DE8420404U1 - Damping device for absorbing or compensating torsional shocks - Google Patents

Description

LuK slats and

Clutch Construction GmbH

Industriestrasse 3

PO Box 13 60 20 June 1984

7580 Bühl (Baden) - 0482 A-D

PT/Schm-vb

Damping device for absorbing or compensating torsional shocks 10

The invention relates to a damping device for absorbing or compensating for rotational shocks, in particular torque fluctuations of an internal combustion engine, with at least two flywheel masses arranged coaxially to one another and mounted so as to be rotatable relative to one another against the effect of damping means, of which the first, one, is rotationally fixed to the internal combustion engine and the other, second, can be connected to the input part of a transmission.

\ In such damping devices, it is known to mount the two flywheel masses so that they can rotate relative to one another via a roller bearing. In many cases, the roller bearings are arranged in such a way that one of their bearing rings is connected to the other flywheel mass in a rotationally fixed manner. Although designs of this type can achieve very good damping of the vibrations occurring between the internal combustion engine and the drive train of a motor vehicle, these designs have not been able to prevail in motor vehicle construction due to the short service life of the bearing between the flywheel masses. This bearing

is the critical point of such a device, since the bearings fail after a short period of operation due to the unfavourable operating conditions. This is probably due to the fact that the bearing rings can only rotate relative to one another within the limited range of the possible angle of rotation between the flywheels. This is particularly disadvantageous when driving under load, where torsional vibrations of very high frequency and small amplitude occur between the bearing rings, since the rolling elements arranged between the bearing rings, such as balls, experience a change in direction of rotation proportional to the frequency of the torsional vibrations and only a very small rolling movement. Another disadvantage, which has a particularly negative effect, is that the high stresses between the rolling elements and the rolling tracks of the bearing rings practically always occur at the same points or in the same, very small areas of the circumference of the rolling tracks, which means that the material is overstressed at these points or in this small area. This excessive stress can lead to the rolling elements working their way into or pressing into the rolling tracks, which leads to rapid destruction of the bearing. Overstressing the material^ can also lead to particles breaking off from the surface of the rolling elements and/or the rolling tracks, and these breakouts also lead to destruction of the bearing.

The object of the present invention was to create a damping device which, compared to the previously known damping devices of the type mentioned at the beginning, has an improved function and an increased service life, and can also be manufactured in a particularly simple and economical manner.

According to the invention, this is achieved in a damping device of the type mentioned at the outset in that the second flywheel mass is supported relative to the first flywheel mass by means of a self-adjusting or self-adjusting roller bearing. 5

It can be assumed that the increase in the service life of the bearing between the two flywheels that can be achieved by this measure is due to the fact that self-adjustment occurs in the rolling bearing and practically no play can occur between the rolling elements and the rolling tracks. As a result, at least no undamped radial or axial relative movement can take place between the rolling elements and the rolling tracks, so that no "hammering" or collision can occur between the rolling elements and the rolling tracks, i.e. no sudden movements can occur between the rolling elements and the rolling tracks that cause the rolling elements to collide with the rolling tracks and thus also increase stress. The reason for the stimulation of such impact or shock movements between the rolling elements and the rolling tracks is probably that the crankshaft is slightly bent due to the individual ignition pulses of the internal combustion engine, so that the flywheel masses at the end of the crankshaft perform both a radial wobbling movement and a wobbling movement in relation to the theoretical axis of rotation. As a result, both alternating moments and radial forces are exerted on the bearing due to the inertia of the flywheel masses.

The self-adjustment can be achieved in a particularly advantageous manner by using a bearing of the type in which at least one of the bearing races is under the effect of a force accumulator that acts on the rolling elements in the axial direction.

The tensioning of the rolling tracks against the rolling elements or of the rolling elements between the rolling tracks achieved by this measure enables a more even load between the individual rolling elements and the rolling tracks, which means that the forces are better distributed over the circumference of the bearing or the rolling elements, so that the peak loads are reduced compared to a non-preloaded bearing used for the same application.

It can be advantageous if the self-adjusting roller bearing comprises a double-row angular contact ball bearing with at least one split bearing ring, whereby it can be appropriate for some applications if the clamping is made possible by the inner ring of this bearing being split and subject to an axial spring force. For other applications, however, it can also be expedient if the outer ring is split and subject to an axial spring force. For other applications, however, it can also be advantageous if both the outer and inner rings are split and subject to an axial spring force. To apply the axial spring force, it can be expedient to use a disc spring which is on the one hand attached to an axially fixed component

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the damping device and on the other hand axially loads a partial ring of the split inner and/or outer bearing ring" in the direction of the other partial ring of the corresponding bearing ring.

