US20140241887A1

US20140241887A1 - Hydrodynamic axial bearing

Info

Publication number: US20140241887A1
Application number: US14/268,466
Authority: US
Inventors: Peter Neuenschwander; Bruno Ammann; Marco Di Pietro; Markus Städeli
Original assignee: ABB Turbo Systems AG
Current assignee: Accelleron Industries AG
Priority date: 2011-11-03
Filing date: 2014-05-02
Publication date: 2014-08-28
Also published as: DE102011085681A1; CN103906936A; BR112014010582A2; HK1199084A1; KR20140083051A; CA2852164A1; EP2773877A1; SG11201401938WA; WO2013064638A1; JP2014533342A

Abstract

A hydrodynamic axial bearing for mounting a shaft which is mounted rotatably in a bearing housing, including an axial stop of the bearing housing and a bearing comb which rotates with the shaft. A lubricating gap, which is loaded with lubricating oil and is delimited by a profiled circular ring face and a sliding face, being formed between the axial stop and the bearing comb. The profiled circular ring face and the sliding face are configured in such a way that the lubricating gap is constricted radially to the outside with regard to the axial direction. As a result, temperature deformations which occur during operation and deformations on account of centrifugal, shearing and other forces in the bearing comb can be compensated for.

Description

RELATED APPLICATIONS

This application claims priority as a continuation application under 35 U.S.C. §120 to PCT/EP2012/071729, which was filed as an International Application on Nov. 2, 2012 designating the U.S., and which claims priority to German Application 102011085681.1 filed in Germany on Nov. 3, 2011. The entire contents of these applications are hereby incorporated by reference in their entireties.

FIELD

The disclosure relates to the field of hydrodynamic axial mounting of rotating shafts, as are used, for example, in turbomachines, for example in exhaust gas turbochargers.

BACKGROUND INFORMATION

If rapidly rotating rotors are loaded with axial shearing forces, load-bearing axial bearings can be used. For example, in the case of turbomachines, such as exhaust gas turbochargers, hydrodynamic axial bearings can be used to absorb axial forces, which can be high as a result of the flow, and to guide the shaft in an axial direction. In order to improve an oblique position compensation capability and wear behavior in applications of this type, disks which float freely in the lubricating oil, known as floating disks, can be used in hydrodynamic axial bearings between a bearing comb which rotates at the shaft rotational speed and a non-rotating axial stop on the bearing housing. The lubricating gaps between a rotating bearing comb and the floating disk and between the floating disk and the stationary axial stop on the bearing housing can be delimited in each case by a profiled circular ring face and a plane sliding face which lies opposite the profiled circular ring face. The profiled circular ring face can serve to optimize the pressure build-up in the lubricating gap. The pressure build-up can be decisive for the load-bearing force of the axial bearing. In order to distribute the lubricating oil which is supplied in the radially inner region of the profiled circular ring face, there are lubricating oil grooves which lead radially to the outside. Wedge faces, which constrict the lubricating gap in the circumferential direction and via which the lubricating oil introduced into the lubricating oil grooves exits, are formed adjacently with respect to the lubricating oil grooves. Here, the lubricating oil is guided into the wedge face as far as possible over the entire radial height of the lubricating oil grooves. The pressure build-up, which is desirable for the load-bearing capability of the axial bearing, takes place substantially in the region of the wedge faces. Rest faces, which include a planar face and are provided by the load-bearing face of the profiled circular ring face, are formed adjacently with respect to the wedge faces in the circumferential direction.
Examples of axial bearings of this type are found, inter alia, in GB 1095999, EP0840027, EP1199486, EP1644647 and EP2042753. The radial guidance of the floating disk takes place either on the rotating body, for example on the shaft or on the bearing comb, by way of a radial bearing which is integrated into the floating disk, as is disclosed, for example, in EP0840027, or else on a stationary bearing collar which surrounds the rotating body concentrically, as is disclosed, for example, in EP1199486. The lubrication of a hydrodynamic axial bearing of this type can take place lubricating oil from a dedicated lubricating oil system or, in the case of exhaust gas turbochargers, via the lubricating oil system of an internal combustion engine which is connected to the exhaust gas turbocharger.
In the cold state, at a standstill, all the load-bearing faces of known axial mountings can lie perpendicularly with respect to the rotational axis of the rotor or else at least parallel to one another. During operation, the load-bearing faces can be deformed on account of temperature gradients, centrifugal, shearing and other forces. A deformation of this type of the bearing load-bearing faces can impair the load-bearing force of the mounting. Temperature gradients over the comb of the comb bearing can have particularly great effects. The comb which protrudes radially with respect to the shaft can be deformed in an umbrella-shaped manner on account of the temperature difference between the load-bearing face and the rear side. This deformation can lead to rubbing of the comb bearing on the floating disk, particularly in the case of a low oil supply pressure. The deformation on account of the temperature gradient can be critical in a known comb bearing construction, because the deformation can cause a lubricating gap which widens to the outside. This can reduce the load-bearing capability for geometric reasons and can reduce the centrifugal force-induced pressure build-up in the radial direction, because the outflow resistance for the lubricating oil radially to the outside is reduced.

