Abstract:
A hydrodynamic axial sliding bearing is disclosed. The bearing includes a substantially planar sliding surface. The bearing includes a bearing surface, wherein the bearing surface separated by a lubricating gap with at least one annularly arranged segment in at least two coaxial ring zones, and wherein each segment comprising a wedge surface and an engagement surface.

Description:
CROSS-REFERENCES TO RELATED APPLICATION 
     This application is a National Stage application which claims the benefit of German Patent Application No. DE 10 2011 077 771.7, filed Jun. 17, 2011, which is hereby incorporated by reference in its entirety. 
     TECHNICAL FIELD 
     The invention relates to a hydrodynamic axial bearing arrangement with a radially outer annual surface and a bearing surface that is separated or can be separated therefrom by a lubricating gap, with annularly arranged segments each comprising a wedge surface and an engagement surface. 
     BACKGROUND 
     Such axial bearings arrangements are known from US 2011/0038716 A1. With this known axial bearing arrangement, the segments are each arranged on both sides, that is on the front and the back of an axial bearing disc, which is axially arranged between two sliding surfaces that are moveable relative to each other, in an annular channel, which during the operation of the axial bearing arrangement is filled with hydraulic lubricant, which because of hydrodynamic effects between the wedge surfaces of the segments and the facing sliding surface of the shaft or of the housing is squashed and then forms a supporting trapped film between the engagement surface following the wedge surface and the sliding surface. This trapped film prevents a direct contact between regions of the axial bearing disc and the adjacent sliding surface of the shaft or of the housing. According to US 2011/0038716 A1, the segments are arranged on the front of the axial bearing disc offset in circumferential direction relative to the segments on the back of the axial bearing disc. This serves to reduce the wear of the tools during the production of the axial bearing disc. 
     In the case of axial bearings of turbochargers, it has to be taken into account that both the turbine wheel as well as the compressor wheel during charger operation can be excited into wobbling movements because of unavoidable imbalances and/or aerodynamic effects. These wobbling movements are transmitted to the shaft with the consequence that the shaft-sided sliding surfaces can wobble relative to the axial bearing disc. Thus, the gap dimensions between the shaft-side sliding surface and the segments changes. This can result in that the supporting force of the trapped film formed by the hydraulic lubricant undesirably drops between a segment and the adjacent shaft-sided sliding surface and can have a direct contact of the sliding surface with the segment as a consequence. 
     SUMMARY 
     It is therefore the object of the invention to create an axial bearing arrangement, wherein a hydrodynamic axial bearing is guaranteed even under extreme conditions. 
     With an axial bearing arrangement of the type stated at the outset, this object is solved in that the segments are arranged in at least two concentric ring zones. 
     Here it is provided, in particular, that the segments of adjacent ring zones are offset to one another in circumferential direction. 
     Through the segments being arranged in a plurality of zones and through their arrangement preferentially offset in circumferential direction a network of hydrodynamic supporting regions is created with the consequence that possible wobbling movements of the shaft-sided sliding surface are effectively dampened and a direct contact of the sliding surface with parts of the segments can be excluded. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       In the drawing it shows 
         FIG. 1  an axial view of an axial bearing disc configured according to the invention, and 
         FIG. 2  a further axial view of a modified embodiment. 
     
    
    
