WO2008006338A1

WO2008006338A1 - Torsional vibration dampers, in particular a dual-mass flywheel between the combustion engine and the gearbox of a motor vehicle

Info

Publication number: WO2008006338A1
Application number: PCT/DE2007/001162
Authority: WO
Inventors: Ludwig Winkelmann; Steffen Dittmer; Rainer Eidloth
Original assignee: Schaeffler Kg
Priority date: 2006-07-10
Filing date: 2007-06-29
Publication date: 2008-01-17
Also published as: DE102006031775A1

Abstract

The invention relates to a dual-mass flywheel (1) that essentially comprises a primary mass (4) directly connected to a drive shaft (2) of the combustion engine and a secondary mass (8) arranged coaxially to this primary mass (4) and connected to the input shaft (7) of the gearbox via a coupling (5). The primary mass (4) and the secondary mass (8) are linked to each other via several dampers (9) arranged therebetween and can be turned in relation to one another in a direction opposite to the effective direction of these dampers (9) via a roller bearing (12) that is arranged between a coaxial bearing flange (11) on the secondary mass (8). According to the invention, the roller bearing (12) is formed by a dual-row spherical roller bearing unit (13) that serves to absorb both radial and axial operating forces. Said spherical roller bearing unit (13) has rolling elements (14) in the form of spherical rollers, each with two side faces (15, 16) in the form of symmetrically flattened balls. The bearing flange (10) on the primary mass (4) has an inner bearing ring (21) with integrated inner tracks (17, 18) for both rows of rolling elements (19, 20). The gap between the bearing flange (10) and a separate outer bearing ring (22) is axially sealed on both sides and forms a grease depot (23) enlarged by the volume, that is reduced by the basic ball shape, of all the rolling elements (14).

Description

Schaeffler KG Industriestr. 1 - 3, 91074 Herzogenaurach

Name of the invention

Torsional vibration damper, in particular dual-mass flywheel between the internal combustion engine and the transmission of a motor vehicle

description

Field of the invention

The invention relates to a torsional vibration damper according to the preamble forming features of claim 1, and it is particularly advantageous on a dual-mass flywheel between the internal combustion engine and the transmission of a motor vehicle feasible.

Background of the invention

It is well known to those skilled in automotive engineering that a dual mass flywheel is a torsional vibration damper, which is connected in particular in passenger vehicles to increase the driving comfort between the internal combustion engine and the transmission. Such a dual mass flywheel consists essentially of a disc-shaped primary mass connected directly to the drive shaft of the internal combustion engine and of a disc-shaped secondary mass arranged coaxially with this primary mass, which is connected via a coupling to the input shaft of the transmission. Both momentum masses are by several between them arranged damping means in the form of usually two opposing, curved helical compression springs coupled together and against the action of these damping means via a coaxial between a bearing flange on the primary mass and a coaxial bearing flange on the secondary mass arranged friction-reducing storage relative to each other rotatable. During operation of the motor vehicle, the primary mass is actively driven via the rotating crankshaft of the internal combustion engine, while the secondary mass, which in turn drives the transmission input shaft, is taken along via the damping elements. On the one hand, the unevenness resulting from imbalances of the moving masses in the drive train and, on the other hand, the rotational irregularities of the internal combustion engine resulting from the piston movements are damped via the damping elements. The occurring during the drive maximum oscillating pivot angle between the flywheels is only about ± 0.4 °, so that the storage during the usual operation of the motor vehicle is hardly dynamically stressed in terms of relative speed.

However, the situation is different when starting the vehicle, where an oscillating swing angle of ± 60 ° and thus a much stronger rotation between the primary mass and the secondary mass can occur. In this case, a bearing system with a very low coefficient of friction of advantage, which can be achieved basically only with rolling bearings between the flywheels advantage. As a result, single-row deep groove ball bearings have been used in the past for the friction-reducing storage between the primary mass and the secondary mass, which, as is known, for example, from DE 41 17 584 A1 and DE 42 14 655 A1, with one of its bearing rings rotatably with a the bearing flanges are connected to the primary or secondary mass and are.

Although with such deep groove ball bearings a very good damping of the vibrations occurring between the internal combustion engine and the drive train of the motor vehicle can be achieved, however, have proven in practice decisive disadvantages. These consist on the one hand that it is These deep groove ball bearings are mostly special bearings that have proven to be very expensive to manufacture due to the design of the bearings as solid bearings and the use of special plastic caps to provide increased fat deposits. On the other hand, their too short service life has repeatedly proven to be a critical point on dual-mass flywheels, as it often came as a result of unfavorable operating conditions after a short period of operation storage failure. The cause of such failures is mainly due to the limited in operation of the motor vehicle slewing angle with high frequency and low amplitude between the flywheel masses through which the bearing balls experience a frequency of the torsional vibrations proportional rotation change and only a very small rolling movement. In addition, these high stresses between the bearing balls and their tracks are practically always at the same place or in very small areas of the circumference of the raceways superimposed by rolling or pivoting movements of the flywheel masses, so that the bearing rings are often overstressed in these small areas and the bearing balls work into the raceways. This leads to the detachment of particles from the surface of the raceways and ultimately to the destruction of the rolling bearing.

