A radial bearing (1a, 1b, 1c, 1d), especially a roller bearing (9) comprising an outer part (3) opposite an inner part (6), which rotate relative to each other about a common longitudinal axis (5) is provided, wherein the radial bearing (1a, 1b, 1c, 1d) includes a bearing seat (8a, 8b, 8c, 8d), which is formed in the outer part (3) or on the inner part (6) and which is generally stationary relative to a radial load (10a, 10b, 10c, 10d) acting on the bearing seat (8a, 8b, 8c, 8d) in a load zone (12a, 12b, 12c, 12d). The bearing seat (8a, 8b, 8c, 8d) features a variable width over its extent in a direction of the longitudinal axis (5), such that, starting from the load zone (12a, 12b, 12c, 12d), the bearing seat (8a, 8b, 8c, 8d) is significantly narrowed outside of the load zone (12a, 12b, 12c, 12d).
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1. radial bearing comprising an outer part opposite an inner part, the outer part and the inner part are rotatable relative to each other about a common longitudinal axis, the bearing comprises a bearing seat having a circumference, which is formed in the outer part or on the inner part and which is generally stationary relative to a radial load acting on the bearing seat in a load zone, the bearing seat has a load carrying width that varies in a direction parallel to the longitudinal axis about the circumference of the bearing seat, such that, starting from the load zone, the bearing seat has a first load carrying width and outside of the load zone the bearing seat has a second width that is narrower than the first load carrying width, whereby a mass and weight of the bearing seat are reduced outside of the load zone:
(i) wherein the outer part comprises a housing, the inner part comprises a shaft mounted in the housing, and the bearing seat of variable width is formed in the housing, and the radial load is generally stationary relative to the housing; or
(ii) wherein the outer part comprises a hub, the inner part comprises an axle bolt supporting the hub, and the bearing seat of variable width is formed in the hub, and the radial load rotates with the hub; or
(iii) wherein the outer part comprises a hub, the inner part comprises an axle bolt supporting the hub, and the bearing seat of variable width is formed on the axle bolt, and the radial load is generally stationary relative to the axle bolt.
0. 2. radial bearing according to
0. 3. radial bearing according to
0. 4. radial bearing according to
0. 5. radial bearing according to
6. radial bearing according to
0. 7. radial bearing according to
0. 8. radial bearing according to
0. 9. radial bearing according to
10. radial bearing according to claim 9 6, wherein a width of the needle bearing generally corresponds to a maximum width of the bearing seat of variable width allocated to the needle bearing, and a minimum width of the bearing seat is smaller than a length of needles of the needle bearing, and for lubricating the needle bearing, only an unpressurized lubricant spray is provided.
11. radial bearing according to claim 9 6, wherein the needle bearing is a needle sleeve formed without an inner ring.
12. radial bearing according to
13. radial bearing according to
14. radial bearing according to
15. radial bearing according to
16. radial bearing according to
17. radial bearing according to
18. radial bearing according to
19. radial bearing according to
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The invention relates to a radial bearing, especially a roller bearing, comprising an outer part opposite an inner part, wherein the outer part and the inner part rotate relative to each other about a common longitudinal axis, wherein the radial bearing comprises a bearing seat, which is formed in the outer part or on the inner part and which is generally stationary relative to a radial load acting on the bearing seat in a load zone.
In principle, such a load case of the radial bearing is known in the state of the art and, in the case of the roller bearing, under the designation of the so-called point load, at which the radial load is generally stationary as a function of the movement relationships relative to the inner ring or the outer ring of the roller bearing. In contrast, the so-called peripheral load involves a load case, in which the radial load rotates relative to the inner ring or to the outer ring of the roller bearing. At this point it should be expressed that with the term “generally,” the radial load, strictly speaking, does not have a point-like effect, but instead can feature a certain radial variation width.
Due to the radial load, the forces to be transferred from the load zone resting on the bearing seat are an essential criteria for setting width dimensions that are stable over the service life for the bearing seat. However, for the required width outside of the load zone, this criteria still plays only a subordinate role, because here the bearing seat is subjected to a clearly lower mechanical load and, in individual cases, to no mechanical load at all. Nevertheless, bearing seats known in the state of the art have a rotationally symmetric shape with constant width and therefore are overdimensioned outside of the load zone in terms of the mechanical loadability for the load case of the point load presented here. This can produce significant disadvantages to the extent that the bearing seat, which is, for the most part, superior in terms of production, has a mass, which is undesired both technically and financially, and which is also avoidable, outside of the load zone. Simultaneously, this leads to unnecessarily high processing costs for the bearing seat, which is to be machined with fine precision completely over its constant width, even outside of the load zone.
