DE102019114669A1 - Spindle motor with fluid dynamic bearing system - Google Patents

Abstract

The invention relates to a spindle motor which comprises a stationary motor component (10, 12, 16) and a motor component (14, 38) which is rotatably mounted relative to this by means of a fluid dynamic bearing system, the fluid dynamic bearing system at least one fluid dynamic radial bearing (24, 26) and has at least one fluid dynamic axial bearing (30). An electrical stator arrangement (42) is arranged on the stationary motor component. A rotor magnet (46) radially opposite the stator arrangement (42) is arranged on the rotatable motor component.
According to the invention, the ratio between the axial height h _{Stat of} the stator arrangement (42) and the axial height h _{Mag of} the rotor magnet (46) is selected to be less than 0.65.

Description

The invention relates to a spindle motor with a fluid dynamic bearing system, such as is used, for example, to drive hard disk drives, fans, laser scanners, etc.

A spindle motor of the type mentioned at the beginning essentially comprises a stator, a rotor and at least one fluid-dynamic bearing system arranged between these two parts. The rotor, driven by an electric motor, is rotatably mounted in relation to the stator with the aid of the fluid dynamic bearing system.

A known embodiment of a spindle motor with a fluid dynamic bearing system is shown in FIG DE 102 39 650 B3 disclosed. A single-plate design, which is frequently used in hard disk motors, is shown, the fluid-dynamic bearing system of which comprises two fluid-dynamic radial bearings and two fluid-dynamic axial bearings arranged in the area of a pressure plate. The spindle motor comprises a base plate into which a bearing bush is inserted. The bearing bush has an axial bearing bore for receiving a shaft. The shaft rotates freely in the fixed bearing bushing, with corresponding bearing surfaces of the shaft and bearing bushing being separated from one another by a thin, concentric bearing gap filled with a bearing fluid, with two fluid dynamic radial bearings being arranged along the bearing gap. A surface structure is incorporated into at least one bearing surface which, as a result of the relative rotational movement between the shaft and the bearing bush, exerts local acceleration forces on the bearing fluid located in the bearing gap. This creates a pumping effect on the bearing fluid located in the bearing gap, which leads to the formation of a homogeneous and uniformly thick lubricant film within the bearing gap, which is stabilized by zones of fluid dynamic pressure. A displacement of the rotor arrangement along the axis of rotation is prevented by appropriately designed fluid dynamic axial bearings. The fluid dynamic axial bearings are preferably formed by the two end faces of the pressure plate, which is preferably arranged at one end of the shaft, one end face of the pressure plate being assigned a corresponding end face of the bearing bush and the other end face of the pressure plate being assigned an inner end face of the cover plate. The cover plate forms a counter-bearing to the pressure plate and closes the relevant side of the bearing system; it also prevents air from entering the bearing gap filled with bearing fluid or the bearing fluid from leaking out. In the storage system shown, a liquid storage fluid, for example a bearing oil, is used. There is an electromagnetic drive system which consists of a stator arrangement arranged on the stationary part of the motor and a rotor magnet arranged on the hub.

In particular with known storage systems in single-plate design, it can happen that, especially with high loads and an unfavorable installation position, for example overhead installation, an unfavorable pressure distribution of the bearing fluid results within the bearing gap during operation. As a result, it can happen that the pressure plate is no longer positioned centrally in the associated recess, but shifts in the direction of the bearing bush or the cover plate and partially touches the bearing bush or the cover plate. This creates unwanted wear on the bearing components, which reduces the service life of the bearing. Furthermore, negative pressure zones can arise in the warehouse, which intensify this effect.

In addition, an axial preload force can occur through the electromagnetic drive system due to an axial displacement of the rotor magnet and the electric stator. It has been found that in particular the axial magnetic pretensioning force which is generated by the axial offset of the rotor magnet in relation to the stator arrangement is subject to very tolerances. Due to component tolerances, assembly tolerances, the use of a magnetic return ring and tolerances in the magnetization of the rotor magnet, there are tolerances in the magnetic preload force which further increase the effect of shifting the pressure plate from the central position.

