JP2003061298A - Motor having hemispherical fluid or gas dynamic pressure bearing balanced with magnetic attractive force at axial end - Google Patents

Abstract

PROBLEM TO BE SOLVED: To achieve and provide a hemispherical fluid and gas dynamic pressure bearing motor that is suited for thinning and reduction in current, has a simple structure, and can reduce costs, by solving the problems of realization of stability in a rotary attitude, a structure that can be assembled easily, an the like that are the problems in the hemispherical fluid and gas dynamic pressure bearing motor. SOLUTION: The motor comprises a shaft where at least one side is hemispherical, a sleeve having a recess that opposes the shaft, lubrication fluid in the gap between the shaft and the sleeve, and a magnetic means that has either a magnet or a magnetic body in the shaft or on the bottom surface of the sleeve and generates magnetic attraction force between a shaft end and a sleeve bottom surface. The motor has a dynamic pressure groove on the hemispherical bearing surface of the shaft or the sleeve, and balances the axial component force in load capacitance that the dynamic pressure groove generates in rotation and the magnetic attraction force for supporting the rotary section, inhibits NRRO for thinning, reduces current, and reduces costs.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a fluid and gas dynamic pressure bearing motor, and more particularly to a fluid and gas dynamic pressure bearing motor which has a hemispherical bearing portion and enables reduction in thickness and cost.

[0002]

2. Description of the Related Art In a rotary type memory device, a cooling fan, etc., a fluid dynamic bearing motor is being adopted due to a demand for quietness or suppression of NRRO (asynchronous shaft runout) of a rotating body. At the same time, along with the development of portable applications, there is a strong demand for thinner devices and lower currents. However, from the viewpoint of suppressing NRRO, the fluid dynamic bearing is difficult to reduce the span between the bearings that support the shaft, and has a limit in thinning. In addition, machining accuracy of sub-micron or less is required to maintain the bearing gap, and cost is reduced. Is in a difficult situation.

In order to make the fluid dynamic pressure bearing thinner, it is possible to separate it from the bearing supported at two points in the axial direction, to reduce the current, the structure in which the bearing sliding area can be reduced, and at a lower cost. In order to realize this, it is necessary to realize a structure that can maintain the bearing gap with the required accuracy even if the processing accuracy of the members is relaxed.

The candidate hemispherical fluid dynamic bearing has been attracting attention since it can support loads in the radial and thrust directions. However, a single hemispherical bearing suitable for thinning is not widely used although there is an example of the gas bearing shown in USP058545424. One of the reasons is that the magnetic attraction force that balances with the hemispherical dynamic pressure bearing is obtained from between the rotor magnet and the stator core, so that the posture restoring force is impaired, and the bearing surface slides when starting and stopping. Therefore, it has a wear problem.

[0005]

The problems with a fluid and gas dynamic pressure bearing motor in the shape of a hemisphere are to realize a stable rotational posture and a structure with less sliding wear. This is the realization of a structure that does not leak easily. An object of the present invention is to solve the above problems and to provide a hemispherical fluid and gas dynamic pressure bearing motor that is suitable for thinning and low current, and has a simple structure and can reduce cost.

[0006]

A fluid dynamic bearing motor according to the present invention comprises a substantially hemispherical shaft, a sleeve having a recess facing the hemispherical shaft, a lubricating fluid in the shaft and a gap between the sleeves, and a shaft end. And a magnetic means for generating a magnetic attraction force between the tops of the sleeve and a taper seal portion which is arranged on the outer peripheral portion of the shaft and faces the outer peripheral wall of the sleeve and has a gap gradually increasing toward the opening. It is composed of an annular barrier and the like. The fluid dynamic bearing motor has a dynamic pressure groove on the hemispherical bearing surface of the shaft or sleeve, and balances the axial component force of the load capacity generated by the dynamic pressure groove during rotation with the magnetic attraction force. Characterized by supporting the rotating part. As a structure in which the boundary surface of the lubricating fluid is arranged on the outer circumference of the sleeve, a stable lubricating fluid sealing structure is realized even at high speed rotation.

A ring-shaped member is fixed to the end portion of the annular barrier arranged on the outer peripheral portion of the shaft, and the inner peripheral portion of the ring-shaped member is arranged so as to be in the annular recess of the outer peripheral wall of the sleeve, and the rotary portion is moved in the axial direction. Regulate the amount. This structure prevents the rotating part from coming off when an excessive impact is applied.

Further, it has been proposed that the electrically conductive magnetic fine powder is mixed in the lubricating fluid and restrained in the magnetic field between the shaft end and the sleeve top to constitute the electrical grounding means of the rotating portion.

A gas dynamic pressure bearing motor according to the second aspect of the present invention is a magnetic bearing pressure motor that produces a magnetic attraction force between a shaft having a substantially hemispherical shaft, a sleeve having a recess facing the hemispherical shaft, and a shaft end and a sleeve top. And has a dynamic pressure groove on the hemispherical bearing surface of the shaft or sleeve, and rotates by balancing the axial component force of the load capacity generated by the dynamic pressure groove during rotation with the magnetic attraction force. Characterized by supporting the club.

A ring-shaped member is fixed to the end surface of the sleeve,
The free end portion side of the ring-shaped member is arranged so as to be in the annular recess portion on the fixed portion side, and the amount of axial movement of the rotating portion is regulated. This structure prevents the rotating part from coming off when an excessive impact is applied.

In the above fluid and gas dynamic pressure bearings, the magnetic means for generating a magnetic attraction force is composed of a permanent magnet and a magnetic body which are respectively disposed inside the shaft or at the opposite sleeve top. The magnetic attraction force at the shaft end forms the posture restoring force together with the load capacity generated by the dynamic pressure groove, and the posture stability can be improved.