Therefore, the objective of the invention is to provide a radial bearing of the above-noted type, such that the cited disadvantages are overcome with simple means. Here, the shaping of the radial bearing should lead, in particular, to advantages in terms of weight and costs relative to bearings known in the state of the art.
According to the invention, this objective is met in that the bearing seat features a width that varies over its extent in a direction of the longitudinal axis, such that the bearing seat, starting from the load zone, becomes significantly narrower outside of the load zone. A radial bearing embodied in this way optimally takes into consideration the circumstance that the bearing seat receiving a point load can become considerably narrower outside of the load zone and thus the potential of known bearings for reducing the rotationally moving mass, weight, processing expense, and costs can be used without negatively affecting the functional properties of the radial bearing. Such a negative effect would be produced when, due to the narrowing of the bearing seat, a critical bearing seat width is reached, below which a given fatigue strength of the radial bearing is no longer guaranteed. In the case of a radial bearing formed as a hydrodynamic sliding bearing, the decisive criteria for this negative effect can be a lubricant film that no longer bears weight or, in the case of a roller bearing, an impermissibly high loading of the bearing seat outside of the load zone.
Furthermore, the invention can be used advantageously, especially for the arrangements described below, in the load cases, in which a point load acts on the bearing seat. In a first load case, the outer part should be formed as a housing, the inner part as a shaft mounted in the housing, and the bearing seat of variable width should be formed on the shaft, wherein the radial load rotates with the shaft. A typical example for this load case is a defined unbalanced weight, which rotates with the shaft and which can be advantageously reinforced by the bearing seat of variable width while simultaneously reducing the mass of the shaft.
In a second load case, the outer part should be formed as a housing, the inner part as a shaft mounted in the housing, and the bearing seat of variable width should be formed in the housing, wherein the radial load is generally stationary relative to the housing. In this case, a bearing seat of variable width can lead to a mass reduction of the housing, wherein a non-uniform mass distribution of the bearing seat over its extent does not require counterbalancing of the stationary housing.
In a third load case, the outer part should be formed as a hub, the inner part as an axle bolt supporting the hub, and the bearing seat of variable width should be formed in the hub, wherein the radial load rotates with the hub. Here, in the case of hub bearing with a defined unbalanced weight on the hub, despite an action of the bearing seat, which is narrower in the unbalanced weight direction, directed against the unbalanced weight, it can be advantageous to form these with variable width, in order to reduce the surface area of the bearing seat in favor of more economical processing.
Finally, in a fourth load case, the outer part should be formed as a hub, the inner part as an axle bolt supporting the hub, and the bearing seat of variable width should be formed on the axle bolt, wherein the radial load is generally stationary relative to the axle bolt. In this case, the processing of the bearing seat with its narrow section is also simplified, while simultaneously a reduction of the mass of the axle bolt can be achieved.
For the first load case, it is also provided that the shaft is formed as an unbalanced weight, whose center of mass is arranged eccentric to the longitudinal axis and results from one or more recesses on the outer periphery of the unbalanced weight shaft. Here, the recesses should run partially or completely on the other side of the longitudinal axis from the unbalanced weight relative to the center of mass of the unbalanced weight shaft and should be directly adjacent to the bearing seat of variable width. With an unbalanced weight shaft formed in this way, for the most part the conflicting goal of a mass that is as small as possible for the largest possible unbalanced weight of the shaft can be solved especially advantageously. This is substantiated essentially in that the recesses now extend up to the narrowed bearing seat and can be used simultaneously as an increased negative mass for reinforcing the unbalanced weight. Nevertheless, in comparison with conventional shafts with bearing seats of constant width, this opens up the possibility not only of increasing the unbalanced weight of the shaft for simultaneous mass reduction, but also makes possible a more extensive mass reduction of the shaft, in the case of a non-variable unbalanced weight, such that the original unbalanced weight increase is equalized by additional recesses, which reduce the mass and which are arranged on the unbalanced weight shaft on the side of the longitudinal axis relative to the center of mass. Obviously, a fit can also be found between these two limiting cases for the unbalanced weight shaft, according to whether the design focus is on a moderate mass reduction with significant increase in unbalanced weight or on a clear mass reduction with unchanged increase in unbalanced weight.
