DE102012007423A1 - Fluid dynamic bearing system of spindle motor for driving e.g. compact disk (CD), has fluid dynamic bearings whose sealing gap is formed adjacent to bearing gap, so that ratio of axial length and gap width satisfies preset value - Google Patents

Abstract

The system has a movable bearing component (12) that is rotatably mounted relative to a fixed bearing component (10) about an axis of rotation (16). A bearing gap (14) formed between the bearing surfaces of the fluid dynamic bearings (18,22,28,30), is filled with storage fluid. A sealing gap (36) for sealing an open end of the bearing gap, is formed adjacent to the bearing gap towards direction of the rotation axis, such that the ratio of axial length and the width of the sealing gap is less than 40. Independent claims are included for the following: (1) a spindle motor; and (2) a method for optimizing shock resistance of fluid dynamic bearing system.

Description

Field of the invention

The invention relates to a fluid dynamic bearing system, in particular for the rotary mounting of a spindle motor, as used for example for driving hard disk drives.

State of the art

Fluid dynamic bearing systems, as used in spindle motors, are known in a variety of designs. As a rule, a fluid-dynamic bearing comprises at least two bearing components rotatable relative to one another, which form a bearing gap filled with a bearing fluid between mutually associated bearing surfaces. The bearing fluid used is preferably a liquid fluid, for example a bearing oil. In a known manner, bearing surfaces assigned to the bearing surfaces and acting on the bearing fluid bearing groove structures are provided, which are usually applied in the form of depressions or elevations on individual or both opposing bearing surfaces. The Lagerillenstrukturen act as storage or pumping structures that exert a pumping action on the bearing fluid in the bearing gap on rotation of the bearing components, so that a hydrodynamic pressure is built up in the bearing gap, whereby the opposite bearing surfaces are separated. This hydrodynamic pressure is responsible for the load capacity and rigidity of the fluid dynamic bearing system.

According to a known design, a fluid-dynamic bearing system comprises a shaft which is rotatably mounted in a bearing bore of a bearing bush. The diameter of the bearing bore is slightly larger than the diameter of the shaft, so that between the surfaces of the bearing bush and the shaft remains filled with a bearing fluid bearing gap. The surfaces of the shaft and / or the bearing bush have pressure-generating bearing groove structures which form at least one fluid dynamic radial bearing along an axial portion of the bearing gap. A free end of the shaft is connected to a rotor component (hub). At one end of the shaft, a pressure plate is attached, forming the two end faces together with opposite surfaces of the bearing bush or a cover plate in the axial direction against each other working fluid dynamic thrust bearing. In each case, a bearing surface of the two thrust bearing is provided with pressure generating Lagerrillenstrukturen.

The bearing gap has at least one open end, which must be sealed from the environment, so that no bearing fluid can escape from the bearing gap. The sealing takes place in a known manner by means of a capillary sealing gap which adjoins the radial portion of the bearing gap above the outer radial bearing. The sealing gap is proportionally filled with bearing fluid. The filled with bearing fluid volume of the sealing gap serves as a fluid reservoir, which provides sufficient storage fluid for the entire life of the fluid dynamic bearing. In addition, the sealing gap serves as a compensating volume for the bearing fluid, for example, due to temperature changes of the bearing fluid and resulting volume changes or due to volume changes in the sealing gap, which can occur between standstill and the operation of the bearing.

In particular, if the fluid dynamic bearing is subjected to strong vibrations or to a shock effect, this leads to a strong axial movement between the movable and the stationary bearing component, so that in particular the width of the sections of the bearing gap running perpendicular to the axial direction changes. This change in the width of the bearing gap leads to a change in the fluid volume between the bearing components. Depending on the direction of movement of the shaft in the bearing bush or the rotor component with respect to the bearing bush, the bearing fluid is forced out of the bearing gap or pulled in at high speed, while the gap width of the axial bearing gap changes.

Since the bearing gap is only connected to the outside environment via the sealing gap, it can occur in the case of an excessively large shock acting on the bearing that bearing fluid escapes from the sealing gap. Of course, fluid dynamic bearings have appropriate requirements for the shock resistance, which must be complied with.