There is a permanent magnet movably held in the shaft with a holding force larger than the magnetic attraction force, and during assembly, the permanent magnet is sufficiently projected from the shaft end and, thereafter, between the shaft and the sleeve is more than the magnetic attraction force. Force is applied to determine the position of the permanent magnet in a state where the sleeve top or the leaf spring arranged on the sleeve top is elastically deformed, and the sleeve top or leaf spring and the end of the permanent magnet make contact when stationary and separate when rotating. The separation distance is smaller or equal to the axial levitation distance between the shaft and the bearing surface of the sleeve. This can reduce the sliding of the shaft and the sleeve on the bearing surface at the time of starting and stopping, and improve the reliability.

Further, dynamic pressure grooves having different angular lengths in the circumferential direction are arranged at substantially the same axial position on both surfaces of the shaft and the sleeve bearing facing each other, and the delay amount from the minimum clearance point in the circumferential direction to the maximum pressure point is dispersed. We have proposed a structure that avoids unstable phenomena such as half whirl.

According to the fluid and gas dynamic pressure bearing motor of the present invention, the load capacity generated by rotation is perpendicular to the hemispherical bearing surface, and the shaft and the sleeve are positioned at a position where the axial component force and the magnetic attraction force are balanced. Rotates without contact. The radial component force of the load capacity is balanced at each point in the circumferential direction and contributes to the centering of the rotating portion.

The problems with the conventional single spherical bearing are attitude stability and sliding wear of the bearing surface. In the present invention, the magnetic attraction force at the shaft end forms the load capacity and the attitude restoring moment force generated by the dynamic pressure groove, and the dynamic pressure grooves having different circumferential angular lengths on both sides of the bearing reduce the half whirl. Can be expected to improve posture stability.
In addition, the sliding of the bearing surface, which tends to occur when starting and stopping, is reduced by the protruding portion at the shaft end, which improves the life of the bearing.

[0015]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments and principles of a fluid and gas dynamic pressure bearing motor according to the present invention will be described below with reference to the drawings.

FIG. 1 shows a sectional structure of a fluid dynamic bearing motor according to a first embodiment of the present invention. The shaft 11 has a hemispherical side surface, and the sleeve 12 disposed so as to face the shaft 11 has a hemispherical concave surface. The gap between the shaft 11 and the sleeve 12 is filled with magnetic fluid oil that is a lubricating fluid,
The gap between the annular barrier 23 provided on the outer peripheral portion of the sleeve 12 and the outer peripheral wall of the sleeve 12 gradually increases in the axial direction to form a tapered seal portion and has a boundary surface 17 for the lubricating fluid. The shaft 11 has a permanent magnet 35 therein and is made of a magnetic material.
A magnetic attraction force is generated between the two tops. Rotating part is axis 1
1, annular barrier 23, hub 41, rotor magnet 44
Etc., the fixed portion is composed of the sleeve 12, the base 43, the stator core 47, the coil 50, and the like.

The hemispherical bearing surface 13 of either the shaft 11 or the sleeve 12 has a set of herringbone dynamic pressure grooves, which will be described later, to form a bearing portion. The set of dynamic pressure grooves pumps the lubricating fluid toward its center to increase the pressure of the lubricating fluid. Since the resulting load capacity is inversely proportional to the gap between the shaft and the sleeve, the gap is determined so that the axial component force of the load capacitance and the magnetic attraction force are balanced,
The shaft 11 is aligned by the radial component force of the load capacity. Therefore, since the size of the load capacity is determined by the magnetic attraction force, the magnetic attraction force is set to generate sufficient load capacity to support the rotating part during rotation, and the gap has a value of about several microns.

The stator core 47 and the coil 50 cooperate with the rotor magnet 44 to rotate the rotating portion. In addition to this, a magnetic disk, an optical disk, or the like is mounted as a load on the rotating portion, and the force applied between the shaft 11 and the sleeve 12 differs depending on the installation form of the storage device such as upright or inverted. That is, the weight of the movable part is added to the magnetic attraction force in the upright state, and the weight of the movable part is subtracted from the magnetic attraction force in the inverted state. Taking these into consideration, the magnetic attraction force should be at least three times the weight of the moving part, and empirically stable posture can be obtained. Precession can be further compressed and posture stability can be increased by increasing the magnetic attraction force and balancing it with a larger load capacity, but on the other hand, the sliding friction at the time of start and stop can be increased to reduce the operating life. Things are known. It depends on the required rotation accuracy, but as a rough guide, in the case of a small magnetic disk device, it is set to generate a magnetic attraction force of about 5 times the weight of the moving part, which is the weight of the rotating part of the fluid dynamic bearing motor plus the load weight. To do.

FIG. 2 shows the shaft 11, the permanent magnet 35, and the sleeve 1.
2 shows a cross section and a magnetic flux path. The shaft 11 is made of a non-magnetic material, and a strong rare-earth permanent magnet 35 is arranged inside the shaft. Assuming that the diameter of the bearing is about 5 mm, the permanent magnet 3
The diameter of 5 can be assigned about 1-2 mm by design. The distance between the tip of the shaft 11 and the sleeve 12 is 20-3.
Since the gap is set to a minute gap of about 0 μm, the magnetic attraction force is almost constant regardless of the gap variation, and the working and assembly tolerance can be increased. The number 55 indicates the magnetization direction of the permanent magnet 35, and the tip of the permanent magnet 35 is a spherical surface to concentrate the magnetic flux, and the magnetic flux 56 flowing into the top of the sleeve 12 passes through the inside of the sleeve 12 as shown by the number 57 and is a hemispherical shape of the sleeve 12. It returns from the bearing surface to the other end of the permanent magnet 35 (number 58). Since the distance from the hemispherical bearing surface to the end of the permanent magnet 35 is large and the area is large, the magnetic flux 58 is dispersed and the magnetic flux density is small, and the hemispherical bearing surface of the sleeve 12 and the permanent magnet 35.
The magnetic attraction of and is small.

