DE102009054341A1 - Bearing surface for fluid dynamic axial bearing, has bearing grooves of fluid-dynamic axial-bearing, where bearing surface is assigned to another bearing surface - Google Patents

Abstract

The bearing surface has bearing grooves of a fluid-dynamic axial-bearing (26). The bearing surface is assigned to another bearing surface. A bearing gap (20) is arranged between the both bearing surfaces. An axial fixture is provided between the bearing surfaces, where the depth of the bearing grooves is calculated. An independent claim is also included for a fluid dynamic axial bearing, which has two bearing components.

Description

Field of the invention

The invention relates to a bearing surface with bearing grooves of a fluid dynamic bearing and a fluid dynamic bearing with such provided with bearing grooves bearing surface, according to the features of the independent claims.

State of the art

Bearing surfaces with bearing grooves of the type described are used for example for fluid dynamic bearings, which in turn are used, for example, for pivotal mounting of spindle motors, fans or pumps. A fluid dynamic bearing comprises at least two bearing components, which are movable relative to one another and are preferably rotatable, which are separated from one another by a very narrow bearing gap filled with a bearing fluid. By a fluid dynamic effect, which generates a pressure build-up in the bearing fluid within the bearing gap during operation of the bearing, the bearing surfaces are kept at a distance and the bearing becomes sustainable. This fluid dynamic effect is generated by the bearing grooves, which are arranged on one or both of the mutually facing bearing surfaces. These bearing grooves generate a pumping action on the bearing fluid during operation of the bearing and thus a pressure build-up in the bearing gap. Both fluid-dynamic thrust bearings and radial bearings are used.

Fluid dynamic bearings generate their bearing effect only during operation of the bearing, that is, only when the two bearing surfaces move relative to each other. When the bearing is stationary, the bearing surfaces lie against each other. When the bearing starts or runs down, the bearing surfaces lie against each other until the bearing has built up sufficient bearing capacity and the resulting hydrodynamic effect separates the bearing surfaces. When starting or running out of the bearing, therefore, there is always a certain amount of material abrasion on the bearing surfaces and corresponding wear. The material abrasion, usually in the form of metal particles, is absorbed by the bearing fluid and can permanently affect the performance of the fluid dynamic bearing and shorten its life.

If such a fluid-dynamic bearing system is used for the rotary mounting of a spindle motor, then the mass of the rotor - depending on the installation position - to a large extent on the bearing surfaces of the one or one of the two thrust bearings. In the area of the bearing surfaces of the thrust bearing, therefore, there is a considerable abrasion of material when starting and stopping the engine.

Disclosure of the invention

It is the object of the invention to improve the design of the bearing surfaces of a fluid dynamic bearing to the effect that the friction between the bearing surfaces and thus wear of the bearing surfaces at low speeds, in particular during start-up and leakage of the bearing can be reduced.

This object is achieved by a bearing surface with bearing grooves according to the features of claim 1. A spindle motor having a bearing surface according to the invention with bearing grooves is specified in independent claim 6.

Preferred embodiments and advantageous features of the invention are set forth in the dependent claims.

The bearing surface according to the invention comprises incorporated in the bearing surface bearing grooves and is part of a fluid dynamic thrust bearing. This bearing surface is associated with a different bearing surface, wherein between the two bearing surfaces arranged with a bearing fluid bearing gap is arranged. The bearing grooves have a defined depth. According to the invention, the depth of the bearing grooves is 8 microns +/- 2 microns, with an axial clearance of 30 microns +/- 10 microns between the bearing surfaces.

Such a thrust bearing can be used for example in a spindle motor, which in turn can be used to drive a 2.5-inch hard disk drive. Heretofore, in bearing surfaces of such a thrust bearing of a spindle motor of the prior art, it has been usual to provide a depth t of the bearing grooves of 17 microns or more, the axial play of the bearing being 35 microns +/- 10 microns.

According to the invention, the depth t of the bearing grooves of an axial bearing of the above-mentioned size is reduced to 8 micrometers +/- 2 micrometers, whereby according to the invention a lower lift-off speed results between the two bearing surfaces of the axial bearing. The reason for this is the relatively shallow bearing grooves, which generate a sufficiently high axial force even at lower rotational speeds of the bearing surfaces, so that the bearing surfaces stand out from each other. As a result, the contact time of the bearing surfaces is reduced when starting and running of the thrust bearing and the mechanical abrasion significantly reduced.