The shock resistance depends on many factors, in particular on the properties, such as the viscosity of the bearing fluid used, the level of the bearing fluid in the sealing gap, the geometry of the sealing gap, the ambient temperature, etc. It should be noted that the fluid dynamic bearing with a minimum amount of bearing fluid must be filled in order to reach the specified minimum life of the bearing, as there is always a loss of bearing fluid due to evaporation, etc. The prescribed filling quantity of bearing fluid must therefore be taken into account when designing the bearing.

So far, attempts have been made to optimize the shock resistance by making the length of the sealing gap in the axial direction as large as possible, so that the bearing fluid has enough space to expand or spread in a shock. An extension of the sealing gap leads However, with a constant total length of the bearing to a shortening of the available bearing distance between the two radial bearings, since the sealing gap and the two radial bearings are arranged axially in series, ie one behind the other.

Disclosure of the invention

The object of the invention is to optimize the shock resistance of a fluid dynamic bearing of the type mentioned, without having to reduce the bearing distance.

This object is achieved by a fluid dynamic bearing system having the features of claim 1.

Preferred embodiment of the invention and further advantageous features are specified in the dependent claims.

The fluid-dynamic bearing system comprises a fixed bearing component and a movable bearing component which is mounted so as to be rotatable relative to the stationary bearing component about a rotation axis. Between bearing surfaces assigned to the two bearing components, a bearing gap is arranged, which is filled with a bearing fluid. At least one fluid-dynamic radial bearing is provided, which is formed by bearing surfaces along an axially extending portion of the bearing gap. At least one fluid dynamic thrust bearing is formed by bearing surfaces along a radially extending portion of the bearing gap. For sealing an open end of the bearing gap, which has a radial bearing gap width d in the range of about two to five microns, a sealing gap is provided, which adjoins the radial portion of the bearing gap and adjacent to the bearing gap minimum radial width b perpendicular to the axis of rotation and Has a running in the direction of the axis of rotation H length.

According to the invention, the ratio between the length H and the width b of the sealing gap is H / b <40.

The invention is particularly concerned with optimizing the shock resistance of a fluid dynamic bearing for a given, predetermined volume of bearing fluid in the bearing, which is determined by the rate of evaporation, the thermal expansion of the bearing fluid and the bearing components, the thrust bearing clearance and other geometrical tolerances of the bearing components. With the help of the invention, the most appropriate structural design of the sealing gap is determined for a given volume of the bearing fluid.

The inventors have found through shock tests and simulations that the radial width of the sealing gap has a much greater influence on the shock resistance than the axial length of the sealing gap or the level of the bearing fluid within the sealing gap. The shock resistance can be optimized by increasing the minimum width of the sealing gap, at the same time reducing the axial length of the sealing gap and thus the bearing distance can be increased accordingly in the case of using two radial bearings, which increases the bearing stiffness. Also, the opening angle of the seal gap or the inclination angle of a boundary surface of the seal gap with respect to the rotation axis plays a role for the shock resistance.

In particular, the transition between the bearing gap and the sealing gap is formed as a step, d. H. the bearing gap abruptly merges into the sealing gap, wherein the bearing gap, which has a width of a few (about 2 to 5) micrometers, abruptly widens to at least 0.04 millimeters, which corresponds to the minimum width b of the sealing gap.

Good shock resistance is achieved if the minimum width b of the sealing gap is preferably between 0.04 mm and 0.16 mm.

The sealing gap extends substantially in the axial direction parallel to the axis of rotation and preferably has an axial length H between 0.7 mm and 1.4 millimeters and is partially filled with bearing fluid.

It follows that the optimum ratio between the length and the width of the sealing gap H / b is <40, but preferably <25. With these values, an optimized shock resistance can be achieved.

Based on the abrupt transition between the bearing gap and sealing gap, at least one surface delimiting the sealing gap widens conically in a continuous opening angle α with respect to the axis of rotation, wherein the opening angle α is between 3 degrees and 20 degrees.

So far, an opening angle α of about 10 degrees was common, but it has been found that in particular opening angles that are less than 10 degrees, may have advantages in shock resistance.

The fluid dynamic bearing system can either be a bearing system with a rotating shaft to which a rotor component is attached, wherein the shaft is rotatably mounted in a stationary bearing bush or a bearing system with a fixed shaft which is arranged in a stationary component and is surrounded by a rotatably mounted bearing bush. In a bearing system with a rotating shaft, the sealing gap is limited by surfaces of the bearing bush and the shaft. However, the sealing gap can also be arranged between surfaces of the shaft and the rotor component without restriction.