SUMMARY

A hydrodynamic axial bearing for mounting a shaft mounted rotatably in a bearing housing, the hydrodynamic axial bearing comprising: an axial stop; a bearing comb for rotation with a shaft when installed; and at least one lubricating gap formed between the axial stop and the bearing comb for receiving lubricating oil, and delimited by a profiled circular ring face and a planar sliding face which lies opposite the circular ring face, the profiled circular ring face being configured so as to rotate around or with the shaft, the profile of the circular ring face having a plurality of segments, each segment including one radially running lubricating oil groove, a wedge face connected to the lubricating oil groove in a circumferential direction, and a rest face which adjoins the wedge face in the circumferential direction, wherein, for the at least one lubricating gap, the rest face and the planar sliding face are configured such that the lubricating gap, delimited by the rest face and the planar sliding face, is constricted radially to the outside with regard to an axis of rotation.
A turbomachine is disclosed, comprising: a shaft mounted rotatably in a bearing housing; and a hydrodynamic axial bearing for mounting the shaft in the bearing housing, the hydrodynamic axial bearing including: an axial stop; a bearing comb for rotating with the shaft; and at least one lubricating gap formed between the axial stop and the bearing comb for being loaded with lubricating oil and delimited by a profiled circular ring face and a planar sliding face which lies opposite the circular ring face, the profiled circular ring face being configured so as to rotate around or with the shaft, the profile of the circular ring face having a plurality of segments, each segment including one radially running lubricating oil groove, a wedge face connected to the lubricating oil groove in a circumferential direction, and a rest face which adjoins the wedge face in the circumferential direction, wherein, for the at least one lubricating gap, the rest face and the planar sliding face are configured such that the lubricating gap, delimited by the rest face and the planar sliding face, is constricted radially to the outside with regard to an axis of rotation.
An exhaust gas turbocharger is disclosed, comprising: a shaft mounted rotatably in a bearing housing; and a hydrodynamic axial bearing for mounting the shaft in the bearing housing the hydrodynamic axial bearing including: an axial stop; a bearing comb for rotating with the shaft; and at least one lubricating gap formed between the axial stop and the bearing comb for being loaded with lubricating oil and delimited by a profiled circular ring face and a planar sliding face which lies opposite the circular ring face, the profiled circular ring face being configured so as to rotate around or with the shaft, the profile of the circular ring face having a plurality of segments, each segment including one radially running lubricating oil groove, a wedge face connected to the lubricating oil groove in a circumferential direction, and a rest face which adjoins the wedge face in the circumferential direction, wherein, for the at least one lubricating gap, the rest face and the planar sliding face are configured such that the lubricating gap, delimited by the rest face and the planar sliding face, is constricted radially to the outside with regard to an axis of rotation.

BRIEF DESCRIPTION OF THE DRAWINGS

In the following text, exemplary embodiments of the disclosure will be explained in detail using drawings, in which:

FIG. 1 shows, in the right-hand part, a section which is guided along the rotational axis of a known axial sliding bearing with a rotating bearing comb, a stationary axial stop and a floating disk, and shows, in the left-hand part, a frontal view in the axial direction of the corresponding floating disk with a profiled circular ring face;

FIG. 2 shows a diagrammatically illustrated axial sliding bearing according to FIG. 1, the bearing comb being shown in each case in the cold state in this figure and in all following figures, and additionally the deformation of the bearing comb in the operating state on account of the heating and the rapid rotation and the resulting lubricating gap being indicated by way of dashed lines;

FIG. 3 shows a diagrammatically illustrated axial sliding bearing according to a first exemplary embodiment according to the disclosure, with a conically shaped bearing comb and a lubricating gap which results therefrom and tapers radially toward the outside;

FIG. 4 shows a diagrammatically illustrated axial sliding bearing according to a second exemplary embodiment according to the disclosure with a floating disk which is shaped conically on the bearing comb side and a lubricating gap which results therefrom and tapers radially toward the outside;

