HUB-BEARING ASSEMBLY FOR A WHEEL OF A MOTOR VEHICLE

Abstract

A hub-bearing assembly includes a stationary radially outer ring, a flanged hub extending into an interior of the radially outer ring and configured to rotate relative to the radially outer ring, and a radially inner ring mounted on the flanged hub and axially abutting a portion of the flanged hub. A toothed sleeve is mounted to the hub axially inward of and adjacent to the radially inner ring and is rotationally fixed relative to the flanged hub. An auxiliary ring is mounted on the flanged hub in contact with the toothed sleeve. The flanged hub includes an axially inner rolled edge in contact with the auxiliary ring and configured to maintain a preload on the radially inner ring, and the auxiliary ring is configured to accommodate deformations induced by the formation of the rolled edge.

Claims

1. A hub-bearing assembly comprising: a rotation axis; a stationary radially outer ring; a flanged hub extending into an interior of the radially outer ring and configured to rotate around the rotation axis and relative to the radially outer ring; a radially inner ring mounted on the flanged hub and axially abutting a portion of the flanged hub; a toothed sleeve mounted axially inward of and adjacent to the radially inner ring, the toothed sleeve being rotationally fixed relative to the flanged hub; an auxiliary ring mounted on the flanged hub in contact with the toothed sleeve, wherein the flanged hub includes an axially inner rolled edge in contact with the auxiliary ring and configured to maintain a preload on the radially inner ring; and wherein the auxiliary ring is configured to accommodate deformations induced by a formation of the rolled edge.

2. The assembly according to claim 1, wherein the auxiliary ring has a radially outer cylindrical surface, a radially inner cylindrical surface parallel to the radially outer cylindrical surface, an axially inner annular surface, an axially outer annular surface parallel to the axially inner annular surface and a curvilinear surface between the axially inner annular surface and the radially inner cylindrical surface.

3. The assembly according to claim 2, wherein the toothed sleeve includes an annular recess forming a seat for the auxiliary ring, the seat having a radially inwardly facing cylindrical surface and an axially inwardly facing annular surface, and wherein the auxiliary ring is mounted in the recess.

4. The assembly according to claim 3, wherein the auxiliary ring and the toothed sleeve are configured such that a radial gap exists between the inwardly facing cylindrical surface of the toothed sleeve and the radially outer cylindrical surface of the auxiliary ring.

5. The assembly according to claim 4, wherein a fit between the radially inner cylindrical surface of the auxiliary ring and a radially outer cylindrical surface of the flanged hub varies from a diameter clearance of 0.01 mm to a diameter interference of 0.05 mm.

6. The assembly according to claim 5, wherein a central opening of the flanged hub includes a first enlarged portion on an axially outer side and a first radial ball bearing in the first enlarged portion and a second enlarged portion on an axially inner side and at least one second radial ball bearing in the second enlarged portion, and a radially outer first relief groove surrounding the second enlarged portion, and wherein the hub-bearing assembly further includes a spacer mounted in the second enlarged portion axially between and in contact with the at least one second radial ball bearing and a shoulder defining an axial end of the second enlarged portion.

7. The assembly according to claim 6, wherein the second enlarged portion includes a second radially inner relief groove located axially outward of the first relief groove.

8. The assembly according to claim 7, wherein the at least one second radially ball bearing is a single ball bearing.

9. The assembly according to claim 7, wherein the at least one second radially ball bearing is exactly two second ball bearings.

10. The assembly according to claim 1, wherein a central opening of the flanged hub includes a first enlarged portion on an axially outer side and a first radial ball bearing in the first enlarged portion and a second enlarged portion on an axially inner side and at least one second radial ball bearing in the second enlarged portion, and a radially outer first relief groove surrounding the second enlarged portion, and wherein the hub-bearing assembly further includes a spacer mounted in the second enlarged portion axially between and in contact with the at least one second radial ball bearing and a shoulder defining an axial end of the second enlarged portion.

11. The assembly according to claim 10, wherein the second enlarged portion includes a second radially inner relief groove located axially outward of the first relief groove.