To avoid these disadvantages, it has therefore been proposed by DE 3 506 818 A1 to form the roller bearing between the primary mass and the secondary mass by a self-adjusting double-row angular contact ball bearing with a split inner bearing ring, which is under the action of an axial spring force. The achievable by this measure increase the life of the rolling bearing between the two masses is due to the fact that no play between the bearing balls and their careers and thus no undamped relative movement between the bearing balls can occur more by the self-adjustment in the rolling bearing, the impact of the Bearing balls on the raceways and thus causes increased stress on the bearing with the consequences mentioned above.

From the point of view that modern internal combustion engines with more and more Performance and torque are designed to cause increasingly higher clutch and on the rolling bearing between the primary mass and the secondary mass of dual mass flywheel release forces, it has been found that the high stresses during operation of the engine no longer with such self-adjusting double row angular contact ball bearings permanently can be handled and the life of the angular contact ball bearings depends mainly on their safe lubricant supply. Since, for design reasons, no relubrication possibilities are provided for rolling bearing in dual-mass flywheels, the service life of such angular contact ball bearings is strongly dependent, above all, on the size or possible service life of the existing fat depot within the bearing. The size of the fat deposit is defined by the free volume of the space between the bearing rings, which, however, is significantly reduced in all types of ball bearings by the nature of the bearing balls and the bearing cage considerably. The available grease reservoir for lubrication of the known angular contact ball bearings is therefore relatively small and can not be extended by the already almost complete utilization of the available space for rolling bearings without complex design changes in the storage environment and / or the bearing dimensions.

Object of the invention

Based on the stated disadvantages of the solutions of the known prior art, the invention is therefore the object of a torsional vibration damper, in particular a two-mass flywheel between the engine and the transmission of a motor vehicle to conceive, the rolling bearing between the primary mass and the secondary mass inexpensive to produce and is adapted to safely absorb the axial and radial forces acting and which has such a large grease reservoir without structural changes in the bearing environment and / or the bearing dimensions that it is characterized by lower wear and increased life compared to previously used bearings. Description of the invention

According to the invention this object is achieved in a torsional vibration damper according to the preamble of claim 1 such that the rolling bearing between the primary mass and the secondary mass is formed by both the recording radial operating forces and the recording axial operating forces serving double-row ball roller bearing unit whose rolling elements as ball rollers with each two symmetrically flattened by a spherical base shape and parallel to each other side surfaces are formed. The bearing flange on the primary mass is at the same time formed as molded inner raceways for both Wälzkörperreihen inner bearing ring of the ball roller bearing unit, while the both sides axially sealed gap between the bearing flange and a separate outer bearing ring forms the grease depot of the ball roller bearing unit, which reduced by the respectively of the ball basic shape Volume of all rolling elements is formed enlarged.

Preferred embodiments and advantageous developments of the inventively embodied torsional vibration damper are described in the subclaims.

Thereafter, it is provided according to claim 2 in the torsional vibration damper according to the invention that trained as ball rollers and preferably in separate Kunststofffensterkäfigen rolling elements both Wälzkörperreihen the ball roller bearing unit preferably each have a width between their side surfaces of about 60% to 70% of the diameter of their ball basic shape and by their shape increase the volume of the fat storage between the bearing rings by about 20% to 30% compared to an identically sized ball bearing. The width of the raceways in the bearing rings is also at least 60% to 70% of the diameter of the spherical base of the ball rollers, so that their treads under the radial / axial load to have a hundred percent line contact with the raceways in the bearing rings. The training Training the ball rollers and the raceways with such a width has been proven in practice with respect to the radial and axial load bearing capacity and corresponds approximately to the contact surface, which also have the balls of conventional angular contact ball bearings to their careers in the bearing rings have. Moreover, it has proven to be advantageous to form the running surfaces of the ball rollers with a logarithmic profile, so as to avoid edge crushing of the ball rollers at the edges of their careers while increasing the carrying capacity of the ball roller bearing unit on.

As a preferred first embodiment of an inventively designed torsional vibration damper, it is further proposed by claim 3, that the ball roller bearing unit is preferably designed as a double-row biased angular contact roller bearing with in X-arrangement against each other employed rows of rolling elements. The outer bearing ring of this angular contact ball roller bearing is formed in a cost effective manner as a deep-drawn part of a rolling bearing material, in which at least one of the outer raceways for the two rows of rolling elements is formed. In particular, 100CR and C80 or also C45 modified as deep-drawable rolling bearing materials have proved to be particularly advantageous, so that the outer bearing ring can still be suitably heat-treated and reworked after being deep-drawn for formation with rolling bearing properties.