FIG. 5 shows a diagrammatically illustrated axial sliding bearing according to a third exemplary embodiment according to the disclosure, with a conically shaped axial bearing and conically shaped bearing comb, and two lubricating gaps which result therefrom and taper radially toward the outside;

FIG. 6 shows a diagrammatically illustrated axial sliding bearing according to a fourth exemplary embodiment according to the disclosure with a conically shaped axial bearing and a floating disk which is shaped conically on the bearing comb side, and two lubricating gaps which result therefrom and taper radially toward the outside;

FIG. 7 shows a diagrammatically illustrated axial sliding bearing according to a fifth exemplary embodiment according to the disclosure with a floating disk which is shaped conically on both sides, and two lubricating gaps which result therefrom and taper radially toward the outside;

FIG. 8 shows a diagrammatically illustrated axial sliding bearing according to a sixth exemplary embodiment according to the disclosure with a conically shaped bearing comb and a floating disk which is shaped conically on the axial bearing side, and two lubricating gaps which result therefrom and taper radially toward the outside;

FIG. 9 shows a diagrammatically illustrated axial sliding bearing according to a seventh exemplary embodiment according to the disclosure, without a floating disk, with a conically shaped bearing comb, and a lubricating gap which results therefrom and tapers radially toward the outside; and

FIG. 10 shows a diagrammatically illustrated axial sliding bearing according to an eighth exemplary embodiment according to the disclosure, without a floating disk, with a conically shaped axial stop, and a lubricating gap which results therefrom and tapers radially toward the outside.