It is the object of the invention to reduce the tolerances of the magnetic preload force in a spindle motor with a fluid dynamic bearing system.

According to the invention, this object is achieved by a spindle motor with a fluid dynamic bearing system using the features specified in claim 1.

Preferred configurations and advantageous developments of the invention are specified in the dependent claims.

The spindle motor comprises a stationary motor component and a motor component mounted rotatably relative to this by means of a fluid dynamic bearing system, the fluid dynamic bearing system having at least one fluid dynamic radial bearing and at least one fluid dynamic axial bearing. At the stationary motor component, an electrical stator arrangement is arranged. A rotor magnet radially opposite the stator arrangement is arranged on the rotatable motor component, the magnetic center of the rotor magnet by an amount V is arranged axially offset to the magnetic center of the stator assembly. This compensates for the loss of magnetic flux caused by the inwardly angled flux return ring, the goal being a balanced axial magnetic preload force F _V of 0 N is sought.

According to the invention is the ratio between the axial height h _Stat the stator assembly and the axial height h _Mag of the rotor magnet selected to be less than 0.65.

It was investigated which sizes and parameters have an influence on the tolerances of the magnetic preload force F _V to have.

Component tolerances and assembly tolerances, in particular the stator arrangement and the rotor magnet, are specified by specifications and can only be reduced with a disproportionately high effort. At least, however, their influence on the tolerances of the magnetic preload force should be reduced by suitable proportions of the component heights.

The use of a magnetic return ring for the rotor magnet leads to additional tolerances for the magnetic preload force F _V . A magnetic return ring is necessary if the hub is made of aluminum or another non-magnetic material. The use of a hub made of a magnetic material, such as. B. steel, does not require a return ring, so that additional tolerances can be avoided. However, it is not always possible to use a magnetic material for the hub.

The reduction in the radial air gap between the rotor magnet and the stator arrangement and a change in the outer diameter of the stator arrangement likewise have no significant influence on the tolerances of the magnetic preload force.

It was recognized that the axial height h _Mag of the rotor magnet has a great influence on the tolerances of the magnetic preload force F _V Has.

In particular, the relationship between the axial height h _Stat the stator assembly and the axial height h _Mag of the rotor magnet should be chosen so that it is less than 0.65. The axial height h _Stat the stator assembly is determined by the axial height of the magnetic core of the stator assembly.
In other words, it is advantageous if the axial height h _Mag of the rotor magnet is chosen to be significantly larger than the axial height h _Stat the stator assembly, thereby reducing the tolerances of the magnetic bias force F _V can be kept low.

The magnetic bias force F _V is essentially proportional to the square of the magnetic flux density B. the stator arrangement: $${\text{F.}}_{\text{V}}={\text{B.}}^2\times \text{S.}/\left(2\times {\mu }_0\right)$$

The magnetic flux density B. the stator arrangement depends on the cross section S. the magnetic core of the stator assembly.
The magnetic permeability becomes with the field constant µ ₀ considered.

The back EMF ( BEMF ) of a phase winding of the spindle motor results from the number of turns n the phase winding, the magnetic flux density B. and the cross section S. of the stator core: BEMF = n × B × S

Da die Toleranz der magnetischen Vorspannungskraft F_V umso größer wird je größer B_rel ist, ist es insbesondere vorteilhaft, das Verhältnis zwischen der axialen Höhe h_Stat der Statoranordnung und der axialen Höhe h_Mag des Rotormagneten kleiner als 0,65 zu wählen, wenn die relative Flussdichte Brei größer ist als 0,065 Vs / (rad m). In diesem Fall ergibt sich eine vertretbare Toleranz der magnetischen Vorspannungskraft.In order to be able to compare the BEMF value of different electrical stator arrangements with different numbers of turns and different heights of the stator cores with the same stator tooth width, a relative flux density B _{rel is} required.
As a relative flux density slurry depending on the number of turns n and the height h _Stat of the core of the stator arrangement can be defined: $${\text{B.}}_{\text{rel}}=\text{BEMF}/\left(\text{n}\times {\text{H}}_{\text{stat}}\right)$$

Because the tolerance of the magnetic bias force F _V the greater, the greater B _rel , the greater the ratio between the axial height h _Stat the stator assembly and the axial height h _Mag of the rotor magnet to be smaller than 0.65, if the relative flux density porridge is greater than 0.065 Vs / (rad m) In this case, there is an acceptable tolerance for the magnetic preload force.