Since the magnetic fluid oil is used as the lubricating fluid, a centripetal force is exerted on the magnetic fluid oil in the vicinity of the top portion having a high magnetic flux density, so that the removal of bubbles from the top portion and the bearing portion can be promoted.
Normal oil will not hinder the functioning of the bearing.

In addition to the means for generating the magnetic attraction force, the rotor magnet 44 and the stator 47 are axially biased.
Alternatively, a magnetic piece may be arranged below the rotor magnet 44. The former has the disadvantage of inducing vibration, and the latter has the disadvantage of increasing the current consumption due to the eddy current generated in the magnetic piece. The magnetic attraction force generating means shown in the embodiment of the present invention can solve the drawbacks of the above means.

FIG. 3 shows the detailed structure of the bearing portion in the first embodiment. FIG. 3 (a) is a sectional view of the shaft 11 and the sleeve 12, and FIG. 3 (b) is a plan view of the sleeve 12. Shown respectively. In order to make it easy to understand the axial position of the dynamic pressure groove, the dynamic pressure grooves 20 and 21 on the surface of the shaft 11 are shown superimposed on the cross-sectional view, and the dynamic pressure grooves 27 and 28 of the bearing surface 13 of the sleeve 12 are shown in plan view. It was shown to. The dynamic pressure grooves 20, 21, 27, 28 are recesses of about several micrometers, and collect the lubricating fluid from the inner peripheral side and the outer peripheral side to the intermediate portion during rotation to increase the pressure of the lubricating fluid, and the shaft 11 to the sleeve 12. On the other hand, levitate and support.
In this embodiment, the pumping capacity from the outer circumference side to the inner circumference side is set to be slightly larger than that from the inner circumference side to the outer circumference side so that the pumping force to the inner circumference side remains, and the pressure of the lubricating fluid on the inner circumference side is set. Is rapidly increased at the start of rotation to reduce sliding friction between the shaft 11 and the sleeve 12. In the dynamic pressure groove shown in the figure, the groove length (20, 27) on the inner peripheral side is shown largely, but the pumping ability is determined by the degree of reduction of the circumferential length of the groove and the radial length of the groove. Therefore, it does not contradict the above description.

The taper seal portion of the lubricating fluid is the sleeve 1
Since it is placed on the outer circumference of 2, it is effective in reducing the thickness of the whole motor, and the axial space of the taper seal part can be sufficiently secured, so that the taper angle can be made a small angle of 10 degrees or less to realize a strong lubricating fluid seal structure. . Further, since the boundary surface 17 of the lubricating fluid is disposed between the outer peripheral wall of the sleeve 12 which is almost vertical and the annular barrier 23, there is little concern that the lubricating fluid leaks due to centrifugal force even during high speed rotation.

The dynamic pressure grooves 20, 21 on the surface of the shaft 11 and the dynamic pressure grooves 27, 28 on the surface of the sleeve 12 are characterized by different angular lengths in the circumferential direction. In this embodiment, the angular length of the dynamic pressure grooves 27, 28 on the surface of the sleeve 12 in the circumferential direction is determined by the dynamic pressure groove 2 on the surface of the shaft 11.
It is set to more than twice as large as 0,21. The arrow number 30 indicates the direction of rotation of the shaft 11.

The dynamic pressure groove pumps the lubricating fluid at the time of rotation to increase the pressure, and the ability to increase the pressure generates a posture restoring force almost in inverse proportion to the bearing gap. Since the dynamic pressure grooves are distributed in the circumferential direction, even if the bearing gap becomes locally small due to eccentricity, there is a delay until the effect is reflected as the pressure distribution difference in the circumferential direction. It is proportional to the circumferential angular length of the dynamic pressure groove. It is known that a control system in which there is a delay from the change of the controlled variable to the control causes some kind of resonance phenomenon, and in the case of a fluid dynamic bearing, it causes instability phenomena such as precession and oil whip.

Therefore, in order to reduce this kind of instability phenomenon, it is necessary to appropriately disperse the delay amount, and for example, the dynamic pressure groove 21 is constituted by dispersing the length in the circumferential direction.
However, if the angular lengths of the dynamic pressure grooves, which are only a few per revolution, are dispersed, uneven posture restoring force and other disadvantages become apparent. In the present invention, the surface of the shaft 11 and the sleeve 12
Dynamic pressure grooves with different circumferential angular lengths are formed on the surface,
The problem was solved by realizing the uniformity of the posture restoring force due to the increased pressure in the circumferential direction and the coexistence of dynamic pressure grooves with different circumferential angular lengths. Machining the dynamic pressure groove is generally not easy, and configuring it on both sides of the bearing part increases the cost. In the present embodiment, since the hemispherical shaft 11 and the hemispherical sleeve 12 can be molded, it does not cause a cost increase and a fluid dynamic bearing motor in which precession hardly occurs can be realized.

The gap between the annular barrier 23 and the outer peripheral wall of the sleeve 12 is gradually increased in the axial direction to form a taper seal portion for sealing the lubricating fluid by surface tension. A ring-shaped member 24 is fixed to an end portion of the annular barrier 23, and an inner peripheral portion of the ring-shaped member 24 is located in an annular recess 26 provided on an outer peripheral wall surface of the sleeve 12 to limit the axial movement of the rotating portion. To do. The ring-shaped member 24 uses elasticity or is formed by notching a part of the ring and preliminarily rotatably fitted in the annular recess 26 at the time of assembly, and after the assembly, the access hole 25 provided in the member facing the annular barrier 23 is provided. The ring-shaped member 24 is fixed to the end of the annular barrier 23 by spot welding, bonding, or the like. There are three access holes 25 in the circumferential direction,
The ring-shaped member 24 is fixed uniformly at three points in the circumferential direction.