DETAILED DESCRIPTION

Exemplary embodiments of the disclosure can improve the load-bearing capability of a hydrodynamic axial bearing for mounting a shaft which is mounted rotatably in a bearing housing.
If the gap, which is formed between the load-bearing faces of the axial bearing, is configured so as to be constricted to the outside in the radial direction, by the load-bearing faces being arranged obliquely relative to one another at least in the radially outer region, a reduction in the relative oblique position of the load-bearing faces results during operation on account of the abovementioned deformation of the rotating load-bearing face. The constriction in the radially outer region is reduced, with the result that the load-bearing faces can rest more uniformly on one another during operation.
If, for example, the bearing comb is manufactured with a conical load-bearing face, that is to say a load-bearing face which is inclined toward the load-bearing face which lies opposite it, the temperature deformation in the comb bearing can be compensated for. During the compensation, the deformations on account of centrifugal, shearing and further forces likewise have to be taken into consideration.
Because the comb bearing deformations are dependent on the operating point, the lubricating gap becomes smaller in the radial direction under certain operating conditions. This situation is more favorable than the current one with a widened lubricating gap, because the load-bearing capability is reduced to a lesser extent and the centrifugal force-induced pressure build-up in the radial direction is aided.
The compensation on account of load-bearing face deformations as a result of temperature gradients, centrifugal, shearing and further forces can also take place at the floating disk, or at the axial stop of the bearing housing in the case of an axial bearing without floating disk. Any temperature-induced deformations which occur in the region of the axial stop on the bearing housing can be carried out in a similar way as on the comb bearing.
If a floating disk, which is conical on both sides, or a very thin floating disk which is adapted to changing geometric conditions during operation, is used, the comb bearing deformation can also be compensated for by way of a conical configuration of the axial stop on the bearing housing.
Due to the compensation for the deformation, the axial mounting can become more robust at the adjacent bearing parts with respect to rubbing of the floating disk or the bearing comb, or, in the case of an axial bearing without floating disk, of the axial bearing. The turbocharger can become more operationally reliable and wear-induced costs can be reduced.
FIG. 1 shows by way of example a known hydrodynamic axial bearing, the three components of the axial bearing being made visible in the right-hand part of the figure in a section which is guided axially along the rotational shaft. The bearing comb 10 is placed on the rotating shaft 40, or is optionally connected in a material-to-material manner to the shaft, or is produced with the shaft from one piece, and rotates with the shaft. A floating disk 30 is arranged axially between an axial stop 21 on the bearing housing 20 and the bearing comb. In each case one lubricating gap is formed firstly between the axial stop and the floating disk and secondly between the floating disk and the bearing comb, in which lubricating gap a thin lubricating oil layer is situated between the load-bearing faces.
In an exemplary embodiment according to the disclosure, the load-bearing face 22 on the axial stop and the load-bearing face 11 on the bearing comb in each case have a sliding face which is of planar configuration in the circumferential direction, whereas the two load-bearing faces of the floating disk are part of a profiled circular ring face. This basic construction of the two lubricating gaps is also adopted in all the exemplary embodiments described in the disclosure of the hydrodynamic axial sliding bearings according to the exemplary embodiment of the disclosure with a floating disk. It applies to all these exemplary embodiments that the sliding faces and the profiled circular ring faces in the case of one or both of the lubricating gaps can also be arranged on the respectively other side of the lubricating gap, with the result that, for example, the floating disk has in each case one planar sliding face on both sides whereas the profiled circular ring face is attached on the load-bearing face of the bearing comb and the axial stop of the bearing housing. In an exemplary embodiment without a floating disk, the profiled circular ring face would correspondingly be arranged on the rotating bearing comb and the planar sliding face would be arranged on the axial stop of the bearing housing or at any rate vice versa, that is to say the planar sliding face on the rotating bearing comb and the profiled circular ring face on the axial stop of the bearing housing.
The construction of the profiled circular ring face can be seen from the left-hand part of FIG. 1, in which the floating disk is rotated by 90°, with the result that one of the end sides of the floating disk can be seen in a plan view.
The profiled circular ring face can optimize the pressure build-up in the lubricating gap between the load-bearing faces, which pressure build-up can be decisive for the load-bearing force of the axial bearing. The profiling of the circular ring face includes a plurality of segments with, in each case, one lubricating oil groove 33 which is led radially to the outside in order to distribute the lubricating oil which is supplied in the radially inner region of the profiled circular ring face. Counter to the rotational direction (indicated by way of the black arrow) of the profiled circular ring face, wedge faces 34 which constrict the lubricating gap in the circumferential direction are formed adjacently with respect to the lubricating oil grooves 33, via which wedge faces 34 the lubricating oil which is introduced into the lubricating oil grooves 33 exits in accordance with the thick arrows. Here, the lubricating oil is guided into the wedge face 34 as far as possible over the entire radial height of the lubricating oil grooves 33. The pressure build-up, which is desirable for the load-bearing capability of the axial bearing, takes place substantially in the region of the wedge faces. Rest faces 35 are formed adjacently with respect to the wedge faces 34 in the circumferential direction, which rest faces 35 include a planar face which is at the smallest spacing from the corresponding contact, as the above-described sliding face. The axial extent (thickness) of the lubricating gap can therefore be described as the spacing between the rest faces 35 and the sliding face which lies opposite. In order to optimize the pressure build-up in the radial direction in the lubricating oil groove and over the wedge faces, the lubricating oil groove and wedge face can be closed radially to the outside by way of a web which constricts the lubricating gap. Here, the web can come to lie as far as the height of the rest face, with the result that the rest face and web lie in one plane.
The configuration of the lubricating oil groove and the wedge face is disregarded for the exemplary embodiments which are described in the disclosure. Accordingly, the expressions of the profiled circular ring face and the sliding face are no longer used in the disclosure. For the practical implementation, however, reference is made to the fact that the lubricating gaps, as described above, are advantageously delimited in each case by a profiled circular ring face and a planar sliding face. The expression used in the following text of the active load-bearing face means that region of the profiled circular ring face which can be called a rest face. The rest faces can be situated so as to adjoin the wedge faces as viewed in the flow direction of the lubricating oil.