The invention equally relates to a spindle motor with a stator and a rotor, which is rotatably mounted relative to the stator by means of the fluid dynamic bearing according to the invention. The rotor is driven by an electromagnetic drive system, wherein the spindle motor is preferably for driving fans, hard drives, or other storage drives, such as for optical media, such as CD, DVD or Blu Ray.

The invention will be described in more detail with reference to several embodiments with reference to the drawings. In this case, from the drawings and their description further advantageous features and advantages of the invention.

Brief description of the drawings:

1 shows a section through a fluid dynamic storage system according to the invention.

2 shows an enlarged section through the storage system of 1 in the area of the sealing gap.

3 shows a diagram of the shock resistance of the fluid dynamic bearing as a function of the minimum gap width b of the sealing gap for bearing fluids of different viscosity.

4 shows a diagram of the shock resistance of a fluid dynamic bearing as a function of the minimum gap width b of the sealing gap with reduced axial length H of the sealing gap for bearing fluids of different viscosity.

5 shows a diagram of the shock resistance of a fluid dynamic bearing as a function of the axial length H of the sealing gap for different widths b of the sealing gap and a bearing fluid of low viscosity.

6 shows a diagram of the shock resistance of a fluid dynamic bearing as a function of the axial length H of the sealing gap for different widths b of the sealing gap and a bearing fluid with high viscosity.

7 shows a diagram of the shock resistance of a fluid dynamic bearing as a function of the minimum width b of the sealing gap for different opening angles α of the sealing gap and the use of a bearing fluid with high viscosity.

8th shows a diagram of the shock resistance of a fluid dynamic bearing as a function of the minimum width b of the sealing gap for different opening angles α of the sealing gap and the use of a bearing fluid with low viscosity.

9 shows a section through a spindle motor with a fluid dynamic bearing according to 1 ,

Description of preferred embodiments of the invention

The storage system after 1 includes a bushing 10 with an axial cylindrical bore, in which a shaft 12 around a rotation axis 16 is received rotatably. Between adjacent surfaces of the bearing bush 10 and the wave 12 are defined bearing surfaces, the two fluid dynamic radial bearings 18 . 22 form in mutual axial distance. The bearing surfaces of the radial bearings 18 . 22 are through a bearing gap 14 separated from each other, which is filled with a bearing fluid, preferably a bearing oil. The radial bearings 18 . 22 are through bearing groove structures 20 . 24 marked on one of the surfaces of bearing bush 10 or wave 12 , but preferably on the bearing bush, are applied. The bearing groove structures 20 . 24 For example, they may be sinusoidal, parabolic or herringbone.

At one end of the wave 12 is an annular pressure plate 32 fixed in an annular recess of the bearing bush 10 is arranged. This recess is made by a cover plate 34 covered, which closes the storage system on this page. The two end faces of the pressure plate 32 form together with the opposite surfaces of the bearing bush 10 and the cover plate 34 each fluid dynamic thrust bearing 28 . 30 out. The thrust bearings 28 . 30 are characterized in a known manner by bearing groove structures (not shown), which may be formed, for example herringbone or spiral. The bearing surfaces of the two thrust bearings 28 . 30 are by radially extending portions of the bearing gap 14 separated, whereas the bearing surfaces of the two radial bearings 18 . 22 through an axially extending portion of the bearing gap 14 are separated from each other. The bearing groove structures 20 . 24 the radial bearing 18 . 22 or the thrust bearing 28 . 30 generate during a rotation of the shaft 12 together with the pressure plate 32 relative to the bearing bush 10 a pumping action on the bearing fluid in the bearing gap. As a result, a hydrodynamic pressure is built up in the bearing gap, which the bearings 18 . 22 . 28 . 30 makes it workable. The pumping action of the bearing groove structures 20 . 24 can be uniform in both directions of the bearing gap 14 or be directed predominantly in one direction. For example, the bearing groove structures of the outer radial bearing pump 18 mainly in the direction of the pressure plate 32 to the bearing fluid in the bearing gap 14 under an overpressure while the bearing groove structures of the inner radial bearing 22 preferably symmetrical to the center line (Apex) 23 are formed and thus produce a non-directional pumping action towards the apex. The bearing groove structures of the two thrust bearings 28 . 30 preferably pump the bearing fluid radially inward in the direction of the axis of rotation 16 ,

To the outside is the bearing gap 14 through a capillary sealing gap 36 sealed, which is proportionally filled with bearing fluid. A fluid meniscus created by capillary forces 38 defines the filling level of the bearing fluid in the sealing gap 36 , The sealing gap 36 keeps a certain supply of bearing fluid to compensate for losses, for example due to evaporation, as well as volume changes of the bearing fluid caused by temperature fluctuations.