Another feature of the thus produced outer bearing ring of the angular ball bearing, it is also according to claim 4, that this has at its primary mass facing axial side in a profile cross-section U-shaped and raised relative to the outer surface of the bearing ring support board, on the axial operating forces from the bearing flange of the secondary mass can be transferred to the angular contact ball bearing. This support board on the outer bearing ring is used according to claim 5 advantageously at the same time for sealing the fat deposits between the bearing rings on an axial side of the bearing by a pressed on the bearing flange of the primary mass double lip seal is pressed into the inner profile of the support board. At the other axial side of the bearing, the sealing of the fat deposits is preferred by a pressed onto the outer bearing ring L-shaped annular cap, are vulcanized to the two also rubbing on the bearing flange of the primary mass sealing lips.

Finally, according to claim 6, a final feature of the first embodiment of the ball-roller bearing unit is that the outer raceway for the second row of rolling elements is preferably designed as a separate annular component, which is fastened therein by being pressed into the outer bearing ring formed with the outer race of the first row of rolling elements. This separate outer raceway is formed in a cost effective manner as a flow press part of the same rolling bearing materials as the outer bearing ring, can be dispensed with after the post-processing due to the resulting from the small rotational movements between the masses low accuracy requirements for the ball roller bearing unit after the heat treatment. To adjust the bias of the angular ball bearing are also inserted between the back of the separate outer race and the support ring of the pressed in the support board double lip sealing ring tuning disks or spring plates formed as bias springs whose bias is fixed during the final beading of the support board.

A second embodiment of the inventively designed torsional vibration damper, however, according to claim 7, on the other hand, that the ball roller bearing unit is preferably designed as a combined radial-axial ball roller bearing with an inclined rolling row and a row of rolling elements. The outer bearing ring of this radial-axial ball roller bearing is also formed as a deep-drawn part of one of the rolling bearing materials, in which, however, in this embodiment, the outer races are formed for both Wälzkörperreihen. The transmission of axial operating forces from the bearing flange of the secondary mass to the radial axial ball roller bearing takes place here according to claim 8 similar to the first embodiment by a profile in cross-section L-shaped and raised against the outer surface of the outer bearing ring support board, pointing to the primary mass facing axial side the outer bearing ring is formed. The axial operating forces are mainly absorbed by the inclined rolling element row and their raceways in the bearing flange on the primary mass and in the outer bearing ring, while the vertically employed Wälzkörperreihe mainly serves the transmission of radial operating forces and for running stabilization of the secondary mass. As sealing of the fat deposits between the bearing rings, in this embodiment, on the one hand, a reinforced double lip seal pressed into the inner profile of the support board on the outer bearing ring and, on the other hand, an L-profile-shaped sealing cap pressed onto the outer bearing ring with a vulcanized sealing lip have proved to be particularly suitable Instead of the above types of seals, other suitable seals are possible.

A third embodiment of a torsional vibration damper designed according to the invention is furthermore characterized in that the ball roller bearing unit is preferably designed as a unidirectional tandem angular ball roller bearing in which a row of rolling elements forms a stepwise arrangement of the inner raceways of the rolling element rows in the bearing flange on the primary mass having larger pitch circle diameter and a larger number of rolling elements than the other row of rolling elements.

In a first variant proposed by claim 11, it is provided that the outer bearing ring of the tandem angular contact ball bearing is preferably formed as extruded part of one of said rolling bearing materials, which has a cylindrical outer circumferential surface substantially and formed in the inside of the outer raceways step-shaped for both Wälzkörperreihen are. An outer bearing ring produced in this way has the advantage, above all with regard to its production costs, that after its heat treatment by means of partial hardening it is possible to dispense with further finishing, since even here, due to the small rotational movements between the flywheels, only low accuracy requirements can be imposed on the tandem. Angular contact ball bearings are provided. According to claim 12 takes place in this tandem angular ball bearing, the transmission of axial operating forces of the secondary mass through a to the bearing flange of the secondary mass molded annular collar which bears against the opposite axial side of the outer bearing ring. To seal the fat deposits between the bearing rings a double lip seal is pressed into the outer bearing ring in this variant on the one hand and on the other hand pressed an L-shaped sealing cap with a vulcanized sealing lip on the outer bearing ring, but here are other suitable seals are conceivable.

In the second variant proposed by claim 14, on the other hand, it is provided that the outer bearing ring of the tandem angular contact ball bearing is preferably made as a deep-drawn part of a steel sheet consisting of said rolling bearing materials, in which the outer raceways are also formed stepwise for both Wälzkörperreihen, the but by the material used has a correspondingly stepped outer shell surface. For the transmission of axial operating forces from the secondary mass to the tandem angular contact ball bearing according to claim 15, the bearing flange of the secondary mass therefore formed with a matched to the step-shaped design of the outer bearing ring cross-sectional profile, which bears against the outer contours of the outer bearing ring. The sealing of the fat deposits between the bearing rings in this variant according to claim 16 on the one hand preferably by a gap seal, which is formed between a pressed into the outer bearing ring U-profile ring and the bearing flange on the primary mass and on the other hand preferably by a on the outer bearing ring Pressed L-profile-shaped sealing cap, on the free leg of a sliding on the bearing flange of the primary mass sealing lip is vulcanized.