As indicated in FIG. 1 and in the detail which is shown on an enlarged scale according to FIG. 2, in the cold state, that is to say at a standstill of the rotor, the load-bearing faces of the axial mountings are configured perpendicularly with respect to the rotational axis of the rotor or else at least parallel to one another. During operation, the load-bearing face in the bearing comb can be deformed on account of temperature gradients, centrifugal, shearing and further forces. The comb which protrudes radially with respect to the shaft can be deformed in an umbrella-shaped manner on account of the temperature difference between the load-bearing face, which is relevant for the axial bearing, and the rear side which faces away from the load-bearing face. This deformation (indicated in FIG. 2 by way of dashed lines) can lead to rubbing of the comb bearing on the floating disk in the radially inner region, because the load-bearing force of the lubricating gap diminishes on account of the radially outwardly diverging load-bearing faces 31 and 11′ of the axial bearing and the associated unimpeded escape of the lubricating oil, for example in the case of a low oil supply pressure, at which sufficient lubricating oil cannot be replenished.
FIG. 3 shows a diagrammatically illustrated hydrodynamic axial sliding bearing according to a first exemplary embodiment according to the disclosure. Here, the active load-bearing face 31 on that side of the floating disk 30 which faces the bearing comb is oriented strictly radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40. In contrast, the load-bearing face 11 of the bearing comb is shaped so as to be inclined toward the floating disk 30, which results in a constriction in the axial direction in the radially outer region of the lubricating gap 52. In this exemplary embodiment, just as in the further exemplary embodiments which will be described in the following text, the inclination of the load-bearing face 11 of the bearing comb can be realized by way of a uniform, straight inclination or by way of a curved inclination. In the figures, the deformations of the rotating components and the constrictions of the lubricating gaps are illustrated in a greatly exaggerated manner. In fact, the inclination angles which are provided according to the disclosure move over the entire radius of the inclined component in the range of a few hundredths of a degree, which results in a constriction of the lubricating gap at the radially outer edge of a few hundredths of a millimeter in the case of a disk with a diameter of 200 mm.
During operation, a deformation of the bearing comb, indicated by way of dashed lines, can result on account of the above-described heating of the bearing comb and as a result of the action of the stated forces. According to the exemplary embodiment of the disclosure, the load-bearing face 11, which is inclined towards the floating disk in the cold state, of the bearing comb stretches in such a way that the angle of the constriction of the lubricating gap 52′ is reduced during nominal operation and the two load-bearing faces 31 and 11′ of the bearing run parallel to one another or, while maintaining a lubricating gap constriction which is less pronounced than in the cold state, run at least virtually parallel to one another. In the cold state, i.e., at a standstill and also at small rotational speeds, the configuration according to the exemplary embodiment of the disclosure of the axial sliding bearing, leads to a constriction of the lubricating gap in the radial outer region. This is not a problem, because the accumulated lubricating oil ensures an additional pressure build-up.
FIG. 4 shows a diagrammatically illustrated hydrodynamic axial sliding bearing according to a second exemplary embodiment according to the disclosure. Here, the load-bearing face 11 of the bearing comb is oriented strictly radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40. For this purpose, the load-bearing face 31 on that side of the floating disk 30 which faces the bearing comb is configured so as to be inclined toward the bearing comb 10 in this exemplary embodiment, which results in the constriction in the axial direction in the radially outer region of the lubricating gap 52. The floating disk is therefore of conical configuration on the side which faces the bearing comb, whereas it is oriented perpendicularly with respect to the rotational axis of the shaft 40 on the other side which faces the axial stop on the bearing housing.
During operation, a deformation of the bearing comb, indicated by way of dashed lines, results on account of the above-described heating of the bearing comb and as a result of the action of the stated forces. The load-bearing face 11, which is oriented perpendicularly with respect to the rotational axis of the shaft 40 in the cold state, of the bearing comb is bent in such a way that, during nominal operation, the angle of the constriction of the lubricating gap 52′ is reduced and the two load-bearing faces 31 and 11′ of the bearing run parallel or virtually parallel to one another.
In the exemplary embodiments of the disclosure according to FIGS. 5 to 8, in addition to the lubricating gap 52 between the floating disk 30 and the bearing comb 10, the lubricating gap 51 between the axial stop 21 and the floating disk 30 can also be configured with a constriction in the axial direction in the radially outer region.
FIG. 5 shows a diagrammatically illustrated hydrodynamic axial sliding bearing according to a third exemplary embodiment according to the disclosure. Here, the load-bearing face 31 is oriented radially on that side of the floating disk 30 which faces the bearing comb, that is to say perpendicularly with respect to the rotational axis of the shaft 40. In contrast, the load-bearing face 11 of the bearing comb is shaped such that it is inclined toward the floating disk 30, which results in a constriction in the axial direction in the radially outer region of the lubricating gap 52. The second lubricating gap which is likewise provided with a constriction in the axial direction in the radially outer region extends between the load-bearing face 32 which is oriented radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40, on that side of the floating disk 30 which faces the axial stop and the load-bearing face 22 of the axial stop 21 on the bearing housing, which load-bearing face 22 is inclined toward the floating disk 30. The floating disk is therefore provided with two sides which run parallel to one another and are oriented perpendicularly with respect to the rotational axis of the shaft 40.
During operation, a deformation of the bearing comb 10, indicated by way of dashed lines, results on account of the above-described heating of the bearing comb and as a result of the action of the stated forces. According to the disclosure, the load-bearing face 11, which is inclined toward the floating disk in the cold state, of the bearing comb, stretches in such a way that, during nominal operation, the angle of the constriction of the lubricating gap 52′ is reduced and the two load-bearing faces 31 and 11′ of the bearing run parallel to one another or virtually parallel to one another.
FIG. 6 shows a diagrammatically illustrated hydrodynamic axial sliding bearing according to a fourth exemplary embodiment according to the disclosure, which hydrodynamic axial sliding bearing differs from the preceding one in that the load-bearing face 11 of the bearing comb is oriented radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40. For this purpose, the load-bearing face 31 is configured so as to be inclined toward the bearing comb 10 on that side of the floating disk 30 which faces the bearing comb. The second lubricating gap, which is likewise provided with a constriction in the axial direction in the radially outer region, extends between the load-bearing face 32 which is oriented strictly radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40, on that side of the floating disk which faces the axial stop and the load-bearing face 22 of the axial stop 21 on the bearing housing, which load-bearing face 22 is inclined toward the floating disk 30. The floating disk is therefore of conical configuration on the side which faces the bearing comb, whereas it is oriented perpendicularly with respect to the rotational axis of the shaft 40 on the other side which faces the axial stop on the bearing housing.
During operation, a deformation of the bearing comb, indicated by way of dashed lines, results on account of the above-described heating of the bearing comb and as a result of the action of the stated forces. According to the disclosure, the load-bearing face 11 of the bearing comb, which is oriented perpendicularly with respect to the rotational axis of the shaft 40 in the cold state, bends in such a way that, during nominal operation, the angle of the constriction of the lubricating gap 52′ is reduced and the two load-bearing faces 31 and 11′ of the bearing run parallel to one another or virtually parallel to one another.