Due to the difference between a maximum and a minimum axial distance of the two bearing surfaces of the thrust bearing axial play s is determined. In particular, according to the invention Ratio between the axial play s of the thrust bearing and the depth t of the bearing grooves chosen larger and is between s / t = 3.33 and 4.0. In known from the prior art thrust bearings of the same size, this ratio was for example between s / t = 1.67 and 2.37.

In a preferred embodiment of the invention, a plurality of bearing grooves are arranged in the same geometric orientation and at a distance from one another on the axial bearing surface. Here, the bearing grooves may be arranged spirally on the bearing surface, or alternatively, depending on the application and design of the thrust bearing, the bearing grooves may also be herringbone. These geometries of the bearing grooves are basically known from the prior art. Compared to the prior art, however, the depth of the bearing grooves is chosen to be much smaller according to the invention.

Due to the shallower bearing grooves also the maximum possible altitude of the thrust bearing, that is, the achievable by the fluid dynamic pressure distance of the two bearing surfaces of the thrust bearing is reduced.

At low rotational speeds of the bearing results in using flat bearing grooves a higher altitude than when using deeper bearing grooves. This is advantageous in particular when using low-viscosity bearing oil or at high temperatures, which lead to a lower viscosity of the bearing fluid, and in the case of a "normal" installation position of the bearing. The flat bearing grooves are therefore more effective under these operating conditions than deep bearing grooves, especially when the axial bearing gap is small. In particular, the bearing begins to lift off at lower speeds ("lift-off"), whereby the wear caused by the start / stop operations of the bearing surfaces is reduced.

At high rotational speeds of the thrust bearing resulting from the shallower bearing grooves a low altitude of the thrust bearing than when using deeper bearing grooves, especially at cold temperatures, d. H. a high viscosity of the bearing fluid and overhead mounting of the bearing. In this case, the flat bearing grooves are less effective than deep bearing grooves, especially when the axial bearing gap is very large.

The reduced flight height of the thrust bearing caused by the shallow bearing grooves increases the pressure in the bearing as well as the flow velocity and flow rate of the bearing fluid through the bearing gap. As a result, the reliability and performance of the bearing is increased, in particular due to the lower risk of negative pressure zones by the overall higher bearing pressure and the improved pumping action of the bearing, which ensures a sufficient flow of the bearing fluid through the bearing gap.

With these bearing grooves according to the invention, even at low speeds, for example in the region of 1000 rpm, a lifting of the axial bearing can be achieved at a reduced flying height. Due to the limited by the flat bearing grooves maximum altitude of the camp, the camp can be operated in a wide speed range, for example, from 5400 U / min to 7200 U / min. Even at higher speeds, the flying height of the thrust bearing is not unacceptably large. Thus, the thrust bearing clearance can be easily designed for a relatively wide speed range.

These bearing grooves are used in particular in a fluid-dynamic thrust bearing which has at least one first bearing component which is arranged to be movable relative to a second bearing component. The bearing grooves are arranged on a corresponding bearing surface of one or both bearing components.

The fluid-dynamic axial bearing may preferably be part of a fluid-dynamic bearing system for rotational mounting of a spindle motor. The spindle motor includes a stator, a rotor, and an electromagnetic drive system for rotationally driving the rotor.

Such a spindle motor may preferably be part of a hard disk drive and rotationally drive at least one disk arranged on the rotor. The hard disk drive includes a read-write device for reading and writing data from and to the storage disk.

The bearing grooves can be applied by various methods on the bearing surface of a fluid dynamic thrust bearing. A first method is an ECM method, ie a method for electrochemical removal. An ECM electrode is preferably used which has the electrically conductive regions corresponding to the bearing grooves.

In particular, the shallow bearing grooves also reduce the process time of the ECM process, resulting in cost savings. The reduced process time of the ECM process for introducing the thrust bearing grooves also improves the accuracy of the groove geometry, in particular results in a lower erosion in the region of the bearing grooves, d. H. the undesired material removal between the bearing grooves caused by the ECM process is reduced.

Another method for applying the bearing grooves provides that the bearing grooves are imprinted. For this purpose, known from the prior art embossing method for impressing the bearing grooves can be used.

Hereinafter, several embodiments of the invention will be explained with reference to the drawings. Here are the drawings and their description further advantages and features of the invention.

