention will now be explained in more detail. Fig. 1 shows a section through a coupling assembly designed according to the invention, Fig. 2 the equalizing device shown on a larger scale and in section, Fig. 3 a view in the direction of the arrow III-III in Fig. 2, Fig. 4 the one against the coupling means in the friction clutch adjacent the adjustment ring in view according to the arrow IV in Fig. 2, Fig. 5 a section along the line VV in ñg. Fig. 6 shows the counter-adjustment ring used in the coupling assembly according to Fig. 1 in view according to the arrow III in Fig. 2, Fig. 7 a section along the line VII-VII in Fig. 6, Fig. 8 a detail of an embodiment variant of the Fig. 2 shows the leveling device, Fig. 9 a further detail in section of a coupling assembly according to the invention, Figs. 10 and 11 wear re-adjustment rings, which can be used with coupling assemblies according to the invention, e.g. according to Fig. 9, Fig. 12 is a section through a coupling assembly according to the invention, Fig. 12a is a circular sector of the sensing spring used in Fig. 12, Fig. 13 is a partial view in the direction of the arrow XIII-XIII in ñg. Fig. 12, Fig. 14 a further design possibility of a friction clutch according to the invention, Fig. 15 a diagrammatically shown disengagement system for a clutch assembly according to the invention, Fig. 16 a further design according to the invention of a friction clutch, which has a brake for the readjustment ring, Fig. 17 a friction coupling according to the invention in plan view, Fig. 18 a section along the line II-II in Fig. 17, Fig. 19 an adjusting ring used in the friction coupling according to Figs. 17 and 18, Fig. 20 a section along the line IV-IV in Fig. 19 Fig. 21 shows a support ring used in the friction clutch according to Figs. 17 and 18, Fig. 22 a section along the line VI-VI in Fig. 21, Figs. 23 and 23a a spring which exerts a torsional force on the adjusting ring, Fig. 24 - 27 diagrams with different curves, from which the interaction between the individual spring and post-adjustment elements of the friction clutch according to the invention is shown, Figs. according to the invention, Fig. 29 showing a section along the line XIII in Fig. 28, Fig. 30 the adjusting ring used in the friction clutch according to Figs. 28 and 29 in plan view, Figs. device, Figs. 34 and 35 diagrams with different curves, of which the interaction between the pressing plate tether and the coating tether as well as the resulting effect on the disengagement force course of the friction clutch is shown, Fig. 36 a further friction clutch according to the invention in partial view, Fig. 36a a partial view in the direction of arrow A in Fig. 36, Fig. 37 a section along the line XXI in Fig. 36, Fig. 38 a partial view of a setting ring useful in a friction coupling according to Figs. 36-37, Figs. 39 and 40 further embodiments of friction couplings according to the invention Fig. 41 is a plan view setting ring which is useful in a friction clutch according to Figs. 28 and 29 or 36-37;

P32724SE0l / LK 10 15 20 25 30 522 574. Figs. 46 to 48 show further embodiments of friction couplings, and Figs. 46 to 48 show details of a further embodiment of a friction coupling according to the invention, wherein ñg. 47 shows a partial view according to the arrow A in Fig. 46 and ñg. 48 a section after arrows B-B in fig. 47.

That i fi g. 1 has a friction coupling 1 with a cover 2 and a pressure plate 3 connected thereto rotatably, but limited axially displaceable pressure plate. and acts on the pressure plate 3 in the direction of a counterpressure plate 6 connected to the lid 2, such as e.g. a flywheel, whereby the friction lining 7 of the coupling disc 8 is clamped between the friction surfaces of the pressure disc 3 and the counter-pressure plate 6. The pressure disc 3 is rotatably connected to the cover 2 by means. in the circumferential direction or tangentially directed leaf springs 9. In the embodiment shown, the coupling disc 8 has so-called Coating spring segment 10, which guarantees a progressive torque build-up during the engagement of the friction clutch 1, in that it over a limited axial displacement of the two friction linings 7 in the direction of each other enables a progressive increase of the axial forces acting on the friction pads 7. However, a coupling disc can also be used, in which the friction pads 7 are practically rigidly applied to a carrier disc.

In the exemplary embodiment shown, the plate spring 4 has a pressing force on the annular basic body 4a, from which actuating tongues 4b extending radially inwards emanate. The disc spring 4 is then built in in such a way that it with radially longer outwardly projecting portions affects the pressure plate 3 and with radially longer inwardly lying portions can be tilted around the pivot bearing 5.

The pivot bearing 5 comprises two pivot bearings 11, 12, between which the disc spring 4 is axially held or clamped. The pivot bearing 11 arranged on the side of the disc plate 4 facing the pressure plate 3 is axially actuated in the direction of the lid 2. For this purpose, the pivot bearing 11 is part of a disc spring or of a disc spring-like structural part 13 2, whereby the radially internally designed pivot bearing 11 is axially actuated against the actuating disc spring 4 and thus also in the direction of the cover 2. The disc spring 13 arranged axially between the actuating disc spring 4 has an annular portion 13b, from 13c, which form the pivot bearing 11.

For the support of the disc-spring-like structural part 13, in the embodiment shown between the lid 2 and the heavy arms 13a of the disc-like p32724SEO 1 / LK 10 15 20 25 30 -. 522 574 n n u u of 24 the construction part 13 arranged a bayonet-like connection or locking.

The disc spring-like structural part or disc spring 13 is designed as a sensor spring, which generates an at least approximately constant force over a predetermined working path. By means of this sensor spring 13 at least substantially the coupling uncoupling force acting on the tongue tips 4c is captured, there always being an at least approximately balance between the force generated by the uncoupling force on the pivot bearing 11 and the counter-bearing 11 exerted by the sensor plate spring 13 on this pivot bearing.

By disengagement force is meant the maximum force which is exerted during the operation of the friction clutch 1 on the tongue tips 4c or on the disengagement portions on the disc spring tongues.

The lid-side pivot bearing 12 is supported against the lid 2 by means of a readjustment device 16; This post-adjusting device 16 guarantees that in the event of an axial displacement of the pivot bearings 11 and 12 in the direction of the pressure plate 3 or in the direction of the counterpressure plate 6, no undesired play can occur between the pivot bearing 12 and the lid 2 or between the pivot bearing 12 and the plate spring 4. This guarantees dead or idle paths occur during the operation of the friction clutch 1, whereby an optimum efficiency and thereby a trouble-free operation of the friction clutch 1 is achieved. The axial displacement of the pivot bearings 11 and 12 takes place during axial wear on the friction surfaces of the pressure plate 3 and the counterpressure plate 3 as well as of the friction pads 7.

The post-adjustment device 16 comprises a spring-actuated post-adjustment element in the form of an annular structural part 17, which has circumferentially extending and axially rising ramp ramps 18, which are distributed over the circumference of the structural part 17.

The post-setting element 17 is mounted in the coupling 1 in such a way that the ramp ramps 18 face the lid bottom 2a.

The post-adjustment ring 17 is spring-loaded in the circumferential direction, namely in the post-adjustment rotation direction, thus in the direction which, by the rise of the ramps 18 on the counter-ramps 19 pressed into the lid bottom 2a, causes an axial displacement of the post-adjustment ring 17 in the axial direction. from the radial lid section 2a.

The mode of operation of the automatic post-adjustment drive of the pivot bearing 5 or the post-adjustment device 16 as well as further design possibilities of the post-adjustment device 16 are described in more detail in connection with ñg. 17-48.

The coupling assembly comprises a leveling device 20, which guarantees that the P32724SEO1 / LK 10 15 20 25 30 522 574. n s n vn 25 through the disc spring tongues 4b formed by the disengagement means playlessly maneuvers the friction clutch 1 in the axial direction and can be displaced by a constant distance 21. The equalizing device 20 is arranged between the disengagement bearing comprising the disengagement 22 and the tongue tips 4c. The disengagement 22 is axially displaceable on a schematically shown guide tube 23 for the operation of the friction clutch 1. The guide tube 23 is supported by a gearbox housing (not shown) and surrounds the gearbox input shaft, on which the clutch disc 8 is also rotatably receivable. The force required for the axial displacement of the disengagement 22 is applied by an operating means 24, which in the embodiment shown is formed by a schematically shown disengagement fork, which can also be mounted on the gearbox side. However, disconnectors which are hydraulically or pneumatically manoeuvrable can also be used, i.e. disconnectors which have a piston / cylinder unit which can be actuated by a pressure means.

The disconnect device 20 is in fi g. 2 and 3 are shown on a larger scale and comprise a post-setting element in the form of an annular construction part 25, which is shown in fi g. 4 and 5. In the exemplary embodiment shown, the annular post-adjustment element 25 has two radially displaced and axially rising sets of ramp ramps 26, 27 which are distributed over the circumferential direction, which are distributed over the circumference of the structural part 25. As especially shown by fi g. 5, the radially inner ramp ramps 26 are offset with respect to the radially outwardly arranged ramp ramps 27 in the circumferential direction, namely by approximately half of a ramp length or a ramp pitch. The readjustment element 25 supports, as shown by fi g. 1 and 2, with its end surface 25a directly against the tongue tips 4c. The ramp ramps 26, 27 are axially turned away from the operating means 4b. The post-adjustment element 25 is loaded with capercaillie in the circumferential direction, namely in the post-adjustment direction of rotation, thus in the direction which, through the rise of the ramps 26, 27 on the counter-ramps 28, 29 on the one in fi g. 6 and 7, the support ring 30 provides an axial displacement of the post-adjustment ring 25 in the direction of the pressure plate 3, i.e. thus in the axial direction away from the disconnector 22.

As shown in fig. 6 and 7, the counter-ramp ramps 28, 29 likewise form two sets of ramp ramps which are mutually offset in the radial direction as well as in the circumferential direction. The ramps 26, 27 of the post-adjustment element 25 and 28, 29 of the support ring 30 are adapted to each other and engage axially with each other. The circumferentially displaced ramps ensure that a trouble-free centered control occurs between the post-adjustment element 25 and the support ring 30. As is particularly apparent P32724SE01 / LK 10 15 20 25 30 522 574 26 o | q c I a. av fi g. 2, the two structural parts 25 and 30 of the leveling device 20 are axially connected to each other. The ramp angle 31 (fi g. 7) of the counter ramp ramps 28, 29 on the support ring 30 corresponds to the angle 32 (fig. 5) of the ramp ramps 26, 27 on the post-adjustment element 25. The support ring 30 can be rotatably connected to the cover 2, but in relation to this in the axial direction limited axial displacement with the clutch actuating distance 21. The axial limitation takes place by means of radial portions 33 on the support ring 30, which in the engaged state of the friction clutch 1 abut against the radially inner portions of the lid bottom 2a. This abutment is effected by the resiliently actuated actuators 4b. When disengaging the friction clutch 1, the road restriction is guaranteed by a plate mold part 34, which is arranged on the side of the support ring 30 facing away from the operating means 4b and is actuatable by the disengagement 22 in the diameter portion 35. This plate shape part 34 also has radial portions 36. can come into abutment against the radially inner portions of the sheet metal bottom 2a.

In the exemplary embodiment shown, the post-adjustment ring 25 as well as the support ring 30 are made of a heat-resistant plastic material, such as e.g. of a thermoplastic, which may also be ñber-reinforced. Thereby, these structural parts can be manufactured in a simple manner as injection molded parts.

The ramp ramps 26, 27 and counter-ramp ramps 28, 29 are designed in the circumferential direction in such a way that they at least enable an angle of rotation between the two structural parts 25 and 30, which over the entire life of the friction clutch 1 guarantees a readjustment to the friction surfaces of the pressure plate 3 and the counterpressure plate 6 as well as the wear occurring on the friction pads 7. This post-adjustment angle can be in the order of between 30 and 90 °, depending on the design of the ramp ramps. In the embodiment shown, the demia is in fig. 3 with the reference numeral 37 indicated the angle of rotation of the order of 75 °. The angle of incidence 31 resp. 32 of the ramps and the counter-ramps can be in the order of between 6 and 14 °, preferably in the order of 8 °, the actual angle 31 and 32 of the ramps and the counter-ramps change over the radial extent of these ramps, since for a given angle of rotation the same height difference must be bridged. This thus means that the ramp angles 31 resp. 32 decreases with increasing diameter.

The force effect required for the retrofitting of the element 25 in the circumferential direction is guaranteed by means of force magazines, which in the embodiment shown are formed by two arcuately arranged, between the support ring 30 and the post-adjusting element 25 P32724SEO1 / LK 10 15 20 25 , 39. These screws fi springs 38, 39 abut against the support ring 30 which is rotatable with the lid 2 and rotates the post-adjustment ring 25 as soon as the actuators or disc spring tongue 4b move away from the lid bottom 2a or from the disconnector 22. as a result of fi g. 3 and 6, the helical springs 38, 39 are contained in a circumferentially channel- or torus-like receiving 40, 41 in the ring 30. As can be seen from ñg. 2, such a receptacle 40 adapted in cross section to the windings of the power magazine 38, 39 extends over more than half the circumference of the cross section of a spring 38 resp. 39, whereby as shown in fi g. 3 and 6, a slit-shaped opening 42, 43 remains on the side facing the actuators 4b and also a slit-shaped opening 44, 45 on the side of the support ring 30 facing away from the actuators 4b. The springs 38, 39 are secured by the receptacles 40, 41 limiting the surfaces in the axial direction relative to the support ring 30. For the fitting of the coil springs 38, 39, the sector-shaped receptacles 40, 41 each have a fitting portion 46, 47 which has a radial insertion width which at least corresponds to the outer diameter of the windings of the coil springs 38, 39. Via these fitting portions 46, 47, the power magazines 38, 39 can be slid obliquely into the sector-shaped receptacles 40, 41. After the still unloaded coil spring tubes 38, 39 have been inserted into the sector-shaped receptacles 40, 41, the post-adjustment element 25 is mounted with the support ring 30. arranged on the post-adjustment ring 25 axial lugs 48, 49, which simultaneously form the actuating portion or support portion of the coil springs 38, 39, in an axially slotted portion 50, 51 adjoining the fitting portions 46, 47 in the circumferential direction, whereby the actuating portions 48, 49 at their end portions will abut against the unloaded coil springs 38, 39. The unloaded position of a power magazine 38 or 39 appears from tig. 3 and is denoted by 39a. The second end portion of the helical springs 38, 39 abuts the circumferential bottom 53, 53a of the sector-shaped receptacles 40, 41. By a rotation between the post-adjustment ring 25 and the support ring 30, the springs 38, 39 can be biased. After a certain relative rotation angle which is greater than the angular extent of the fitting portions 46, 47, the biasing portions 48, 49 of the post-adjustment ring 25 will be axial over an end portion of a slot 44, 45, so that the post-adjustment ring 25 and the support ring 30 can be moved towards each other. , until the ramp ramps 26, 27 and the counter ramp ramps 28, 29 touch each other. The slots 44, 45 and the axial lugs 48, 49 are adapted to each other in such a way that between the two structural parts 25, 30 there is a snap connection which acts in the axial direction. For this purpose, the axial lugs 48, 49 at their end portion have a hook-like section 48a which can abut radially extending portions of the support ring 30.

