JP2001069719A - Dynamic pressure bearing motor - Google Patents

Abstract

PROBLEM TO BE SOLVED: To obtain a thin dynamic pressure bearing which has an aerodynamic pressure radial bearing, displays an excellent performance, and has a long life, when the motor is used for shaft-horizontal posture. SOLUTION: An aerodynamic pressure bearing is formed by providing a shaft 1 arranged with its shaft center in a horizontal posture, and having groove patterns 1a, 1b carved in its periphery, and fitting a bearing bush 2 at a specified gap freely-rotatably. One side of the aerodynamic pressure bearing is made to be an air intake side, and the other side is a tightly sealed side. The groove patterns 1a, 1b are carved in the shaft 1 so as to incline as against the shaft center direction of the shaft 1 at least on the air taking-in side.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a motor having a dynamic pressure radial bearing, for example, a motor for rotating a disk, a motor for driving a polygon mirror, etc., which can be driven continuously for a long time at high speed and high rotational accuracy.

[0002]

2. Description of the Related Art Generally, a thin and thin motor, in which the motor is mounted in a horizontal position, and a quartz disk or the like that drives a quartz disk or the like continuously for a long time with high rotation accuracy and long time, has a motor mounted in a horizontal position. Since the bearing may be in a posture, a ball bearing is often used as the bearing.

[0003] In this case, since the bearing is a ball bearing, there are great restrictions in terms of speeding up, high rotational accuracy, low vibration, low noise, and bearing life in long-time continuous driving. Therefore, if the bearing is a dynamic pressure radial air bearing, these quality characteristics can be greatly improved, but in that case, it is a major obstacle that the motor should be thin and the motor should be mounted horizontally. It has become.

For example, an example in which a dynamic pressure radial air bearing is used as a bearing and applied to a scanner motor for driving a polygon mirror will be described with reference to FIG.

[0005] Groove patterns 71a, 71b, 71c, 7 constituting two sets of herringbone grooves for air dynamic pressure on the outer peripheral surface.
The shaft 71 engraved with 1d is rotatably inserted into the bearing bush 72 in a shaft vertical posture to form a dynamic pressure air bearing. The bearing bush 72 is held by a bearing support 73. The bearing support 73 and the mounting flange 85 are integrated, and a fixed yoke 84 is provided on the mounting flange 85.
And the substrate 82 are fixed by screws.
, A coil 81 is fixedly bonded.

A rotor magnet 80 and a rotor yoke 79 provided at a position facing the coil 81
1 is screwed and fixed to a flange 88 that is shrink-fitted and fixed to 1 and rotates together with the shaft 71. Further, a polygon mirror 75 is fixed to the shaft at the upper end of the shaft by a mirror retainer 74b. A thrust receiver 83 is provided at the lower end of the shaft.

[0007]

However, the hydrodynamic bearing motor having the above-described structure has the following disadvantages.

(A) It is difficult to adapt the motor, which is increasingly required in information equipment and the like, to a thinner motor.

That is, for example, in the case of a motor for a polygon mirror scanner used in a laser printer,
The arrangement of the optical systems involved in the arrangement is almost in a planar state. In addition, since the laser printer itself aims to be thinner, the optical system is also required to be thinner, and accordingly, the motor used for it is also desired to be as thin as possible.

In the above-described dynamic air bearing, a polygon mirror or a disk is fixed at a position overhanging from the bearing bush.
A large bearing load capacity is required, and a large shaft diameter (1
0 mm) to improve the efficiency of the bearing. Therefore, a structure is formed in which the rotating body is supported by two bearings that satisfy a relation of shaft diameter ≦ bearing length.

As a result, a bearing length for two bearings is required, and if the space for fixing the thrust receiving space and the polygon mirror or the disk is added, the total length of the motor height is inevitably increased. . It is difficult for them to meet the above-mentioned requirement of reducing the motor height to 20 mm or less.

(B) When the motor is used in the horizontal position, the weight of the rotating body is also applied to the radial load.
It is necessary to make the bearing with a load capacity that can withstand this, but the structure as it is does not sufficiently cope with the usage in such a posture, and in many cases it can not withstand long-time use In particular, a serious problem occurs when starting and stopping the motor.

(C) The fixed yoke type thin motor is
The thrust receiving structure and the motor structure itself are simple, and the ratio of the bearing length to the motor thickness can be large.
The large thrust load due to the magnetic attraction force of the fixed yoke shortens the bearing life, and the increase in power consumption due to eddy current loss increases the temperature rise of the motor. The range of application is limited as a high-speed, high-performance dynamic pressure air bearing in which the bearing performance changes greatly with time.

The present invention has been made based on the above-mentioned circumstances, and when a motor having a dynamic pressure air radial bearing is used in a horizontal position of a shaft, the motor is thin and has good performance and a long life. It is intended to provide.