Brief description of the drawings

1 shows a section through a spindle motor for driving a hard disk drive with a thrust bearing according to the invention with bearing grooves.

2 shows a plan view of a bearing surface with bearing grooves of the thrust bearing according to the invention 1 ,

3 : shows an enlarged view of the area of the thrust bearing 1 ,

4 shows a plan view of a storage area with herringbone bearing grooves according to the invention.

5 shows a section through a second embodiment of a spindle motor with a thrust bearing and bearing grooves according to the invention

6 : shows a graphical representation of a comparison of the take-off speeds of a conventional and a thrust bearing according to the invention.

7 Fig. 1 is a graph showing the axial clearance ratio of the thrust bearing to the bearing groove depth for a conventional thrust bearing and an thrust bearing according to the present invention.

Description of preferred embodiments of the invention

1 shows a spindle motor with a fluid dynamic bearing system, in particular a fluid dynamic thrust bearing according to the invention. Such a spindle motor can be used to drive disks of a hard disk drive.

The spindle motor comprises a base plate 10 having a substantially central cylindrical opening in which a cup-shaped bearing member 16 is included. The first bearing component 16 is approximately pot-shaped and comprises a central opening, in which a shaft 12 is attached. At the free end of the fixed shaft 12 is a disc-shaped bearing component 18 arranged, preferably annular and integral with the shaft 12 is trained. The named components 10 . 12 . 16 and 18 form the fixed component of the spindle motor. The bearing comprises a bearing bush 14 in one by the shaft 12 and the two bearing components 16 . 18 formed intermediate space is rotatably arranged relative to these components. The upper bearing component 18 is in an annular recess of the bearing bush 14 arranged. Adjacent surfaces of the shaft 12 , the bearing bush 14 and the bearing components 16 . 18 are by a bearing gap open on both sides 20 separated from each other. The bearing gap 20 is filled with a bearing fluid, such as a bearing oil

The bearing bush 14 has a cylindrical bore on its inner circumference two cylindrical radial bearing surfaces 22a and 22b are formed, which by a circulating Separatornut in between 24 are separated. The radial bearing surfaces surround the standing wave 12 at a distance of a few microns to form an axially extending portion of the bearing gap 20 , The radial bearing surfaces are provided with suitable bearing grooves, so that they each with opposite bearing surfaces of the shaft 12 two fluid dynamic radial bearings 22a and 22b form.

To the lower radial bearing 22b closes a radially extending portion of the bearing gap 20 on, by radially extending bearing surfaces of the bearing bush 14 and corresponding opposite bearing surfaces of the first bearing component 16 is formed. These bearing surfaces form a fluid dynamic thrust bearing 26 in the form of a to the axis of rotation 46 vertical circular ring. The fluid dynamic thrust bearing 26 is for example by spiral bearing grooves 27 marked either on the front side of the bearing bush 14 , the first bearing component 16 or both parts can be attached. The bearing grooves 27 of the thrust bearing 26 preferably extend over the entire end face of the bearing bush 14 that is, from the inner edge to the outer edge. This results in a defined pressure distribution in the entire thrust bearing gap and vacuum zones are avoided, since the fluid pressure continuously increases from a radially outer to a radially inner position of the thrust bearing.

Advantageously, all are for the radial bearings 22a . 22b , the thrust bearing 26 and possibly a pump seal 36 necessary storage or pumping grooves 27 at the bearing bush 14 arranged what the production of the bearing, in particular the shaft 12 and the bearing component 16 simplified.

At the radial portion of the bearing gap 20 in the area of the thrust bearing 26 joins in proportionally filled with bearing fluid sealing gap 34 on, by opposing surfaces of the bearing bush 14 and the first bearing component 16 is formed. The sealing gap 34 seals the bearing gap 20 off on this page. The sealing gap 34 includes one opposite the bearing gap 20 widened radially extending portion that merges into a conically opening nearly axially extending portion of an outer peripheral surface of the bearing bush 14 and an inner peripheral surface of the bearing member 16 is limited. In addition to the function as a capillary seal, the sealing gap is used 34 as a fluid reservoir and provides the required for the life of the storage system fluid amount. Furthermore, filling tolerances and a possible thermal expansion of the bearing fluid can be compensated. The two of the conical section of the sealing gap 34 forming surfaces on the bearing bush 14 and the bearing component 16 can each relative to the axis of rotation 46 to be inclined inwards. As a result, the bearing fluid in a rotation of the bearing due to the centrifugal force inward in the direction of the bearing gap 20 pressed.