By a further relative rotation between the two parts 25 and 30 corresponding to the angle 37 (fi g. 3), the springs 38, 39 are burned to their pre-tensioned angular position 54 corresponding to the new state of the friction clutch 1. In this position the two parts 25, 30 can be secured by a average. This agent can e.g. comprise a mold bonding, which is effective between the two structural parts 25 and 30 and which after the mounting of the friction clutch 1 on the back pressure plate 6 can be removed, whereby the equalizing device 20 is activated. The possible post-adjustment angle for the smoothing, especially of a coating wear, corresponds to that in fi g. 3 with the specified angle of rotation of 37. After this angle of rotation 37, the axial lugs 48, 49 come into abutment against the end portions of the slots 44, 45 present in the post-adjustment direction of the ring 25. A prestressed position corresponding to this position for a screw cap 38, 39 is indicated in fi g. 3 with 38a.

In the new state of the friction clutch 1, the ramp ramps 26 and 27 and 29, 29 forming the axial ramps and the counter-ramp ramps engage each other substantially axially. This means that the abutting rings 25 and 30 need minimal axial building space.

In the exemplary embodiment shown, the limitation of the operating distance in the direction of disengagement of the friction clutch 1 through the plate front part 34 is guaranteed. According to an embodiment not shown, the abutment portions required for this, e.g. cooperates with the cover 2, is also arranged on the disconnector 22, namely on the bearing ring rotating with the friction clutch or on a structural part connected thereto. The axial limitation of the operating distance of the friction clutch 1 in at least one of the axial directions can also be formed by at least one axial stop for the disconnector 22 arranged on the guide tube 23.

Furthermore, the disengagement 22 can act directly on the actuating means 4b and a corresponding equalizing device is arranged between the disengaging means 22 and the disengaging means 24.

It is suitable if the disengagement 22 is actuated by a function of the friction clutch 1 and the equalizing device 20 not affecting bias in the direction of the operating means 4b.

As can be seen from Figs. 2 to 4, the post-adjustment ring 25 has radially internal lugs 55, which form engaging portions for a rotating or restraining means, which if necessary can be applied partly against the cover 2 or against the support ring 30 for rotary protection. Such restraining means can be provided during the manufacture or during the assembly of the friction clutch 1 or the leveling device 20 and after the mounting of the friction clutch 1 on the flywheel 6 is removed.

The i ñg. 8 constitutes an embodiment variant of the lower half of it in ñg. 1 and 2, in the variant according to Fig. 8, the axial limitation between the disengagement device 120 and the cover 102 takes place in the engaged state for the friction coupling via hook-like axial arms 133, which are formed in one piece with the plate mold part 134. The arms 133 are formed at the outer edge of the plate-shaped plate portion 134 serving and gripping axially through the lid 102. At its free ends facing the plate edge 104, the arms 133 have radially outwardly extending portions 133a which radially engage behind the lid 102 on their side facing the plate spring 104. Such a design ensures that the axial forces exerted by the plate spring 104 on the leveling device 120 can be supported by the sheet metal pressure piece 134, so that greater axial forces can be absorbed by the leveling device 120, than in the leveling device 20 according to Fig. 2, in which the stops are formed. through the portions 33 of the support ring 30 made of plastic material. 120 can occur i.a. during transport, ie with unmounted friction clutch, then in this condition the main disc spring 4 resp. 104 by means of the capercaillie tongues axially support against the support ring resp. the plastic material leveling element 30, 130. The sheet metal pressure piece 134 may have at least two, preferably three or more hook-shaped arms 133, which are preferably symmetrically or uniformly distributed over the circumference. The plate thickness of the thrust piece 134 can be designed corresponding to the axial forces to be supported. The ring 130 consisting of plastic material is rotatably connected to the pressure piece 134. As with ñg. 2, the thrust piece or plate front portion 134 has radially outwardly likewise portions 136 which, viewed in the circumferential direction, extend between the hook-shaped arms 133 and serve to limit the disengagement path or to avoid an impermissibly large distance by abutment against the lid 102.

The i fi g. The detail of a friction clutch 201 shown in Fig. 9 has substantially a similar structure to the right lower portion of the friction clutch 11 of Fig. 1. In Fig. 9, the clutch cover 202, the pivot bearing 205 for the disc spring 204, the post-setting device 216 and the equalizing device 220 are partially shown. With regard to the function of the post-adjustment device 216 and the equalization device 220, reference is made to the description of Figs.

P32724SE0l / LK 10 15 20 25 30 30 For the design variant according to ñg. 9, a rotary fuse 260 is provided for the post-adjustment element in the form of a post-adjustment ring 217.

In the case of unmounted friction clutch 201, the adjusting or rotary fuse 260 guarantees a defined position for the post-adjustment element 217 in relation to the other structural parts, such as especially the cover 202. In particular, the adjustment fuse 260 can guarantee that in the new condition practically in the zero position, in which no post-adjustment has taken place, and this although plate dem edema 204 does not affect the post-adjustment ring 217 in the area of the pivot or support bearing 212. The latter can be traced back to the non-mounted friction clutch 201 or to the friction cup ling 201 the main disc spring 204 axially abuts by means of its spring tongues against the leveling device 220, as this also appears in connection with tig. 1 and 2. As a result of this support, the main disc spring 204 pushes the force sensor in the form of a disc spring 213 axially in the direction away from the cover 202 or the adjusting ring 217, whereby an axial bias of the adjusting ring 217 is no longer guaranteed in the direction of the cover 202.

Without the rotary fuse 260, the ring 217 could thus be adjusted. When mounting the friction clutch 201 on the drive shaft of an internal combustion engine, the ring 217 would thus not have the desired retracted position, a post-adjustment of the wear, which is particularly guaranteed on the friction pads of the clutch disc. I ñg. 9, the position of the structural parts, which corresponds to the friction clutch mounted on a flywheel, is completely plotted. Dashed is the position of the disc spring 204 and the sensor spring 213, which corresponds to a new unmounted friction clutch. As can be seen, in the unassembled state of the friction clutch 201 between the post-adjustment element 217 or the annular support 212 and the disc spring 204, an axial distance or an air gap is arranged. The i.a. for the transport of the friction clutch 201, the post-adjustment device 216 arranged for the post-adjustment device 216 has at least one securing element 261, which at least in the unassembled state of the friction clutch 201 and possibly in the engaged state of the mounted friction clutch 201 is rotatably held in relation to of the post-setting element 217 cooperates with this. the fuse element 261 may, for example, as shown in ñg. 10, have individual radially extending arms 262 which are radially externally fixedly connected to the post-adjustment ring 217 and are radially internally clampable between the stop 233 arranged on the side of the plate teeth 204 and the cover 202. Thereby a force-connected connection is provided P32724SE01 / LK 10 15 20 25 30 522 574 31 the adjusting element 217 and the cover 202. The arms 262 may be formed similar to blades and radially internally by means of an annular portion 263 be connected to each other. In the embodiment shown, the arms 262 are screwed together with the post-adjustment element 217, but the arms 262 can also be riveted together with the post-adjustment element 217 or even have portions which, for the torsionally fixed connection with the post-setting element 217, are embedded in the plastic material forming this element.

The equalizing device 220 or its restraining stop 233 forms, in connection with the cover 202 and the axially clampable portions of the securing element 261, a brake or clutch for the post-setting element 217, which is active in non-actuated friction clutch 201. When disengaging the clutch or the tabs 262, so that the readjustment element or the tuning ring 217 can be readjusted if necessary.

It is advantageous if the securing element 261 or the leaf spring-like flaps thus forming in the axial direction have a low degree of spring or spring stiffness, but in the circumferential direction are relatively rigid or fi-spring stiff.

At the i fi g. 11, the tabs 362 are directly sprayed onto the plastic material ring 317. In the axial direction, the tabs 362 can be similarly sprayed on the plastic material ring 317. In a similar manner, as shown in connection with fig. 10, be radially internally connected by means of a circular ring-shaped portion.

The i fi g. The coupling unit or friction coupling 401 shown in Figs. 12 and 13 has a lid formed by a plate cover 402, a pressure plate 403 connected to it rotatably but axially limited, as well as a pressing plate spring 404 clamped therebetween. stored as a two-armed lever, in that it is pivotally or pivotally held in a pivot bearing 405.

With in relation to the annular pivot bearing 405 radially outwardly extending portions, the disc spring 404 acts on the pressure plate 403 in the direction of the friction linings 407 on a clutch plate 408 which can be clamped between the pressure plate 403 and a flywheel. in the lowering direction of the pressure plate 403 may be biased from the friction liners 407.

The disc spring 404 has an annular base body 404a, as well as from this outwardly and radially inwardly directed tongues 404b.

The pivot bearing 405 comprises two pivot bearings 411, 412, between which the disc spring 404 is axially held or clamped. The pivot bearings 411 and 412 are arranged in a similar manner and designed and have the same function as those in connection with ñg. 1 describes the swing bearings 11 and 12 in relation to the structural parts acting on the swing bearings 411, 412 as well as an automatic post-adjustment function. 1 as well as to ñg. 17-48.

The disengagement means formed by the disc spring tongues 404b of the friction clutch 401 is axially operable by a disengagement device 422, whereby the conicity of the disc spring 404 is changeable. The disconnect device 422 may, as described in connection with ñg. 1-7, have a leveling device 20. However, in the case of clutch disconnection systems with self-adjusting disconnection bearings, such a leveling device 420 is not required. In such a disconnection system, the disconnection device 422 may be connected to the disconnection layer rotating with the clutch 401 at least during the disconnection process. In order to prevent an impermissibly large disengagement distance for the coupling disengaging means formed by the disc spring tongues 404b, a road limiting means 436 for the disc spring tongues 404b is arranged on the coupling 401 or on the cover 402. The path limiting means 436 limits the pivot path or pivot angle of the disc spring 404 by axially supporting the disc spring tongues 404b and thereby by axially capturing the disengagement force acting on the disengagement device 422.

In the embodiment shown, the road limiting means 436 is formed by an annular abutment portion 436 formed by the radially inner section of the lid 402, against which the tongue tips 404c come into abutment after a predetermined axial distance 421. The annular abutment portion 436 is designed at least so that approximately lies on the disengagement diameter of the disc spring tongues, i.e. the diameter on which the disengagement device 422 comes into abutment against the disc spring tongues 404b. The abutment portion 436 is arranged axially between the spring tongues 404b or the spring tongue tips 404c and the coupling disc 408.

The annular abutment portion 436 is connected to the lid body 402a by radially extending cams or webs 437. As shown in fi g. 13, six such lives are provided in the exemplary embodiment shown. For many use cases, however, only three such lives can be provided. In connection with coupling designs, in which particularly large disengagement forces are required, your life, Lex. nine are arranged.

The web 437 extends radially inwardly and axially inclined towards the pressure plate 403 or the coupling plate 408 starting from the lid bottom 402b or from the lid body 402a.

Stop portions 436 are offset relative to the lid bottom 402b axially into the lid space.

The spring tongues 404b engage through the openings formed between the annular abutment portion 436, the radially further outwardly extending lid body 402a and the connecting chambers 437. 438. In the embodiment shown, for this purpose the disc spring tongues 404b are radially inwardly directed over a part of their length in the axial direction opposite to the processes of the webs 437 bent down or upwards. As shown in fi g. 13, the disc spring tongues 404b form three grooves, which are arranged in an opening or recess 438. Between the individual groups of three, a slot 439 is each arranged for receiving the webs 437. The slots 439 and the webs 437 are thereby adapted to each other in such a way that a problem-free oscillation of the disc spring 404 is possible.

An adjustment of the disc spring tongues 404b in the openings 438 takes place during the mounting of the friction coupling 401. For this purpose, the disc spring 404 is in the unloaded condition, as in fig. 12 shows a dotted dot, in the region of the tongue tips 404c an inner diameter 440 which is larger than the outer diameter 441 of the annular abutment portion 436. Thereby the disc spring 404, at least in a fully relieved condition, with its disc spring tongues 404b can be axially inserted into the lid 438. During the mounting of the friction clutch 401 or most recently during the mounting of the friction clutch 401, e.g. on a flywheel, the disc spring 404 pivots, thereby reducing the inner diameter 440 defined by the disc spring tongues 404b. In the case of a friction clutch mounted on a flywheel, the disc spring 404 has its operating position, and the tongue tips 404c define an inner diaphragm 442 which is smaller than the outer diameter 441 of the abutment portion 436. The disc spring 404 is pivotally held so that the lid after reciprocating the swing path 421, the inner diameter defined by the tongues is smaller than the outer diameter of the abutment portion 436.

The axially limited, maximum possible operating distance 421 is dimensioned in such a way that the coupling 401, after achieving the maximum permissible wear on its bearings 407, at least still has the full cage coupling path required for the trouble-free function, i.e. a trouble-free separation of the coupling assembly. 401.

The clutch 401 or the sensor spring spring 1313 guaranteeing an automatic coating wear compensation in the clutch and the readjusting device 416 are designed in such a way that in the new condition of the friction clutch 401 no incorrect axial adjustment of the pivot bearing 405 occurs even when the section 421 is completely passed.

P32724SED1 / LK 10 15 20 25 30 574 34 522 In the following, by means of a calculation example, the mode of action or interaction between the abutment portion 436 and the disc spring tongues 404b will be explained and demonstrated: The above-described disengagement distance for the friction clutch 401 4 - 10mm. The coupling 401 is designed in such a way that in the new condition an incorrect axial displacement of the pivot bearing 405 would be possible only at disengagement distances over 14 mm. The abutment 436 is designed or positioned in such a way that in the new state of the friction coupling the portion coming into contact with the abutment 436, namely the tongue tips 404c, can travel an axial distance 421 of 12.5 mm. When the disc spring tongues abut against the abutment 436 and apply the maximum disengagement force, the lid can spring in axially by approximately 0.5 mm, so that in total a maximum axial distance 421 of 13 mm is possible.

Assuming that on the pads 407 a maximum coating wear of 3 mm is possible, during the life of the friction clutch 401 the disc spring is displaced by axial displacement of its pivot bearing 405 with these 3 mm in the direction of the clutch disc. The maximum possible disengagement distance 421 is thus reduced from approximately 13 mm to approximately 10 mm, so that the coupling at the end of its service life is still within the required disengagement distance tolerance of 8.4-10 mm.

In the embodiment shown, the stop 436 is formed in one piece with the cover 402. However, this stop could also be formed by a separate structural part, which is connected to the cover 402. The webs 437 can also be formed by separate construction parts or be formed in one piece with the abutment 436 designed as a separate structural part.

The i fi g. The coupling unit 401 shown in Figs. 12 and 13 further has a device or means which during the operation of the coupling unit 401 at least in sub-areas of the speed range, in which the coupling unit 401 rotates during use, causes an axial support force increase on the disc spring 404. due to at least interference factors occurring within a certain speed range in the operation of the coupling unit 401, the occurrence of an unauthorized readjustment due to an undesired, non-abrasive axial deflection or resilience of the friction linings 407 which may occur in the co-acting arrangement of sensor spring 413.

I fi g. 12 is for the increase 'of the axial force acting on the unwinding bearing 411 arranged speed or centrifugal force dependent means 450. The centrifugal force dependent P32724SE01 / LK 10 15 20 25 30 522 574 n | . . The means 450 is formed by tongues 450 formed on the outer periphery of the sensor disc 413 and raised axially in the direction of the lid 412. As can be seen from Fig. 12a, the disc-like sensor spring 413 has radially outwardly extending tongue-like arms 413a, as can be seen from fi g. 12 and 13, axially abut against the cover 402. Between the arms 413a and these axially supporting portions 451 of the cover 402 a bayonet-like connection or lock 452 is arranged. The bayonet-like connection 452 is designed so that by axial joining of the sensor spring 413 and the lid 402 and by a subsequent relative rotation between these two components, the arm 413a will be axial over the support portions 451 of the lid 402. cap 402 is pre-biased before rotation of these two structural members by the sensor spring 413 first axially resiliently and relieved after rotation, whereby the biased arm 413a abuts the cover 402.