A common feature of the two aforementioned variants of the tandem angular contact ball roller bearing is, according to claim 17, furthermore, that between the two rows of rolling elements an angle disc pressed into the outer bearing ring is arranged, which divides the fat depot between the bearing rings into two chambers. The arrangement of such also inexpensive producible by deep drawing angle disc has the advantage that thus a centrifugally induced migration of grease from the radially lower chamber arranged with the partially smaller row of rolling elements in the radially higher arranged chamber prevents the part of a circle larger row of rolling elements and thus also a sufficient lubrication of the part smaller rolling element row of such a tandem angular contact roller bearing is ensured over its maximum service life.

By claim 18, finally, a fourth embodiment of an inventively designed torsional vibration damper is proposed, in which the ball roller bearing unit is also designed as a single-acting tandem angular contact ball bearings, in which, however, by level alignment of the inner races of Wälzkörperreihen in Lagerflansch on the primary mass both Wälzkörperreihen the same pitch circle diameter and have the same number of rolling elements. A special feature of this embodiment, it is according to claim 19, that the outer bearing ring of the tandem angular contact ball roller bearing preferably consists of two separate as Tiefzieh- parts of one of said rolling bearing materials formed part rings, each forming the outer raceways for both rows of rolling elements. Both partial rings of the outer bearing ring in this case have substantially radially level-parallel cylindrical outer circumferential surfaces, so that in this embodiment according to claim 20, the transmission of axial operating forces from the secondary mass on the tandem angular contact ball bearings can be done again by a molded onto the bearing flange of the secondary mass annular collar, the the opposite axial side of a partial ring of the outer bearing ring rests. As a seal of the fat deposits between the bearing rings, in this embodiment, on the one hand, a labyrinth seal, the see between a pressed onto a partial ring of the outer bearing ring U-profile ring and a circumferential groove in the bearing flange is formed on the primary mass, and on the other hand, one on the other part ring of the outer bearing ring pressed L-profile-shaped sealing cap with two vulcanized sealing lips proved to be particularly advantageous, which, however, can be replaced by other suitable seals here.

In summary, the inventively designed torsional vibration damper thus in all described embodiments opposite the known from the prior art torsional vibration dampers on the advantage that its storage between the primary mass and the secondary mass through the use of a largely spanens producible double row ball roller bearing unit both inexpensive to produce and is suitable to safely absorb the axial and radial forces acting. In this case, the ball roller bearing unit used without constructive changes in the bearing environment and / or the bearing dimensions formed by ball rolling as a rolling compared to previously used ball bearings enlarged grease depot, so that the rolling bearing is characterized by a lower wear and increased overall life.

Brief description of the drawings

Preferred embodiments of the inventively embodied torsional vibration damper will be explained in more detail with reference to the accompanying drawings. Showing:

Figure 1 is a partial view of a cross section through a dual mass flywheel inventively designed roller bearing between the primary mass and the secondary mass;

Figure 2 is an enlarged view of the cross section through a first embodiment of the double-row ball roller bearing unit used for rolling bearing according to detail X in Figure 1;

Figure 3 is an enlarged view of the cross section through a second embodiment of the double row ball roller bearing unit used for rolling bearing;

Figure 4 is an enlarged view of the cross section through a first variant of a third embodiment of the double row ball roller bearing unit used for rolling bearing; Figure 5 is an enlarged view of the cross section through a second variant of the third embodiment of the double-row ball roller bearing unit used for rolling bearing;

Figure 6 is an enlarged view of the cross section through a fourth embodiment of the double row ball roller bearing unit used for rolling bearing.

Detailed description of the drawings

From Figure 1 is clearly a known per se dual mass flywheel 1, which consists essentially of a connecting screws 3 directly connected to the drive shaft 2 of the internal combustion engine disc-shaped primary mass 4 and a coaxial with this primary mass 4 arranged disc-shaped secondary mass 8, in turn is connected via the friction linings 6 of a clutch 5 to the input shaft 7 of the transmission of the motor vehicle. Furthermore, it can be seen from FIG. 1 that the primary mass 4 and the secondary mass 8 are coupled together by a plurality of damping elements 9 arranged between them and designed as semicircularly curved coil springs and counteracted by the action of these damping means 9 via a coaxial bearing flange 10 on the primary mass 4 and a coaxial bearing flange 11 arranged on the secondary mass 8 roller bearing 12 are rotatable relative to each other. This arrangement causes the driving force of the internal combustion engine initially acts on the primary mass 4 via the rotating drive shaft 2 during operation of the motor vehicle and is then transmitted via the damping means 9 to the input shaft 7 of the transmission driving secondary mass 8, via the damping means 9 the resulting from imbalances in the drive train resulting inequalities are attenuated.