On the other side of the fluid bearing system is the bearing bush 14 following the upper radial bearing 22a designed so that it forms a radial extending surface, which with a corresponding opposite surface of the second bearing component 18 forms a radial gap. At the radial gap, an axially extending sealing gap closes 32 which terminates the fluid bearing system at this end. The sealing gap 32 is made by opposing surfaces of the bearing bush 14 and the bearing component 18 limited widened at the outer end with preferably conical cross-section. The sealing gap 32 preferably comprises a pumping seal 36 passing through pumping structures 36a is marked. A cover 30 closes the sealing gap 32 , The cover 30 is at a stage 38 the bearing bush 14 held and there, for example, glued, pressed and / or welded. The inner edge of the cover 30 can be together with the outer circumference of the shaft 12 form a gap seal. This increases the safety against leakage of bearing fluid from the seal gap 32 ,

The electromagnetic drive system of the spindle motor is formed in a known manner by a on the base plate 10 arranged stator arrangement 42 and an annular rotor magnet surrounding the stator assembly at a distance 44 which is on an inner circumferential surface of a hub 48 is arranged. In principle it is also possible to use the hub 48 and to form the bearing bush in one piece. The rotor magnet 44 can with a yoke 50 be combined. When the spindle motor is used to drive a hard disk drive, the hub carries one or more disks 52 ,

Since the spindle motor only a fluid dynamic thrust bearing 26 having a force in the direction of the second bearing component 18 generated, a corresponding counterforce or biasing force must be provided on the movable bearing part, which holds the bearing system axially in balance. For this purpose, the base plate 10 a ferromagnetic ring 40 have, the rotor magnet 44 axially opposite and is magnetically attracted by this. This magnetic attraction acts against the force of the thrust bearing 26 and keeps the bearing axially stable. Alternatively or in addition to this solution, the stator assembly 42 and the rotor magnet 44 axially offset from each other, in such a way that the magnetic center of the rotor magnet 44 axially further away from the base plate 10 is arranged as the magnetic center of the stator assembly 42 , As a result, an axial force is built up by the magnet system of the motor, which is opposite to the thrust bearing 26 acts.

In order to ensure a continuous flushing of the storage system with bearing fluid, in a known manner, a recirculation channel 28 intended. The recirculation channel 28 is inventively as axially or slightly obliquely extending channel in the bearing bush 14 formed, preferably at an acute angle with respect to the axis of rotation 46 of the warehouse is arranged. The recirculation channel 28 connects the two radial sections of the bearing gap 20 between the bearing areas and the sealing areas directly to each other and preferably ends in the radially outer portion of the thrust bearing 26 in that the axial gap distance is greater than the portion of the radial bearing gap which is located closer to the shaft. Due to the directed pumping action of the bearing groove structures of the thrust bearing 26 and the radial bearing 22a . 22b results in the bearing gap 20 preferably a flow of the bearing fluid in the direction of the upper sealing gap 32 , In addition, the bearing fluid in the recirculation channel 28 due to the effect of centrifugal force in the inclined channel downwards in the direction of the thrust bearing 26 promoted, so that sets a stable fluid circuit.

2 schematically shows the top view of a storage area 15 of the fluid dynamic thrust bearing 26 according to 1 , The storage area 15 is an end surface of the bearing bush 14 and has approximately the shape of a circular ring, being distributed on the bearing surface 15 several similar, spiral bearing grooves 27 are applied. The bearing grooves 27 preferably extend from one on the storage area 15 provided, outer gutter 15a up to an inner gutter 15b , The outer gutter 15a as well as the inner gutter 15b are as a chamfer or a radius respectively at the edge of the bearing surface 15 the bearing bush 14 educated.

3 shows an enlarged view of the area of the thrust bearing 1 , The representation is only to be regarded as schematic, wherein the dimensions of the bearing groove structures and the width of the bearing gap or the axial play are exaggerated in relation to the dimensions of the bearing components.

One recognizes those on the front side or the bearing surface 15 the bearing bush 14 arranged Lagerrillenstrukturen 27 , where the storage area 15 a radially extending bearing surface 17 of the bearing component 16 opposite. The two storage areas 15 . 17 are through the bearing gap 20 separated from each other. The bearing gap 20 is filled with bearing fluid.