As can be seen from Fig. 12a, a tongue 450 is arranged on both sides of a radial arm 413a.

At the rotary coupling unit 401, due to the centrifugal force acting on the tongues 450, a force is generated which is superimposed on the force applied by the sensor spring 413 as a result of its clamping, and thus is added, whereby the support force for the actuating plate spring 404 increases in the area further on the swing bearing 411 through the tongues 450 generated force becomes larger with increasing speed. However, this force growth can be limited by deforming or pivoting the tongues 450 from a certain speed level due to the centrifugal force acting on them in such a way that they abut radially externally against the cover 402, so that then no or practically no further increase of it by the centrifugal force depending on the means 450 generated the additional support force is present in the area of the pivot bearing 411.

In considering the axial force ratios or force equilibrium between the pivot bearing 411 and the disc spring 404, the leaf spring-like torque transfer means 409 must again be taken into account. by the torque means 409 forcefully biased axially against the disc spring 404. Thus, the axial force applied by the torque transmission means 409 acts against the force exerted by the disc spring 404 on the pressure plate 403 and is thereby added by the force spring 413 on the disc spring. the decoupling force acting on the tongue tips 404c. The actual sensor force, which is opposed to an axial displacement of the disc spring 404 at non-rotating coupling unit 401, is formed thereby by the resulting action on the plate 40 generated by the torque transmission means 409 and by the sensor spring 413. the power. Upon rotation of the clutch unit 401, this resulting force is also superimposed on a speed or centrifugal force dependent force, which is generated by the tongues 450.

In the case of a clutch disc 408 with a device, for example in the form of a coating suspension 453, which during the renovation of the friction clutch unit 401 is guaranteed over a partial portion of the pressure plate 403 lifting distance a gradual lowering or a gradual construction of the torque transferable by the clutch disc 408, this device supports 450, until the release of the friction liners 407 or of the clutch plate 408 through the pressure plate 403, the axial support of the disc spring 404 against the post-adjustment element in the form of a post-adjustment ring 417. This ensures that the plate teeth 404 and the lid or lid 402 and thereby no readjustment can take place.

When the coupling unit 401 is disengaged, the pressure plate 403 is lifted up from the liners 407, so in a coupling unit without the tongues 450, only the resultant force generated by the leaf-spring-like torque transmission means 409 and the sensor disc spring 413 acts as an axial biasing force. the disconnection force. In certain speed ranges, especially at higher engine speeds, e.g. oscillations generated by the motor occur, which cause an axial oscillation of the pressure plate 403. If the pressure plate 403 oscillates axially, this pressure plate 403 can be lifted briefly from the main or disc spring 404, whereby the resulting sensor force briefly falls as it then The axial force generated no longer acts on the disc spring 404. This has the consequence that the force ratio required for a intended readjustment of the device 416 between the disc spring 404 or the disengaging force acting on it and the resulting supporting force acting on this disc spring 404 is disturbed during operation. the coupling unit 401 the axial support force acting on the disc spring 404 is too low, whereby the coupling is set too early or undesirably, and thus the operating point of the disc spring 404 is displaced in the direction of the disc spring minimum. Furthermore, at certain operating conditions of the engine, especially at higher engine speeds, particularly high crankshaft peripheral accelerations may occur which, due to the inertia of the post-adjustment ring 417, generate peripheral forces which, due to the post-adjustment ramps 24 10 15 20 25 30 522 574 37 yu: unon 404, which is opposite to the resulting sensor force, whereby likewise an undesired post-reset can take place. Furthermore, due to the oscillations occurring, the frictional engagement between the ramp ramps 418, 419 can be reduced, so that the axial force produced by the post-adjustment ring 417a acting in the circumferential direction 417 and circumferentially acting on the disc edge 404 is increased accordingly.

In order to obviate the above-mentioned disadvantages of a coupling unit 401 without centrifugal force-dependent support means 450, in the embodiment shown, according to fi g. 12-13, these centrifugal force-dependent means compensate for the speed-dependent interference effects by generating a speed- or centrifugal force-dependent increasing support force acting in parallel with the force generated by the sensor spring 411.

The centrifugal force-dependent means can then be designed in such a way that a readjustment required for the coating wear in the coupling unit 401 is only possible at standstill or at low speeds for the coupling unit 401. At rotary coupling unit 401 or above a speed at which critical oscillations can occur. the readjustment device 416 is practically blocked.

At the i fi g. 14, the sensor spring 513 is arranged radially inside the disc spring pivot bearing 505. The sensor spring 513 has an annular basic body 513a, from which radially inwardly directed tongues 513b emanate. By means of these tongues 51b, the sensor cap 513 abuts the annular abutment portion 536, which is arranged and formed similarly to the annular abutment portion 436 according to ñg. 12 and 13.

The sensor tongues 513b support the side of the abutment portion 536 facing the disc spring tongue tip 504c.

The mounting of the sensor spring 513 on the cover 502 can take place by deforming it in the prestressing direction until the inner diameter 540 limited by the inner tongues 513 is larger than the outer diameter 541 of the abutment portion 536. as described in connection with the tongues 404b and the openings 438 in ñg. 12 and 13. After insertion of the tongues 51b into the openings 538, the sensor spring 513 can be relieved, whereby the inner end portion of the tongues 51b is displaced towards a smaller diameter and comes into abutment against the abutment portion 536.

P32724SE0l / LK 10 15 20 25 30 522 574 38, u: »uon: A further possibility consists in that for the mounting of the sensor spring 513 on the lid 502 at least partial portions of the inner tongues 513b are axially bent up in the direction of the lid 502, so that they limits a larger inner diameter 540 than the outer diameter of the abutment portion 536. After the sensor tether or tongues 513b are inserted into the apertures 538 of the lid, the tongues 513b can be bent back in such a way that they come into abutment with their radially inner portions with prestress against the abutment portion 536. By the back bend the tongues 513b are pivoted from it in fi g. 14 in broken lines shows the position of the position shown in solid lines by plastic defonations in the disc spring material. For the plastic deformation of the sensor tongues 513b, these can axially abut the tongues 504b or the tongue tips 504c of the plate tines 504. For the bending operation of the tongues 513b, a tool can be used which supports the tongues 504b of the actuating plate tines 504 from above and affects sensor 5 the diarrheal portion, at which the tongues 513b are bent down. The stops 436 and 536 for limiting the disengagement distance or pivot angle of the plate tadpoles 404 and 504 have the advantage that these are integrated in the corresponding coupling 401 or 501 and become active in the area of the plate spring tongue 404b or 504b, whereby the plate can be guaranteed 404b, 504b against the abutments 436, 536 the disc spring tongues can not be deformed at all or only insignificantly in the axial direction. Thereby it can also be guaranteed that the disc tongue tongues 404b, 504b themselves in their position corresponding to the disengaged state of the friction clutch 401, 501 do not come into contact with a structural part of the clutch disc 408. In ñg. 12, the corresponding position of the disc spring 404 corresponding to the disengaged clutch is shown in broken lines and indicated by the numeral 450. Thereby it can thus be avoided that the disc spring tongue 404b in the disengaged state of the friction clutch 401 comes into contact with or can grind against it then in relation to this clutch. 401 rotary clutch plate 408.

In the embodiments shown according to fi g. 12 to 14, the stops 436, 536 are arranged in the area of the disc tongue tongue tip 404c, 504c. However, these stops can also be designed in another way and be displaced radially outwards relative to the inner tongue tips 404c, 504c. In such a design, however, it is suitable if the radial lever present between the tongue tips 404c, 504c and the radially longer abutments then is chosen in such a way that an impermissible deflection of the disc spring tongue 404b, 504b due to the actuating force acting on them and support through appropriations does not take place.

The said excessively large or impermissibly large disconnection distance can be caused P32724SE01ILK 10 15 20 25 30. Q u ø i; 522 574 39 through the disengagement system or the operating system, which acts on the clutch operating means, which in the embodiments shown and described are formed by the disc spring tongues. This actuating system usually contains a disengagement bearing which acts on the actuating means of the friction clutch, an actuating means, such as e.g. a clutch pedal, and a power transmission line arranged between the disengagement bearing and the actuator. This power transmission line can have a receiver as well as a sensor cylinder. In the case of disconnection systems with a sensor and a pick-up cylinder, an unauthorized disconnection distance lying above the normal disconnection distance can be caused by the fact that due to a quick connection and disconnection of the friction clutch the pick-up cylinder cannot be reset quickly enough, ie. it does not reach its end position, so a short subsequent reconnection of the pick-up cylinder itself, although the returner to the normal disconnection distance, corresponds to a total disconnection distance for the clutch, which corresponds to the sum of the normal disconnection distance and the remaining untrained distance. As a result, a total operation of the friction clutch can occur, which considerably exceeds the intended maximum permitted disengagement distance for the clutch. This means that the overshoot reserve for the actuation arranged during a friction clutch can thus also be exceeded.

Thus, by the measures according to the invention or the abutments 36, 436, 536, an impermissibly large disengagement distance or overstretch can be prevented during the operation of the friction clutches, while still guaranteeing the normal disengagement distance required and arranged over the life of the clutch.

Thus, according to the invention, in general in the case of couplings, and in particular in the case of couplings with a post-adjusting device which compensates at least for the wear of the friction plate of the clutch plate, at least one stop can be provided in the clutch actuating line which avoids overstretching of the clutch actuating means when actuating the same. Such a grant can e.g. limit the disengagement distance of the disengagement bearing or the pivot distance of the disc spring. However, such a grant can also be arranged elsewhere. Furthermore, the operating distance for the friction clutch can be limited to a defined constant value, by arranging a defined limitation, such as abutment, both in the uncoupling direction and also in the clutch direction.

Advantageously, such a restriction can take place in the area of the disengagement bearing, since in this part the tolerance chain between the operating means, such as the disc spring tongue of the friction clutch and the structural part limited to a certain distance, is small.

P32724SE01 / LK 10 15 20 25 30 522 574- 40 u = | n v fo In the presence of such a restriction or such a stop, the fl is shifted when disengaged against a practically rigid restriction, whereby an over-stressing of structural parts, especially of those which may occur in disengagement systems or in foot-operated systems for operation, may also be undesirable. Therefore, according to a further embodiment of the invention, a resilient or elastically resilient means and / or a pressure limiting means in the actuation system for the friction clutch is provided, this means having a biasing force or needing a minimum deformation force or opening force which is at least slightly greater than the required maximum. the force or the maximum pressure required to operate the clutch. This ensures that when the stop is activated, the clutch pedal can be pushed further down or the control motor up to a defined position can perform a pre-movement. The resilient means arranged in the actuation line of the friction clutch can be arranged between the clutch actuators and the disengagement bearing or between the latter and the disengagement actuator, such as e.g. clutch pedal or disengagement motor.

I fi g. 15, a disengagement system 601 is shown, showing various possibilities for the arrangement of a means for limiting the maximum force exertable through the disengagement bearing 622 to the clutch actuator 604 and / or the clutch cover 602. I fi g. 15, an axial stop 636 is provided, which after a certain distance for the coupling bearing 622 comes into abutment against the cover 602, similar to that described in connection with the stop 36 according to ñg. 1 and 2. However, the disconnection distance limitation could also take place in an amate manner, e.g. as described in connection with fi g. 12-14. The disconnection system 601 has a sensor cylinder 650 and a pick-up cylinder 651, which are connected via a line 652.

The piston 653 of the receiving cylinder 651 carries the disengagement bearing 622 and is axially displaceable received in a lid 654. The pressure chamber 655 is supplied via the line 652 with hydraulic medium, such as e.g. oil. The cylinder unit 650 has a lid 656 which, in connection with the piston 657 arranged therein, forms a volume-adjustable pressure space 658. The pressure space 658 is connected via the line 652 to the pressure space 655. In the pressure space 658 a resetting spring 659 for the piston 657 is arranged. The piston 657 is axially displaceable by means of a clutch pedal or an actuating motor, such as e.g. an electric motor or pump.

The pressure medium circuit of the disengagement system 601 communicates with a pressure medium reservoir 660. Preferably, the sensor cylinder 650 is directly connected via a line 661 to the pressure medium reservoir 660.

For the limitation of the P32724SE0l / LK 10 15 20 25 30 522 574 41 ø | n n n u. the clutch actuating force is in the embodiment according to fi g. At least one means is provided in the pressure medium circuit of the disengagement system 601, which limits the pressure arising in the pressure medium circuit during the operation of the friction clutch to a defined value. In the exemplary embodiment according to fi g. 15, this means is formed by at least one pressure relief valve. I fi g. 15, various device possibilities for such a pressure relief valve are shown. Such a pressure relief valve 662 may, for example, be provided in the conduit system 652 and have a return 663 to the pressure medium reservoir 660. However, instead of the pressure relief valve 662, a pressure relief valve 664 may also be provided, which is supported by the cover 654 or even may be integrated therein, and communicates with the pressure space 655 and is connected via a return line 665 to the pressure relief reservoir 660.

I fi g. 15 shows a further alternative device possible for a pressure relief valve 666. The pressure relief valve 666 communicates with the pressure space 658 of the sensor cylinder and can be supported by the lid 656 or integrated therein. Furthermore, the pressure relief valve 666 has a return line to the pressure medium container 660. For this purpose, the pressure relief valve 666 may have its own line, or a connection to the line 661.

A further possibility for arranging a pressure relief valve 667 consists in integrating it in the piston 657 of the sensor cylinder 650. On the relief side, this valve 667 must also have a connection with the pressure medium reservoir 660 or at least with an intermediate magazine.

Instead of an overpressure valve, a hydromagazine could also be arranged in the pressure manifold circuit, which limits the maximum pressure occurring in the disengagement system, by relieving the system after activation of the stops of the disengagement section and thus in practice acting as a buffer or spring reservoir.

The i fi g. 16, as described in connection with the preceding figures, has a post-adjustment device 716 for the automatic smoothing of the wear occurring on the friction pads 707 of the coupling disc 708. In the embodiment shown, the basic structure and mode of operation of the post-resetting device 716 corresponds to that in fi g. 12 and 13. The post-setting element or post-setting ring 717 has abutment or abutment portions 770, which can co-operate with disc plate 704 during a disengagement process for the coupling unit 701. The axial relative arrangement of the abutment portions 770 in the abutment portions 770 with these cooperating portions 771 on disc edge 704 are designed in such a way that during a disengagement process the disc spring portions 771 support at least indirectly, axially against the abutment portions 770 supported by the adjusting ring 717. This mutual support occurs corresponding to the pivot angle of the disc spring 704 in the groove of the tongue tips 704c. Such exceeding of the cage disengagement distance 772 may occur due to an incorrect or incorrectly set disengagement system. By the axial support of the disc spring 704 against the abutment portions 770, the post-adjustment ring 717 is secured against an undesired rotation.

The disc spring 704 thus acts in practice as a brake for the post-adjustment ring 717 when exceeding a predetermined disengagement distance 772.