     DETAILED DESCRIPTION 
     The axial bearing disc  1  shown in  FIG. 1  has a central opening  2  in the usual manner, which with assembled axial bearing arrangement is penetrated by the shaft (not shown) to be axially mounted. Here, the face end of the axial bearing disc  1  visible in  FIG. 1  can face a shaft-sided collar with shaft-sided sliding surface, while the face end of the bearing disc  1  that is not visible in  FIG. 1  faces a relatively stationary sliding or bearing surface in particular on the housing side. It is to be understood that the aforementioned shaft is additionally mounted radially, wherein the corresponding radial bearings in the top view of  FIG. 1  can be arranged above the visible face end of the axial bearing disc  1 . 
     Radially between the central opening  2  and a radially outer annular surface  3  a segment zone  4  that is deepened relative to the ring surface  3  is provided, within which segments  5  are arranged in two ring zones which are concentric to the central opening  2 . These segments  5  each have an engagement surface  5 ′ and a wedge surface  5 ″. The engagement surfaces  5 ′ extend in a common radial plane, while the wedge surfaces  5 ″ rise ramp-like to the engagement surfaces  5 ′ in arrow direction P from a level that is lower relative to the radial plane of the engagement surfaces  5 ′. 
     During the rotational operation of the shaft, the segment zone  5  is filled with fluidic, in particular hydraulic lubricant. When the shaft with its collar now rotates relative to the axial bearing disc in the arrow direction P, lubricant is dragged along between the wedge surfaces  5 ″ and the facing shaft collar within the gap narrowing in arrow direction P, with the consequence that between the shaft collar and the engagement surfaces  5 ′, load-bearing trapped films of the lubricant form each between the shaft collar and the engagement surfaces  5 ′. 
     In the event that the shaft and thus the shaft collar should perform wobbling movements relative to the axial bearing disc  1  because of imbalances or other interference forces, this necessarily results in that the thicknesses of the trapped films assume different values. In the case of conventional axial bearing arrangements, wherein the segments  5  are each arranged in only a single ring zone, the wobbling movements can result in that the trapped film undershoots a permissible minimum thickness on one of the segments  5  and thus loses its load capacity so that a direct contact of a shaft part with a part of the axial bearing disc  1  is made possible. 
     With the invention, by contrast, the segments are provided in a plurality of ring zones. Surprisingly, this results in that an adequately thick and thus load-bearing trapped film is always available on a sufficient number of segments  5  and contacts between shaft and axial bearing disc are excluded. Otherwise, an improved damping of relative movements between shaft and axial bearing disc is achieved through the wide support base that is possible through the invention by way of the segments being arranged in a plurality of ring zones. 
     The offset arrangement of the segments in the two ring zones also evident from  FIG. 1  has also proved to be advantageous for avoiding direct contact between axial bearing elements. In radial direction, a wedge surface of a segment of the one ring zone each lies next to an engagement surface of a segment of the adjacent ring zone. Otherwise, the segments in adjacent ring zones are each arranged within a same circumferential angle, i.e. the segments in a radially outer ring zone have a correspondingly greater length in circumferential direction. With respect to the wedge surfaces, this results in correspondingly reduced wedge angles. 
       FIG. 2  shows a particularly preferred embodiment, wherein the segments  5  are arranged in three concentric ring zones. 
     Otherwise, the segments  5  of adjacent ring zones in both embodiments are arranged offset in circumferential direction relative to one another, for example in such a manner that the wedge surface of one segment is located radially adjacent to the engagement surface of the segment in the neighboring ring zone. This means at the same time that the segments of adjacent ring zones each have the same circular measure but have different lengths in circumferential direction. 
     The embodiments shown as drawings have proved to be advantageous. At the same time, the shown feature combinations are not compulsory. 
     Initially, the number of the segments  5  can be changed while maintaining the (shown) area conditions between the engagement and wedge surfaces  5 ′ and  5 ″, i.e. the total area of the engagement surfaces  5 ′ on the one hand and the total area of the wedge surfaces  5 ″ on the other hand remain unchanged and are merely divided over a different number of the segments  5 . Apart from this, the number of the segments  5  on the segment paths which are concentric to one another can differ, while the segments furthermore can also have different circular measures on different segment paths. In addition to this, it is also possible within one or a plurality of the segment paths to provide segments  5  with circular measures deviating from one another, wherein the area ratio between wedge and engagement area of a respective segment in the segment path can also be different. Because of this, the load bearing behavior can be optimized, specifically with minimal acoustic excitations or instability at the same time. The circular spacings between consecutive segments in circumferential direction can also be varied, if required such that the number of the wedge and engagement surfaces for each pitch circle is different. Finally, all function surfaces, in particular the wedge and engagement surfaces can be embodied with a defined surface structure, e.g. as honed, precision-turned or punched surfaces or such with cross-grind or profilings, such as longitudinal, transverse, diagonal grooves or angular or oval depressions or elevations. By doing so, the trapped film flow between the function surfaces of an axial bearing can be deliberately influenced with respect to a dampening of acoustic excitations or an increase of the stability.