It can also be seen, shown schematically here, the axial play s of the thrust bearing, which according to the invention, for example, 30 microns +/- 10 microns. The axial clearance s is determined by the difference between a maximum and a minimum axial distance of the two bearing surfaces 15 . 17 or the difference between the maximum and minimum widths of the thrust bearing gap 20 , The depth t of the bearing groove structures 27 is according to the invention 8 microns +/- 2 microns.

Thus, there is a nominal ratio between the axial clearance and the depth of the bearing grooves of s / t = 3.75.

In 4 is shown on the storage area 15 alternatively to the spiral bearing grooves 27 herringbone bearing grooves 27a ("Herringbone grooves") may be provided, the basic shape of which is known from the prior art.

5 shows a longitudinal section through a spindle motor with a fluid dynamic thrust bearing according to the invention. The spindle motor includes a fixed bushing 114 having a central bearing bore and forming the fixed component of the bearing system. In the bearing bore of the bearing bush 114 is a wave 112 used, whose diameter is slightly smaller than the diameter of the bore. Between the surfaces of the bearing bush 114 and the wave 112 there remains a narrow bearing gap 120 a few microns wide. The opposite surfaces of the shaft 112 and the location socket 114 form two fluid dynamic radial bearings 122a . 122b out, by means of which the shaft 112 around a rotation axis 146 rotatable in the bearing bush 114 is stored. The radial bearings 122a . 122b are characterized by bearing groove structures that rest on the surface of the bearing bush 114 or the wave 112 are applied. The bearing gap 120 is filled with a suitable bearing fluid, such as a bearing oil. The bearing groove structures of the radial bearings 122a . 122b practice with rotation of the shaft 112 a pumping action on the in the bearing gap 120 between wave 112 and bearing bush 114 located bearing fluid. As a result, a pressure is built up in the bearing gap, the radial bearings 122a . 122b makes it workable.

A free end of the wave 112 is with a hub 148 connected, which one the bearing bush 114 partially surrounding cylindrical neck. The bearing gap 120 includes an axial section extending along the shaft 112 and the two radial bearings 122a . 122b extends, and a radial portion extending along the end face of the bearing bush 114 and the thrust bearing 126 extends. A lower, flat surface of the hub 148 forms together with a storage area 115 trained end face of the bearing bush 114 a fluid dynamic thrust bearing 126 out. The storage area 115 the bearing bush 114 or a trained as a bearing surface surface of the opposite hub 148 are with bearing grooves 127 provided during rotation of the shaft 112 a pumping action on the in the bearing gap 120 between the hub 148 and the end face of the bearing bush 114 located bearing fluid exerts, so that the thrust bearing 126 becomes sustainable. The bearing grooves 127 of the thrust bearing 126 are preferably as helical bearing grooves 127 educated. The shape of the storage area 115 and the bearing grooves 127 corresponds essentially to the illustration in 2 ,

At the bottom of the shaft 112 is a one-piece with the shaft or a separately formed stopper ring 113 arranged, which has an enlarged outer diameter compared to the shaft diameter. The stopper ring 113 prevents the shaft from falling out 112 from the bushing 114 , The bearing is on this side of the bearing bush 114 through a flat cover plate 130 locked. Between the surfaces of the stopper ring 113 and the surfaces of the bearing bush 114 or the cover plate 130 remains a filled with bearing fluid recess, with the bearing gap 120 connected is. The stopper ring 113 turns together with the shaft 112 inside the recess between bearing bush 114 and cover plate 130 in the bearing fluid.

At the radially outer end of the radial portion of the bearing gap 120 a gap is arranged with a larger gap distance, which partially as a sealing gap 132 acts. The gap initially extends from the bearing gap 120 radially outwardly and merges into an axial section extending along the outer circumference of the bushing 114 between the bearing bush 114 and a cylindrical hub approach 148 extends and the sealing gap 132 forms. The outer surface of the bearing bush 114 and the inner surface of the hub 148 form the limit of sealing gap 132 , Thus, the sealing gap runs 132 approximately parallel to the axis of rotation 146 ,

In the bearing bush 114 can be a recirculation channel 128 be provided, the one at the outer edge of the thrust bearing 126 located section of the storage gap 120 with one below the lower radial bearing 122b located section of the storage gap 120 interconnects and supports a circulation of the bearing fluid in the camp.