In designs where there is little space for the bearing, it may be appropriate for the bearing to be formed by a so-called wire bearing, in which at least one row of rolling elements can roll on four rolling supports formed by wire rings and at least one of the rolling supports or one of the wire rings is subject to an axial spring force. Depending on the application and the space available, it may be advantageous if either one of the outer rolling supports or one of the inner rolling supports or both are subject to an axial spring force, whereby the rolling elements are clamped between the rolling supports.

According to a further development of the invention, it can also be appropriate if the bearing is formed by a so-called four-point bearing, in which at least one of the bearing rings is split and is subject to the effect of an axial spring force, i.e. that the two partial rings of the split bearing ring are clamped towards each other or against the rolling elements. Even when using such a bearing, either the inner ring or the outer ring or both the inner and outer rings can be split and subject to the effect of an axial spring force.

According to another embodiment of the invention, the self-adjusting bearing can also be formed by a bearing in which the self-adjustment takes place in that at least one of the bearing rings

- exerts an axial clamping force in the direction of the rolling elements. In other words, this means that when the bearing is assembled, at least one of the bearing rings is elastically clamped in such a way that it presses or clamps the rolling elements in the direction of the raceways onto the other bearing ring. Such clamping is advantageous for double-row angular contact ball bearings with non-split rings, since the rings used there are relatively wide and can therefore have a larger elasticity reserve.

Although it may be advantageous for some applications if the load on the rolling support causing an adjustment is carried out by a force accumulator that is effective in the disengagement direction of the friction clutch connecting the second flywheel mass to the input part of a transmission, it may be useful for other applications if the load on the rolling support causing an adjustment is carried out by a force accumulator that is effective in the disengagement direction of the friction clutch connecting the second flywheel mass to the input part of a transmission.

\ with the friction clutch connecting the input part of a gearbox effective energy storage device. The force applied by the tensioned energy storage device can correspond to between one and three times the maximum actuation force.

In order to ensure perfect guidance and centering of the two flywheel masses relative to one another, it can be expedient according to a further feature of the invention if the bearing ring acted upon by the energy accumulator has a press fit or transition fit on its seat that can be overcome by the energy accumulator.

It can also be useful if the roller bearing is accommodated on a projection or shoulder connected to the first flywheel and extending axially away from the internal combustion engine. Such a shoulder can be integral with the first flywheel or can also be attached to it. It can also be advantageous if the force accumulator that acts on the rolling elements in the axial direction at least one bearing race is a disc spring that is accommodated on the projection or shoulder of the first flywheel. The disc spring can be supported radially on the outside of the first flywheel and can act radially on the inside of the inner split bearing ring of a ball bearing, such as in particular a double-row angular contact ball bearing. The outer bearing ring of a bearing braced in this way can advantageously be accommodated in a receiving bore of the second flywheel.

The invention is described in more detail with reference to Figures 1 to 6.

It shows:
25

Figure 1 shows a section through a damping device according to the invention, Figure 2 shows a view according to arrow II of Figure 1,

Figures 3 to 6 show further embodiments of bearings for a damping device according to the invention.

The device 1 for compensating torsional shocks shown in Figures 1 and 2 has a flywheel 2 which is divided into two flywheels 3 and 4. The flywheel 3 is attached to a crankshaft 5 of an internal combustion engine (not shown in detail) via fastening screws 6. A friction clutch 7 is attached to the flywheel 4 via means not shown in detail. A clutch disc 9 is provided between the pressure plate 8 of the friction clutch 7 and the flywheel 4 and is received on the input shaft 10 of a transmission (not shown in detail). The pressure plate 8 of the friction clutch 7 is actuated in the direction of the flywheel 4 by a plate spring 12 pivotably mounted on the clutch cover 11. By actuating the friction clutch 7, the flywheel 4 and thus also the flywheel 2 of the transmission input shaft 10 can be coupled and uncoupled. Between the flywheel 3 and the flywheel 4, damping means in the form of a damping device 13 and a slip clutch 14 connected in series with the latter are provided, which enable a limited relative rotation between the two flywheel masses 3 and 4.