When the bearing system is used to pivotally mount a spindle motor, it is on a free end of the shaft 12 a hub 26 of the spindle motor which is driven by an electromagnetic drive system. The bearing bush 10 is in a base plate 40 held the spindle motor.

2 shows an enlarged view of the sealing gap 36 of the camp of 1 , The sealing gap 36 begins at the end of the storage gap 14 with an approximately right-angled step 46 ie the bearing gap 14 widens abruptly from a width d to a width b of the sealing gap 36 on which the minimum width of the sealing gap 36 represents. This minimum width b is according to the invention by an order of magnitude greater than the width d of the bearing gap.

While the width d of the bearing gap is a few micrometers, for example two to five micrometers, the minimum width b of the sealing gap is 36 at least 40 microns or more.

Starting from this stage 46 between bearing gap 14 and sealing gap 36 the sealing gap widens 36 in the direction of its opening at an opening angle α, wherein this opening angle α can be between 3 degrees and 20 degrees. As a rule, only a single boundary surface of the sealing gap widens 36 , For example, the surface of the bearing bush 10 at an angle α, while the other boundary surface, for example, the surface of the shaft 12 , parallel to the axis of rotation 16 runs. However, it can be provided that both boundary surfaces of the sealing gap 36 at an acute angle to the axis of rotation 16 run.

The sealing gap 36 is directly with the bearing gap 14 connected and proportionally filled with bearing fluid, which is due to the capillary forces between the bearing fluid and the wall of the sealing gap 36 a fluid meniscus 38 forms the filling height of the bearing fluid in the sealing gap 36 displays. The sealing gap 26 has an axial length H, which is so dimensioned that it ensures on the one hand a sufficient shock resistance of the bearing, so unfolds a sufficient sealing effect, and on the other hand, the bearing clearance, ie the distance between the center lines 19 . 23 the two radial bearings 18 . 22 not shortened unnecessarily. It has been found that not - as previously suspected - the axial length H of the sealing gap 36 the greatest impact on the shock resistance of the seal, but the width b of the sealing gap 36 , The larger one chooses the width b (in a reasonable frame), the shorter can be the axial length H of the sealing gap with the same sealing gap volume and the greater is then the bearing clearance between the two radial bearings 18 . 22 , As a result, not only the shock resistance of the seal or the bearing is optimized, but also the bearing clearance and thus the rigidity of the bearing. The invention is based on optimum values for the width b and the height H of the sealing gap 36 to find.

3 Fig. 10 is a graph similar to the shock resistance in [g] of a fluid dynamic bearing 1 depending on the minimum width b of the sealing gap 36 , The axial length H of the sealing gap 36 is 1.6 millimeters in this example. In previous camps was the typical width b of the sealing gap 36 only about 0.012 millimeters.

The upper curve 100 shows the shock resistance of a bearing with a bearing fluid of high viscosity, it can be seen that a maximum of the shock resistance of about 700 g with a width b of about 0.08 millimeters is reached.

The lower curve 102 shows the same representation for a bearing with a bearing fluid of low viscosity. As expected, the shock resistance is lower because the bearing fluid flows more easily due to the low viscosity. It can be seen, however, that here too the optimum width b of the sealing gap is about 0.08 millimeters, so that there the shock resistance reaches its maximum value of about 575 g reached. The axial length H is still 1.6 millimeters.

The 4 shows similar diagrams as the 3 , here also the shock resistance of a bearing as a function of the width b of the sealing gap 36 is shown. However, lower values for the axial length H of the sealing gap are selected here. The upper curve 104 shows a bearing with a sealing gap 36 with the axial length H = 1.32 millimeters and a bearing fluid of high viscosity. It can be seen that the shock resistance reaches its maximum value of about 550 g at about b = 0.12 millimeters, while in the curve 106 with a bearing fluid in low viscosity and a slightly longer axial length H = 1.4 millimeters, the optimum shock resistance of about 500 g at about b = 0.10 millimeters is reached.