In FIG. 4, the pressure distribution in the lubricating fluid generated during rotation, the shaft 11 and the sleeve 12 as a result of the pressure distribution.
It will be explained that the restoring force of the rotational posture can be obtained by showing the load capacity appearing in between. FIG. 4 (a) also shows the pressure distribution during rotation in the above dynamic pressure groove structure. The number 73 indicates the coordinate in the axial direction, the number 74 indicates the pressure value, and the numbers 75 and 7
Reference numerals 6, 77, 78 and 79 represent average pressures in the circumferential direction at the respective axial positions. Since the atmospheric pressure is subtracted, the pressure value 75 at the outer peripheral portion shows zero, the pressure 76 increases due to the dynamic pressure groove 21, and becomes substantially constant at the central portion as indicated by reference numeral 77. The pressure decreases at a portion corresponding to the position of the dynamic pressure groove 20 as indicated by reference numeral 78, and becomes slightly higher than the atmospheric pressure at the top 14 as indicated by reference numeral 79.

Basically, the posture of the rotating portion is maintained by the high pressure 77 at the central portion, which will be described in detail with reference to FIG. 4 (b). Pressure distribution 7 shown in FIG.
5,76,77,78,79 are the average values in the circumferential direction,
When the shaft 11 and the sleeve 12 are eccentric or tilted, the pressure distribution in the circumferential direction also differs locally. FIG. 4B shows a situation where the rotating sleeve 12 is tilted to the left and is rotating. The load capacities generated by the grooves 20 and 21 are not uniform in the circumferential direction, and the load capacities 68 on the right side and 67 on the left side.
As a representative example, 67 on the side where the bearing gap is small becomes large. These load capacities 67 and 68 work perpendicularly to the bearing surface and generate a magnetic attraction force 83 at the shaft end and a moment force for restoring the posture. In this description, only the left and right moment forces are taken up for simplification, but in reality, the moment forces at each point in the circumferential direction and the axial direction are balanced.

The viscosity of oil used as a lubricating fluid is generally small at high temperatures, and the load capacity decreases. In the conventional design, the upper limit of the operating temperature range is set so that the load capacity can be secured with a margin, but as a result, at low temperatures, excessive load capacity and excessive current are plagued. According to the invention, axis 1
Since the gap between 1 and the sleeve 12 is determined at a position where the axial component force of the load capacities 67 and 68 and the magnetic attraction force are in equilibrium, the load capacity is kept substantially constant regardless of temperature. In other words, temperature compensation is performed automatically, and the load capacitance set value in the design can be kept constant over the entire temperature range. It is possible to avoid excessive load capacitance, excessive current, etc. at low temperatures, enabling low current design. Become.

Further, since the axial loss in the fluid dynamic pressure bearing is mainly caused by the frictional force between the lubricating fluid and the surface of the shaft 11, the sleeve 12, etc. in the narrow gap portion having the dynamic pressure groove, the axial loss in the dynamic fluid dynamic bearing is the same as that of the present invention. From this point as well, if the pressure groove has a minimum configuration of one set, the axial loss is small and a low current can be achieved.

FIG. 5 further illustrates that the magnetic attraction force at the shaft end is effective for restoring the rotational posture. FIG. 5A shows an embodiment having a magnetic attraction force at the shaft end, and FIG. 5B shows an example in which a rotor magnet is used as the magnetic attraction force generating means. In these drawings, an example in which the upper part of the shaft 11 is tilted to the left is shown in FIG.
As described in (b), the load capacities due to the generation of dynamic pressure are represented by the two left and right points and are shown by the numbers 67 and 68. The load capacity 67 with a small bearing gap is larger than the load capacity 68, and as described with reference to FIG. In FIG. 5A, the magnetic attraction force 8 at the shaft end is further added.
It is easy to see that 3 generates the load capacities 67 and 68 and the restoring moment force of the rotational posture. In the example of FIG. 5B in which one of the rotor magnets 44 is used as the magnetic attraction force generating means, the left and right magnetic attraction forces 84 and 85 are almost balanced,
Since the magnetic attraction forces 84 and 85 between the rotor magnet 44 and the magnetic piece 53 are inversely proportional to the gap, the magnetic attraction force 84 with a smaller gap becomes larger than the magnetic attraction force 85 and rather hinders the restoration of the rotational posture. Even a factor. In this way, the structure having the magnetic attraction force at the shaft end further contributes to the stabilization of the rotational posture.

6 and 7 show the detailed structure of the embodiment of FIG. 1 in which the shaft and the sleeve are not completely in contact with each other at the bearing surface when stationary. In FIG. 6, a movable permanent magnet 35 that contacts the top of the sleeve 12 when stationary is provided in the shaft 11, and a leaf spring 33 is provided at the top of the sleeve 12. The shaft 11a shown by the dotted line shows the position at rest, and the solid line shaft 11 shows the position at the time of rotation. The gap between the end of the permanent magnet 35 and the leaf spring 33 at the time of rotation is d, the sleeve 12, and the shaft 11 respectively. The amount of protrusion of the permanent magnet 35 is determined so that f is always equal to or larger than d, where f is the axial flying height on the bearing surface between them. As a standard, (f−d) is set to about 5 μm if the flying height f is in the range of 10−20 μm, because the flying height f of the sleeve 12 changes with temperature. In this way, at the time of rotation, the end of the permanent magnet 35 is floated by at least about 5 μm and the bearing surface is floated by 10-20 μm in the axial direction, which does not affect the rotational posture.

Generally, in the fluid dynamic bearing, there is a high possibility that the bearing portion will slide and wear at the time of starting and stopping, and the life or reliability will be impaired. In this embodiment, however, the end of the permanent magnet 35 and the leaf spring 33 are reduced.
However, if the end of the permanent magnet 35 and the opposing leaf spring 33 are provided with a ceramic material, a plating treatment, or other wear countermeasures, the performance can be stabilized for a long period of time.