As a new feature of this dual-mass flywheel 1, it is also clear from Figure 1 that the rolling bearing 12 between the primary mass 4 and the secondary mass 8 according to the invention by a double-row ball roller lenlagereinheit 13 is formed, which serves both the recording of radial operating forces and the inclusion of axial operating forces. The enlarged representation of the detail X in FIG. 1 shown in FIG. 2 makes it clear, as do the illustrations of FIGS. 3 to 6, that the rolling elements 14 of this ball roller bearing unit 13 are formed as two rows 19, 20 of ball rollers arranged side by side, each two have symmetrically from a basic spherical shape flattened and arranged parallel to each other side surfaces 15, 16. The bearing flange 10 on the primary mass 4 is at the same time formed as an inner bearing ring 21 of the ball roller bearing unit 13, in which the inner races 17, 18 are formed for both Wälzkörperreihen 19, 20. Another novel feature of the inventively embodied torsional vibration damper is that the axially sealed intermediate space between the bearing flange 10 and a separate outer bearing ring 22 forms the grease depot 23 of the ball roller bearing unit 13, which has the volume of all the rolling elements reduced in each case by the ball basic shape 14 is formed enlarged. Only hinted in this regard Figures 2 to 6 can be removed that trained as ball rollers and guided in separate Kunststofffensterkäfigen 24, 25 rolling elements 14 both Wälzkörperreihen 19, 20 of the ball roller bearing unit 13 preferably each have a width between their side surfaces 15, 16 of about 60% to 70 % of the diameter of their basic ball shape and enlarge by their shape, the volume of the fat deposits 23 between the bearing rings 21, 22 by about 20% to 30% compared to an identically sized ball bearing.

The illustrated in Figure 2 first embodiment of the used double-row ball roller bearing unit 13 is characterized in particular by the fact that it is designed as a double-row biased angular contact roller bearing 26 with in X arrangement against each other employed Wälzkörperreihen 19, 20, the outer bearing ring 22 as a deep-drawn part of a rolling bearing material a molded outer raceway 27 is formed for the outer row of rolling elements 19. The outer race 28 for the inner row of rolling elements 20, however, is formed as a separate annular component, which is inserted by means of sliding fit into the outer bearing ring 22. Clearly visible, this outer bearing ring 22 shows Moreover, on its axial side facing the primary mass 4, a supporting board 29, which is U-shaped in profile cross-section and opposite the outer lateral surface of the bearing ring 22, can be transferred via the axial operating forces from the bearing flange 11 of the secondary mass 8 to the angular ball roller bearing 26. The sealing of the grease reservoir 23 between the bearing rings 21, 22 takes place in this embodiment, on the one hand by a pressed into the inner profile of the support board 29 on the outer bearing ring 22 double lip seal 30 and the other by a pressed onto the outer bearing ring 22 L-shaped annular cap 31 with two vulcanized sealing lips 32. To adjust the bias of the angular ball bearing 26 are also between the back of the separate outer raceway 28 and the double lip seal 30 trained as Tellerfedem, in the drawing only suggestively recognizable biasing springs 33 are inserted, the bias is fixed during the final flanging of the support board 29.

In the illustrated in Figure 3 second embodiment, however, is provided that the double-row ball roller bearing unit 13 used is formed by a combined radial-axial ball roller bearing 34 with an inclined rolling row 20 and a vertically salaried Wälzkörperreihe 19, the outer bearing ring 22 as a deep-drawn part of a Rolling bearing material with molded outer raceways 27, 28 for both rows of rolling elements 19, 20 is formed. Also in this embodiment, the outer bearing ring 22 of the radial-axial ball roller bearing 34 clearly visible on its side facing the primary mass 4 axial side here in the profile cross-section, however, L-shaped and relative to the outer circumferential surface of the bearing ring 22 raised support board 35 over which the Operating forces from the bearing flange 11 of the secondary mass 8 to the radial-axial ball roller bearing 34 are transferable. The axial operating forces are thereby mainly absorbed by the inclined rolling element row 20 and its raceways 18, 28 in the bearing flange 10 on the primary mass 4 and in the outer bearing ring 22, while the vertically employed rolling element row 19 is primarily used for transmitting radial operating forces and for running stabilization the secondary mass 8 is used. For sealing the fat depot 23 between the bearing rings 21, 22, this embodiment has on one side in the inner profile of the support board 35 on the outer bearing ring 22 pressed reinforced double lip seal 36, while the other side is sealed by a pressed onto the outer bearing ring 22 L-shaped sealing cap 37 with a vulcanized sealing lip 38.