12. A method of assembling a bearing hub assembly comprising: inserting a flanged hub into a stationary radially outer ring; mounting a radially inner ring on the flanged hub in axial abutment with a portion of the flanged hub; mounting a toothed sleeve on the flanged hub axially inward of and adjacent to the radially inner ring such that the toothed sleeve is rotationally fixed relative to the flanged hub; mounting an auxiliary ring on the flanged hub in contact with the toothed sleeve, forming a rolled edge on the flanged hub such that the rolled edge contacts the auxiliary ring and maintains a preload on the radially inner ring and accommodates deformations induced by a formation of the rolled edge.

13. The method according to claim 12, wherein the auxiliary ring has a radially outer cylindrical surface, a radially inner cylindrical surface parallel to the radially outer cylindrical surface, an axially inner annular surface, an axially outer annular surface parallel to the axially inner annular surface and a curvilinear surface between the axially inner annular surface and the radially inner cylindrical surface.

14. The method according to claim 13, wherein the toothed sleeve includes a recess; and wherein the method includes mounting the auxiliary ring in the recess.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

[0015] Non-limiting embodiments of the invention are described below with reference to the attached drawings.

[0016] FIG. 1 is a cross sectional view of a hub-bearing assembly that can be provided with a system for disconnecting the assembly from a transmission shaft according to one embodiment of the present disclosure.

[0017] FIG. 2 is a detail view of a first portion the hub-bearing assembly of FIG. 1.

[0018] FIG. 3 is a detail view of a second portion of the hub-bearing assembly of FIG. 1.

[0019] FIG. 4 is a detail view of a radially inner relief groove of a flanged hub of the hub-bearing assembly in FIG. 1.

DETAILED DESCRIPTION

[0020] By way of non-limiting example, the present disclosure is described below with reference to a hub-bearing assembly for motor vehicles provided with a bearing unit.

[0021] With reference to FIG. 1, a hub-bearing assembly according to a preferred embodiment of the disclosure is denoted as a whole using reference sign 10. The figure shows a detail of an example arrangement.

[0022] When in use, the hub-bearing assembly 10 is interposed between a wheel and a frame, both of which are known and not illustrated, of a vehicle, and may be selectively coupled to a constant velocity joint, which is known and not illustrated, via a transmission device, which is also known and not illustrated, to transmit or otherwise the drive torque to the respective wheel (not illustrated).

[0023] The hub-bearing assembly comprises a rotary flanged hub 20 and a bearing unit 30 having a central axis of rotation X, a stationary radially outer ring 31, and a radially inner ring. The radially inner ring is defined by a portion 20a of the flanged hub 20 and a further radially inner ring 34 mounted on and rigidly connected to the flanged hub 20, both of the radially inner rings 20a, 34 being rotatable with respect to the radially outer ring 31 as a result of the interposition of two rows of rolling bodies 32, 33, in this case balls between the inner and outer rings.

[0024] Throughout the present description and in the claims, terms and expressions indicating positions and orientations, such as radial and axial, are to be understood with reference to the central axis of rotation X of the bearing unit 30. On the other hand, expressions such as axially outer and axially inner refer to the mounted state of the hub-bearing assembly, and in this case preferably refer to a wheel side and to a side opposite the wheel side respectively.

[0025] To simplify the graphical representation, reference signs 32, 33 are used to denote both individual balls and rows of balls. Again for the sake of simplicity, the term ball shall be used by way of example in the present description and in the attached drawing instead of the more generic term rolling body (and the same reference signs shall be used).

[0026] The flanged hub 20 has a central through-hole 21 that extends along the axis X that is configured to be engaged by the constant velocity joint and comprises, on an axially outer side thereof, a flange 25 for fastening the hub-bearing assembly 10 to a wheel of a vehicle, and, on an axially inner side thereof, a rolled edge 24 (obtained, e.g., by orbital roll forming) that is designed to axially preload both the inner ring 34 and a toothed sleeve 55 against a radially outer shoulder 22 of the flanged hub 20. The toothed sleeve 55, which is mounted close to the inner ring 34, is coupled to a toothed profile 23 of the flanged hub 20.