[0015]

According to the means of the present invention, a thrust bearing is provided at one end and a cylindrical bearing bush and a shaft are hermetically sealed on the thrust receiving side. A bush is rotatably arranged through a predetermined gap, and a thrust bearing is formed by the shaft tip and the thrust bearing. The shaft at the other end of the bearing bush in a portion where the shaft faces the bearing bush. A dynamic pressure bearing motor comprising a dynamic pressure radial bearing by providing a groove pattern inclined with respect to the axial direction on either the outer peripheral surface or the inner peripheral surface of the bearing bush.

Further, according to the means of the present invention, of the portion where the shaft and the bearing bush oppose each other, either one of the shaft outer peripheral surface on the one end side of the bearing bush or the bearing bush inner peripheral surface is provided with the second surface. A dynamic pressure bearing motor characterized in that the groove pattern is provided.

According to the means of the present invention, the second
Wherein the groove pattern is formed in parallel with the axial direction of the shaft.

According to the invention, there is provided a hydrodynamic bearing motor, wherein the thrust receiver is a member coated with resin, metal, ceramic or the like.

According to the means of the present invention, at least one of the shaft and the bearing bush is made of iron-based steel.
A hydrodynamic bearing motor comprising a nickel-based alloy of Invar alloy.

Further, according to the means of the present invention, the shaft is a fixed shaft arranged in a horizontal direction, and a ring-shaped magnetic material is provided on the outer peripheral surface of the bearing bush. And a magnet for generating a magnetic attraction force with the magnetic material is provided above the fixed shaft arranged in the horizontal direction at a position facing the portion.

Further, according to the means of the present invention, the shaft is a fixed shaft arranged in a horizontal direction, and a ring-shaped second shaft which is magnetized on the outer peripheral surface of the bearing bush so as to have an NS pole in the axial direction. A second magnet provided with the first magnet and magnetized so as to have an NS pole in the axial direction at a position excluding a predetermined range below the shaft on the outer periphery of the ring-shaped magnet and shifted in the axial direction. And a dynamic pressure bearing motor.

Further, according to the means of the present invention, the displacement amount of both magnets is about one third of the axial length of the first or second magnet. is there.

According to the means of the present invention, the second
Of the magnet at 30 degrees to 45 degrees below the axis
A hydrodynamic bearing motor characterized in that a portion corresponding to a degree is formed in a substantially C shape with a cutout.

[0024] According to the measure according to the invention, in the formula, bearing number Λ = (6μω / pa) × (R / Cr) 2 However, mu is the viscosity count of the air, omega is angular velocity, _{p a} is the air Specific gravity (atmospheric pressure, normal temperature), R is the radius of the shaft, Cr is the radial gap. If Λ ≦ 4.8, the rotating body is cantilevered by one hydrodynamic air bearing. If Λ ≧ 4.8, A dynamic bearing motor is characterized in that a rotating body is cantilevered by a pair of dynamic pressure air bearings.

Further, according to the means of the present invention, the dynamic pressure air bearing has a relationship between the total bearing length L and the shaft diameter D when the rotating body is used in a cantilevered state with the shaft in a horizontal position. However, when 1.2 ≧ L _T / D, the groove pattern is provided symmetrically with respect to the center of the bearing length, and the axial length of one groove is 1/3 or less of the bearing length. 4 or more, the number of grooves is 10 or more and 15 or less, the inflow angle of the grooves is 20 ° or more and 30 ° or less, and the relationship between the groove width and the ridge width is the ratio of the groove width to the ridge width. Is 3: 7 to 4: 6.

According to the means of the present invention, in the dynamic pressure air bearing, the groove pattern for air dynamic pressure provided on the outer peripheral surface of the shaft is provided at a base side of the shaft from a center of a bearing length. , The axial length of the groove pattern is 1/1 of the bearing length.
2 or less, 1/3 or more, and the number of grooves is 15 for 10 or more.
The groove inflow angle is not less than 20 ° and not more than 30 °, and the relationship between the groove width and the ridge width is such that the ratio of the groove width to the ridge width is 3: 7 to 4: 6. Dynamic bearing motor.

Further, according to the means of the present invention, in the dynamic pressure air bearing, the groove pattern for air dynamic pressure provided on the outer peripheral surface of the shaft is arranged such that a root side of the shaft from a center of the bearing length and a bearing length are provided. The axial length of the groove is 1/3 or less and 1/4 or more, respectively, the number of grooves is 10 or more and 15 or less, and the groove width is Between the groove width and the ridge width is 3: 7 to
4: 6, the inflow angle of the groove is 20 ° or more and 30 ° or less, and the direction of the inflow angle is the same, or
A hydrodynamic bearing motor characterized in that only the inflow angle of a groove provided on the shaft tip side is parallel to the axis.