Figures 4 and 5 further show two variants of a third embodiment of the double-row ball roller bearing unit 13 used, which differs from the aforementioned embodiments in that it is designed as a unidirectional tandem angular ball bearing 39, in which by stepwise arrangement of the inner raceways 17, 18th the Wälzkörperreihen 19, 20 in the bearing flange 10 on the primary mass 4 which has a Wälzkörperreihe 19 a larger pitch circle diameter and a larger number of Wälzkörperanzahl than the other row of rolling elements 20.

In the variant shown in Figure 4, the outer bearing ring 22 of the tandem angular contact roller bearing 39 is formed as a flow press part of a rolling bearing material, in the inner circumferential surface of the outer raceways 27, 28 are formed stepwise for both Wälzkörperreihen 19, 20 and has a substantially cylindrical outer circumferential surface , The transmission of axial operating forces from the secondary mass 8 to the tandem angular ball bearing 39 therefore takes place via an integrally formed on the bearing flange 11 of the secondary mass 8 annular collar 40, which, as can be clearly seen in the drawing, rests against the opposite axial side of the outer bearing ring 22. The sealing of the grease reservoir 23 between the bearing rings 21, 22 takes place here on the one hand by a pressed into the outer bearing ring 22 double lip seal 41, while on the other hand pressed onto the outer bearing ring 22 L-profile-shaped sealing cap 42 is used with a vulcanized sealing lip 43.

In contrast, the outer bearing ring 22 of the tandem angular contact roller bearing 39 in the variant shown in Figure 5, however, formed as a deep-drawn part of a rolling bearing material, which also stepped molded outer raceways 27, 28 for both Wälzkörperreihen 19, 20 but a corresponding Having stepped outer outer surface. For transmitting axial operating forces from the secondary mass 8 to the tandem angular ball roller bearing 39, the bearing flange 11 of the secondary mass 8 therefore has a cross-sectional profile adapted to the step-shaped design of the outer bearing ring 22, which bears against the outer contours of the outer bearing ring 22. The sealing of the grease reservoir 23 between the bearing rings 21, 22 takes place in this variant on the one hand by a gap seal 45, which is formed between a pressed into the outer bearing ring 22 U-profile ring 44 and the bearing flange 10 on the primary mass 4, and on the other hand by a on the outer bearing ring 22 pressed L-profile-shaped sealing cap 46, at the free leg of a grinding on the bearing flange 10 of the primary mass 4 sealing lip 47 is vulcanized.

A common feature of the two aforementioned variants of the tandem angular contact roller bearing 39, it is also that between the rows of rolling elements 19, 20 an additional angle plate 50 is pressed into the outer bearing ring 22, the fat depot 23 clearly divided into two chambers 48, 49 , By such an inexpensive producible by deep drawing angle plate 50 is avoided that due to the centrifugal force acting grease from the radially lower chamber 48 moves with the teilkreiskleineren Wälzkörperreihe 19 in the radially higher chamber 49 with the pitch circle roller row and thus possibly causes a one-sided lack of lubrication ,

Finally, a fourth embodiment is shown in Figure 6, in which the ball roller bearing unit 13 is also designed as a single-acting tandem angular contact roller bearings 51, but in which by level alignment of the inner races 17, 18 of the rows of rolling elements 19, 20 in the bearing flange 10 on the primary mass 4 both rows of rolling elements 19, 20 have the same pitch circle diameter and the same number of rolling elements. A special feature of this embodiment is that the outer bearing ring of the tandem angular contact roller bearing 51 preferably consists of two separate partial rings 52, 53, which as deep-drawn parts of a rolling bearing material in the Substantially cylindrical outer circumferential surface are formed and each form the outer races 27, 28 for both rows of rolling elements 19, 20. Similar to the first variant of the described third embodiment, the transmission of axial operating forces from the secondary mass 8 to the tandem angular ball roller bearing 51 is effected here by an annular collar 54 integrally formed on the bearing flange 11 of the secondary mass 8, which at the opposite axial side of the partial ring 52 of the outer Bearing ring 22 is present. For sealing the fat deposits 23 between the bearing rings 21, 52, 53, however, a labyrinth seal 56 is provided in this embodiment, the see between a pressed onto the sub-ring 52 of the outer bearing ring 22 U-profile ring 55 and an unspecified designated a circumferential groove in the Bearing flange 10 is formed on the primary mass 4, while on the other hand turn on a pressed onto the sub-ring 53 of the outer bearing ring 22 L-profile-shaped sealing cap 57 is used, are vulcanized to the two on the bearing flange 10 of the primary mass 4 grinding sealing lips 58.