[0027] Bearings are required between the hub and the constant velocity joint to enable the flanged hub 20 to rotate independently of the constant velocity joint when the flanged hub 20 is disengaged from the constant velocity joint. In the present embodiment, the flanged hub 20 has a first radially inner shoulder 26 and a second radially inner shoulder 27 formed inside the central through-hole 21, close to which are mounted at least two sets of radial ball bearings 59, 60 (exactly two sets in the configuration illustrated in FIG. 1). The two sets of radial ball bearings 59, 60 are mounted in a radially inner position with respect to the flanged hub 20. In particular, the first radial ball bearings 59 are mounted in an axially outer position against the first shoulder 26 and the second radial ball bearings 60 are mounted in an axially inner position against the second shoulder 27. There may also be an additional set of radial ball bearing in the axially inner position.

[0028] According to the present disclosure and also with reference to FIG. 2, a ring 50 of rectangular section, sometimes referred to as an auxiliary ring, is provided between the rolled edge 24 of the flanged hub 20 and the toothed sleeve 55. Therefore, according to the disclosure, the rolled edge 24 of the flanged hub 20 axially preloads and blocks three axially adjacent components against the shoulder 22 of the flanged hub 20, namely, the radially inner ring 34, the toothed sleeve 55 and the ring 50.

[0029] The ring 50, which has a solid rectangular section, supports the material of the rolled edge 24. Indeed, whereas in the prior art the material of the rolled edge is deformed and pushed against the discontinuous surface of the toothed sleeve, according to the present disclosure the material of the rolled edge 24 is pressed against an axially inner annular surface 50a of the ring 50, the annular surface resulting from the fact that the ring 50 has a rectangular section, and not for example a circular section. The annular surface 50a is a solid surface and is continuous for 360, i.e. it has no gaps. Consequently, the ring 50 of rectangular section is able to absorb the deformations induced by the orbital roll forming process without these deformations being transferred to the toothed sleeve 55.

[0030] Advantageously and more specifically, the ring 50 includes the axially inner annular surface 50a, a radially inner cylindrical surface 51, and an axially inner, radially inner curvilinear surface 50b that has a radius of curvature R and that is located between the annular surface 50a and to the cylindrical surface 51.

[0031] Advantageously, the connection radius R is selected so that an axially outer, radially outer profile 24 of the plastically deformed rolled edge 24 can be fitted to the radially inner, axially inner profile of the ring 50 formed by the annular surface 50a and the cylindrical surface 50b. Indeed, the material of the rolled edge 24 is deformed and pressed against the annular surface 50a and the curvilinear surface 50b, both of which are continuous and have no gaps around the full 360 of the circumference. The combination of the two surfaces 50a, 50b provides a broad support area for the material of the rolled edge 24, and therefore reduces a concentration of stresses on the rolled edge 24 of the flanged hub 20 as compared to the stresses that would occur without the use of the ring 50.

[0032] The ring 50 is accommodated in a seat (recess) 56 of the toothed sleeve 55, the seat 56 comprising a radially inner cylindrical surface 56a and an axially inner annular surface 56b that blocks axial movement of the ring 50.

[0033] The fit between the cylindrical surface 56a of the toothed sleeve 55 and a radially outer cylindrical surface 52 of the ring 50 is preferably a clearance fit. This clearance fit enables the ring 50, after being pressed axially by the orbital roll forming process, to freely expand radially without the radial expansion being transferred to the toothed sleeve 55. This prevents the external toothed profile 57 of the toothed sleeve 55 from expanding radially, thereby avoiding potential coupling problems between the toothed sleeve and the ring gear of the disconnection system during operation.

[0034] Advantageously, the coupling between the ring 50 and the flanged hub 20, in particular between the radially inner cylindrical surface 51 of the ring 50 and a radially outer cylindrical surface 20 of the flanged hub 20, may be a coupling with slight interference (on the borderline, or even with minimum clearance) on the basis of the actual dimensions of the two components according to the respective specified tolerances. Preferably, the coupling may vary between a diametric clearance of 0.01 mm to an interference of 0.05 mm (in all cases relative to the diameters of the two coupling surfaces). Indeed, greater interference would excessively deform the ring 50 and, in particular, would increase the external diameter thereof with a consequent impact on the coupling between the cylindrical surface 56a of the toothed sleeve 55 and the cylindrical surface 52 of the ring 50, for which adequate clearance is required, as mentioned above. Furthermore, low interference values require low press-fitting forces and consequently simplify assembly, while guaranteeing that the components are centered and that the stresses caused by press-fitting both on the ring 50 and on the neighboring areas of the flanged hub 20 are reduced.