FIG. 7 shows a diagrammatically illustrated hydrodynamic axial sliding bearing according to a fifth exemplary embodiment according to the disclosure. Here, the load-bearing face 11 of the bearing comb is oriented radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40. In contrast, on that side of the floating disk 30 which faces the bearing comb, the load-bearing face 31 is configured so as to be inclined toward the bearing comb 10, which results in a constriction in the axial direction in the radially outer region of the lubricating gap 52. The second lubricating gap, which is likewise provided with a constriction in the axial direction in the radially outer region, extends between the load-bearing face, which is oriented strictly radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40, of the axial stop 21 on the bearing housing and the load-bearing face 32 which is inclined toward the axial stop on that side of the floating disk which faces the axial stop. The floating disk 30 is therefore configured so as to be conical on both sides.
During operation, a deformation of the bearing comb 10 which is once again indicated by way of dashed lines, results on account of the above-described heating of the bearing comb and as a result of the action of the stated forces. According to the disclosure, the load-bearing face 11, which is oriented perpendicularly with respect to the rotational axis of the shaft 40 in the cold state, of the bearing comb bends in such a way that, during nominal operation, the angle of the constriction of the lubricating gap 52′ is reduced and the two load-bearing faces 31 and 11′ of the bearing run parallel to one another or virtually parallel to one another.
FIG. 8 shows a diagrammatically illustrated hydrodynamic axial sliding bearing according to a sixth exemplary embodiment according to the disclosure, which hydrodynamic axial sliding bearing differs from the preceding one in that the load-bearing face 31 on that side of the floating disk 30 which faces the bearing comb is oriented radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40. For this purpose, the load-bearing face 11 of the bearing comb is shaped so as to be inclined toward the floating disk 30, which results once again in a constriction in the axial direction in the radially outer region of the lubricating gap 52. The second lubricating gap which is likewise provided with a constriction in the axial direction in the radially outer region once again extends between the load-bearing face 22, which is oriented radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40, of the axial stop 21 on the bearing housing and the load-bearing face 32 which is inclined toward the axial stop on that side of the floating disk which faces the axial stop. The floating disk is therefore of conical configuration on the side which faces the axial stop on the bearing housing, whereas it is oriented perpendicularly with respect to the rotational axis of the shaft 40 on the other side which faces the bearing comb.
During operation, a deformation of the bearing comb, indicated by way of dashed lines, results on account of the above-described heating of the bearing comb and as a result of the action of the stated forces. According to the disclosure, the load-bearing face 11, which is inclined toward the floating disk in the cold state, of the bearing comb stretches in such a way that, during nominal operation, the angle of the constriction of the lubricating gap 52′ is reduced and the two load-bearing faces 31 and 11′ of the bearing run parallel to one another or virtually parallel to one another.
FIG. 9 and FIG. 10, in each case show a hydrodynamic axial sliding bearing without a floating disk, in which a load-bearing face 12 is arranged on the rotating bearing comb 10 and a load-bearing face 22 is arranged on the axial stop 21 of the bearing housing 20. According to the disclosure, the lubricating gap 53 which results between them is configured so as to converge radially to the outside, that is to say the lubricating gap tapers in the radially outer region.
The seventh exemplary embodiment according to the disclosure (shown in FIG. 9) of a hydrodynamic axial sliding bearing has a load-bearing face 12 of the bearing comb 10, which load-bearing face 12 is shaped so as to be inclined toward the axial stop 21 of the bearing housing 20, which results in the constriction in the axial direction in the radially outer region of the lubricating gap 53. The load-bearing face 22 of the axial stop 21 of the bearing housing 20 is oriented strictly radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40, in this exemplary embodiment.
During operation, a deformation of the bearing comb, indicated by way of dashed lines, results on account of the above-described heating of the bearing comb and as a result of the action of the stated forces. According to the disclosure, the load-bearing face 12 of the bearing comb, which is inclined toward the load-bearing face of the axial stop 21 in the cold state, stretches in such a way that, during nominal operation, the angle of the constriction of the lubricating gap 53′ is reduced and the two load-bearing faces 12′ and 22 of the bearing run parallel to one another or virtually parallel to one another.
The eighth exemplary embodiment according to the disclosure (shown in FIG. 10) of a hydrodynamic axial sliding bearing has a load-bearing face 12 of the bearing comb 10, which load-bearing face 12 is oriented radially, that is to say perpendicularly with respect to the rotational axis of the shaft 40. For this purpose, in this exemplary embodiment, the load-bearing face 22 of the axial stop 21 on the bearing housing 20 is configured so as to be inclined toward the bearing comb 10, which results once again in the constriction in the axial direction in the radially outer region of the lubricating gap 53. The axial stop is therefore of conical configuration on the side which faces the bearing comb.
During operation, a deformation of the bearing comb, indicated by way of dashed lines, results on account of the above-described heating of the bearing comb and as a result of the action of the stated forces. According to the disclosure, the load-bearing face 12 of the bearing comb 10, which is oriented perpendicularly with respect to the rotational axis of the shaft 40 in the cold state, bends in such a way that, during nominal operation, the angle of the constriction of the lubricating gap 53′ is reduced and the two load-bearing faces 12′ and 22 of the bearing run parallel to one another or virtually parallel to one another.
In all the exemplary embodiments, in each case one of the load-bearing faces is described as deviating from the plane which is oriented perpendicularly with respect to the rotational axis of the shaft and the other load-bearing face is described as running radially, that is to say along a plane which is oriented perpendicularly with respect to the rotational axis of the shaft. According to the disclosure, the narrowing lubricating gaps can also be realized by the respective load-bearing faces both deviating from respective planes which are oriented perpendicularly with respect to the rotational axis of the shaft, but being at an angle with respect to one another. For example, in the exemplary embodiment with a floating disk, both the load-bearing face on that side of the floating disk which faces the bearing comb and the load-bearing face on the bearing comb can run so as to be inclined toward the lubricating gap in comparison with the plane which is oriented perpendicularly with respect to the rotational axis of the shaft, and can thus delimit the narrowing lubricating gap.
Even if in each case only load-bearing faces were mentioned in all the above-mentioned exemplary embodiments, it is to be noted once again here that, if one or both of the components which delimit a respective lubricating gap have a profiled surface with a lubricating oil groove, wedge faces and rest faces, the expression load-bearing face means in each case that region of the profiled surface which is called rest face. In the absence of a rest face, the load-bearing face extends along the maximum elevation of the wedge faces in the transition region to the respectively next lubricating oil groove.
Thus, it will be appreciated by those skilled in the art that the present disclosure can be embodied in other specific forms without departing from the spirit or essential characteristics thereof. The presently disclosed embodiments are therefore considered in all respects to be illustrative and not restricted.