In the exemplary embodiment shown, the abutment or abutment portions 770 are formed by an annular projection 773 formed radially outside the pivot bearing 705 against the ring 717. Instead of an annular radial projection, radially distributed radial arms 773 can also be used. The projection or arms 773 extend in the illustrated embodiment up to the outer edge of the disc spring 704. As soon as the intended disengagement distance 772 has been reached, the disc spring 704 extends with its outer portion 771 towards the extension portions 72 of the post-adjustment ring 717. 704, as it is displaced by the diameter of the pivot bearing 705 on the contact diameter between the portions 771 of the disc spring 704 and the abutment portions 770. This displacement also reduces the disengagement force required in the region of the ends 704c on the tongues. , namely, because the first in front of the disengagement section 772 as the two-armed lever stored the disc springs when exceeding the distance 772 is practically pivoted as a single-armed lever. Through this reduction in disconnection force, it is also guaranteed that disc edema 704, among others. a. axial resulting support force or actuating force applied through the sensor spring 713 and the leaf spring ring 709 is pressed in the direction of the cover 702 or the post-adjustment ring 717. The disc spring 704 can thus not be displaced away from the post-adjustment ring 717 or from the cover 702. 713 axially resilient, namely while the disc spring then in the area of the pivot bearing 705 is lifted up from the readjustment ring 717.

The protrusion or arms 773 can advantageously be sprayed on that of P32724SE01 / LK 10 15 20 25 30 522 574 43 u = | The maximum force acting axially on the arms 773 is obtained from the difference between the minimum disengagement force in the region of the disc spring tongues 704c and the axial sensor or support force of the disc spring 704, which is applied through the sensor spring element 713 and the leaf spring element 713. The springs 773 are designed in such a way that they can withstand the maximum force without significant deformation.

A further significant advantage consists in that the maximum lifting of the pressure plate 703 in practice remains constant and thus when exceeding the distance 772 the axial force applied by the leaf springs 709 on the main disc spring 704 does not fall further. Since the force applied by the leaf springs 709 forms part of the resulting sensor force, due to the residual residual stress of these leaf springs, the tensile strength of the friction clutch 701 is increased. in passenger car couplings in the area of the tongue tips 704c, an overstretch of about 0.5 to 2 mm is achieved, without affecting the function of the readjustment device.

The lifting limitation of the pressure plate 703 can also take place by exceeding the pressure plate 703 axially against the sensor spring 713 when a certain disconnection distance is exceeded. cams, projections or the like.

The i ñg. 17 and 18 have a cover 2 and a pressure plate 3 which is rotatably connected to it but axially limited but axially limited. actuates the pressure plate 3 in the direction of a counter-pressure plate 6 fixedly connected to the cover 2 by means of screws 6a, such as e.g. a flywheel whereby the friction linings 7 of the coupling disc 8 are clamped between the friction surfaces of the pressure disc 3 and the counter-pressure plate 6.

The pressure plate 3 is rotatably connected to the cover 2 via circumferentially or tangentially directed leaf springs 9. In the exemplary embodiment shown, the coupling plate 8 has a so-called Coating spring segments 10, which, as is known per se, guarantee a progressive torque build-up at the engagement of the friction clutch 1, in that over a limited axial displacement of the two friction linings 7 in the direction of each other enables a progressive increase of the effects on the friction pads 7. the axial forces. However, a coupling disc can also be used, in which case the friction linings 7 would be axially practically rigidly mounted on a carrier disc. In such a case, a "coating spring replacement" could thus be used a suspension in series with the disc spring, e.g. a suspension between cover and flywheel, between cover and P32724SE0l / LK 10 15 20 25 30 522 74 44 b? lid-side support as well as between disc teeth and pressure plate or through the lid elasticity.

In the exemplary embodiment shown, the disc spring 4 has a compressive force on the annular basic body 4a, from which actuating tongues 4b extending radially inwards emanate. The disc spring 4 is then inserted in such a way that it with radially longer outwardly lying portions affects the pressure plate 3 and with radially longer inwardly lying portions can be tilted around the pivot bearing 5.

The pivot bearing 5 comprises two pivot supports 11, 12, which here are formed by wire rings and between which the disc teeth 4 are axially held or clamped. The pivot support 11 arranged on the side of the disc spring 4 facing the pressure plate 3 is actuated in the direction of the lid 2 by means of a power magazine 13. The power magazine 13 is formed by a disc spring or by a disc-spring-shaped structural part 13 and with radially longer inwardly facing sections of the pivot support 11 axially acting against the actuating disc spring 4 and thus also in the direction of the cover 2. The disc spring 13 arranged between the pressure plate 3 and the actuating disc spring 4 has an outer annular edge portion 13b, from which the inner edge extends radially inwards , which support the swivel support 11.

For the support of the plate-like structural part 13, additional means 14 are attached to the shown exemplary embodiment of the lid 2, which form a pivot support for the plate-like construction part 13. These further means may be formed by stapled or riveted segment-shaped separate parts 14, which may be uniformly distributed over The perimeter. However, the means 14 can also be formed by a circular annular, inherently closed structural part. Furthermore, the support means 14 can be formed directly from the lid 2, e.g. by indentations inserted on axial portions of the lid 2 or by tongue-shaped cut-outs, which after the insertion and clamping of the spring-like structural part 13 are pressed into the outer edge portion of this structural part 13 by material deposition. a bayonet-like connection or locking, so that the disc-spring-like structural part 13 is first biased and its radially outer portions can be axially displaced over the support means 14. Thereafter, by a corresponding rotation of the disc-spring-like construction part 13 relative to the lid 2 the aid funds 14.

Support portions on the disc-tooth-like structural part 13 can then be formed by projectors radially outwardly projecting on the annular base body 13b.

P32724SE0l / LK 10 15 20 25 30 nu no a o oo av in 1 aa oo nu o I 0 I! n on aa en oaa en: un: l 'nu u- n ha: a: az e: I: v Q 0 anv oo I nu nsa: nsuaaonanv »nn nu i» 45 For torsional protection of operating disc spring 4 and possibly the disc spring-like structural part 13 as well as for centering wire ring ridges 11, 12, axially extending centering means in the form of rivet elements 15 is attached to the lid 2. The rivet elements 15 each have an axially extending stem 15a, which extends axially through a cut-out arranged between adjacent disc spring tongues 4b and which can be partially enclosed by portions 13d formed on the associated tongues 13c of disc spring 13.

The disc-spring-like structural part or the disc-spring 13 is designed as a sensor spring, which over a predetermined working path generates an at least substantially approximately constant force. The coupling disengagement force applied to the tongue tips 4c is captured over this sensor spring, an at least approximately equilibrium always prevailing between the force generated by the disengagement force on the pivot support 11 and the counter-force exerted by the sensor disc spring 13 on this pivot support 11. By the disengagement force is meant the maximum force which, during the operation of the friction clutch 1, is exerted on the tongue tips 4c or on the disengagement arm of the disc spring tongues and thus counteracts the sensor fi edema 13.

The lid-side pivot support 12 is supported by means of a post-adjusting device 16 arranged in the axial space between disc spring 4 and lid 2 against the lid 2. This post-adjustment device 16 guarantees that in the event of an axial displacement of the pivot supports 11 and 12 in the direction of the pressure plate 3 or in the direction against the back pressure plate 6, no undesired play can occur between the pivot support 12 and the lid 2 or between the pivot support 12 and the disc spring 4.

This ensures that no undesired dead or idle paths occur during the operation of the friction clutch 1, whereby an optimum efficiency and thereby a trouble-free operation of the friction clutch 1 is achieved. The axial displacement of the pivot supports 11 and 12 takes place in the case of axial wear on the friction surfaces of the pressure plate 3 and the counterpressure plate 6 as well as the friction lining 7. However, in devices according to the invention a wear of the disc spring in the region of the pressure plate support cams (at 3a) or the opposite portions of the disc spring.

The mode of operation of the automatic post-adjustment of the swing bearing 5 is explained in more detail in connection with the diagrams according to FIG. 24-27.

The post-adjustment device 16 comprises a spring-actuated post-adjustment element in the form of an annular structural part 17, which is shown in Figs. 19 and 20. The annular structural part 17 has circumferentially extending and axially rising ramps P32724SE01 / LK 10 15 20 25 30 | f | n u. 522 574 46 18, which are distributed over the circumference of the structural part 17. The post-setting element 17 is inserted into the coupling 1 in such a way that the ramp ramps 18 face the lid bottom 2a. On the side of the post-setting element 17 facing away from the ramp ramps 18, the pivot support 12 formed by a wire ring is centered positioned in a groove-shaped receptacle 19 (fl g. 18). The receptacle 19 can then be designed in such a way that the pivot support 12 on the post-setting element 17 is also secured in the axial direction. This can e.g. This can be done by clamping the pivot support 12 or forming a snap connection for the pivot support 12 at least in sections of the recess element 19. large temperature changes occur the expansion differences, if the pivot support 12 designed as a wire ring is open, i.e. divided over the circumference at least in one place, whereby a displacement of the wire ring 12 in relation to the receptacle 19 in the circumferential direction is possible and thus the wire ring 12 can be adapted to the diameter changes of the receptacle 19.

In the exemplary embodiment shown, the post-adjustment element 17 is made of plastic material, such as e.g. made of a heat-resistant thermoplastic, which can also be ñber-reinforced. As a result, the post-setting element 17 can be manufactured in a simple manner as a spray part.

A post-setting element of plastic material with a lower specific weight obtains, as already mentioned, a lower mass inertia weight, thereby also reducing the sensitivity to pressure fluctuations. The swivel support could also be formed directly by the plastic material ring. However, the post-setting element 17 can also be produced as a sheet metal part or by sintering. Furthermore, with the corresponding choice of material, the pivot support 12 can be formed in one piece with the post-setting element 17. The pivot support 11 can be formed directly by the sensor spring 13. For this purpose, the tips of the tongues 13c can have corresponding indentations or information, such as e.g. falsar.

The post-adjustment ring 17 is centered through the axially extending portions 15a on the circumferentially distributed rivets 15. For this purpose, the post-adjustment ring 17 has centering contours 20 formed by circumferentially extending recesses 21 which lie radially inside the pivot support 11. To form the recesses 21, the post-adjustment ring 17 on the inner edge portion radially inwardly extending cams 22 which define radially inner contours on the recesses 21.

As shown in fig. 19 is, viewed in the circumferential direction, between the uniformly distributed recesses 21 arranged five ramp ramps 18 each. The recesses 21 P32724SE01 / LK 10 15 20 25 30 522 574 47 I 1. | a of are designed in the circumferential direction in such a way that they enable at least one angle of rotation of the post-adjustment ring 17 relative to the cover 2, which during the entire life of the friction clutch 1 guarantees a post-adjustment of it on the friction surfaces of the pressure plate 3 and the counterpressure plate 6 as well. at the friction pads 7 the wear occurring as well as possibly the wear of the coupling itself, thus Lex. the supports 11, 12, the intermediate disc spring portions, the pressure plate cams (at 3a) or the opposite portions of the disc spring 4. °. In the exemplary embodiment shown, the angle of rotation is in the range of 12 °, the upward inclination angle 23 of the ramp ramps 18 also being in the range of 12 °. This angle 23 is chosen in such a way that when pressed against each other by the ramp ramps 18 of the post-adjustment ring 17 and the counter ramp ramps 24 on it in fi g. 21 and 22, the friction arising from the riser 18 and 24. Depending on the material pairing in the area of the ramps 18 and the ramps 24, the angle 23 can be in the range between 5 and 20 °.

The post-adjustment ring 17 is spring-loaded in the circumferential direction, namely in the post-adjustment rotation, i.e. in the direction which, by the rise of the ramps 18 on the counter-ramps 24 of the support ring 25, causes a displacement of the post-adjustment ring in the direction of the pressure plate 3, i.e. thus in the axial direction away from the radial lid section 2a. At that in fi g. 17 and 18, the spring load of the post-adjustment ring 17 is guaranteed by at least one annular leg spring 26, such as e.g. may have two windings and has at one of its ends a radially extending leg 27, which is rotatably fixed with the post-adjustment ring 17, and at its other end has an axially extending leg 28, which is rotatably fixed in the cover 2. The spring 26 is resiliently biased inserted .

The support ring 25 shown in Figs. 21 and 22 is likewise formed by an annular structural member having counter-ramp ramps 24 which form complementary surfaces to the surfaces bounded by the ramp ramps 18, the surfaces bounded by the ramp ramps 18 and the counter ramp ramps 24 may also be congruent.

The angle of inclination 29 of the counter-ramp 24 corresponds to the angle 23 of the ramp 18. As can be seen from a comparison between fl g. 19 and 21, the ramp ramps 18 and the counter ramp ramps 24 are evenly distributed in the circumferential direction. Support ring P32724SE01lLK 10 15 20 25 30 522 574 48 1. A .- 25 are rotatably connected to the cover 2. For this purpose, the support ring 25 has recesses 30 distributed over the circumference, through which the riveting shoulders of the rivets 15 extend.

Fig. 18 shows in dash a further annular leg spring 26a, which, like the leg spring 26, can be bent down at its end portion, in order to guarantee a torsionally fixed connection with partly the lid 2 and partly the post-adjustment element 17. This spring 26a is also resiliently biased inserted, so that it exerts a torque on the post-setting element 17. The use of two leg springs 26, 26a can be advantageous for many use cases, since upon rotation of the friction clutch 1 as a result of the centrifugal forces acting on the spring 26 or 26a, a spring force amplification occurs. Through the use of two leg springs, it can e.g. the force amplification occurring on the spring 26 is compensated by the force applied by the shank spring 26a. For this purpose, the leg springs 26 and 26a are wound in such a way that, at least during the centrifugal force fl surface on the post-setting element 17, they generate forces which act in the opposite direction in the circumferential direction. The two leg springs 26, 26a may have one or more windings, and furthermore these leg springs 26, 26a may have different winding diameters, as shown in ñg. 18, whereby the centrifugal forces connected thereto and acting on the springs 26, 26a, which would generate different peripheral forces on the post-setting element 17, can be at least approximately equalized by correspondingly shaping the wire thicknesses and / or the number of windings of the individual springs 26, 26a. I fi g. 18, the spring 26 is arranged radially inside the post-adjustment element 17 and the spring 26a radially outside this post-adjustment element 17. However, both springs can by the corresponding design also be arranged radially inside or radially outside the post-adjustment element 17.

In Fig. 23, the leg spring 26 is shown in plan view. In the unloaded condition of the leg spring 26, the legs 27, 28 are displaced by an angle 31, which can be in the order of magnitude between 40 and 121 °. In the embodiment shown, this angle is 31 in the order of 85 °.

32 shows the relative position of the leg 27 relative to the leg 28, which it assumes in the case of a new friction lining 7 in the friction clutch 1. 33 shows the position of the leg 27, which corresponds to the maximum permissible wear on the friction lining 7. Post-adjustment angle 34 is in the embodiment shown in the order of 12 °. The spring 26 is designed in such a way that in the relieved state of this spring 26 only one wire winding 35 extends between the two legs 27, 28. In the other peripheral area two wire windings lie axially above each other. The spring 26a is designed in the same way as the spring 26, but nevertheless has a larger winding diameter and a different biasing direction in relation to the post-setting element 17 according to fi g. 18. However, the force exerted by the spring 26 on the post-adjustment ring 17 is greater than that of the spring 26a.

In the new state of the friction clutch 1, the ramp ramps 18 and the counter ramp ramps 24 forming axial cams 18a, 24a engage substantially axially with each other, which means that the superimposed rings 17 and 25 need the least axial installation space.

In the exemplary embodiment according to fi g. 17 and 18, the counter-ramp ramps 24 or the cam-shaped projections 24a forming them are formed by a separate structural part.