[0035] Advantageously, the profile of the rolled edge 24 remains absolutely unchanged from rolled edges formed on conventional wheel hubs. In particular, the diameter of the cylindrical surface 20 on which lies the bending point F from which the rolled edge 24 bends radially outwards remains unchanged. This feature enables the internal geometry of the flanged hub 20 to remain unchanged, and does not reduce the diameter of the central hole 21 on which the shoulders 26, 27 are formed for the radial ball bearings 59, 60. Consequently, the disclosed arrangement does not adversely affect the structural strength of the flanged hub 20 or the available internal radial space. Consequently, the size of the radial ball bearings also remains unchanged.

[0036] With reference to FIG. 3, the axial stop for the second radial ball bearing 60 is advantageously formed by a spacer 70 located radially inside the flanged hub 20 and axially outside the radial ball bearing 60. The spacer 70 is in turn axially blocked against the second axial shoulder 27 of the flanged hub 20.

[0037] More specifically, the two radial bearings 59, 60 and the spacer 70 can be pressed by a ring (known and not illustrated) against the shoulders 26, 27, thereby exerting axial compression forces through the outer rings of the bearings 59, 60 and the spacer 70 onto the flanged hub 20 that, in response to the stresses transmitted by the wheel, generate reactive forces on the shoulders 26, 27 and in particular on the axially inner second shoulder 27, since it is located in the part of the flanged hub 20 with a radial section SR of reduced thickness. Indeed, this radial section SR coincides with a radially outer relief groove 29 of the flanged hub 20 that is required to machine the toothed coupling profile 23 between the flanged hub 20 and the toothed sleeve 52, and therefore necessarily has a reduced thickness.

[0038] The presence of the spacer 70 helps to improve the performance under stress of the geometry of the flanged hub 20. Indeed, to enable grinding operations to be carried out on the radially inner seat 28 of the flanged hub 20, which accommodates the second radial ball bearing 60 and the spacer 70, a relief groove 80 has to be defined between the seat 28 and the second axial shoulder 27 of the flanged hub 20. The relief groove 80 is therefore axially outside the seat 28 and radially inside the entire flanged hub 20. The presence of the spacer 70 enables the relief groove 80 to be positioned sufficiently far away from the radial section SR of reduced thickness of the flanged hub 20 (a more critical section in terms of stress). This enables the relief groove 80 to be designed with greater freedom, for example with a sufficiently large radius R, but primarily does not concentrate stresses in the vicinity of the narrow section SR of the flanged hub 20.

[0039] Consequently, the inclusion of the spacer 70, which is used as an axial shoulder for the axially inner radial ball bearing 60, is primarily intended to enable the provision of a radially inner relief groove 80 that is different in shape and position compared to the relief groove that would be possible to provide without the spacer 70.

[0040] This resolves several structural criticalities. Indeed, if there were no spacer, the radially inner relief groove would coincide (be aligned) with the radial ball bearing 60 and would therefore be in a position axially facing the radially outer relief groove 29. Furthermore, if the radially inner relief groove coincides with the radial ball bearing, the shape therefore would be conditioned by the geometry of the radial ball bearing, with very limited axial length and radius (for example, axial length in the order of 2 mm and radius in the order of 0.8 mm). All of this creates a very reduced narrow section SR of the flanged hub, and a notch effect caused by the presence of the radially inner relief groove.

[0041] The flanged hub is subjected to high bending loads in this zone, which generate high stresses that have an adverse effect on the service life of the flanged hub itself.