• 1 zeigt einen Schnitt durch einen Spindelmotor, der mittels eines fluiddynamischen Lagersystems drehgelagert ist. Der Spindelmotor umfasst eine feststehende Basisplatte 10 mit einer Öffnung mit hochgezogenem Rand, in welcher eine zylindrische Hülse 16 befestigt ist. In der Hülse 16 ist eine Lagerbuchse 12 aufgenommen, die eine axiale zylindrische Lagerbohrung aufweist, in welcher eine Welle 14 drehbar aufgenommen ist. Zwischen dem Innendurchmesser der Lagerbohrung und dem etwas kleineren Außendurchmesser der Welle 14 ist ein Lagerspalt 18 vorgesehen, der mit einem Lagerfluid, beispielsweise einem Schmieröl, gefüllt ist. Einander zugeordnete Lagerflächen der Lagerbuchse 12 und der Welle 14 bilden zusammen zwei fluiddynamische Radiallager 24, 26, die durch entsprechende Radiallagerrillen 24a, 26a gekennzeichnet sind. Die Radiallagerrillen 24a, 26a sind auf der Oberfläche der Lagerbohrung und/oder der Oberfläche der Welle 14 angeordnet. Sobald sich die Welle 14 in der Lagerbuchse 12 um die Drehachse 40 dreht, üben die Radiallagerrillen eine Pumpwirkung auf das im Lagerspalt 18 befindliche Lagerfluid aus. Auf diese Weise entsteht im Lagerspalt 18 ein hydrodynamischer Druck, wobei sich ein homogener und gleichmäßig dicker Schmiermittelfilm innerhalb des Lagerspalts 18 ausbildet, der die Radiallager 24, 26 tragfähig macht. Solange sich die Welle 14 in der Lagerbohrung dreht, wird diese durch den durch die Radiallagerrillen 24a, 26a erzeugten fluiddynamischen Druck stabilisiert und läuft berührungslos in der Lagerbohrung getrennt durch den Lagerspalt 18. Die beiden Radiallager 24, 26 sind durch einen Bereich mit vergrößerter Lagerspaltbreite, dem sogenannten Separatorspalt 28 axial voneinander getrennt.

Description of a preferred embodiment of the invention:

• 1 shows a section through a spindle motor which is rotatably mounted by means of a fluid dynamic bearing system. The spindle motor includes a fixed base plate 10 with a raised rim opening in which a cylindrical sleeve 16 is attached. In the sleeve 16 is a bearing bush 12 added, which has an axial cylindrical bearing bore in which a shaft 14th is rotatably received. Between the inside diameter of the bearing bore and the slightly smaller outside diameter of the shaft 14th is a bearing gap 18th provided which is filled with a bearing fluid, for example a lubricating oil. Associated bearing surfaces of the bearing bush 12 and the wave 14th together form two fluid dynamic radial bearings 24 , 26th , by means of corresponding radial bearing grooves 24a , 26a Marked are. The radial bearing grooves 24a , 26a are on the surface of the bearing bore and / or the surface of the shaft 14th arranged. Once the wave 14th in the bearing bush 12 around the axis of rotation 40 rotates, the radial bearing grooves exert a pumping effect on the one in the bearing gap 18th located storage fluid from. This creates in the bearing gap 18th a hydrodynamic pressure, with a homogeneous and evenly thick lubricant film within the bearing gap 18th that trains the radial bearings 24 , 26th makes sustainable. As long as the wave 14th rotates in the bearing bore, this is caused by the radial bearing grooves 24a , 26a generated fluid dynamic pressure is stabilized and runs without contact in the bearing bore separated by the bearing gap 18th . The two radial bearings 24 , 26th are characterized by an area with an enlarged bearing gap width, the so-called separator gap 28 axially separated from each other.