According to the means of the present invention, one end of the shaft is engaged with the shaft fixed to the shaft support at a predetermined interval to form a dynamic pressure air bearing with the shaft. In a hydrodynamic bearing motor including a sealed bearing bush and a rotating body fixed to the bearing bush, the rotating body has a disk-shaped rotor magnet and a magnetic focusing yoke facing each other, and both are the same. A rotating axial magnetic gap rotating yoke structure is formed, and the dynamic pressure air bearing communicates with a gap formed between the rotor magnet and a coil fixed to a substrate facing the rotor magnet. The magnetic focusing yoke and an air inflow path formed by a gap formed along the outer surface of the shaft support, and all of the magnetic focusing yoke except for the air inflow path are sealed. It is dynamic bearing motor, characterized in being formed by granulation.

[0029] According to the means of the present invention, the closed structure is provided at a front end portion of the shaft opposite to a side supported by the shaft support, at a central portion having substantially the same diameter as the diameter of the shaft. A disk having a protrusion having a height of about 0.5 mm to 1.0 mm and a disk made of an elastic material having substantially the same axial diameter as that of a disk are combined with the protrusion being on the inside and tightening the connection screw. Accordingly, a dynamic pressure bearing motor is provided with a thrust receiving holder having a structure capable of fixing a two-layer disk at an arbitrary position in a bearing bush.

Further, according to the means of the present invention, the thrust receiving holder has a polyimide resin,
A hydrodynamic bearing motor comprising a thrust bearing provided with a sliding bearing material in which iron, metal or ceramic is coated with Teflon (registered trademark) resin.

According to the means of the present invention, the bearing bush is made of a hard-to-work material, and the fixing means for directly inserting and fixing the inner diameter of a precision component, which is difficult to be post-processed, to the outer diameter of the bearing bush. Is a thin ring on the outer periphery of the bearing bush and 0.5-1.
Two parts having a recess of about 0 mm are mounted on top of each other with the recess side being the flange side, and the two parts are connected with several screws and pressed against the end face of the precision part. ,
By tightening the screw, the flange-side ring is fastened and fixed to the outer peripheral portion of the bearing bush at that position, and furthermore, the elastic component of the precision component directly inserted into the outer periphery of the bearing bush by elastic deformation of the flange-side ring. A hydrodynamic bearing motor having a structure in which an end face is pressurized and fixed.

According to the means of the present invention, the shaft is
From the end face of the hydrodynamic air bearing on the side supported by the shaft support, to the surface within a range of about の of the bearing length,
At an angle of 0 ° or more and 30 ° or less, the depth extending obliquely in the rotation direction is substantially equal to the radial gap, and the ratio of the groove width to the ridge width is 3: 7 to 4: 6. A hydrodynamic bearing motor comprising 15 or less.

According to the means of the present invention, the shaft is
From the end face of the hydrodynamic air bearing on the side supported by the shaft support, to the inner surface of the bearing bush within a range of about 1/4 of the bearing length, in an anti-rotation direction at an angle of not less than 20 ° and not more than 30 ° with the end face of the bearing. The depth that extends obliquely is almost equal to the radius gap,
10 grooves having a groove width to ridge width ratio of 3: 7 to 4: 6
A hydrodynamic bearing motor characterized in that at least 15 and at most 15 are provided.

According to the means of the present invention, there is provided a hydrodynamic bearing motor, wherein an end face of the coil fixedly provided on the substrate is a flat end face made of resin and molded.

Further, according to the means of the present invention, a gap formed between one end face of the coil or the substrate and the end face of the rotor magnet is formed between the other end face of the coil or the end face of the substrate and the magnetic face. A hydrodynamic bearing motor characterized in that it is smaller than a gap formed by an end face of a converging yoke.

Further, according to the means of the present invention, the rotator is made of a non-magnetic material except for the rotor yoke and the magnetic converging yoke, and a part of the outer peripheral portion of the rotator is made of a magnetic material. A hydrodynamic bearing motor characterized by having a ring made of

According to the means of the present invention, the rotator is made of a non-magnetic material except for the rotor yoke and the magnetic converging yoke, and a part of the outer periphery of the rotator is made of a magnetic material. A dynamic pressure bearing motor characterized in that a magnet is fixedly provided within a range where a magnetic attraction effect is exerted on the ring above the ring when the ring is used and the shaft is in a horizontal attitude. is there.

According to the means of the present invention, there is provided a hydrodynamic bearing motor characterized in that the magnetic center of the magnet is provided at a position shifted toward the base of the shaft from the center of the axial length of the ring. .