The bearing bush 114 is in a base plate 110 arranged the spindle motor. The hub 148 has on its outer circumference on a peripheral edge. At the base plate 110 one is the bearing bush 114 surrounding stator assembly 142 arranged, which consists of a ferromagnetic laminated stator core and corresponding stator windings. This stator arrangement 142 is at a radial distance surrounded by an annular rotor magnet 144 , The rotor magnet 144 is on the inner circumference of the circumferential edge of the hub 148 attached. When the spindle motor is used to drive a hard disk drive, the hub carries one or more disks 152 ,

The drive system has an axial offset between the magnetic center of the rotor magnet 144 and the magnetic center of the laminated stator core 142 on. This results in a static, downward in the direction of the base plate 110 directed magnetic force. This magnetic force is opposite to the bearing force of the thrust bearing 126 directed and serves the axial bias of the bearing system or the thrust bearing 126 , It can also be a ferromagnetic ring 140 below the rotor magnet 144 be arranged, which together with the rotor magnet 144 an additional axial force to bias the thrust bearing 126 generated.

The bearing grooves of the radial bearings 122a . 122b as well as the bearing grooves 127 of the thrust bearing 126 can be introduced in a known manner and according to a preferred embodiment of the invention by means of an electrochemical removal process (ECM process) in the corresponding storage areas. For this purpose, an ECM electrode is used, which has on its surface an image of the bearing grooves to be introduced. By means of the ECM process, the bearing grooves 127 of the thrust bearing 126 with a depth of 8 μm +/- 2 μm into the surface of at least one of the opposite bearing components, preferably into the bearing bush 114 brought in. According to the invention, a separator groove is now preferably also used in the same working process 124 introduced into the bearing component, namely between the corresponding bearing grooves of the two radial bearings 122a . 122b , Because the separator groove 124 is relatively narrow, this can be very well realized by means of an ECM method.

Also the bearing grooves of in 4 shown bearing can preferably be introduced by ECM method in the storage areas.

6 shows a graphical representation of a comparison between the altitude of a conventional thrust bearing and a thrust bearing with the flat bearing grooves according to the invention plotted against the rotational speed. The curve 62 shows the altitude of a conventional thrust bearing, it can be seen that the thrust bearing surfaces only stand out at a speed of 3500 rev / min. The altitude increases approximately linearly with increasing speed and is for example at a speed of 4200 rev / min 1.2 microns. The curve 60 shows the flying height of a thrust bearing according to the invention. It can be seen clearly that the thrust bearing surfaces already stand out from one another at a speed of just over 1000 rpm and then the flight altitude increases substantially linearly and is already just under 2 micrometers at about 2000 rpm. This comparison shows that the shallow bearing grooves with a groove depth of 8 microns +/- 2 microns allow a relatively high flying height of the thrust bearing even at low speeds. This raises the thrust bearing already at a speed of 1000 rev / min and the mechanical abrasion on the bearing surfaces is reduced accordingly. This is an essential aspect of the invention.

7 shows a graphical representation of the relationship between the axial clearance s and the bearing groove depth t of the thrust bearing. The left illustration, denoted by the reference numeral 64 Fig. 13 shows the range of this ratio for a conventional thrust bearing of the prior art. In a conventional axial bearing of a spindle motor for driving 2.5 inch hard disk drives, the minimum ratio is 1.67 and the maximum ratio is 2.37.

The right curve 66 shows the corresponding relationship between thrust bearing clearance s and depth t of the bearing grooves in a thrust bearing with flat bearing grooves according to the invention. The minimum ratio is 3.33 and the maximum ratio is 4.0. The curves 64 and 66 illustrate the difference between a prior art thrust bearing with conventional deep thrust bearing grooves and a thrust bearing according to the invention with substantially flatter thrust bearing grooves.

LIST OF REFERENCE NUMBERS

10, 110

baseplate

12, 112

wave

113

stopper ring

14, 114

bearing bush

15, 115

storage area

15a

gutter

15b

gutter

16

pot-shaped bearing component

17, 117

storage area

18

disc-shaped bearing component

20, 120

bearing gap

22a, 22b

radial bearings

122a, 122b

24, 124

Separatornut

26, 126

thrust

27, 27a, 127

raceways

28, 128

recirculation

30, 130

cover

32, 132

seal gap

34

seal gap

36

pump seal

36a

Pumping groove structures

38

step

40, 140

ferromagnetic ring

42, 142

stator

44, 144

rotor magnet

46, 146

axis of rotation

48, 148

hub

50

yoke

52, 152

disk

60

Curve (bearing groove depth 9 μm)