LIST OF REFERENCE NUMBERS

Dual mass flywheel 30 Double lip seal

Drive shaft 31 ring cap

Connecting screws 32 sealing lips

Primary mass 33 preload spring

Coupling 34 Radial-axial ball roller bearing

Friction linings 35 Support board

Input shaft 36 double lip seal

Secondary mass 37 sealing cap

Damping means 38 sealing lip

Bearing flange 39 Tandem angular contact ball bearing

Bearing flange 40 collar

Rolling bearings 41 Double lip seal

Ball roller bearing unit 42 Sealing cap

Rolling element 43 Sealing lip

Side surfaces 44 U-profile ring

Side surfaces 45 gap seal

Inner race 46 sealing cap

Inner race 47 sealing lip

Rolling element series 48 chamber

Series of rolling elements 49 Chamber inner bearing ring 50 Angular disk outer bearing ring 51 Tandem angular contact ball bearing

Fettdepot 52 partial ring

Plastic window cage 53 partial ring

Plastic window cage 54 Ring collar

Angular ball bearing 55 U-profile ring

Outer race 56 Labyrinth seal

Outer race 57 sealing cap

Support shelf 58 sealing lip

Claims

Schaeffler KG Industriestr. 1 - 3, 91074 HerzogenaurachPatentansprüche

1. Torsionsschwingungsdämpfer, in particular two-mass flywheel between see the internal combustion engine and the transmission of a motor vehicle, consisting essentially of a directly to the drive shaft (2) of the internal combustion engine connected disc-shaped primary mass (4) and arranged coaxially to this primary mass (4) and about a clutch (5) connected to the input shaft (7) of the transmission disc-shaped secondary mass (8), wherein the primary mass (4) and the secondary mass (8) coupled by a plurality of damping means (9) arranged between these and against the action of these Damping means (9) via a between a coaxial bearing flange (10) on the primary mass (4) and a coaxial bearing flange (11) on the secondary mass (8) arranged rolling bearing (12) are rotatable relative to each other, characterized in that the rolling bearing (12) between the primary mass (4) and the secondary mass (8) by one of the up is formed, which Wälz- body (14) as ball rollers with two symmetrically flattened by a spherical base shape and parallel to each other side surfaces (15, 16) are formed, wherein the Bearing flange (10) on the primary mass (4) at the same time as with molded inner raceways (17, 18) for both Wälzkörperreihen (19, 20) provided inner bearing ring (21) of the ball roller bearing unit (13) is formed and the axially sealed on both sides gap between the bearing flange (10) and a separate outer bearing ring (22) than in each case reduced by the ball basic shape men of all rolling elements (14) enlarged fat depot (23) is formed.

2. torsional vibration damper according to claim 1, characterized in that the ball rollers designed as and preferably in separate Kunstofofffensterkäfigen (24, 25) guided rolling elements (14) both Wälzkörperreihen (19, 20) of the ball roller bearing unit (13) preferably each have a width between their Side surfaces (15, 16) of about 60% to 70% of the diameter of their ball base shape and by their shape increase the volume of the fat deposits (23) between the bearing rings (21, 22) by about 20% to 30% compared to an identically sized ball bearing ,

3. Torsional vibration damper according to claim 2, characterized in that the ball roller bearing unit (13) is preferably designed as a double-row prestressed angular ball roller bearing (26) in an X-arrangement against each other ten rolling elements (19, 20) whose outer bearing ring (22) as a deep-drawn part is formed of a rolling bearing material with at least one molded outer raceway (27) for a row of rolling elements (19).

4. Torsionsschwingungsdämpfer according to claim 3, characterized in that the outer bearing ring (22) of the angular ball bearing (26) on its primary mass (4) facing axial side in the profile cross-section U-shaped and with respect to the outer circumferential surface of the bearing ring (22) raised support board (29 ), via which axial operating forces from the bearing flange (11) of the secondary mass (8) to the angular ball roller bearing (26) are transferable.

5. torsional vibration damper according to claim 4, characterized in that the sealing of the fat deposits (23) between the bearing rings (21, 22) on the one hand by a in the inner profile of the support board (29) on the outer bearing ring (22) pressed double Lipendichtring (30) and on the other hand by an L-profilför-mige ring cap (31) pressed onto the outer bearing ring (22) with two vulcanized sealing lips (32).

6. Torsionsschwingungsdämpfer according to claim 5, characterized in that the outer race (28) for the second row of rolling elements (20) is preferably formed as a separate, in the outer bearing ring (22) pressed-ring member and for adjusting the bias of the Schrägkugelrollenla- gers (26) between the separate outer raceway (28) and the Doppelellip- pendichtring (30) Abstimmscheiben or biasing springs (33) are inserted.

7. Torsionsschwingungsdämpfer according to claim 2, characterized in that the ball roller bearing unit (13) preferably as a combined radial-axial ball roller bearing (34) with an obliquely staffed Wälzkörperreihe (20) and a vertically grooved Wälzkörperreihe (19) is formed, the outer bearing ring ( 22) as a deep-drawn part of a rolling bearing material with molded outer raceways (27, 28) for both rows of rolling elements (19, 20) is formed.

8. torsional vibration damper according to claim 7, characterized in that the outer bearing ring (22) of the radial-axial ball roller bearing (34) at its primary mass (4) facing axial side in a profile cross-section L-shaped and relative to the outer circumferential surface of the bearing ring (22) has received support board (35) via which axial operating forces from the bearing flange (11) of the secondary mass (8) to the radial-axial ball roller bearing (34) are transferable.