However, the counter-ramp ramps 24 may be formed directly through the cover 2, e.g. by indentations of cam-shaped approaches, which may extend into the lid interior. The indentation is particularly advantageous in the case of plate lids or lids which are formed in one piece.

In order to keep the adjusting ring 17 before mounting the friction coupling 1 in its retracted position, it has in the area of the cams 22 engaging portions 36 for a rotating or holding means, which on the other hand can support the cover 2. Such holding means can be arranged in the manufacture or assembly of the friction clutch 1 and is removed from the clutch after the mounting of the friction clutch 1 on the flywheel 6, whereby the post-adjustment device 16 is activated. In the embodiment shown, for this purpose elongate recesses 37 located in the cover or cover 2 are arranged in the circumferential direction and the post-adjustment ring 17 is a recess or shoulder 38. The elongate recesses 37 located in the circumferential direction must at least have such an extent that the return ring 17 can be retracted against the largest possible wear and tear adjustment angle. After the assembly of the friction coupling 1, a turning tool can also be inserted axially through the slots 37 of the lid and advanced in / to the recesses 38 of the adjusting ring 17. Thereafter, the ring 17 can be rotated back by means of the tool so that it is displaced in relation to this portion 2a occupies its smallest axial distance. In this position, the reset ring 17, e.g. through a clamp or a pin which engages in a aligned recess in the lid and the post-adjustment ring 17 and prevents a rotation of these two structural parts. This pin can be removed from the recess after mounting the friction clutch 1 on the flywheel 6, so that, as already mentioned, the post-adjustment device 16 is released. The slots 37 in the cover 2 are designed in such a way that during the disassembly or after the disassembly of the friction clutch 1 from the flywheel 6, the post-adjustment ring 17 can be brought to its retracted position. For this, the clutch 1 is first disengaged, so that the actuating disc brake 4 does not exert any axial force on P32724SE01 / LK. .-., .u 10 15 20 25 30 5 2 2 5 7 ¿, 'ii? ~ lf the swivel support 12 and thus a trouble-free rotation of the post-adjustment ring 17 is guaranteed.

A further possibility of bringing the structural parts of a friction clutch 1 already fastened to an internal combustion engine into a functional position consists in turning back or resetting the post-setting element or the post-setting ring 17 only after mounting it together on the internal combustion engine or on the flywheel. For this, e.g. by means of an auxiliary tool the friction coupling 1 is operated and the ring 17, which is then practically relieved, is set in its position retracted in relation to the pressure plate. Thereafter, the friction clutch 1 is engaged again, so that the ring 17 first maintains this retracted position.

In connection with those in the diagrams according to fi g. 24-27 plotted curves, the modes of operation of the friction clutch 1 described above will now be explained in more detail. _ Line 40 in tig. 24 shows the axial force generated in dependence on the change of conicity of the disc spring 4, namely when deforming the disc spring 4 between two supports, the radial distance of which corresponds to the radial distance between the pivot bearing 5 and the radially outer ice support diameter 3a on the pressure plate 3. the two supports and on the ordinate the force generated by the disc spring. Point 41 represents the plane position of the disc spring, which is suitably selected as the mounting position of the disc spring 4 with closed coupling 1, i.e. the position in which the disc spring 4 exerts the maximum pressing force on the pressure plate 3 for the corresponding mounting position.

The point 41 can be displaced along the line 40 upwards or downwards by changing the conical mounting position, thus the arrangement, for the disc spring 4.

The line 42 constitutes the axial disengagement force applied by the coating spring segments 10, which acts between the two friction linings 7. This axial disengagement force counteracts the axial force exerted by the disc spring 4 on the pressure plate 3. It is advantageous if the axial force required for the possible axial deformation of the cap segments 10 corresponds at least to the force exerted by the disc spring 4 on the pressure plate 3. When disengaging the friction clutch 1, the capercaillie segments 10 are relieved, namely over the section 43. Over this also a corresponding axial displacement of the pressure plate 3 corresponding to the section 43, the disengagement process of the clutch 1 is supported, which means at the non-occurrence of the coating tar segment 10 (when no coating tar thus occurs). When the point 44 is exceeded, the friction pads 7 are released, whereby due to the degressive curve portion of the disc teeth 4 the then still applicable disengagement force is considerably reduced in relation to that which would correspond to the point 41. The disengagement force P32724SE01 / LK 10 15 20 25 30 522 oc ø | o now for the coupling 1 decreases until the minimum or valley point 45 on the sinusoidal curve line 40 has been reached. When the minimum 45 is exceeded, the required disengagement force rises again, the disengagement distance in the area of the tongue tips 4c being chosen in such a way that even when the minimum 45 disengagement force is exceeded, the maximum disengagement force present at point 44 exceeds, but preferably remains below. Point 46 must therefore not be exceeded.

The spring 13 serving as a force sensor has a stretch-force course corresponding to the line 47 in fi g. 25. This curve 47 corresponds to that generated when the disc-like structure part 13 changes from the unloaded position in its conicity, namely between two pivot supports, which have a radial distance corresponding to the radial distance between the pivot supports 11 and 14. the disc-spring-like structural part 13 a spring path 48, over which the axial force generated by them remains virtually constant. The force generated in this lot 48 is then chosen in such a way that it “at least roughly corresponds to that in point 44 in ñg. 24 adjacent the disengagement force for the coupling. The support force applied by the sensor spring 13 is reduced in relation to the force corresponding to point 44 of disc edema 4 in accordance with the lever exchange of this disc spring 4. This gear ratio is in most cases in the order of magnitude between 1: 3 and 1: 5. , however, for many use cases it can also be larger or smaller.

The said disc spring gear corresponds to the ratio between the radial distance of the pivot bearing 5 to the support 3a and the radial distance between the pivot bearing 5 and the abutment diameter 4c, e.g. for a disconnect bearing.

The mounting position of the disc-like element 13 in the friction clutch 1 is chosen in such a way that in the area of the pivot bearing 5 it can undergo an axial capercaillie path in the direction of the friction lining 7, which corresponds to at least the axial readjustment path of the pressure plate 3 towards the counterpressure plate 6. , which occurs as a result of the friction surface and friction coating wear, which also has an at least approximately constant axial support force for the pivot bearing 5. This means that the linear area 48 of the curve line 47 must have at least a length corresponding to said wear path, and is preferably greater than this. wear path, as this also allows installation tolerances to be at least partially equalized.

In order to obtain a practically constant or defined release point 44 for the friction pads 7 during the disengagement of the friction clutch 1, a so-called double-segment suspension is used between the friction linings 7, i.e. a lining suspension, in which individual pairs of spring springs are arranged back to back, the individual pairs of segments being able to have a certain axial mutual pretension, so that the total axial force applied by the coating suspension at non-clamped clutch disc 8 corresponds at least to the decoupling force corresponding to point 44 on disc edgings 4, or is preferably slightly higher. By the biasing of the spring means arranged between the coatings, it can be achieved that the embedding losses occurring during the operating time of the segments are at least substantially equalized or compensated for on the back of the coatings. Embedding losses are understood to mean the losses that arise from the incorporation of the segments in the back of the coatings.

It is suitable if the prestress of the suspension arranged between the coatings is in the order of 0.3 mm to 0.8 mm, preferably in the order of 0.5 mm.

By a corresponding limitation of the axial spring path between the two friction linings 7 as well as by a defined prestressing of the suspension acting between the friction liners it can further be achieved that at least when the friction clutch 1 is disengaged the pressure plate 3 is pressed back by a defined distance 43 . In order to obtain a defined distance 43, the axial distance between the friction linings can be limited by corresponding abutments both in the relief direction and also in the insertion direction of the lining suspension 10. As lining springs can in advantageously be used in connection with the present invention such as those known from German patent application P 42 06 880.0.

In Figure 26, line 49 shows the force required to disengage the coupling through a disengagement element engaging the plate 4c portion 4c, to move the pressure plate from point 41 to point 44 (fi g. 24). Line 49 further shows the distance of the tongue tips on the disc spring in the portion 4c.

In order to ensure an optimal function for the friction clutch 1 or for the automatic adjustment of the coating wear guaranteeing the post-adjustment device, it is suitable that - considered over the actual deceleration force course 49 according to ñg. 26 - the forces exerted first by the coating spring 10 and sensor spring 13 on the disc spring 4 and the increasing forces are greater than the force exerted by the disc spring 4 on the support 11. Even after the pressure plate 3 is lifted from the friction pads 7, the force exerted by the sensor spring 13 on the disc spring 4 must then be greater, or at least as great, as that attacking in the area 4c on the disc spring tongue tips and in accordance with fi g. 26 over the disconnection path required and changing the disconnection force (according to line 49). The force exerted on the support 11 by the sensor disc spring 13 on the support 11 must furthermore be so adapted that a rotation of the ring 17 under the force of the spring 26 and thus an axial displacement of the disc spring is prevented, at least approximately until the point 41 of the rising branch on the curve line 40 corresponding to the mounting position of the disc spring is not exceeded.

The consideration hitherto corresponds to a completely definite installation position for disc edema 4, and no wear on the friction pads 7 has been taken into account.

In the case of axial wear, especially of the friction linings 7, the position of the pressure plate 3 is displaced in the direction of the counterpressure plate 6, whereby a change in the conicity of the disc springs (tongue tip 4c travels, as seen by the viewer, to the right) and thus a change in condition of the pressing force applied to the friction clutch 1 arises, namely in the direction of an increase. This change causes point 41 to move in the direction of point 41 ', and point 44 in the direction of point 44'. This change disturbs the force equilibrium originally present when disengaging the clutch 1 in the area of the swivel support 11 between the actuator plate fi edge 4 and the sensor spring 13. . The resulting disengagement force course is shown in Fig. 26 by the broken line 50. axial distance, which mainly corresponds to the wear of the friction lining 7. During this penetration phase of the sensor spring 13, the disc spring 4 abuts against the actuating region 3a on the pressure plate 3, so that this disc spring 4 changes its conicity and thus also the energy stored therein and the torque stored therein. consequently also the force exerted by the disc spring 4 on the swivel support 11 or the sensor spring 13 and on the pressure plate 3. This change takes place, as can be seen in connection with Fig. 24 in the direction of a reduction of the forces applied by the disc spring 4 to the pressure plate 3. This change takes place as long as until the axial force exerted by the disc spring 4 in the area of the pivot support 11 on the sensor spring 13 is in equilibrium with the counter-force generated by the sensor spring 13. This means that in the diagram according to Fig. 24 the point diagrams 41 'and 44' travel again in the direction of the points 41 and 44. After this equilibrium has been set again, the pressure plate 3 can be lifted up again from P32724SE01 / LK 10 15 20 25 30 522 574 During the post-adjustment phase of the wear, while thus in a disengagement process for the friction clutch 1 the sensor spring 13 yields, the post-adjustment element 17 of the post-adjustment device 16 is rotated by the prestressed spring 26, whereby also the pivot support 12 travels after the corresponding coating and a game-free pivot bearing 5 of the disc spring 4. After the readjustment process, the disengagement force sequence again corresponds to the line 49 according to fi g. 26 lines 50 and 51 in fi g. 26 resp. the axial distance of the pressure plate 3 at a decoupling force-distance course corresponding to the lines 49, 50.

In the diagram according to fi g. 27, the force course over the disengagement distance of the force exerted during a disengagement process on the lid 2 or on the disc tether 13 is shown, the extreme values having been cut. Based on the switched on position according to fi g. 17 acts on the cover 2 and thus on the pressure plate 3 first a force corresponding to the installation point 41 (fi g. 24) of the disc spring 4. During the disengagement process the axial force exerted by the disc spring 4 on the cover 2 or the pivot support 12 corresponding to line 52 in ñg decreases. 27, namely up to point 53. Exceeding point 53 in the direction of disengagement, in the case of a conventional coupling in which the disc spring is axially fixedly pivotably mounted on the cover, i.e. the pivot support 11 would be axially resiliently connected to the cover 2, through the disc spring 4 on the lid 2 take place at the radial height of the pivot bearing 5. In the coupling according to the invention, the force generated by the pivot bearing 54 lifts the disc spring 4 up from the impact area 3a 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 applied by the coating spring 10, since it acts against the disc spring force. The force applied by the coating spring 10 then decreases with increasing disengagement distance in the area 4c of the tongue tips or with increasing axial disengagement stroke for the pressure plate 3.

The line 52 thus shows a result of the uncoupling force observed over the disengagement process, partly in the tongue tip portion 4c and partly in the axial force exerted in the radial portion 3a of the disc spring 4 by the coating spring 10. When the point 54 in the disengagement direction is exceeded, the axial force exerted by the disc spring 4 on the pivot support 11 is captured by the counter-force applied by the sensor disc spring 13, these two forces at least after relieving the friction pads 7 through the pressure disc 3 the sensor spring 13 in the area P32724SE01 / LK 10 15 20 25 30 55 of the axial force applied by the pivot bearing 5 is preferably slightly greater than the adjacent disengagement force. Subpart 55 of the curve line 52 of the diagram according to ñg. 27 shows that with increasing disengagement path, the disengagement force or the force exerted by disc edgings 4 on the swivel support 11 becomes smaller compared with the disengagement force applied at point 54.

The dashed line 56 corresponds to a state of the friction clutch 1, in which a wear has occurred in the area of the friction pads 7, but no readjustment has yet taken place in the area of the pivot bearing 5. Here again it appears that the change in the mounting position caused by the wear the disc spring 4 causes an increase in the forces exerted on the cover 2 and on the swivel support 11 or on the sensor spring 13. This has the particular consequence that the point 54 travels in the direction of the point 54 ', which causes that in the case of renewed disengagement of the friction clutch 1 the axial force exerted by the disc spring 4 on the sensor spring 13 in the region of the pivot support 11 is greater than the counter force of the sensor spring 13. the process of post-resetting already described by axial dilution of the sensor dam 13 takes place. Due to the post-adjustment process effected by the spring 26, thus by the rotation of the ring 17 and the axial displacement of the support 12, the point 54 'is displaced again in the direction of the point 54, whereby the desired state of equilibrium in the region of the pivot support 5 between the disc spring 4 and the sensor spring 13 is maintained.

In practice, the described readjustment takes place continuously or in very small steps, so that the large point offset series shown in the diagrams for a better understanding of the invention and the curve line offsets never normally occur.

During the operating time of the friction clutch 1, some operating parameters or operating points may change. So can Lex. due to an incorrect operation of the friction clutch 1 an overheating of the coating suspension 10 takes place, which may result in a settlement, i.e. a reduction of the axial suspension of the coating suspension or the coating segments 10. By a corresponding design of the curve line 40 for the disc spring 4 and corresponding adjustment of the course 47 for the sensor 13 spring 13, however, a reliable function for the friction clutch can be guaranteed. An axial setting of the coating suspension 10 would only have the effect that the disc spring 4 would assume a position pressed back relative to the position shown in Fig. 17, whereby the pressing force exerted by the disc spring 4 on the pressure plate would be somewhat lower, as can be seen in connection with curve line 40 according to ñg. 24. Furthermore, a corresponding axial deformation of the sensor spring 13 and thus a corresponding axial displacement of the pivot support 11 would take place.