LIST OF DESIGNATIONS

10 Bearing comb
11, 12 Load-bearing face on the bearing comb
11″, 12″ Load-bearing face on the bearing comb (in the operating state)
20 Bearing housing
21 Axial stop
22 Sliding face
30 Floating disk
31, 32 Load-bearing face of the floating disk
33 Lubricating oil groove
34 Wedge face
35 Rest face
40 Shaft
51 Lubricating gap between the axial stop and the floating disk
52 Lubricating gap between the floating disk and the bearing comb
52″ Lubricating gap between the floating disk and the bearing comb (in the operating state)
53 Lubricating gap between the axial stop and the bearing comb
53″ Lubricating gap between the axial stop and the bearing comb (in the operating state)

Claims

What is claimed is:

1. A hydrodynamic axial bearing for mounting a shaft mounted rotatably in a bearing housing, the hydrodynamic axial bearing comprising:

an axial stop;

a bearing comb for rotation with a shaft when installed; and

at least one lubricating gap formed between the axial stop and the bearing comb for receiving lubricating oil, and delimited by a profiled circular ring face and a planar sliding face which lies opposite the circular ring face, the profiled circular ring face being configured so as to rotate around or with the shaft, the profile of the circular ring face having a plurality of segments, each segment including one radially running lubricating oil groove, a wedge face connected to the lubricating oil groove in a circumferential direction, and a rest face which adjoins the wedge face in the circumferential direction, wherein, for the at least one lubricating gap, the rest face and the planar sliding face are configured such that the lubricating gap, delimited by the rest face and the planar sliding face, is constricted radially to the outside with regard to an axis of rotation.