The two flywheel masses 3 and 4 are mounted so that they can rotate relative to one another via a bearing 15. The bearing 15 comprises a roller bearing in the form of a double-row angular contact ball bearing 16 with a split inner ring 17. The outer bearing ring 18 of the roller bearing 16 is arranged in a bore 19 of the flywheel mass 4 and the inner split bearing ring 17 of the roller bearing 16 is arranged on a central cylindrical pin 20 of the flywheel mass 3 which extends axially away from the crankshaft 5.

The two partial rings 21, 22 of the inner ring 17 are axially braced by a force accumulator in the form of a disc spring 23. To secure the partial ring 22, a locking disk 24 is attached to the front side 20a of the pin 20 by means of screws 24a. The locking disk 24 extends radially outward beyond the pin 20, so that the partial ring 22 can be axially supported on it. The disc spring 23 arranged on the pin 20 axially between the partial ring 21 and the flywheel 3 acts on the partial ring 21 axially in the direction of the partial ring 22 by supporting itself with radially outer areas on the flywheel 3 and acting with radially inner areas against the partial ring 21. As a result of this structure, the rolling elements 25 of the bearing 16 are clamped between the rolling tracks 18a, 21a and 18b, 22a assigned to them. To ensure that the bearing 16 remains clamped even when the friction clutch 7 attached to the flywheel 4 is actuated, the plate spring 23 applies a force that is greater than the maximum force required to actuate the friction clutch 7.

In order to ensure a friction-free guidance and centering of the two flywheel masses 3 and 4 relative to one another, partial rings 21, 22 of the inner ring 17 are applied to the pin 20 with a press or transition fit. The fit between the partial ring 21, which is acted upon by the plate spring 23, and the pin 20 is designed in such a way that it can be overcome by the axial force applied by the plate spring 23, whereby the partial ring 21 has the possibility of axial displacement on the pin 20 despite the pairing.

The flywheel 3 has an axial, ring-shaped extension 27 radially on the outside, which forms a chamber 28 in which the damping device 13 and the slip clutch 14 are essentially accommodated. The input part of the slip clutch 14 is formed by two disks 29, 30 provided at an axial distance from one another, which are rotationally fixed to the flywheel 3. The ring-shaped disk 30 is fastened to the end face 27a of the extension 27 by means of rivets 31 and axially delimits the chamber 28 with inward-reaching areas 30a. The disk 29 accommodated in the chamber 28 has axial projections which are formed by tabs 29a integrally formed on the outer circumference. To prevent the disk 29 from rotating relative to the disk 30, the tabs 29a engage in recesses 32 in the disk 30. The recesses 32 and the tabs 29a are designed in such a way that the disk 29 can be displaced axially relative to the disk 30. Radial arms 33 of the flange 34 are clamped axially between the two disks 29 and 30, in that a force accumulator in the form of a disc spring 35 provided axially between the disk 29 and the flywheel 3 acts on the disk 29 in the direction of the disk 30. For this purpose, the disk spring 35 rests with radially outer areas on the pressure plate 3.

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and with radially inner areas on the disk 29. The disc spring 35 is supported radially with its radially outer periphery on a circular shoulder 36 of the flywheel 3. Between the arms 33 of the flange 34 and the two disks 29 and 30 there is a friction lining 37, which is made up of individual lining segments 37a glued to the arms 33. In the area between the arms 33 of the flange 34 there are recesses 38, 39 in the disks 29 and 30, which are axially aligned and accommodate energy storage devices 40. In the example shown, these energy storage devices are formed by coil springs 40. The energy storage devices 40 serve as end stops for the arms 33 of the flange 34 and thus limit the angle of rotation of the slip clutch 14.

The flange 34, which forms the output part of the slip clutch 14, simultaneously represents the input part for the damping device 13. The damping device 13 has two disks 41, 42 arranged on both sides of the flange 34, which are connected to one another in a rotationally fixed manner at an axial distance via spacer bolts 43. The spacer bolts 43 also serve to attach the two disks 41, 42 to the flywheel 4. Recesses 41a, 42a and 34a are made in the disks 41 and 42 as well as in the areas of the flange 34 lying radially inside the arms 33, in which energy accumulators in the form of coil springs 44 are accommodated. The energy accumulators 44 counteract a relative rotation between the flange 34 and the two disks 41, 42.