P32724SE0l / LK 10 15 20 25 30 u uno ln no nu nu nu nu en 'I q nu en nu av 0 c O: uo: o to or o: co I u un »un:» se nu n la o: nu: nn no snv V ca | noo nu o 0 ou naoocoras: oo Is vu I co u: According to a further inventive concept, the resulting support force acting on the actuating disc springs 4 may increase with increasing wear of the friction pads 7. The increase may be limited to a portion of the overall maximum allowable wear path for the friction pads 7. The increase in the support force for the actuating disc brake 4 can then take place by correspondingly designing sensor fi edema 13. I fi g. Is indicated by a dashed line and the reference numeral 47 indicates a corresponding curve line course over the area 48. By increasing the support force for the actuating plate spring 4 with increasing wear, a compressive force decrease of the actuating plate spring 4 for the pressure plate 3, depending on e.g. by embedding the segments in the coatings, at least partially compensated. It can be particularly advantageous in this case if the support force for the actuating plate spring 4 increases in proportion to the setting of the padding suspension or in proportion to the segment embedding in the pads. This means that with the reduction of the disc thickness in the area of the coatings, i.e. a reduction of the distance between the friction surfaces of the coatings due to segment embedding and / or a settling of the coating suspension and / or the coating wear, the said support force shall increase. It is particularly advantageous if the increase in force takes place in such a way that this over a first sub-portion is greater than in an adjoining second sub-portion, the two sub-portions being within the area 48 according to fi g. The latter design is advantageous, since the greater part of the said embedding between the spring segments and the pads mainly takes place within a small time in relation to the total life of the friction clutch, and thereafter the relationships between spring segments and friction pads are practically stabilized. This means that from a particular embedding there is no longer any significant change with regard to the embedding. An increase in the support force for the actuating disc spring can also take place over at least a part of the abrasion wear of the friction pads.

In the above description of the post-adjustment process for the smoothing of the friction lining wear, no account has been taken of the axial forces possibly applied by the leaf springs 9. When the leaf spring 9 is prestressed in the direction of a lifting of the pressure plate 3 from the corresponding friction lining 7, i.e. in the direction of a pressing of the pressure plate 3 against the disc spring 4, a support of the disengagement process takes place. It is superimposed on the axial force applied by the leaf spring 9 with the forces applied by the sensor spring 13 and disc spring 4 as well as by the disengagement force. This has so far not been taken into account due to better understanding in the description of the diagrams according to fi g. 24 - 27. the leaf spring segments 9, by the sensor spring 13 and by the prevailing disengagement force on the actuating plate spring 4. The leaf spring elements 9 can then be assembled between the cover 2 and the pressure plate 3 in such a way that with increasing wear of the friction linings 7 the spring spring 9 becomes larger. Then e.g. over the section 48 according to fig. And also over the wear compensation path of the post-setting device 16 the axial force applied by the leaf spring 9 has a course along the line 47b. Av fi g. It also appears that with increasing suspension of the sensor spring 13, the restoring force exerted by the leaf spring 9 on the pressure plate 3, which also acts on the actuating disc spring 4. By adding the force course according to the curve lines 47b and the plate spring curve line, the resulting force course direction for pressing the plate edema 4 against the lid-side pivot support 12. In order to obtain a course along the line 47a, whereby at the beginning of the setting area 47d there is first an incipient increase in force, which transitions to an approximately constant force range, it is suitable to design the sensor disc spring such that it has a curve line sequence corresponding to the line 47c in fi g. 25. By adding the force course according to line 47c and the force course according to line 47b, the force course according to line 47a is then obtained. By means of a corresponding prestressing of the leaf fi springs 9, the support force exerted by the sensor spring or the support force course can thus be reduced. By corresponding design and arrangement of the leaf spring elements 9, a lowering of the coating suspension and / or an embedding of the coating spring segments in the coatings can also be at least partially compensated. This ensures that the disc spring 4 maintains substantially the same operating point or the same mode of operation, so that the disc spring 4 exerts a substantially constant pressing force on the pressure plate 3 during the service life of the friction clutch. by the resulting axial force generated by the resetting spring 26 and / or 26a acting on the resetting element 17, account is taken of which counteract the sensor spring 13 and / or the leaf springs 9.

When designing the friction clutch 1 with prestressed leaf springs 9, it must also be taken into account that the axial force exerted by the pressure plate 3 on the friction linings 7 is affected by the prestressing of the leaf springs 9. This means that in the case of a biasing of the leaf springs 9 in the direction of the actuating disc spring 4, the pressing force exerted by the disc spring 4 now decreases with the biasing force of the leaf springs 9. In such a friction clutch 1 a resultant compressive force course is thus formed for the pressure plate 3 or for the friction linings 7, which are obtained by superimposing the compressive force course of disc spring 4 with the prestressing course of the leaf springs 9. Assuming that - considered over the friction clutch 24 shows the resulting force course from the actuating disc spring 4 and prestressed leaf springs 9 in the new state of the friction clutch 1, with the lowering of the distance between the pressure plate 3 and the back pressure plate 6 due to the coating wear a displacement of the resulting course towards a reduction would be obtained. In Fig. 24, the line shows a line 40a, which corresponds, for example, to a total coating wear of 1.5 mm. Due to the displacement of the line 40 in the direction of the line 40a occurring during the service life of the friction clutch, the axial force exerted by the disc spring 4 on the sensor spring 13 on the sensor spring 13 is lowered, namely due to increasing wear by the spring springs 9 on the spring. This counter-torque occurs due to the radial distance between the pivot bearing 5 and the actuating diameter 3a between actuating disc spring 4 and pressure plate 3. When designing the friction clutch 1, it is especially important that the increase of the blade the disengagement force increase occurs, which causes the pressure oscillation of the sensor spring 13 required for the post-adjustment.

The friction clutch 101 shown in Figs. 28 and 29 differs substantially from that in Figs. 17 and 18 show the friction clutch 1 in that the post-adjustment ring 117 is loaded in the circumferential direction by means of coil springs 126. As regards its function and mode of action with regard to the wear compensation of the friction linings, the post-adjustment ring 117 corresponds to the post-adjustment ring 17 according to Figs. 18-20. In the embodiment shown, three coil springs 126 are provided, which are uniformly distributed over the circumference and prestressed between the coupling cover and the post-adjustment ring 117.

As shown in particular in Fig. 30, the adjusting ring 117 on the inner circumference has radial projections or ledges 127, against which the arcuately arranged coil springs 126 can abut in the circumferential direction with their ends for actuating the adjusting ring 117. The other end portions of the springs 126 support the coupling stop 128. In the embodiment shown, these stops 128 are formed by screw-like connecting elements, which are connected to the cover 2. However, these stops 128 can also be formed. by axial embodiments, which are formed in one piece with the coupling cover 2. Thus, e.g. the stops 128 are formed by extrudes or fl ices axially formed from a sheet metal lid 2. As can be seen in particular from Figs. 29 and 30, the ring 117 on the inner circumference can be designed in such a way that at least substantially in the area of the extent of the springs 126 and preferably also over the angle of rotation required for the post-adjustment of wear 117 or over the relief path for the springs 126 a guide 129, which guarantees an axial holding and radial support of the springs 126. In the embodiment shown, the spring guides 129 are formed by, in cross-sectional view, substantially semicircularly shaped depressions, the limiting surfaces of which are substantially adapted to the cross-sections of screws 12.

Such a design has the advantage that during rotary friction clutch a trouble-free control of the springs 126 is given, so that these cannot be bent out axially. For further securing of the helical springs 126, as shown in Fig. 2913, the lid 2 on its radially inner edge portion may have axial configurations 130, which overlap the springs 126 in the shaft direction. Instead of individual configurations 130, the lid 2 can also have a circumferential and axial inner edge 130 which extends over the circumference. The inner edge 130 can serve to limit the relief of the disc spring 4.

A control of the post-adjusting springs 126 according to Figs. the reset ring 117 generated by the frictional resistors is lowered or empty repealed. The springs 126 can also be held practically rigidly when the friction clutch 101 is rotated (due to the frictional forces suppressing the spring action).

Thereby it can be achieved that at least at speeds above the idle speed of the internal combustion engine the post-setting ring 117 can not be rotated through the springs 126. Thereby it can be achieved that a smoothing of the friction lining wear only occurs when operating the friction clutch 101 at idle speed or at least idle speed. However, the post-adjustment ring 117 can also be blocked in such a way that only with a stationary internal combustion engine, i.e. non-rotating friction clutch 101, a post-adjustment can take place due to the wear of the coating.

P32724SE0l / LK 10 15 20 25 30 522 574 eo A blocking of the post-setting process when rotating the friction clutch 1 or when exceeding a certain speed can also be advantageous in an embodiment according to fl g. 17 and 18. For this purpose, means can for instance be arranged on the cover 2, which under the action of centrifugal force on the post-setting element 17 provides a torsional protection, namely against the adjusting force generated by the leg jig 26 and / or 26a. The blocking means can then be formed by at least one weight radially outwardly displaceable under the action of centrifugal force, which, for example, abuts the inner edge of the ring 17 and there can generate a friction which on the ring 17 produces a braking torque greater than the torque applied by the adjusting springs. .

For the radial support of at least a part of the extension of the springs 126, support means carried by the cover 2 can also be provided. These aids can at the execution form according to fi g. 28 and 29 may be formed integrally with the stops 128. For this purpose, the stops 128 may be formed angularly so that they each have a portion extending in the circumferential direction, which at least over a partial section of the extension of a spring 121 projects into it. As a result, at least a part of the capercaillie windings can be guided and at least supported in the radial direction.

As can be seen from Fig. 29, the wire ring 11 provided in Fig. 18 has been omitted and replaced by moldings 111 applied in the tongue tip portion of the sensor cap 113.

Figs. 31 to 33 show a further design variant of a wear post-adjustment according to the invention, in which instead of an annular post-adjustment ring individual post-adjustment elements 217 are used. These post-adjustment elements are uniformly distributed over the circumference of the lid 202. The post-setting elements 217 are formed by knob-shaped or disc-shaped construction parts, which have a circumferentially extending and axially ascending ramp 218. The annular post-setting elements 217 have a central recess or bore 219, through which the axial pin-like axial struts 21 are supported by the lid. that the annular post-adjustment elements 218 are rotatably mounted on these shoulders 215a. On the cover 202, extensions 225 are arranged, which form counter-ramp ramps 224 for the ramps 218. Between a post-adjustment element 217 and the cover 202, a spring element 226 is clamped, which influences the post-adjustment element 217 in the direction of rotation which provides a post-adjustment. The spring element 226 can extend, as can be seen from Fig. 31, around an axial shoulder 215a, i.e. be screw-spring-like P32724SED1 / LK 10 15 20 25 30 522 574 61 designed. At the end portions of a spring 226, embodiments such as e.g. deflections or legs arranged to support one capercaillie end against the lid 202 and the other capercaillie end against the corresponding post-adjustment element 217. Upon an axial displacement of the disc spring 204 or of the sensor spring 213 in the region of the pivot support 205, the post-adjustment element 217 is displaced by the frame opening on the ramps 224.

The axial support of the sensor plate spring 213 against the cover 202 takes place by means of plates 214, which are formed from the axially extending portion of the cover 202 and radially inwardly pressed under the outer portions of the sensor spring 213.

The annular post-adjustment elements 217 have the advantage that they can be designed substantially centrifugally force-independent with respect to their post-adjustment effect.

Instead of that in ñg. 31 rotating or rotating post-adjustment element 217, separate wedge-like post-adjustment elements can also be used, which are displaceable in radial direction and / or in the circumferential direction for wear post-adjustment. These wedge-like readjustment elements may have a longitudinal recess through which an axial shoulder 215a for guiding the corresponding readjustment element may extend. Due to the centrifugal force acting on them, the wedge-shaped post-adjustment elements can have a post-adjustment effect. However, the power magazines can also be arranged, which actuate the wedge-shaped post-adjustment element in the post-adjustment direction. For the trouble-free control of the wedge-like post-adjustment elements, the cover 202 can have attachments. The wedge surfaces of the post-setting elements which extend perpendicularly to the axis of rotation of the friction clutch with a certain running angle can be arranged on the lid side and / or on the side of the actuating disc spring. When using such wedge-shaped individual elements, it is convenient to make these from a light material, in order to reduce the centrifugal forces acting on it to a minimum.

The material pairing between the post-adjustment ramps forming the structural parts is suitably selected in such a way that during the entire service life of the friction coupling, no post-adjustment preventing adhesion between the run-up ramps and the counter-run-up ramps can occur. To avoid such adhesion, for example, such a structural part may be provided with a coating at least in the area of the ramps or counter-ramps. Due to such coatings, corrosion in particular can be avoided when using two metallic structural parts. An adhesion or gluing between the structural parts forming the post-ramp ramps can further be avoided in that the supporting parts and the structural parts forming the ramps as well as the counter-ramps are made of a material with different coefficient of expansion, so that as a result of the temperature fluctuations occurring during the operation of the friction clutch, the contacting surfaces, which form the post-adjustment ramps, perform a movement relative to each other. Thereby, the ramp ramps and the counter-ramp ramps forming the structural parts are always kept mutually movable. An adhesion or gluing between these parts can thus not take place, as due to the various extensions these parts are always broken off or detached from each other again. A loosening of the post-adjustment ramps can also be achieved by, due to different strengths and / or design of the parts, the centrifugal forces acting on these parts cause different expansions or movements, which in turn prevent an adhesion or gluing of the parts.

In order to avoid an adhesive connection between the ramp-up ramp and the counter-ramp-up ramp, at least one device can also be provided which, during the disengagement of the friction clutch or during the wear-off resetting, exerts an axial force on the post-setting element or post-setting elements. For this purpose, the post-adjusting element 17, 117 can be axially coupled to a structural part, which has portions which, in the event of wear, are displaced axially. This coupling can take place especially in the area of the pivot bearing 5, namely with the actuating plate spring 4 and / or the sensor spring 13.

In the diagram according to ñg. 34, a pressing plate spring curve line 340 is shown, which has a valley point or a minimum 345, in which the force applied by the pressing plate spring is relatively low (about 450. Nm). The maximum number of disc springs with the road-power curve line 340 is in the order of 7600 Nm. The curve line 340 is generated by deformation of the disc spring between two radially spaced supports, namely as described in connection with the curve line 40 according to fi g. 24 and in connection with disc spring 4.

Disc plate curve line 340 can be combined with a coating fi vein curve line 342.

As shown in fl g. 34, the path force of the coating spring segment curve line 342 is approached to the pressing plate spring curve line 340, or the two curve lines extend only a small distance from each other, so that the corresponding friction clutch can be operated with very low force. In the field of activity of the evidence, the theoretical decoupling force is obtained from the difference between two vertically overlapping points of lines 340 and 342. Such a difference is indicated by 360. The really required decoupling force is reduced by the corresponding hx. of the operating elements, such as e.g. the disc spring tongues. This has also been described in connection with the embodiments according to fl g. 17 and 18 as well as the diagrams according to fi g. 24-27.

I fi g. 34 is shown in phantom an additional actuating disc spring line 440, which has a minimum or a valley point 445, in which the force applied by the disc spring is negative, i.e. not in the engagement direction of the corresponding friction clutch, but acts in the disengagement direction. This means that when point 461 is exceeded during the disengagement phase, the friction clutch automatically opens. The disc spring curve line 440 can be provided with a coating fi spring curve line corresponding to the line 442. In order to obtain minimal disengagement forces, the aim is as parallel as possible between the coating der spring curve line 442 and the disc spring curve line 440.

I ñg. For the disengagement of the corresponding friction clutch on the actuating lever, such as the disc tongue tongs, the disengagement force course can be applied in the over disengagement distance of the associated curve lines 340 and 342, respectively. 440 and 442.