2. The hydrodynamic axial bearing as claimed in claim 1, wherein the axial stop comprises:

the planar sliding face, the planar sliding face delimiting the lubricating gap constricted radially to the outside, the planar sliding face being configured to be inclined toward the bearing comb at least in a radially outer part in a manner which deviates from a plane which lies perpendicularly with respect to the axis of rotation.

3. The hydrodynamic axial bearing as claimed in claim 1, wherein the bearing comb comprises:

the planar sliding face, the planar sliding face delimiting the lubricating gap constricted radially to the outside, the planar sliding face being configured to be inclined toward the axial stop in a manner which deviates from a plane which lies perpendicularly with respect to the axis of rotation.

4. The hydrodynamic axial bearing as claimed in claim 1, comprising:

a floating disk arranged axially between the axial stop and the bearing comb, the lubricating gap being formed between the floating disk and the bearing comb.

5. The hydrodynamic axial bearing as claimed in claim 2, comprising:

6. The hydrodynamic axial bearing as claimed in claim 3, comprising:

7. The hydrodynamic axial bearing as claimed in claim 6, wherein the floating disk comprises:

the profiled circular ring face, and the planar sliding face of the bearing comb and the circular ring face of the floating disk delimit the lubricating gap which is constricted radially to the outside, the planar sliding face of the bearing comb being configured to be inclined toward the axial stop, at least in a radially outer part, in a manner which deviates from the plane which lies perpendicularly with respect to the axis of rotation.

8. The hydrodynamic axial bearing as claimed in claim 7, comprising:

a second lubricating gap delimited by a planar sliding face of the axial stop and the floating disk, the planar sliding face of the axial stop being configured to be inclined toward the floating disk in a manner which deviates from the plane which lies perpendicularly with respect to the axis of rotation.

9. A turbomachine, comprising:

a shaft mounted rotatably in a bearing housing; and

a hydrodynamic axial bearing for mounting the shaft in the bearing housing, the hydrodynamic axial bearing including:

an axial stop;

a bearing comb for rotating with the shaft; and

at least one lubricating gap formed between the axial stop and the bearing comb for being loaded with lubricating oil and delimited by a profiled circular ring face and a planar sliding face which lies opposite the circular ring face, the profiled circular ring face being configured so as to rotate around or with the shaft, the profile of the circular ring face having a plurality of segments, each segment including one radially running lubricating oil groove, a wedge face connected to the lubricating oil groove in a circumferential direction, and a rest face which adjoins the wedge face in the circumferential direction, wherein, for the at least one lubricating gap, the rest face and the planar sliding face are configured such that the lubricating gap, delimited by the rest face and the planar sliding face, is constricted radially to the outside with regard to an axis of rotation.

10. A turbomachine as claimed in claim 9, wherein the axial stop comprises:

11. A turbomachine as claimed in claim 9, wherein the bearing comb comprises:

the planar sliding face, the planar sliding face delimiting the lubricating gap constricted radially to the outside, the planar sliding face configured to be inclined toward the axial stop in a manner which deviates from a plane which lies perpendicularly with respect to the axis of rotation.

12. A turbomachine as claimed in claim 9, comprising:

13. A turbomachine as claimed in claim 10, comprising:

14. A turbomachine as claimed in claim 11, comprising:

15. A turbomachine as claimed in claim 14, wherein the floating disk comprises:

16. A turbomachine as claimed in claim 15, a second lubricating gap delimited by a planar sliding face of the axial stop comprising:

the floating disk, the planar sliding face of the axial stop being configured to be inclined toward the floating disk in a manner which deviates from the plane which lies perpendicularly with respect to the axis of rotation.

17. An exhaust gas turbocharger, comprising:

a shaft mounted rotatably in a bearing housing; and

a hydrodynamic axial bearing for mounting the shaft in the bearing housing the hydrodynamic axial bearing including:

an axial stop;

a bearing comb for rotating with the shaft; and

18. The exhaust gas turbocharger as claimed in claim 17, wherein the bearing comb and the shaft are connected in a material-to-material manner or are manufactured from one piece.