The damping device 13 also has a friction device 13a, which is effective over the entire possible angle of rotation between the two flywheel masses 3 and 4. The friction device 13a is arranged axially between the disk 41 and the flywheel mass 3 and has a force accumulator 45 formed by a disc spring, which is arranged between

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the disc 42 and a pressure ring 46 is held in place, whereby the friction ring 47 arranged between the pressure ring 46 and the flywheel 3 is clamped. The force exerted on the disc 41 by the plate spring 45 is absorbed by the bearing 16.

On its radially inner periphery, the flange 34 has inwardly open cutouts 48 through which the spacer bolts 43 protrude axially. These cutouts form radially inward-pointing teeth 49 which - viewed in the circumferential direction - engage between the spacer bolts 43 and interact with them as stops for limiting the angular deflection of the damping device 13.

The bearing 115 shown in Figure 3 differs from that shown in Figure 1 in that the disk spring 123 is arranged on the other side of the bearing 116, namely between the locking disk 24 and the inner partial ring 22 of the inner ring 17. The second partial ring 21 of the inner ring 17 is supported on a shoulder 3a of the flywheel 3. The disk spring 123 supported on the locking disk 24 acts on the partial ring 22 in the direction of the shoulder 3a of the flywheel 3, whereby the rolling elements 25, 26 are clamped between the rolling tracks of the inner and outer bearing rings 17, 18. In this embodiment, the direction of action with which the disk spring 123 acts on the inner ring 17 corresponds to the

The release direction of the friction clutch mounted on the flywheel 3 corresponds to Figure 1, so that the plate spring does not have to absorb the force required to actuate the friction clutch. When the friction clutch is released, the release force is transferred from the outer bearing ring 18, which axially supports the pressure plate 3, via the rolling elements 25 to the partial ring 21, which is supported on the shoulder 3a of the flywheel 3. The release force is therefore not transmitted via the plate spring 123.

In the embodiment according to Figure 4, the bearing between the two flywheel masses 3 and 4 is carried out by means of a two-row angular contact ball bearing 216 with a split outer ring 218. The split outer ring 218 is arranged in the bore 19 of the flywheel mass 14 and the inner bearing ring 217 on a pin 20 of the flywheel mass 3. The plate spring 223 is supported on a shoulder 4a of the flywheel mass 4 and acts on the partial ring 222 axially in the direction of the partial ring 221, whereby the bearing 216 is clamped. The partial ring 221 is axially secured against the flywheel mass 4 by being axially engaged behind by the disk 41 which is firmly connected to the flywheel mass 4 via the spacer bolts 43.

In the embodiment according to Figure 4, the disc spring 223, which is used to actuate the friction clutch according to Figure 1, which is mounted on the flywheel 4, must absorb the force. However, this can be avoided by arranging the disc spring 223 on the other side of the bearing 218, i.e. between the disk 41 and the partial ring 221.

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In the embodiment shown in Figure 5, the bearing 315 is formed by a so-called four-point bearing 316 with a split inner ring 317. The two partial rings 321 and 322 are axially clamped towards one another by a disc spring 323, whereby the rolling elements 325 are clamped between the contact points 318a, 318b of the outer ring 318 and the contact points 321a, 322a of the inner ring 17.

In order to ensure that the bearing 316 remains tensioned even when the friction clutch mounted on the flywheel 4 is actuated, the plate spring 323 must be dimensioned and preloaded in such a way that it exerts a force from the partial ring 321 that is greater than the maximum force required to actuate the friction clutch.

In order to avoid that the disc spring 323 has to counteract the release force of the clutch mounted on the flywheel 4, the disc spring 323 can be arranged axially on the other side of the bearing 316 in such a way that it acts on the partial ring 322 in the direction of the partial ring 321.

\ In the embodiment shown in Figure 6, the bearing 415 is formed by a so-called wire bearing 416. The bearing 416 has a series of balls 425 that can roll between four rolling supports 418a, 418b, 421, 422 that are formed by wire rings. The radially inner wire ring 422 is axially loaded by a disc spring 423 in the direction of the radially inner ring 421, whereby the balls 425 are clamped between the four wire rings 418a, 418b, 421, 422.

In order to enable a perfect adjustment of the bearing 416, at least the wire ring 422 is cut at one point on its circumference.

Instead of applying pressure to a radially inner wire ring by means of a disc spring 423, a radially outer wire ring 418a or 418b could also be applied by means of a
The corresponding force accumulator is subjected to axial load.