According to the invention, the axial length H of the sealing gap can be further reduced, the achievable shock resistance still being significantly above the specified shock resistance of 300 g.

5 shows the shock resistance of a bearing as a function of the axial length H of the sealing gap 36 , The curve 108 shows the shock resistance of a fluid dynamic bearing with a previously common width of the sealing gap 36 of b = 0.012 millimeters, where it can be seen that the shock resistance reaches its maximum value of about 800 g with an axial length of H = 1.4 millimeters. The curve 110 shows in comparison the shock resistance of a bearing according to the invention, in which b = 0.080 millimeters was chosen. Here you can see that the maximum shock resistance of about 850 g has already been achieved with an axial length of H = 0.9 millimeters. D. h., The axial length H of the sealing gap 36 can be chosen much smaller for larger widths b as usual, so that a larger radial bearing span is available, which increases the storage stability. The illustrated shock resistance curves apply to a spindle motor with a bearing fluid of low viscosity.

6 shows the shock resistance as a function of the axial length H of the sealing gap 36 for a fluid dynamic bearing with a larger viscosity bearing fluid. Compared to 5 it can be seen that the shock resistance is generally greater, since the bearing fluid flows worse due to the greater viscosity and thus better in the sealing gap 36 is held.

The curve 112 again shows the behavior of a fluid dynamic bearing of conventional design with a sealing gap 36 with the typical width b = 0.012 millimeters. Here, the greatest value of the shock resistance of about 950 g is achieved with an axial length of about H = 1.4 millimeters. In comparison, the curve shows 114 a fluid dynamic bearing with b = 0.08 millimeters, whereby here the greatest value of the shock resistance of 1025 g at H = 0.9 millimeters is achieved.

From the 5 and 6 It can be seen that the optimum values for the width b and the axial length H of the sealing gap 36 in both examples are substantially the same. The absolute shock resistance is determined by the viscosity of the bearing fluid used.

Not only the width b and the axial length H of the sealing gap 36 have an influence on the shock resistance of a bearing, but also the opening angle α of the sealing gap or the inclination angle α of the boundary surface of the sealing gap 36 with respect to the axis of rotation, more specifically the inclination angle α with respect to the direction of the course of the adjacent sealing gap 14 , The hitherto customary opening angle α in fluid dynamic bearings of the Applicant was 10.38 degrees.

7 shows a representation of the shock resistance as a function of the minimum width b of the sealing gap 36 for different opening angles α of the sealing gap with a constant axial length of the sealing gap of H = 1.6 millimeters and a constant filling volume with bearing fluid within the sealing gap. The family of curves shows from bottom to top a decreasing opening angle α of the sealing gap of initially 20 degrees to 3 degrees. It can be seen that with large opening angles above 15 degrees, the shock resistance at different widths b of the sealing gap does not change significantly. At an opening angle α smaller than 15 degrees, however, it can be seen that the shock resistance changes relatively strongly depending on the value of the width b. For smaller opening angles α the shock resistance increases and there is for each opening angle α a maximum of the shock resistance for a certain width b of the sealing gap, which at opening angles is smaller than α <10 ° at about b = 0.1 millimeters. However, it can be seen that at small aperture angles α of less than 7 ° for small minimum widths b of the seal gap, the shock resistance decreases rapidly as the bearing fluid becomes close due to the constant axial length H of the seal gap and a constant fill volume with bearing fluid within the seal gap the upper edge of the sealing gap 38 passes and thus relatively easily escape from the sealing gap under shock action.

The family of curves shown here applies to a fluid dynamic bearing with a bearing fluid of high viscosity.

8th shows a comparable set of curves for a bearing with a bearing fluid of lower viscosity, in which case the values for different opening angles and different widths b with 7 are comparable. It is only the maximum achievable shock resistance due to the lower viscosity of the bearing fluid overall lower.

From the 7 and 8th it can be seen that, when the width of the sealing gap is suitably dimensioned 36 of about b> 0.06 mm, the shock resistance at smaller opening angles α of, for example, 5 degrees to 7 degrees significantly increased. In turn, then could the axial length H of the sealing gap 36 be further reduced and the storage margin increased, the specified shock resistance would still be met.