62

Curve (bearing groove depth 17 μm)

64

Ratio axial clearance / bearing groove depth (so far)

66

Ratio axial clearance / bearing groove depth (invention)

s

Axiallagerspiel

t

groove depth

Claims (12)

Storage area ( 15 ; 115 ) with bearing grooves ( 27 ; 127 ) of a fluid dynamic thrust bearing ( 26 ; 126 ), this storage area ( 15 ; 115 ) another storage area ( 17 ; 117 ), and between the two storage areas ( 15 ; 115 ; 17 ; 117 ) a bearing gap filled with a bearing fluid ( 20 ; 120 ), wherein the bearing grooves ( 27 ; 127 ) have a defined depth, characterized in that the depth t of the bearing grooves ( 27 ; 127 ) Is 8 μm +/- 2 μm and between the bearing surfaces ( 15 ; 115 . 17 ; 117 ) An axial clearance s of 30 microns +/- 10 microns is provided. Bearing surface according to claim 1, characterized in that by the difference between a maximum and a minimum axial distance of the two bearing surfaces ( 15 . 115 ; 17 ; 117 ) the axial play s of the fluid dynamic thrust bearing ( 26 ; 126 ), wherein the ratio between the axial play s of the thrust bearing ( 26 ; 126 ) and the depth t of the bearing grooves ( 27 ; 127 ) is between 3.33 and 4.00 Bearing surface according to one of claims 1 or 2, characterized in that a plurality of bearing grooves ( 27 ; 127 ) in the same geometric orientation at a distance from each other on the bearing surface ( 15 ; 115 ) are arranged. Bearing surface according to one of claims 1 to 3, characterized in that the bearing grooves ( 27 ; 127 ) are arranged spirally on the bearing surface. Bearing surface according to one of claims 1 to 3, characterized in that the bearing grooves ( 27a ) are arranged herringbone on the bearing surface. Fluid dynamic thrust bearing with a first bearing component ( 14 ; 148 ), which relative to a second bearing component ( 16 ; 114 ) is movable, preferably rotatable, is arranged, wherein bearing grooves ( 27 ; 127 ) on a storage area ( 15 ; 115 ) of one of the bearing components are arranged, wherein this bearing surface ( 15 ; 115 ) another storage area ( 17 ; 117 ) of the other bearing component is assigned, and between the two bearing surfaces filled with a bearing fluid bearing gap ( 20 ; 120 ), wherein the bearing grooves ( 20 ; 120 ) have a defined depth, characterized in that the depth of the bearing grooves ( 27 ; 127 ) Is 8 μm +/- 2 μm and between the bearing surfaces ( 15 ; 115 ; 17 ; 117 ) An axial clearance s of 30 microns +/- 10 microns is provided. Fluid-dynamic thrust bearing according to claim 6, characterized in that the ratio between the axial play s of the thrust bearing ( 26 ; 126 ) and the depth t of the bearing grooves ( 27 ; 127 ) is between 3.33 and 4.00 Fluid dynamic thrust bearing according to one of claims 6 or 7, characterized in that a plurality of bearing grooves ( 27 ; 127 ) in the same geometric orientation at a distance from each other on the bearing surface ( 15 ; 115 ) are arranged. Fluid dynamic thrust bearing according to one of claims 6 to 8, characterized in that the bearing grooves ( 27 ; 127 ) are arranged spirally on the bearing surface. Fluid dynamic thrust bearing according to one of claims 6 to 8, characterized in that the bearing grooves ( 27a ) are arranged herringbone on the bearing surface. Fluid-dynamic thrust bearing according to one of claims 6 to 10, characterized in that it is part of a fluid dynamic bearing system for pivotally mounting a spindle motor, wherein the spindle motor comprises a stator, a rotor and an electromagnetic drive system ( 42 ; 44 ; 142 ; 144 ) to the rotary drive of the rotor. Fluid-dynamic thrust bearing according to one of claims 6 to 11, characterized in that it is part of a fluid dynamic bearing system for pivotally mounting a spindle motor, wherein the spindle motor is part of a hard disk drive and at least one arranged on the rotor disk ( 52 ; 152 ) drives, and the hard disk drive further a read-write device for reading and writing data from and to the disk ( 52 ; 152 ) having.

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