In an improvement of the invention, the unbalanced weight shaft belongs to a device for balancing the mass forces and/or mass moments of a reciprocating piston internal combustion engine with a crankshaft arranged parallel to the longitudinal axis of the unbalanced weight shaft and indirectly driving the unbalanced weight shaft for the reciprocating piston internal combustion engine. Such a balancing device is known to someone skilled in the art in the field of reciprocating piston internal combustion engines, especially for in-line or V arrangements, as an effective measure for reducing vibrations due to oscillating mass forces. However, especially for reciprocating piston internal combustion engines used in the field of vehicles, the requirements on the light construction quality of the reciprocating piston internal combustion engine are continuously rising, so that the previously mentioned possibilities for reducing the mass of the unbalanced weight shaft can be used particularly advantageously for this application. The lower mass moment of inertia of the unbalanced weight shaft also leads to improved dynamics of the reciprocating piston internal combustion engine, because less resistance is counteracted by a high rotational speed gradient. In addition, especially for diesel engines with pronounced rotational vibrations of the crankshaft in the lower range of rotational speeds, the peak mechanical load in the driving range of the balancing device can be reduced with a decreased moment of inertia of the unbalanced weight shaft.
The previously stated considerations apply to a greater degree when the balancing device includes two unbalanced weight shafts, which rotate in opposite directions at twice the crankshaft rotational speed. This arrangement, which is also known to someone skilled in the art as a Lancaster balance, is used for balancing the second order free mass forces in a four-cylinder, in-line engine.
In an especially advantageous continuation of the invention, it is further provided that for radial bearing of the unbalanced weight shaft, there is at least one roller bearing formed as a needle bearing without an inner ring and preferably as a needle sleeve. In addition to the favorable friction properties of the roller bearing in comparison with a hydrodynamic sliding bearing, which can lead to considerable friction power losses in the drive of the unbalanced weight shaft, especially at low temperatures and/or high viscosity hydraulic means, the advantages of the mass reduction and/or the unbalanced weight increase of the unbalanced weight shaft can be even further expanded by its radial bearing formed as a needle bearing, because the bearing seat included by a needle bearing as a rule can be narrowed more than would be possible in the case of the hydrodynamic sliding bearing based on consideration of a weight-bearing lubricant film. Thus, the use of a needle sleeve, which is known to someone skilled in the art as a component with very small radial installation height with a non-cutting shaped outer ring and a needle collar, permits above all an especially installation space-saving and economical radial bearing for the unbalanced weight shaft with sufficient fatigue strength.
Finally, in another construction of the invention, it can be advantageous that the width of the needle bearing generally corresponds to a maximum width of the variable width bearing seat allocated to the needle bearing, while a minimum width of the bearing seat is smaller than the length of the needles of the needle bearing. Simultaneously, for lubricating the needle bearing, only one unpressurized lubricant spray should be provided. With this construction of the invention, the lubricating conditions on the needle bearing can be improved in that the needles projecting locally and temporarily over the bearing seat can be exposed to the unpressurized lubricant spray to a greater degree. Conversely, this can also be beneficial for a discharge of abrasive particles from the track region of the needle bearing in favor of increased wear resistance of the radial bearing. In order to guarantee total contact of the length of the bearing seat with the needles in the region of the minimal width of this bearing seat, for this construction the use of a one-row needle bearing is to be preferred relative to the use of a multi-row needle bearing.
Additional features of the invention will become apparent from the following description and from the drawings, in which the radial bearing according to the invention is shown simplified basically for the previously mentioned load cases and, as an example, with reference to an unbalanced weight shaft of a device for balancing the masses of a reciprocating piston internal combustion engine. Shown are:
In
The load case shown in
In the load case shown in
Finally, the load case shown in
The load case shown in
The bearing of one of these unbalanced weight shafts 23 appears in detail from its longitudinal representation according to
The unbalanced weight 11a acting in the direction of the arrow for the unbalanced weight shaft 23 is based on a center of mass 29 eccentric to its longitudinal axis 5 and represented symbolically in
A structural construction in the area of the second limit case is shown symbolically in the present embodiment in the form of an unbalanced weight 11d equalizing the additional unbalanced weight 11c. As is also visible from
List of reference numbers and symbols
1a, b, c, d
Radial bearing
2a, b
Housing
3
Outer part
4a, b
Shaft
5
Longitudinal axis
6
Inner part
7a, b, c, d
Bearing seat
8a, b, c, d
Bearing seat
9
Roller bearing
10a, b, c, d
Radial load
11a, b, c, d
Unbalanced weight
12a, b, c, d
Load zone
13a, b
Hub
14a, b
Axle bolt
15
Sliding bearing
16
Device
17
Reciprocating piston internal combustion engine
18
Cylinder
19
Piston
20
Connecting rod
21
Crankshaft
22
Intermediate shaft
23
Unbalanced weight shaft
24
Ball bearing
25
Needle sleeve
26
Maximum width
27
Minimum width
28
Needle
29
Center of mass
30
Recess
31
Recess
E
Plane
u
Direction of unbalanced weight
v
Direction orthogonal to unbalanced weight
Tisch, Siegfried, Solfrank, Peter
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