As can be seen, the decoupling force course 349, which is coordinated with the curve lines 340, 342, is always in the positive force range, i.e. that in order to keep the clutch in the disengaged state, a force in the disengagement direction is always required. The disengagement force course 449, which belongs to the curve lines 440 and 442, has a sub-portion 449a in which the disengagement force first decreases and then passes from the positive to the negative force range, so that the corresponding friction clutch in the disengaged state does not need any holding force.

At the i fi g. 36, 36a and 37 of the friction clutch 501, the sensor plate spring 513 is axially supported against the clutch cover 502 by a bayonet-like connection 514, for this the sensor spring 513 has radially from the outer circumference of the annular base body 513b extending radially 502a, in the form of utform ices formed from the lid material. The lid edges 502a are formed by the substantially axially extending edge portion 502b of the lid, whereby it may be suitable if for this the edges 502a are at least partly first formed by a free cut 502c or 502d of the lid material. By at least partially circumcising the ices 502a, these can be more easily defined to their intended positions. As especially shown by fi g. 37, the 50s 502a and the outriggers or tongues 5l3d are so adapted to each other that a centering of the sensor spring 513 can take place relative to the cover 502. In the embodiment shown, the 50s 502a do not have an axial shoulder 502c for this purpose. "To guarantee a trouble-free positioning the sensor cap 513 relative to P32724SE01 / LK 10 15 20 25 30 522 574 64 In the case of the cover 502 during the manufacture of the bayonet-like locking connection 514, at least three preferably over the circumference of the cover 502 are uniformly distributed tabs 502a relative to the second cover portion. so adapted that after a defined relative rotation between sensor sensor 513 and cover 502 the corresponding outriggers 513d come into contact with a circumferential stop 502f and thus a further relative rotation between sensor spring 513 and cover 502 is avoided. The stop 502f is of fi g. 36a formed by an axial shoulder on the cover 50 2. Av ñg. 36a it further appears that at least individual, at least three 50s 502a form an additional torsional restriction 502g between the lid 502 and the tongues 513d of the sensor spring 513d. In the example shown, the equals 502a form the torsion bar 502f and 502g for both directions of rotation. The stops 502g preventing the unlocking between the sensor spring 513 and the cover 502 are formed by axial, radially extending bends of the tongues 502a. The circumferential stops 502f and 502g provide a defined positioning in the circumferential direction of the sensor spring 513 relative to the cover 502. For the production of the locking connection 514, the sensor spring 513 is axially tensioned in the direction of the cover 502, so that the tongues 513d axially project into the free cuts 5020 and 502d. lie axially over the cover supports 502a. Thereafter, the cover 502 and the sensor spring 513 can be rotated relative to each other until some of the tongues 513d come into contact with the rotation restrictions 502f. Then a partial relief of the sensor spring 513 takes place, so that some of the tongues 513d, viewed in the circumferential direction, will lie between the corresponding stops 502f and 502g and all the tongues 513d are stored on the lid-side supports 502a. The design according to the invention of the bayonet-like reading 514 guarantees that during mounting at the friction clutch 501, the tongues 513d will not lie next to the lid-side supports 502a.

In the embodiments shown so far, the actual spring force for the sensor spring 513 is applied to the circular annular basic body, e.g. 513b, arranged radially outside the actuating portion or the support portion between the pressure plate and the actuating plate.

For many use cases, however, it may also be suitable if the circular annular basic body of the sensor disc spring is arranged radially inside the actuating diameter between the pressure plate and the operating disc spring. It thus means for an embodiment according to ñg. 17 and 18 that the axial biasing force of the sensor spring 13 applying to the base body 13b is arranged radially inside the actuating portion 3a between actuating plate springs 4 and pressure plate 3.

In the embodiment according to fig. 36-37, the lid-side counter-ramp ramps 524 P32724SE0lfLK 10 15 20 25 30 ns 526 through guide pins 528, which are integrally formed with the post-adjustment ring 517 and extend in the circumferential direction. These guide pins 528 may, as shown in particular in Fig. 37, have in the axial direction an elongate cross-section adapted to the inner diameter of the springs 526. The guides 528 extend over at least a part of the longitudinal extension of the springs 526 into them. As a result, at least a part of the spring windings can be guided and at least supported in the radial direction. Furthermore, a deflection or a suspension of the springs 526 in the axial direction can be avoided. Through the pins 528 the mounting of the friction clutch can be considerably facilitated.

In Fig. 38, the readjustment ring 517 is shown in part. The post-adjustment ring 517 has radially inwardly extending configurations 527 which support the pin-shaped circumferentially extending guide portions 528 of the coil springs 526. In the embodiment shown, the spring receiving portions 528 are integral with the resilient portion 52, respectively. be formed by individual structural members and / or all jointly by a single structural member, which or which is connected to the post-adjustment ring 517, e.g. by means of a snap lock. Thus, all the guide portions 528 may be formed by a ring also open over the circumference, which is connected to the post-adjustment ring 517 over at least three connecting points, preferably designed as a snap lock.

Similar to what is described in connection with fi g. 28 and 29, the coil springs 526 may radially abut further, e.g. due to centrifugal force action, against correspondingly shaped portions on the cover 502 and / or the post-adjustment ring 517.

The cover-side supports for the helical springs 526 are formed by extensions 526a formed by the cover material and extending in the axial direction or by extensions 526a formed by axial walls. These support portions 526a for the spring frames 526 are then mandatorily designed in such a way that the corresponding ends of the springs are guided and thus secured against an impermissible displacement in the axial and / or radial direction.

In the embodiment of a coupling 601 shown in Fig. 39, the sensor spring 613 is arranged on the side of the cover 602 facing away from the pressure plate 603. the risk of a setback of this spring 613 due to a thermal overload is avoided. The same applies to the outer side P32724SE0l / LK 10 15 20 25 30 n n | n n now 522 574 66 of the cover 602 a better cooling of the spring 613.

The support 611 arranged on the pivot side 611 of the actuating disc spring 604 is supported by means of spacer rivets 615, which extend axially through corresponding recesses in the disc spring 604 and the cover 602 and are axially connected to the sensor spring 613. sensor spring 613. Instead of spacer rivets 615, other means can also be used which form a connection between the unwinding support 611 and the sensor spring 613. Thus, e.g. the sensor spring 613 in the radially inner portion have axially extending fls, which support the unwinding support 6111 with corresponding radial portions or even directly this unrolling support 611 forms by corresponding embodiments. Instead of the elements 615 fixed to the sensor spring, they can also be designed in other ways, e.g. articulated at sensor attached elements can be used.

In the embodiment according to ñg. 40, the sensor spring 713 extends radially within the pivot bearing 715 for actuating disc ead 704. Sensor spring 713 is supported at its radially inner portion against the cover 702. To this end, the cover 702 has axially extending through corresponding slots or recesses in disc spring 704 extending axially. the sensor disc spring 713. According to a second embodiment variant, the sensor spring 713 could also have on its inner edge portion fl axes, which extend in the axial direction axially through corresponding openings in the disc spring 704 and are supported on the lid side.

The i ñg. 41 can be used in a friction clutch according to fi g. 36-37. The post-adjustment ring 817 has radially inner configurations 827 which extend radially. The configurations 827 have radial projections 827a, which form the support portions for the helical springs 826 prestressed in the circumferential direction between the coupling cover and the adjusting ring 817. The ring 828 is connected to the radial configurations 827a. For this, the configurations 827a may have circumferentially extending depressions or grooves which are designed in such a way that they form a snap connection in connection with the ring 828. The cover-side supports for the post-adjustment springs 826 are formed by axial axes 826a on the coupling cover. The axial members 826a have an axial notch 826b for receiving the ring 828. The notches 826b are then designed in such a way that the ring 828 has an axial displacement possibility relative to the members 826a, at least corresponding to the wear path of the friction clutch. For this, it is particularly suitable if the depressions inserted in the radial configurations 827a for receiving the ring 828 and the cut-outs 826b, viewed in the axial direction, are oppositely formed, or in other words, that the depressions in the configurations 827a are open in one axial direction and the cutouts 826b are open in the other axial direction.

At the i fi g. 42, the friction clutch 904 is supported in the disengagement direction in a central portion of the base body 904 of the disc spring 904. Radially outside, the base body 904a abuts the thrust plate 903 and extends radially inwardly beyond the pivot bearing 905. This means that the pivot bearing 905 is at a relatively large distance from the inner edge of the base body 904a of disc edge 904 or slit end artefacts which form tongues on the disc spring 904. known disc spring couplings. In the exemplary embodiment shown, the radial width ratio between the basic body portion arranged radially inside the pivot bearing 905 and the basic body portion located radially outside the pivot bearing 905 is in the order of 1: 2. It is appropriate if this ratio is between 1: 6 and 1: 2. By such a support of the actuating disc spring 904, a damage or an overloading of the disc spring base 904a in the area of the pivot bearing 905 can be avoided.

I tig. 42 further indicates an axial attachment 903a, which is arranged on the pressure plate 903. Over such arrangements 903a arranged on the pressure plate 903, especially in the area of the support cams 903b, the actuating plate spring 904 can be centered relative to the coupling 901. The actuating ring in the radial direction relative to the cover 902, so that in fi g. 40 also shown the centering rivets or pins 915 may be omitted. Although this is not shown, the outer diameter centering can also take place through flaps or extensions formed from the material of the lid 902.

At the friction clutch 901, the sensor spring 913 is formed in such a way that the force applying to the base body 913a is arranged radially inside the cams 903b. For supporting the actuating disc spring 904 on the one hand and for its own support against the cover 902 on the other hand, sensor 9 edem 913 has radial outriggers or tongues, which extend partly from the base body 913a radially inwards and partly from the base body 913a radially outwards.

At the i fi g. The embodiment variant of a friction clutch 1001 shown in Fig. 43 is the disengaging force of the friction clutch or the pivoting force of the actuating disc spring 1004 applied by a sensor spring 1013, which is axially clamped between the cover 1002 and the pressure plate 1003. In such an embodiment, the actuator 68 1004 in the pivot ~ or tipporrlrådet 1005 not by any pivot bearing in the disengagement direction.

The abutment of the disc spring 1004 against the lid-side pivot support or the support support 1012 is guaranteed by the biasing force of the sensor spring 1013. This sensor spring is designed in such a way that during the disengagement process of the friction clutch 1001, the axial force applied by this sensor spring 1013 to the disc spring 1004 becomes greater than the required disengagement force of the friction clutch 1001. It must be guaranteed that when there has been no wear on the friction pads, the disc spring 1004 always abuts against the lid-side support or the swing supports 1012. For this, in a similar manner, as described in connection with the previous embodiments, an adjustment must take place between the individual in the axial direction the acting and superimposing forces. These forces, which are generated by the sensor spring 1013, by the coating spring through the leaf spring elements possibly arranged between the pressure plate 1003 and the cover 1002, by the actuating disc spring 1004, 1001 the post-adjustment ring 1017 acting on the post-adjustment spring elements must be adapted by the disengagement force.

At the friction clutch 1101 according to fi g. 44 then supports the sensor spring 113 radially outside the lid-side annular support portion 1112. In the illustrated embodiment, the opposite support between the actuating plate spring 1104 and the sensor spring 1113 is also arranged radially outside the support diameter 1103a of the actuating plate spring 1102 on the compression spring 1101 external configurations in the form of radially outwardly facing arms 1113b, which in a similar manner, as described in connection with ñg. 36-37, over a bayonet lock 1514 axially abut the cover 1102 and are secured against rotation. For mounting the sensor spring 1113, the cover 1102 has corresponding axial recesses 1502b, corresponding axial recesses 1502b, in which the radially outer supports of the sensor spring 1113 can be axially inserted to form the bayonet lock 1514. The abutment of the disc spring 1104 against the lid-side support is supported the biasing force of the sensor spring 1113.

In connection with fi g. 43, the function of the coupling shall be explained in more detail. In this case, the force of the sensor spring is designed in such a way that it corresponds to the disengagement force at the post-set point. If it is disconnected after occurring coating wear (or wear elsewhere) and thus changed disc angle and thereby higher disc kraft force, the disc spring first swings around the support 1012 to in the vicinity of the post-adjustment point. Then at this point when the disengagement force becomes equal to the sensor force together with the coating fi edema - P32724SE0l / LK and through the residual force in 10 15 20 25 30 522 574 69 between the disengagement force and the sensor force. Thereby, the plate edema is lifted up from the lid-side support and releases this for the post-adjustment. During the further disengagement distance, the disengagement force decreases again, the sensor force overrides and presses through the pressure plate the disc spring against the lid-side support 1112, around which then the further pivoting of the disc spring takes place. With the transition of the disc spring from the lid-side support to the pressure plate-side support, the disc spring tends to change its function as a two-armed lever. It rests against the pressure plate transiently with the present decoupling force against the pressure plate and thereby is temporarily lifted from the lid-side support.

After further disengagement path, due to the associated power drop, the force from the sensor spring prevails and presses the disc spring again against the lid-side support, whereby the post-adjustment device is blocked and the post-adjustment process is completed. The plate tether is then again used as a two-armed lever for the further disconnection path.

The disc spring can be designed taking into account all the capercaillie forces, which act directly or indirectly against the disc spring. These include in particular the forces generated by the actuating disc spring and the axially displaceable components of the corresponding leveling or readjusting device.

The embodiment according to fi g. 44 further has the advantage that in the engaged state of the friction clutch, the disc spring 1104 is practically biased or acts as a two-armed lever and the disc spring 1104 is thus biased between the lid-side support 1112 and the pressure plate-side support 1103a. taken only against the sensor spring 1113 and pivots about the support portion 1113a, at the same time as the axial displacement of the support portion 1113a, so that it then practically acts as a single-lever lever.

The sensor cap 1113 according to ñg. 44 can - like the sensor disc springs in the other figures - with a corresponding design or adjustment support an arbitrary diameter of the actuating disc spring 1104. Thus, the support of the sensor spring 1113 against the disc spring 1104 can take place at a diameter located between the lid-side pivot portion 1105 and the pressure plate side support diameter 1103a. Furthermore, the support of the sensor spring 1113 against the disc spring 1104 could also be arranged radially opposite the lid-side support diameter 1105. Thus, the axial support force applied by the sensor spring 1113 tends to be greater the smaller its support diameter 1113a becomes on the disc spring portion 1104. Furthermore, the sensor spring force must be constant. 1113 become larger the further away from the support diameter P32724SE0l / LK 10 15 20 25 30 522 574 70 ø. now 1113a between the springs 1104 and 1113 from the lid side support diameter 1105 the disc spring 1104 is located.

The embodiment according to ñg. 45 has a post-adjustment device 1216 which operates in a manner similar to that described with the preceding figures, especially in connection with fi g. 17-30. The actuator plate spring 1204 is pivotally mounted between two annular unrolling supports 1211 and 1212. The support 1211 located adjacent the pressure plate 1203 is actuated by the sensor spring 1213. The friction clutch 1201 has a device 1261 which ensures that during the friction clutch arranged countermeasures. In the exemplary embodiment shown, the counter-ramps, similar to that described in connection with Fig. 18, are arranged on a support ring 1225 which is rotatable on the lid.