The bearing groove structures 24a of the upper radial bearing 24 are preferably configured asymmetrically, ie they do not generate a uniform pumping action in both directions of the bearing gap, but a directed pumping action, which is predominantly downwards in the direction of the second radial bearing 26th is directed. The second radial bearing 26th includes bearing groove structures 26a , which are arranged symmetrically, for example, so that the second radial bearing 26th a uniform pumping action on the bearing fluid in both directions of the bearing gap 18th generated.

Due to the influence of the upper radial bearing 24 is a direction of flow of the bearing fluid in the bearing gap 18th given down.

At one end of the shaft 14th is a printing plate 20th arranged on the shaft 14th pressed on or alternatively in one piece with the shaft 14th is trained. The pressure plate 20th is rotatably received in a recess through the bearing bush 12 and the downwardly extended sleeve 16 is formed. The sleeve 16 is down through a cover plate 22nd locked. The two faces of the pressure plate 20th form together with opposite end faces of the bearing bush 12 or the cover plate 22nd two fluid dynamic thrust bearings 30th , 32 out. The thrust bearings 30th , 32 are characterized by thrust bearing grooves on the bearing surfaces of the pressure plate 20th or the bearing bush 12 or the cover plate 22nd are arranged. Once the wave 14th in the bearing bush 12 is set in rotation, a fluid dynamic pressure builds up in the radial sections of the bearing gap due to the axial bearing grooves 20th between pressure plate 20th and bearing bush 12 as well as printing plate 20th and cover plate 22nd so that the thrust bearings 30th , 32 become sustainable.

The two axial bearings 30th , 32 are characterized by bearing groove structures, which are preferably designed in the shape of a spiral groove or a herringbone.
It is preferred here if the bearing groove structures of the axial bearings 30th , 32 produce a uniform pumping action in both directions of the radial sections of the bearing gap.

The open end of the bearing gap 18th is through a seal, for example a capillary sealing gap 36 , sealed. The sealing gap 36 is formed by an outer peripheral surface of the shaft 14th and an inner peripheral surface of the bushing 12 . The inner circumferential surface is in the area of the sealing gap 36 beveled so that the sealing gap 36 has a substantially conical cross section.

The free end of the shaft protruding from the bearing bush 14th is with a hub 38 connected. The hub 38 is designed according to the purpose of the spindle motor. If the spindle motor is intended to drive a hard disk drive, the hub 38 one or more storage disks (not shown) of the hard disk drive arranged and secured.

On an inner lower edge of the hub is an annular rotor magnet 46 arranged with a plurality of pole pairs. The rotor magnet 46 is due to a return ring 44 at. Opposite the rotor magnet is on the base plate 10 a stator assembly 42 attached by a radial air gap from the rotor magnet 46 is separated. The stator assembly 42 has corresponding stator windings which, when energized accordingly, generate an alternating electric field, so that the rotor consists of the hub 38 and the wave 14th , is set in rotation.

The lower fluid dynamic thrust bearing 30th creates an axial force F ₁ which the pressure plate 20th and the associated wave 14th and the one with the wave 14th connected hub 38 pushes upwards, ie the rotor is removed from the base plate 10 pushed away.

The upper fluid dynamic thrust bearing 32 creates an axial force F ₂ which the pressure plate 20th and the associated wave 14th and the one with the wave 14th connected hub 38 pushes down, ie the rotor towards the base plate 10 presses.

The magnetic center of the stator assembly 42 is arranged axially offset by an amount v to the magnetic center of the rotor magnet, namely the magnetic center of the rotor magnet 46 further away from the surface of the base plate 10 as the magnetic center of the stator assembly 42 . This creates an axial restoring torque, ie an axial magnetic preload force F _V generated that is opposite to the axial force F ₁ of the lower fluid dynamic thrust bearing 30th and in the same direction as the force F ₂ of the upper thrust bearing 32 is directed. The axial offset v is, for example, 0.2 mm, which results in a balanced magnetic pretensioning force in the spindle motor shown F _V of 0 Newtons, because the inwardly overlapping return ring 44 part of the magnetic flux is reduced.