FIG. 7 is a diagram for explaining the position adjustment of the permanent magnet 35 in FIG. The permanent magnet 35 is movably held by a holding force larger than the magnetic attraction force in the cylinder 32 in the shaft 11 by a clearance fit. At the time of assembly, the sleeve 12 is assembled by preliminarily projecting the permanent magnet 35 onto the shaft 11, and a pressing force larger than the magnetic attraction force is applied to the sleeve 12 and the shaft 1.
In addition to the interval 1, the top of the sleeve 12 is brought into contact with the leaf spring 33, and the pressing force is removed while the leaf spring 33 arranged on the top of the sleeve 12 is elastically deformed. The dotted line 11b indicates the shaft under the pressing force, and the shaft 11 shown by the solid line returns to the original shape of the leaf spring 33 after removing the pressing pressure, and a gap is provided on the bearing surface of the shaft 11 and the sleeve 12. The state is shown. The same effect can be obtained by utilizing the elastic deformation of the top portion of the sleeve 12 instead of utilizing the elastic deformation of the leaf spring 33.

If the permanent magnet 35 is adjusted in position and then fixed by adhesion or welding, it is possible to resist an excessive impact force.
The cylinder 32 serves as a through hole for holding a permanent magnet with a clearance fit or screwing, and the sleeve 1 is sufficiently protruded during assembly.
2 is brought into contact with the top 14 and then through the through hole (f-
d) It is possible to adjust and fix the position by projecting a considerable amount. The permanent magnet 35 shown in the embodiment of FIG.
Although it may be fixed to 1, the protrusion amount of the permanent magnet 35 and the diameter dimensions of the shaft 11 and the sleeve 12 must be controlled in that case. N as a fluid dynamic bearing motor
Dimensional management is easy when performance requirements such as RRO are relatively loose, but difficult when performance requirements are strict, and the structure for adjusting the position of the permanent magnet as in this embodiment is low in total cost. .

FIG. 8 shows an embodiment in which the axial position of the ring-shaped member can be adjusted, and FIG. 8 (a) is a sectional view of the bearing portion.
FIG. 8B is an enlarged view of the cross-section portion 89 around the ring-shaped member. In the embodiment, the end portion of the annular barrier 23 has the protrusion 8
6, the ring-shaped member 24 has a through hole that engages with it. The ring-shaped member 24 is fitted in the annular recess 26 on the outer peripheral portion of the sleeve 12 in advance and combined with the shaft 11, and the projection 86 and the ring-shaped member 24 are inserted through the access hole 25.
And the inner peripheral portion of the ring-shaped member 24 is abutted against the end portion 87 of the annular recess 26 by the jig 88, and the ring-shaped member 24 is elastically deformed and fitted and fixed to the protrusion 86. Let

In the assembly process, the ring-shaped member 24
The amount of elastic deformation in the axial direction is about 20 microns,
If the strength of the fitting between the ring-shaped member 24 and the protrusion 86 is sufficiently large, the amount of axial movement of the movable portion including the hub 12 is limited to a minute amount of about 20 μm even when an impact is applied. it can. In the example of HDD and the like, there is a strong demand to limit the amount of axial movement of the magnetic disk to a very small amount, but by using the elastic deformation of the ring-shaped member 24 as in this embodiment, the tolerance of each member is set strictly. The demand can be satisfied without doing. After the ring-shaped member 24 and the projection 86 are assembled, the joint strength can be further improved by means such as adhesion and welding to improve the impact resistance.

In the first embodiment of the present invention, the structure in which the joint between the members is eliminated at the portion in contact with the lubricating fluid is adopted. In the conventional structure, caulking, bonding,
Alternatively, although sealing is performed by laser welding or the like, a fatal failure was caused due to leakage of the lubricating fluid due to poor jointing during mass production. In the fluid dynamic bearing according to the present invention, as shown in the embodiment, the joint portion in contact with the lubricating fluid can be eliminated, and a fluid dynamic bearing motor without fear of oil leakage can be realized.

When conductive fine magnetic powder is mixed in the lubricating fluid, the shaft end where the magnetic field is concentrated and the sleeve can be bridged to each other to facilitate the conduction between the rotating portion and the fixed portion. .

FIG. 9 shows a sectional structure of a gas dynamic bearing motor according to a second embodiment of the present invention. The shaft 11 has a hemispherical shape, and the sleeve 12 arranged so as to face the shaft 11 has a hemispherical concave surface. The shaft 11 has a permanent magnet 35 therein and generates a magnetic attractive force between the shaft 11 and the top of the sleeve 12 made of a magnetic material. A bearing portion is formed on the hemispherical bearing surface 13 of either the shaft 11 or the sleeve 12 with a dynamic pressure groove. The dynamic pressure groove pumps air toward its center to increase the pressure of the air. Since the resulting load capacity is inversely proportional to the clearance between the shaft and the sleeve, the clearance is determined so that the axial component force of the load capacity and the magnetic attraction force are balanced, and the radial component force of the load capacity adjusts the shaft 11. The wick is done.
Therefore, since the magnitude of the load capacity is determined by the magnetic attraction force, the magnetic attraction force is set so that sufficient load capacity can be generated to support the rotating part during rotation.

Reference numeral 34 indicates a vent hole, which allows the air pressurized on the top portion 14 to leak to the outer peripheral portion of the bearing through the gap on the side surface of the permanent magnet 35. Adjust the flow path resistance by reducing the diameter of the vent hole 34 or filling it with fibrous material, porous material, etc.
By leaving the pressure of the top portion 14 to an appropriate level, the levitation at the time of startup is swiftly performed, and when impact vibration is applied, the pressurized air is released to control the degree of damping.

The rotating portion is composed of the sleeve 12, the rotor magnet 46, the magnetic disk 91, etc., and the stationary portion is the shaft 11,
It is composed of a base 43, a stator core 49, a coil 52 and the like. Reference numeral 90 is a magnetic material and constitutes a magnetic shield.