The device 1261 forms a loose friction mechanism which, when disengaging the friction clutch 1201 and in the event of wear on the friction linings 1207, can exert an axial force on the post-adjustment ring 1217, whereby the possible adhesion connection between the ramps and the counter-ramps is loosened. The mechanism 1261 includes an axially resilient member 1262 which, in the illustrated embodiment, is axially connected to the disc spring 1204. The member 1262 has an annular diaphragm or disc-like resilient base body 1262a which is radially externally connected to the disc-shaped inner ring 1204. From the the base body 1262a extends circumferentially distributed, axial fls 1263, which extend through axial recesses in plate dem edema 1204. At their free end portions, fls 1263 have abutment contours in the form of deflections 1264, which cooperate with counter-abutment contours 1265 on the post-adjustment abutment. formed by recesses radially inserted into the ring 1217 or by a circumferential groove. The distance between the abutment contours 1264 and the counter-abutment contours 1265 in the engaged state of the friction clutch is adjusted so that over at least a larger part of the clutch disengagement phase there is no contact between the contours 1264 and the counter-contours 1265. the element 1262 can be elastically biased between the post-adjustment ring 1217 and the disc tether 1204. This ensures that, as soon as due to coating wear an axial displacement of the swivel support 1211 occurs, the post-adjustment ring 1217 is forcibly lifted up. ; : ß -: fix - ... II - 71 from the lock-side ramp ramps. Furthermore, the mechanism 1261 shall prevent, in the event of an excessively large disconnection path, Lex. due to an incorrect basic setting of the disconnection system, a readjustment of the ring 1217 takes place. This is achieved by, in the event of too great an angle of rotation of the disc spring 1204 in the direction of disengagement, the resilient element 1262 prestresses the post-adjustment ring 1217 against the disc spring 1204, whereby a pivot lock of the post-adjustment ring 1217 takes place relative to the disc spring 1204. according to fi g. 24 in the direction of disengagement, the post-setting ring 1217 is kept rotationally fixed relative to the plate edge 1204, when the retention force of the sensor spring 1213 is exceeded when point 46 is exceeded, whereby even if there is no wear on the clutch disc, a post-adjustment would take place. This would result in a change in the operating point, thus a change in the mounting position of the disc spring 1204, namely in the direction of a lower pressing force.

This means that in ñg. 24 the operating point 41 would travel along the curve line 40 in the direction of the minimum specified in medu45.

In an embodiment of a friction clutch, which is designed in accordance with the details according to fi g. 46-48, the individual screw caps 1326 are received on tabs 1328, which are formed integrally with the coupling cover 1302. The tabs 1328 are formed from the sheet material of the cover 1302 by forming Lex. a punched U-shaped circumferential 1302a.

The flaps 1328 extend, viewed in the circumferential direction, arcuately or tangentially and are suitably at least approximately at the same axial height as the directly adjacent lid portion. Av fl g. 48 it can be seen that in the embodiment shown the 13s 1328 are displaced by approximately half the material thickness in relation to the bottom portions 1302b of the lid. .

The width of a fl ik 1328 is adapted so that the coil spring 1326 arranged thereon is guided both in radial and also in axial direction.

The post-adjustment ring 1317 actuated by the springs 1326 in the post-adjustment direction has radially inwardly facing configurations or extenders 1327 extending on its inner circumference which extend between the lid 1302 and the disc spring 1304. The extensions 1327 have a fork or U-shape 13 the shaft direction directed pimps 1327b on both sides grip the capercaillie guides 1328. For this purpose the two pins 1327b extend axially through the cut-out 1302a in the lid 1302. On the recesses 1327a or on their pins 1327b the post-adjustment springs 1326 support.

The readjustment ring 1317 similarly supports, by means of its ramp-up ramps, the counter-ramp-up ramps 1324 pressed into the lid 1302, as described in connection with the preceding figures. However, the counter-recesses 1324 forming the counter-ramp ramps are designed in such a way that they form an air passage opening 1324 in the direction of rotation of the coupling. In particular, the post-adjustment ring 1317 made of plastic material is also cooled in this way, whereby the thermal load of this component can also be significantly reduced. According to a further design variant, the sensor force, which acts on the actuating plate edge of the friction coupling, can be applied through, for example, blade spring elements arranged between the coupling cover and the pressure plate, these leaf spring elements being able to couple the pressure plate and the cover rotatably but axially limited. In such an embodiment, no special sensor springs would thus be required, but e.g. the blade spring elements 9 of the friction clutch 1 could be designed according to fi g. 17 and 18 in such a way that they also take over the function of the sensor pulley spring 13. i As a result, both the sensor spring 13 and also the unwinding ring 11 can be omitted.

The leaf spring elements 9 must then be designed in such a way that during an operation of the friction clutch 1 and without the coating wear occurring, the operating disc spring 4 becomes abutting against the lid-side unwinding support 12. However, as soon as a corresponding wear occurs on the friction pads 7, , the leaf spring elements 9 must enable a wear adjustment of the disc spring 4 corresponding to the wear. It thus means that the leaf spring elements 9, as described in connection with fi g. 25, shall have a curve line portion 48 according to the curve lines 47 or 47a.

P32724SE01 / LK

Claims (46)

10 15 20 25 30 522 574 73 Patent claims Friction coupling, in particular for motor vehicles, with a pressure plate which is rotatably but axially limitedly displaceably connected to a housing, a pressing plate spring being arranged between the housing and the pressure plate, which is pivotable about one of the housing bores, through on both sides of the pressing plate spring pivot bearing formed pivot bearing and partly - in mounted condition for the opp coupling on a counterpressure plate - the pressure plate acts in the direction of a between this and the counterpressure plate such as a flywheel, cleavable clutch plate, characterized by the two-sided pivot bearing (412) is axially displaceable The compensating, automatic post-adjustment device (416) and pressing plate edema (404) of the coupling discs of the clutch disc are under the influence of a support force i directed in the direction of the reciprocating pivot bearing (412), the size of which can be changed depending on the speed. 2. A friction clutch according to claim 1, characterized in that an at least wear of the friction lining of the clutch disc is automatically compensated for setting the actuating device, which provides a practically constant effect on the pressure plate by the pressure clamping means and the friction clutch which can be operated by means of a disconnector, wherein means (2a, 36; 436) are arranged for limiting the disengagement movement of the operating means (4b, 404b). Friction coupling according to one of Claims 1 to 2, characterized in that the talc spring (404) is pivotally supported at the housing and has an annular basic body (404a), from which radially inwardly directed actuating tongues (404b) emanate, the friction coupling having abutment means ( 436), which limits the angle of oscillation of the disengagement tongues (404b) when disengaging the sk coupling and is operative in the radial extent of the actuating tongues (404b), preferably in the radially inner region of the actuating tongues (404b) or at least adjacent area. Friction coupling according to one of Claims 1 to 3, characterized in that an adjusting device provides a practically constant force effect on the pressure plate through the compression spring, and the friction coupling has an actuating means (4b) for the uncoupling and coupling, which by means of a axially displaceable disconnector (22) can be actuated by a disengaging means (24), the actuating means (4b) being axially displaced in the direction of the disengaging movement due to the wear of at least the friction pads and the disc spring having a spring pattern and minimum \ 10- \ 103272402 \ 03l106 requirements final 10 10 20 20 25 30 o L. 52 574? 4 adjoining area with crater increase which, when disconnected from the clutch within this crane shaft in the setting device, contains a sensing spring for axially sissing and supporting material. direction towards the minimum releases the clutch disc and storage of speech lriks fi ädern at a house (such as e.g. described in DE patent application 42 39 291.8 or FR patent application 2,684,730), wherein in the path of movement of the clutch disengagement means, such as the actuating means (4b), the disengagement disengagement means or similar abutment means (2a + 36, 436) are arranged, which after passing a minimum prevents a following lcra increase over the value, which has resulted in a release of sensing edema. Friction coupling according to one of Claims 2 to 4, characterized in that the restricting or abutment means prevents an impermissibly high impact and / or deformation of the actuating means (4b). Friction coupling according to one of Claims 2 to 5, characterized in that the abutment or restraining means (2a, 436) is arranged at least at the housing. Friction coupling according to one of Claims 2 to 6, characterized in that the restraining means has at least one restraining bearing (436) carried by the housing. Friction coupling according to one of Claims 2 to 7, characterized in that the abutment or limiting means (2a + 36) is operative between at least one structural part (34) of the disconnector and the coupling housing. Friction coupling according to one of Claims 2 to 8, characterized in that the restriction of the disengagement movement takes place by abutment of the actuating means (4b) against a restriction stop (43 6) carried by the housing. Friction coupling according to one of the preceding claims, characterized in that the edging talc edema has an annular basic body (4a), from which radially inwardly directed tongues (4b) emanate, which form operating means. 11. ll. Friction coupling according to one of Claims 2 to 10, characterized in that the restricting means (436) limits the angular oscillation of the pressing plate ladder. Friction clutch according to one of Claims 2 to 1, characterized in that the restricting means (436) is integrated in the friction clutch. Friction coupling according to one of Claims 2 to 12, characterized in that the restriction of the disengagement movement takes place by abutment of the radially inner disc spring tongue tips against an abutment carried by the housing (436). Friction coupling according to one of Claims 2 to 13, characterized in that the new housing (436) of the housing buma limiting stop (436) is arranged axially. between the disc tray rings and the pressure plate. Friction coupling according to one of Claims 2 to 14, characterized in that the boom limit stop of the housing is formed by an annular stop portion (436) which is connected to the housing via radially extending life (437). Friction coupling according to claim 15, characterized in that the web (437), starting from the housing, extends radially inwards and axially inclined in the direction of the pressure plate (403) and thereby extends through slots (439) arranged between the disc spring tongues. Friction clutch according to one of the preceding claims, which is operable by means of a disengagement system (601) acting on the actuator, characterized in that means (662, 664, 666, 667) are provided for limiting it on the actuator (601). influencing disconnection crane. Friction coupling according to one of the preceding claims, characterized in that a coating suspension (10) is arranged between the friction linings of the coupling disc. 19. A friction clutch according to any one of the preceding claims, characterized in that it is a printed clutch. 20. A friction clutch according to any one of claims 1-19, characterized in that the friction clutch has actuating means (604) for disengaging and engaging which can be actuated by means of a disengaging system (601) acting thereon, wherein in the disengaging system means (662, 664, 666, 667) are provided for limiting the maneuvering force exertable on the operating means. 21. A friction clutch according to claim 17 or 20, characterized in that the limitation of the actuating force is effected by means of a releasable means arranged in the power transmission path of the disengagement system when exceeding a certain actuating force. Friction clutch according to claim 21, characterized in that the releasable means is resilient or elastic. Friction coupling according to one of Claims 17 to 22, characterized in that the disengagement system has a pressure medium circuit in which a maximum disengagement circuit (662, 664, 666, 667) is arranged. Friction clutch according to one of Claims 17 to 23, characterized in that the means for limiting the maximum actuating force is arranged in the force transmission path between or a clutch pedal, an actuating motor and the actuating means (4B) for the friction clutch, such as a disc plate. K: \ Patent \ l0- \ l03272402 \ 030827 new claims.doc 10 15 20 25 30 '- 522 574 76 Friction coupling according to one of Claims 17 to 24, characterized in that the disengagement system has a pressure medium circuit in which a maximum pressure limiting means (662, 664) is arranged. Friction coupling according to one of Claims 17 to 25, characterized in that the means arranged for limiting the actuating means (4B) acting on the friction coupling sets a maximum force which is at least slightly greater than the maximum required for the actuation of the coupling. the power. 27. A friction clutch according to any one of claims 20-26, characterized in that the collar limiting means is formed by a buffer, such as e.g. hydraulic or resilient bearings. Actuating device with a drive motor for a clutch according to any one of claims 1-27, in particular for a motor vehicle clutch, wherein for the connection and disconnection of the sk coupling between it and the drive motor an operating line is arranged, characterized in that the drive motor is provided with a safety device , by which, when a predeterminable actuating force fl or a predetermined torque of the drive motor is exceeded, the torque delivered by the drive motor or the actuating force is limited to a predeterminable amount. Friction coupling according to one of Claims 1 to 27, characterized in that the air setting device (416) is operative between the lid (402) and the plate spring (404). Friction coupling according to Claim 1 or 29, characterized in that the support collar loads the pivot bearing (41 1) arranged on the side facing away from the housing of the pressing plate (41 1). Friction coupling according to one of Claims 1 or 29 to 30, characterized in that the support force is centrifugally dependent. 32. Friction coupling according to one of Claims 1 to 27 or 29 to 31, characterized in that the pressure plate edema (404) against the disengagement collar is supported only by the collar attached to the housing. Friction coupling according to one of Claims 1 or 29 to 32, characterized in that the support collar is generated by at least one plate-like spring force-laying element (41), which, in the case of a wear-dependent post-adjustment of the pressure plate spring (404) or its lid, 34. Friction clutch according to one of Claims 1 or 29 to 33, characterized in that the support force increases with increasing speed. K: \ Patent \ 10- \ l03272402 \ 030827 new claims.doc 10 15 20 25 30 522 574 77 Friction coupling according to one of Claims 1 or 29 to 34, characterized in that central force-dependent means (450) are provided, which produce a support collar fi which acts in parallel with the base support collar av generated by a disc-spring-like element (413). According to claim 35, the centrifugal dependent means (450) enables a wear re-adjustment only in the case of a friction clutch characterized in that it is stationary or at low speeds of the friction clutch (401). Friction clutch according to one of Claims 1 or 29 to 36, characterized in that the setting device of the setting device (416) is practically blocked above a certain speed. Friction coupling according to one of Claims 1 or 29 to 37, characterized in that the centrifugal dependent means are formed by tongues (450) which are buoyed by a base support force or a generating plate-like element. Friction clutch according to claim 38, characterized in that the tongues (450) are integral with the disc-spring (413) element. A friction clutch according to any one of claims 1-27 or 29-39, characterized in that at least one compression spring, such as e.g. a disc spring, is operative between housing and pressure plate, furthermore an at least wear of the iction pads of the clutch disc is automatically compensated for a setting device which in the non-pressure plate mounted state for the iction clutch is in a constant clutch position and further After-actuation of the pressure plate by the pressing fi vein, and the sk clutch has actuating means (204) for disengaging and engaging, characterized in that the post-setting device (216) in the unmounted state of the friction clutch (201) is secured against an impermissible (unintentional) . Friction coupling according to Claim 40, characterized in that the securing of the setting device (216) is at least revocable after the mounting of the friction coupling (201) has been mounted on the receiving plate. Friction coupling according to one of Claims 40 or 41, characterized in that the adjusting device (216) has a rotatable ring (217) which is pivotally secured by means of securing means (261) at least in the unmounted state of the friction coupling. Friction coupling according to one of Claims 40 to 42, characterized in that the adjusting ring (717) is arranged axially between the pressing plate spring (704) and the housing, and further has radial outriggers (773), against which the pressing talc spring (704) K: \ Patent \ 10- \ l03272402 \ 030827 new requirements.doc 10 522 574 78 come into play when exceeding a certain conicity. A friction clutch according to any one of claims 40-43, characterized in that there is an impermissible displacement of the setting device in the case of an unmounted friction clutch preventing the coupling means (773) even when the friction clutch is mounted and exceeding a certain disengagement distance (70). ) blocks the e-setup device. Friction coupling according to Claim 44, characterized in that the locking device is blocked by the installation of the adjusting device (770) (717) on the pressure plate edema (704). batches against batches of an e-setting ring on A friction clutch according to claim 45, characterized in that the outer edge portion of the press plate edge (704) abuts radial outriggers (773) of the setting ring (717) when exceeding a predeterminable conicity on the press plate (70). K: \ Patent \ l0- \ 103272402 \ 030827 nya claimsxloc