9. torsional vibration damper according to claim 8, characterized in that the sealing of the fat deposits (23) between the bearing rings (21, 22) on the one hand by a in the inner profile of the support board (35) on the outer bearing ring (22) pressed reinforced double lip seal (36) and on the other hand by an on the outer bearing ring (22) pressed L-profile-shaped sealing cap (37) with a vulcanized sealing lip (38).

10. Torsionsschwingungsdämpfer according to claim 2, characterized in that the ball roller bearing unit (13) is preferably formed as a unidirectional tandem angular ball bearing (39), in which by Stufenförmi- ge arrangement of the inner raceways (17, 18) of the Wälzkörperreihen (19, 20) in the bearing flange (10) on the primary mass (4) having a Wälzkörperreihe (19) has a larger pitch circle diameter and a larger number of Wälzkörperanzahl than the other row of rolling elements (20).

H .Torsionsschwingungsdämpfer according to claim 10, characterized in that the outer bearing ring (22) of the tandem angular ball bearing (39) preferably as Fließpressteil of a rolling bearing material with step-shaped outer raceways (27, 28) for both Wälzkörperreihen (19, 20) and substantially cylindrical outer circumferential surface is formed.

12.Torsionsschwingungsdämpfer according to claim 11, characterized in that for transmitting axial operating forces of the secondary mass (8) on the tandem angular ball bearing (39) on the bearing flange (11) of the secondary mass (8) on an axial side of the outer bearing ring (22) fitting annular collar (40) is integrally formed.

13.Torsionsschwingungsdämpfer according to claim 12, characterized in that the sealing of the fat deposits (23) between the bearing rings (21, 22) on the one hand by a in the outer bearing ring (22) pressed double lip seal (41) and on the other hand by a on the outer bearing ring ( 22) pressed L-profile-shaped sealing cap (42) with a vulcanized sealing lip (43).

14.Torsionsschwingungsdämpfer according to claim 10, characterized in that the outer bearing ring (22) of the tandem angular ball bearing (39) preferably as deep-drawn part of a rolling bearing material with step-shaped outer raceways (27, 28) for both Wälzkörperreihen (19, 20) and correspondingly stepped Outer jacket surface is formed.

15.Torsionsschwingungsdämpfer according to claim 14, characterized in that for transmitting axial operating forces of the secondary mass (8) on the tandem angular contact ball roller bearing (39) of the bearing flange (11) of the Se kundärasse (8) has a to the step-shaped design of the outer bearing ring (22) adapted cross-sectional profile.

16.Torsionsschwingungsdämpfer according to claim 15, characterized in that the sealing of the fat deposits (23) between the bearing rings (21, 22) on the one hand by a between the outer bearing ring (22) pressed U-profile ring (44) and the bearing flange (10 ) formed on the primary mass (4) gap seal (45) and on the other hand by a on the outer bearing ring (22) pressed L-shaped sealing cap (46) with a vulcanized sealing lip (47).

17, torsional vibration damper according to claim 13 and 16, characterized in that between the Wälzkörperreihen (19, 20) arranged in the outer bearing ring (22) and the grease depot (23) in two chambers (48, 49) dividing angle plate (50) is with which a centrifugally induced migration of grease from the radially lower chamber (48) arranged in the radially higher chamber (49) is avoidable.

18.Torsionsschwingungsdämpfer according to claim 2, characterized in that the ball roller bearing unit (13) is preferably designed as a unidirectional tandem angular ball roller bearing (51) in which by level alignment of the inner raceways (17, 18) of the Wälzkörperreihen (19, 20) in the bearing flange (10) on the primary mass (4) both rows of rolling elements (19, 20) have the same pitch circle diameter and the same number of rolling elements.

19.Torsionsschwingungsdämpfer according to claim 18, characterized in that the outer bearing ring (22) of the tandem angular ball bearing (51) preferably consists of two separate formed as deep-drawn parts of a rolling bearing material partial rings (52, 53), each of the outer races (27 , 28) for both rows of rolling elements (19, 20) form and have substantially cylindrical outer circumferential surfaces.

20.Torsionsschwingungsdämpfer according to claim 19, characterized in that for transmitting axial operating forces of the secondary mass (8) on the tandem angular ball bearing (51) to the bearing flange (11) of

Secondary mass (8) on one axial side of a partial ring (52) of the outer bearing ring (22) abutting annular collar (54) is integrally formed.

21.Torsionsschwingungsdämpfer according to claim 20, characterized in that the sealing of the fat deposits (23) between the bearing rings (21, 22) on the one hand by a pressed onto the part ring (52) U-profile ring (55) and the bearing flange ( 10) formed on the primary mass (4) labyrinth seal (56) and on the other hand by a on the sub-ring (53) pressed L-shaped sealing cap (57) with two vulcanized sealing lips (58).