9 shows a section through a spindle motor with a fluid dynamic bearing according to the invention. Regarding 1 are the same components designated by the same reference numerals. The spindle motor includes a fixed base plate 40 , in which the bearing bush 10 is attached to the fluid dynamic bearing. The free end of the shaft 12 is with the hub 26 connected. The hub 26 is designed according to the purpose of the spindle motor. If the spindle motor is intended to drive a hard disk drive, be on the hub 26 one or more storage disks (not shown in the drawing) of the hard disk drive are arranged and fastened. At an inner, lower edge of the hub is an annular permanent magnet 44 arranged with a plurality of pole pairs. In the case of an aluminum hub 26 the permanent magnet is surrounded by a ferromagnetic yoke. Opposite the permanent magnet 44 is at the base plate 40 a stator assembly 42 attached by a radial air gap from the permanent magnet 44 is disconnected. The stator arrangement 42 has a plurality of stator windings, which act accordingly energized generate an alternating electric field, so that the rotor of the spindle motor, consisting of the hub 26 and wave 12 , is set in rotation.

LIST OF REFERENCE NUMBERS

10

bearing bush

12

wave

14

bearing gap

16

axis of rotation

18

radial bearings

19

Center line (Apex)

20

Bearing groove structures

22

radial bearings

23

Center line (Apex)

24

Bearing groove structures

26

Rotor component (hub)

28

thrust

30

thrust

32

Pressure plate, stopper plate

34

cover

36

seal gap

38

fluid meniscus

40

baseplate

42

stator

44

rotor magnet

46

step

b

Minimum width of the sealing gap

H

axial length of the sealing gap

d

Width of the storage gap

α

opening angle

Claims (10)

Fluid dynamic bearing system, comprising: a fixed bearing component ( 10 ), a movable bearing component ( 12 ), which relative to the fixed bearing component about an axis of rotation ( 16 ) is rotatably mounted, a bearing gap ( 14 ), which is arranged between mutually associated bearing surfaces of the two bearing components and filled with a bearing fluid, at least one fluid dynamic radial bearing ( 18 ; 22 ) by bearing surfaces along an axially extending portion of the bearing gap ( 14 ) is formed, at least one fluid dynamic thrust bearing ( 28 ) by bearing surfaces along a radially extending portion of the bearing gap ( 14 ) is formed, a proportionately filled with bearing fluid sealing gap ( 36 ) for sealing an open end of the storage gap ( 14 ), wherein the sealing gap ( 36 ) to the bearing gap ( 14 ) adjacent and adjacent to the bearing gap minimum width b and one in the direction of the axis of rotation ( 16 ) has axial length H, characterized in that the relationship between the axial length H and the width b of the sealing gap is H / b <40. Fluid dynamic bearing system according to claim 1, characterized in that the transition between the bearing gap ( 14 ) and the sealing gap ( 36 ) is formed as a stage. Fluid dynamic bearing system according to claim 1 or 2, characterized in that the minimum width b of the sealing gap ( 36 ) is between 0.04 mm and 0.16 mm. Fluid dynamic bearing system according to one of claims 1 to 3, characterized in that the axial length H of the sealing gap ( 36 ) is between 0.7 mm and 1.4 mm. Fluid dynamic bearing system according to one of claims 1 to 4, characterized in that at least one of the sealing gap ( 36 ) limiting surface in a constant opening angle α with respect to the axis of rotation ( 16 ), wherein the opening angle α is between 3 ° and 20 °. Fluid dynamic bearing system according to one of claims 1 to 5, characterized in that the fixed bearing component is a bearing bush ( 10 ) and the movable bearing component a shaft ( 12 ) and a rotor component connected to the shaft ( 26 ). Fluid dynamic bearing system according to one of claims 1 to 5, characterized in that the fixed bearing component comprises a shaft and a stationary member connected to the shaft and the movable bearing member comprises a bearing bush. Spindle motor with a stator and a rotor which is rotatably supported by means of a fluid dynamic bearing according to one of the claims 1 to 7, relative to the stator, and an electromagnetic drive system for driving the rotor. Method for optimizing the shock resistance of a fluid dynamic bearing system according to one of claims 1 to 7, wherein for a given total filling volume of bearing fluid in the bearing, the minimum width b and the axial length H of the sealing gap ( 36 ) of the fluid dynamic bearing are dimensioned such that the sealing gap is proportionally filled with bearing fluid and where: H / b <40. A method according to claim 9, characterized in that the opening angle α of the sealing gap ( 36 ) is selected between 3 ° and 20 °.

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