[0042] Conversely, the inclusion of the spacer 70 provides numerous advantages. For one, the radially inner relief groove 80 is provided in a more axially outer position, which is therefore further away from the radially outer relief groove 29 and the narrow section SR. In addition, the geometry of the relief valve 80 is characterized by larger radii and a greater overall length, with a consequent reduction in the notch effect. When required by the application, it is also possible to include an axially inner second radial ball bearing beside the first ball bearing without modifying the flanged hub but by merely reducing the axial dimension of the spacer 70. The second radial ball bearing may be useful where one ball bearing is not enough to withstand the stresses coming from the constant velocity joint.

[0043] With reference to FIG. 4, one key geometric parameter for optimizing the relief groove 80 and eliminating the structural problems of the flanged hub is the geometry of the groove itself. The radius R of the relief groove 80 should preferably be between 1.8 mm and 2.2 mm, while the axial length L can preferably vary between 3.8 mm and 4.2 mm. A second geometric parameter is the axial position of the relief groove 80. The limit positions, both on the axially inner side and the axially outer side, can advantageously be identified to avoid structural criticalities. In particular, the axially inner limit position of the relief groove 80 could coincide with the projection on the axis X of rotation of the pressure center C between the rolling bodies 33 and the raceway 34 of the radially inner ring 34, since the forces coming from the external loads are mainly discharged in this zone.

[0044] Furthermore, the axially outer limit position of the relief groove 80 could be defined at a distance D of between 2.4 mm and 2.6 mm in an axially inward direction with respect to the shoulder 22 of the flanged hub 20 that forms the stop of the radially inner ring 34 on the flanged hub 20. This prevents the relief groove 80 from being aligned with the shoulder 22 and with the adjacent connection 22b of the flanged hub 20, so as not to generate increased stress at the radius RA of the connection 22b, as deduced from the results of the structural analyses carried out.

[0045] Consequently, the axial position of the relief groove 80 is inside the stretch RP defined by the two limit positions mentioned above and indicated by the double arrow in FIG. 3.

[0046] In short, the adoption of the ring 50 between the rolled edge 24 of the flanged hub 20 and the toothed sleeve 55 provides the a plurality of advantages. These include the rolled material being deformed and pressed against a continuous surface with no gaps (about the full 360 of the circumference). This prevents discontinuities in the material that could cause a concentration of stresses and potential starting points for cracks on the rolled edge 24, in particular in the vicinity of the bending point F from which the rolled edge 24 bends radially outwards. In addition, no deformations are induced on the radially inner ring 34 that could alter the raceway 34 thereof and/or modify the osculation thereof. Such deformations may also increase the fatigue stresses between the rolling bodies and the raceway and reduce the service life of the hub-bearing assembly. Also, as a result of the clearance fit between the ring 50 and the toothed sleeve 55, no radial expansion is induced in the external toothed profile 57 of the toothed sleeve 55, thereby avoiding any malfunction (failed or difficult coupling) between the toothed sleeve and the ring gear of the disconnection system during operation. Finally, there is a substantial reduction in stresses and consequently deformations of the seat of the axially inner radial ball bearing of the flanged hub.

[0047] The use of the spacer 70 together with the ring 50 provides further advantages. These include a greater robustness of the flanged hub 20 in the zone between the shoulder 27 and the rolled edge 24, i.e. in the most structurally stressed zone, improving the capacity thereof to withstand loads. Also, the zone of the relief groove does not require local thermal treatment, which has a positive impact on the process and resulting costs. Furthermore, the solution is flexible in that the length of the spacer can be modified to enable two axially inner radial ball bearings to be inserted without modifying the design of the flanged hub. Indeed, the decision to use one or two axially inner radial ball bearings (naturally in addition to the axially outer radial ball bearing) can be taken at any stage of development without resulting in design modifications to the flanged hub.

[0048] In addition to the embodiment of the disclosure as described above, it is to be understood that there are numerous other variants. It is also to be understood that the embodiments are solely exemplary and do not limit the scope of the disclosure, its applications, or its possible configurations. Indeed, although the above description enables the person skilled in the art to carry out the present disclosure according to at least one example embodiment thereof, many variants of the described components can also be used without thereby departing from the scope of the disclosure as defined in the attached claims, which should be understood literally and/or according to the legal equivalents thereof.