Claims (21)

Patent claims 1. Damping device for absorbing or compensating for rotational shocks, in particular torque fluctuations of an internal combustion engine with at least two flywheel masses arranged coaxially to one another and mounted so as to be rotatable relative to one another against the effect of damping means, of which the first, first, is rotationally fixed to the internal combustion engine and the other, second, can be connected to the input part of a transmission, characterized in that the second flywheel mass (4) is mounted relative to the first (3) via a self-adjusting or self-adjusting rolling bearing (15, 115, 215, 315, 415). \ 2. Damping device according to claim 1, characterized in that the bearing (15, 115, 215, 315, 415) is formed by such a bearing (16, 116, 216, 316, 416) in which at least one of the bearing raceways (21, 22, 222, 321, 422) is under the effect of a force accumulator (23, 123, 223, 323, 423) acting on it in the axial direction on the rolling elements (25, 26, 325, 425). &Ggr;: 3. Damping device according to claim 2, characterized in that the bearing (16, 116, 216) is a double-row angular contact ball bearing with at least one split bearing ring (17, 218). 4. Damping device according to claim 3, characterized in that the inner ring (17) is divided and is subject to the action of an axial spring force. 5. Damping device according to claim 3 or 4, characterized in that the outer ring (218) is divided and is subject to the action of an axial spring force. 6. Damping device according to claim 1 or 2, characterized in that the bearing (415) is formed by a so-called wire bearing (416), in which at least one row of rolling elements (425) can roll on four rolling supports (418a, 418b, 421, 422) and at least one of the rolling supports (422) is subject to the effect of an axial spring force (through 423). 7. Damping device according to claim 6, characterized in that at least one of the outer rolling supports (418a, 418b) is under the action of an axial spring force. 8. Damping device according to claim 6 or 7, characterized in that at least one of the inner rolling supports (421, 422) is under the action of an axial spring force. 9. Damping device according to at least one of the preceding
Claims, characterized in that the bearing (315) is
a so-called four-point bearing (316) is formed, in which at least one of the bearing rings (317) is divided and under the action
an axial spring force (through 323). 10. Damping device according to claim 9, characterized in that the inner ring (317) is divided and is subject to the action of an axial spring force. 11. Damping device according to claim 9 or 10, characterized in that the outer ring (318) is divided and under the action
an axial spring force. 12. Damping device according to claim 1 or 2, characterized in that the self-adjusting bearing is formed by a bearing in which the self-adjustment takes place in that at least one of the bearing rings exerts an axial tensioning force in the direction of the rolling elements. ^X 13. Damping device according to at least one of claims 1 to 11, characterized in that the loading of the rolling support (22, 222, 422) causing an adjustment is carried out by a
The force accumulator (123, 223, 423) effective in the disengagement direction of the friction clutch (7) connecting the second flywheel (4) with the input part of a transmission takes place. 14. Damping device according to one of the preceding claims 1 to 11, characterized in that the rolling support (21, 321) effecting an adjustment is acted upon by a force accumulator (23, 323) acting against the disengagement direction of the friction clutch (7) connecting the second flywheel mass (4) to the input part of a transmission. 15. Damping device according to claim 13 or 14, characterized in that the force of the energy accumulator (23, 123, 223, 323, 423) is between one and three times the maximum actuation force. 16. Damping device according to at least one of the preceding claims, characterized in that the bearing ring (21, 22, 222, 321) has a press fit on its seat that can be overcome by the force accumulator to transition fit. 17. Damping device according to one of claims 1 to 16, characterized in that the roller bearing is accommodated on a projection (20) or shoulder connected to the first flywheel mass (3) and extending axially away from the internal combustion engine. 18. Damping device according to claim 17, characterized in that the projection or extension (20) is integral with the first flywheel mass (3). 19. Damping device according to at least one of the preceding claims, characterized in that the force accumulator acting on at least one bearing raceway (21, 22, 321, 422) in the axial direction on the rolling elements (25, 26, 325, 425) is a disc spring (23, 123, 323, 423) which is received on the projection or shoulder (20) of the first flywheel mass (3). 20. Damping device according to claim 19, characterized in that the disc spring (23, 123) is supported radially on the outside on the first flywheel mass (3) and radially on the inside acts on the inner, split bearing ring (17) of a double-row angular contact ball bearing (16, 116). 21. Damping device according to at least one of the preceding Claims, characterized in that the second flywheel mass (4) has a receiving bore (19) for the outer bearing ring (18, 218, 318).