FIG. 10 shows the detailed structure of the bearing portion in the second embodiment shown in FIG. 9, and FIG. 10 (a) shows the sleeve 1
2 is a plan view, and FIG. 10 (b) is a shaft 11 and a sleeve 12
The cross-sectional structures of are shown respectively. As shown in FIG. 10A, the bearing surface of the sleeve 12 is provided with a set of herringbone-shaped dynamic pressure grooves 18, and the surface of the shaft 11 is provided with a set of dynamic pressure grooves 19 facing each other. The dynamic pressure grooves 18 and 19 are recesses of about several micrometers, and air is introduced from the inner peripheral side and the outer peripheral side at the center of the dynamic pressure grooves 18 and 19, that is, the dynamic pressure grooves 18 and 19 during rotation.
The air pressure is increased by gathering it at the bent portions of the sleeve 12 to levitate and support the sleeve 12 with respect to the shaft 11. In this embodiment, the pumping ability from the outer peripheral side to the inner peripheral side is set to be slightly larger than that from the inner peripheral side to the outer peripheral side so that the pumping force to the inner peripheral side remains, and the air pressure on the inner peripheral side is set. At the time of starting the rotation, the sliding friction of the shaft 11 and the sleeve 12 is reduced quickly to reduce the sliding friction. The circumferential angular length of the dynamic pressure groove 19 is about half that of the dynamic pressure groove 18. A ring-shaped member 24 that restricts the axial movement amount of the rotating portion is provided on the outer peripheral end of the sleeve 12.
Since other structures and functions are the same, the same members as those in the first embodiment are designated by the same reference numerals, and the description thereof will be omitted.

FIGS. 11 and 12 show the detailed structure of the second embodiment in which the shaft and the sleeve are not completely in contact with each other at the bearing surface when stationary. In FIG. 11, the shaft 11 has a movable permanent magnet 35 that contacts the top of the sleeve 12 when stationary. The sleeve 12a shown by the dotted line shows the position at rest and the sleeve 12 shown by the solid line shows the position at the time of rotation. The end of the permanent magnet 35 and the sleeve 1 at the time of rotation are shown.
The protrusion amount of the permanent magnet 35 is determined so that f is always equal to or larger than d, where d is the gap from the top of 2 and f is the axial levitation amount on the bearing surface between the sleeve 12 and the shaft 11. As a standard, (f−d) is set to about 5 μm if the flying height f is in the range of 10−20 μm, because the flying height f of the sleeve 12 changes with temperature. In this way, at the time of rotation, the end of the permanent magnet 35 is at least about 5 μm, and the bearing surface is 10
It floats by -20 microns and does not affect the rotation posture.

FIG. 12 is a diagram for explaining the position adjustment of the permanent magnet 35 in FIG. Cylinder 32 in shaft 11
The permanent magnet 35 is movably held by a holding force larger than the magnetic attraction force by the clearance fitting. At the time of assembly, the permanent magnet 35 is sufficiently protruded from the shaft 11 in advance to assemble the sleeve 12, and a pressing force larger than the magnetic attraction force is applied to the sleeve 12,
In addition to the space between the shafts 11, the shaft 12 is brought into contact with the top of the sleeve 12, and the pressing pressure is removed while the top of the sleeve 12 is elastically deformed. The dotted line 12b shows the sleeve under the pressing force, and the solid line shows the sleeve 12 after the pressing pressure is removed, and the top of the sleeve 12 returns to its original shape.
And a state in which a gap is provided on the bearing surface of the sleeve 12. The same effect can be obtained by utilizing the elastic deformation of the leaf spring instead of utilizing the elastic deformation of the top of the sleeve 12.

FIG. 13 shows an embodiment in which a ring-shaped member is provided at the end of the sleeve to regulate the axial position of the movable part.
FIG. 13A is a sectional view of the right half of the motor, and FIG. 13B is an enlarged sectional view 89 around the ring-shaped member. In this embodiment, the end of the sleeve 12 has a protrusion 86a and the ring-shaped member 24 has a through hole that engages with the protrusion. The ring-shaped member 24 is fitted in the annular recess 26 on the outer peripheral portion of the shaft 11 in advance and combined with the sleeve 12, and the access hole 25 is formed.
The protrusion 86a and the through-hole of the ring-shaped member 24 are engaged with each other via the jig, and the inner peripheral portion of the ring-shaped member 24 is pressed against the end 87 of the annular recess 26 by the jig 88, and the ring-shaped member 24 is elastically deformed. The protrusion 86a is fitted and fixed in this state.

In the assembly process, the ring-shaped member 24
The amount of elastic deformation in the axial direction is about 20 microns,
If the strength of the fitting between the ring-shaped member 24 and the protrusion 86a is sufficiently large, the amount of axial movement of the movable portion including the hub 12 is limited to a minute amount of about 20 μm even when an impact is applied. it can. In the example of HDD and the like, there is a strong demand to limit the amount of axial movement of the magnetic disk to a very small amount, but by using the elastic deformation of the ring-shaped member 24 as in this embodiment, the tolerance of each member is set strictly. The demand can be satisfied without doing. After the ring-shaped member 24 and the protrusion 86a are assembled, the joint strength can be further improved by means such as adhesion and welding to improve the impact resistance.

In the first and second embodiments, the shaft and sleeve bearings are made of stainless steel,
Materials conventionally used for fluid and gas dynamic pressure bearings such as copper alloys and ceramics can be used. Forming a thin film of nickel, titanium, DLC, molybdenum disulfide, or the like on one surface of the bearing surface is effective in reducing wear at the time of starting and stopping, and is desirable.

As to the method of manufacturing the bearing portion, as shown in the embodiment, the bearing member according to the present invention is not limited to the convex shaft, but the sleeve having the concave surface has a shape in which the inclination is expanded upward. It is easy to remove, and it is possible to perform simultaneous molding, including dynamic pressure grooves, using techniques such as pressing or injection molding. Therefore, it is possible to perform molding using ceramics, sintered alloys, etc., and molding using a resin material with excellent abrasion resistance such as PPS (polyphenylene sulfide resin) containing carbon fiber, which is suitable for low-cost manufacturing. ing.