The sum of the axial force F ₂ of the thrust bearing 32 is preferably about the same as the axial force F ₁ of the thrust bearing 30th so that the pressure plate 20th essentially in the middle of the recess provided between the sleeve 16 , the bearing bush 12 and the cover plate 22nd is positioned.

The axial height h _Stat the magnetic core of the stator assembly 42 is selected according to the invention significantly smaller than the axial height h _Mag of the rotor magnet 46 , in such a way that the ratio between the axial height h _Stat the stator assembly and the axial height h _Mag of the rotor magnet is less than 0.65.
This can reduce the tolerance of the axial magnetic preload force F _V limited to less than +/- 0.2 N.

In the spindle motor shown as an example, the axial height of the magnetic core of the stator arrangement is, for example, h _Stat = 2.8 mm and the axial height of the rotor magnet, for example, h _Mag = 4.4 mm. The ratio between the axial height h _Stat the stator assembly 42 and the axial height h _Mag of the rotor magnet 46 is therefore 0.64.
The BEMF is, for example, 0.01 Vs / rad. The number of turns is 36.
This results in a relative flux density B _rel of approximately $${\text{B.}}_{\text{rel}}=\text{BEMF}/\left(\text{n}\times {\text{H}}_{\text{Stat}}\right)=0.01\text{Vs}/\text{wheel}/\left(36\times 0.0028\text{m}\right)\approx 0.1\text{Vs}/\left(\text{wheel m}\right)$$

List of reference symbols

10

Base plate

12

Bearing bush

14th

wave

16

Sleeve

18th

Bearing gap

20th

printing plate

22nd

Cover plate

24

Radial bearing

24a

Radial bearing grooves

26th

Radial bearing

26a

Radial bearing grooves

28

Separator gap

30th

Thrust bearings

32

Thrust bearings

36

Sealing gap

38

hub

40

Axis of rotation

42

Stator assembly

44

Return ring

46

Rotor magnet

V

Amount [mm]

F _V

Preload force [N]

F ₁

axial force [N]

F ₂

axial force [N]

h _Stat

Height of stator arrangement [mm]

h _Mag

Height of rotor magnet [mm]

B.

Flux density [T]

S.

Cross section stator core [mm ² ]

BEMF

Back EMF [Vs / (rad m)]

n

Number of turns

µ ₀

magnetic field constant

QUOTES INCLUDED IN THE DESCRIPTION

This list of the documents listed by the applicant was generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Patent literature cited

• DE 10239650 B3 [0003]

Claims (7)

Spindle motor with a stationary motor component (10, 12, 16) and a motor component (14, 38) rotatably mounted relative to this by means of a fluid dynamic bearing system, the fluid dynamic bearing system at least one fluid dynamic radial bearing (24, 26) and at least one fluid dynamic axial bearing (30 ), wherein an electrical stator arrangement (42) is arranged on the stationary motor component and a rotor magnet (46) radially opposite the stator arrangement (42) is arranged on the rotatable motor component, characterized in that the ratio between the axial height h _{Stat of} the magnetic core of the stator arrangement ( 42) and the axial height h _{Mag of} the rotor magnet (46) is less than 0.65. Spindle motor after Claim 1 , characterized in that the relative flux density Mrei of the stator arrangement (42) is greater than 0.065 Vs / (rad m), where B _rel = BEMF / (n × h _Stat ). Spindle motor according to one of the Claims 1 or 2 , characterized in that a second fluid dynamic axial bearing (32) is provided, the axial force F _{2 of} which is directed in the opposite direction as the axial force F _{1 of} the first fluid dynamic axial bearing (30). Spindle motor according to one of the Claims 1 to 3 , characterized in that the magnetic center of the rotor magnet (46) is arranged axially offset by an amount V to the magnetic center of the stator arrangement (42), an axial magnetic biasing force F _{V being} generated which is opposite to the axial force F _{1 of} the first fluid dynamic Axial bearing (30) is directed Hard disk drive with a spindle motor according to one of the Claims 1 to 4th . Fan with a spindle motor according to one of the Claims 1 to 4th . Laser scanner with a spindle motor according to one of the Claims 1 to 4th .

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