[0051]

As described above with reference to the embodiments, according to the fluid and gas dynamic pressure bearing motor of the present invention, the bearing portion has the dynamic pressure groove on the hemispherical bearing surface and is enhanced. It has a simple structure that balances the load capacity due to the pressure of fluid or gas and the magnetic attraction force, and it has the problems of stabilizing the rotational posture, lowering the mass production cost by molding, and thinning the fluid and gas dynamic pressure bearing motor. , Temperature compensation of load capacity of rotating part support,
A low current can be realized at the same time, and the object of the present invention can be sufficiently achieved. In particular, it can be used for a small-sized magnetic disk device, a rotary storage device such as an optical disk device, a CPU cooling fan, and the like.

[Brief description of drawings]

FIG. 1 is a sectional view of a fluid dynamic bearing motor according to a first embodiment of the present invention.

FIG. 2 shows a magnetic attraction force generating means and a magnetic flux path in the embodiment shown in FIG.

3A and 3B show details of the bearing portion. FIG. 3A is a sectional view of a shaft and a sleeve, and FIG. 3B is a plan view of the sleeve.

FIG. 4 (a) shows the cross section and pressure distribution of the shaft and sleeve, and FIG. 4 (b) shows the load capacity when the shaft is eccentric.

FIG. 5 compares the posture restoring force between an embodiment having a magnetic attraction force generating means at the shaft end (FIG. 5A) and an example having a rotor magnet as the magnetic attraction force generating means (FIG. 5B). The figure for is shown.

FIG. 6 is a detailed cross-sectional view of a bearing portion in which a permanent magnet is movably provided in the shaft in order to regulate contact between the shaft and the sleeve when stationary.

FIG. 7 is a diagram for explaining a method of adjusting the position of the permanent magnet in FIG.

8A and 8B are views for explaining an embodiment in which the axial position of the ring-shaped member can be adjusted. FIG. 8A is a sectional view of a bearing portion, and FIG. 8B is an enlarged sectional view of the periphery of the ring-shaped member. And show it.

FIG. 9 shows a sectional view of a gas dynamic bearing motor according to a second embodiment of the present invention.

FIG. 10 shows the details of the bearing portion, FIG. 10 (a) is a plan view of the sleeve, and FIG. 10 (b) is a sectional view of the shaft and the sleeve.

FIG. 11 is a detailed cross-sectional view of a bearing portion in which a permanent magnet is movably provided in the shaft in order to regulate contact between the shaft and the sleeve when stationary.

FIG. 12 is a view for explaining a method of adjusting the position of the permanent magnet in FIG.

13A and 13B are views for explaining an embodiment in which the axial position of the ring-shaped member can be adjusted. FIG. 13A is a sectional view of the bearing portion, and FIG. 13B is an enlarged sectional view of the periphery of the ring-shaped member. And show it.

[Explanation of symbols]

11, 11a, 11b ... Shaft, 12, 12
a, 12b ... Sleeve, 13 ... Bearing surface,
14 ... Top part, 17 ... Boundary surface of lubricating fluid, 18 ... Dynamic pressure groove, 19 ... Dynamic pressure groove, 20, 21 ... Dynamic pressure groove,
23 ... annular barrier, 24 ...
Ring-shaped member, 25 ... Access hole,
26 ... Annular recesses, 27, 28 ... Dynamic pressure grooves, 30
... Sleeve rotation direction, 32 ... Cylinder, 33 ... Leaf spring, 34 ...
..Ventilation holes, 35 ... Permanent magnets, 41 ... Hubs,
43 ... Base, 44, 46
..Rotor magnets, 47,49..Stator cores, 50,52..Coils, 5
3 ... Magnetic piece, 55 ... Magnetization direction,
56, 57, 58 ... Magnetic flux, 67, 68 ... Load capacity, 73 ... Coordinates in the axial direction, 7
4 ... Pressure value, 75, 76, 77, 78, 79 ... Pressure, 83, 84, 85 ... Magnetic attraction force, 86 ...
..Projections at the end of the annular barrier, 86a ... Projections at the end of the sleeve, 87 ... Ends of the annular recess, 88 ...
Jig, 89 ... Cross-section around ring-shaped member, 90 ... Magnetic plate,
91 ... Magnetic disk

─────────────────────────────────────────────────── ─── Continuation of front page (51) Int.Cl. ⁷ Identification code FI theme code (reference) H02K 7/08 H02K 7/08 A 21/22 21/22 MF term (reference) 3J011 AA08 AA11 AA20 BA10 CA03 DA02 JA02 JA03 KA04 LA05 MA06 MA21 MA22 PA02 QA20 RA04 RA10 5H605 AA13 BB05 BB14 BB19 CC04 DD03 EB04 EB06 5H607 BB01 BB17 BB25 DD03 FF04 GG04 GG09 GG12 GG14 5H621 GA01 HH01 JK08 JK17 JK19 JK17 JK19 JK17 JK17 J19

Claims (16)

[Claims] 1. A shaft having a hemispherical shape on at least a side surface, a sleeve having a concave portion facing the shaft, a lubricating fluid in a gap between the shaft and the sleeve, and a magnet and a magnetic body in the shaft or on the sleeve top, respectively. And a magnetic means for generating a magnetic attraction force between the shaft end and the sleeve top, and a taper seal having a gap gradually increasing toward the opening, facing the sleeve outer peripheral wall, disposed on the shaft outer periphery. In a fluid dynamic bearing motor composed of an annular barrier forming a portion, a shaft or a sleeve has a dynamic pressure groove on a hemispherical surface, and the dynamic force groove generates an axial component force of the load capacity and A fluid dynamic bearing motor characterized by supporting a rotating part by balancing magnetic attraction force. 2. The fluid dynamic bearing motor according to claim 1, wherein the magnetic means is movably held in the shaft by a holding force larger than the magnetic attraction force and a sleeve-side magnetic body. And the permanent magnet at the end of the permanent magnet when in a stationary state, contacting the sleeve when stationary, and separating at the time of rotation so that the distance is not larger than the axial levitation distance between the shaft and the bearing surface of the sleeve. Fluid dynamic bearing motor characterized by enabling position adjustment and fixing 3. The fluid dynamic bearing motor according to claim 1, wherein the magnetic means is movably held in the shaft by a holding force larger than the magnetic attraction force and a sleeve-side magnetic body. The state in which the permanent magnet is sufficiently protruded from the shaft end and then a force greater than the magnetic attraction force is applied between the shaft and the sleeve to cause elastic deformation of the sleeve top portion or the leaf spring arranged on the sleeve top portion. The sleeve is assembled so that the position of the permanent magnet is fixed, and the top or leaf spring of the sleeve is in contact with the end of the permanent magnet when stationary, and is separated during rotation, and the distance is the axial levitation distance between the shaft and the bearing surface of the sleeve. Fluid dynamic bearing motor characterized by being configured so as not to become larger 4. The fluid dynamic bearing motor according to claim 1, wherein a dynamic pressure groove is provided on both surfaces of the hemispherical bearing of the shaft and the sleeve at positions facing each other in the axial direction, and the dynamic pressure on each surface is provided. Fluid dynamic bearing motor characterized in that the grooves have different angular lengths in the circumferential direction 5. The fluid dynamic bearing motor according to claim 1, further comprising a ring-shaped member fixed to the end of the annular barrier,
A fluid dynamic bearing motor characterized in that an axial movable distance of a rotating portion is limited between the ring-shaped member and an annular recess provided in an outer peripheral wall of the sleeve. 6. The fluid dynamic bearing motor according to claim 5, wherein the ring-shaped member is fixed to the end of the annular barrier by means such as fitting, bonding or welding, and is required for fixing work. A fluid dynamic bearing motor characterized in that an access hole is provided in a member on a fixed portion side or a rotating portion side facing the end portion of the annular barrier. 7. The fluid dynamic bearing motor according to claim 6, further comprising means for fitting and fixing the annular barrier end portion and the ring-shaped member, and the inner periphery of the ring-shaped member through the access hole. The part is pressed and elastically deformed, but is pressed against the end face of the annular recess to fit and fix the ring-shaped member to the end of the annular barrier, and the axial movement amount of the movable part is adjusted as the elastic deformation amount of the ring-shaped member. Fluid dynamic bearing motor characterized by setting 8. The fluid dynamic bearing motor according to claim 1, wherein magnetic fluid oil is used as the lubricating fluid to facilitate elimination of bubbles in the lubricating fluid. 9. The fluid dynamic bearing motor according to claim 1, wherein conductive magnetic fine powder is mixed in the lubricating fluid, and bridged between the shaft end and the top of the sleeve for electrical conduction. Fluid dynamic bearing motor characterized by 10. An axis having a hemispherical shape on at least a side surface,
It is composed of a sleeve having a concave portion facing the shaft, and a magnetic means having a magnet and a magnetic body inside the shaft or at the sleeve top, respectively, to generate a magnetic attraction force between the shaft end and the sleeve top. In a gas dynamic bearing motor,
The shaft or the sleeve has a dynamic pressure groove on the hemispherical surface, and the rotating portion is supported by balancing the axial component force of the load capacity generated by the dynamic pressure groove during rotation with the magnetic attraction force. Gas dynamic pressure bearing motor 11. The gas dynamic pressure bearing motor according to claim 10, wherein the magnetic means is movably held in the shaft by a holding force larger than the magnetic attraction force and a sleeve-side magnetic body. And the permanent magnet at the end of the permanent magnet when in a stationary state, contacting the sleeve when stationary, and separating at the time of rotation so that the distance is not larger than the axial levitation distance between the shaft and the bearing surface of the sleeve. Gas dynamic pressure bearing motor characterized by enabling position adjustment and fixing 12. The gas dynamic pressure bearing motor according to claim 10, wherein the magnetic means is movably held in the shaft by a holding force larger than the magnetic attraction force and a sleeve-side magnetic body. The state in which the permanent magnet is sufficiently protruded from the shaft end and then a force greater than the magnetic attraction force is applied between the shaft and the sleeve to cause elastic deformation of the sleeve top portion or the leaf spring arranged on the sleeve top portion. The sleeve is assembled so that the position of the permanent magnet is fixed, and the top or leaf spring of the sleeve is in contact with the end of the permanent magnet when stationary, and is separated during rotation, and the distance is the axial levitation distance between the shaft and the bearing surface of the sleeve. Gas dynamic pressure bearing motor characterized by being configured so as not to become larger 13. The gas dynamic pressure bearing motor according to claim 10, wherein dynamic pressure grooves are provided at axially opposite positions on both bearing surfaces of the shaft and the sleeve, and the dynamic pressure grooves on each surface are Gas dynamic pressure bearing motor characterized by different circumferential angular lengths 14. The gas dynamic pressure bearing motor according to claim 10, wherein the motor rotates between a ring-shaped member fixed to the sleeve end and an annular recess provided on the fixed side corresponding to the ring-shaped member. Gas dynamic bearing motor characterized by limiting the axial movable distance 15. The gas dynamic pressure bearing motor according to claim 14, wherein said ring-shaped member is fixed to said sleeve end by means such as fitting, adhering or welding, and access required for fixing work is performed. A gas dynamic pressure bearing motor characterized in that a hole is formed in a member on the fixed portion facing the end of the sleeve. 16. The gas dynamic bearing motor according to claim 15, wherein the sleeve end and the ring-shaped member have means for fitting and fixing each other, and the free end side of the ring-shaped member is provided through the access hole. Is pressed and elastically deformed to abut the end face of the annular recess to fit and fix the ring-shaped member to the sleeve end, and the axial movement amount of the movable part is adjusted and set as the elastic deformation amount of the ring-shaped member. Gas dynamic pressure bearing motor characterized by