[0039]

Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 is a sectional side view of a thin dynamic pressure bearing motor for a long-time continuous operation used in a disk drive according to a first embodiment of the present invention.

The shaft 1 having the first and second groove patterns 1a and 1b for the air dynamic pressure bearing formed on the outer peripheral surface has the shaft center fixed to the shaft support 3 in the horizontal direction. The other side of the shaft 1 is in contact with the thrust receiver 13. Furthermore, a bearing bush 2 made of an invar-type alloy of an iron-nickel alloy containing a small amount of manganese, silicon, and carbon is rotatably fitted to the shaft 1 to form a rotating body. The bearing bush 2 includes a rotor yoke 9 having a rotor magnet 10 fixed to one end of the shaft 1 at the base (shaft support 3) side. On the other hand, the coil 11 and the substrate 12 are fixed to the shaft support 3 together with the fixed yoke 14 at a position facing the rotor magnet 10.

Further, on the opposite side of the bearing bush 2, a disk 5 and a disk holding and thrust receiving holder 4 are fixed, and a shaft-fixed type dynamic pressure air bearing is formed. The disc holding / thrust receiver 13 is formed by coating resin, metal, ceramic, or the like with resin. A ring-shaped magnetic body 6 is provided on the outer peripheral portion of the bearing bush 2, and a magnet 7 is supported by a magnet holder 8 at a position facing the upper side of the ring-shaped magnetic body 6. The magnet holder 8 is fixed to the mounting plate 15.

It should be noted that the ring-shaped magnetic body 6 and the fixedly installed magnet 7 pass through at least the center of gravity of the rotating body.
It is provided at a position that can include a plane perpendicular to the rotation axis. Thus, the bearing load due to the weight of the rotating body is reduced by the action of the upward moment due to the magnetic attraction applied to the rotating body and the downward moment due to the weight of the rotating body. The center of rotation of the rotating body with respect to the axis 1 minimizes the moment that causes inclination.

In this structure, the shaft is fixed.
Since the mounting position of the disk 5 held by the disk holder 4 does not necessarily have to be a position overhanging from the bearing, the overall length of the motor can be reduced relative to the length of the dynamic pressure bearing. As a result, the rotating body such as the disk 5 is formed so as to be easily rotatably supported in the cantilevered state by the shaft 1 in the shaft horizontal posture.

FIGS. 2 and 3 (a) and (b) show a case where the rotating body is supported by one dynamic pressure air bearing in the motor and a case where two rotating bodies are supported by a pair of dynamic pressure air bearings. ). FIG. 2 is a graph showing the relationship between the number of bearings and the load capacity (dimensionless). FIG. 3 (a) is an explanatory diagram showing the relationship between one dynamic pressure air bearing and the length of the bearing portion.
(B) is an explanatory view of the relationship between two pairs of dynamic pressure air bearings and the length of the bearing portion.

Normally, the motors related to the information equipment as described above are usually used at a rotation speed of 8,000 to 12,000 rpm.
The case of 0 rpm will be described. Also, considering the allowable bearing space of this motor, the shaft diameter is set to φ10 mm,
Bearing allowable length (total bearing length _{L T)} to 10 mm.

First, the load capacity in each case will be described with reference to FIG.

In the case of one bearing, the total bearing length L _T = bearing length; L = 10.0 mm, and L / D (shaft diameter) = 1.
It becomes 0. Also, if the herringbone grooves, which are the groove patterns 1a and 1b, are provided at both ends of the bearing length; L = 10.0 mm so that Lg = L / 3, the bearing length will be: A flat portion 19 of about 3.3 mm is formed at the center of L = 10.0 mm. This flat portion 19 becomes a perfect circular dynamic pressure air bearing.

The hydrodynamic radial air bearing with a herringbone groove is relatively small and is used where high rotational accuracy is required under a light load. The high-pressure air layer is formed in the gap, and the central flat portion 19 is formed into a perfect circular dynamic pressure air bearing, that is, into a wedge-shaped gap generated by the eccentricity of the shaft 1, and air is drawn into the wedge-shaped gap due to its viscous drag. Forced to create pressure.

The axial length, the number of grooves, the groove inflow angle, and the ratio between the groove width and the ridge width of the pattern groove 1a are determined by the intake of the surrounding air in the hydrodynamic air bearing and the pressure generation by the wedge-shaped gap. This is a theoretical value for optimizing the synergistic effect (maximum load capacity) within a limited bearing length, and a value obtained by experimental verification.

That is, the axial length of the groove on one side is 1/3 or less of the bearing length and 1/4 or more. Also,
The number of grooves is 10 or more and 15 or less. The inflow angle of the groove is not less than 20 ° and not more than 30 °. The relationship between the groove width and the ridge width is such that the ratio of the groove width to the ridge width is 3: 7 to 4: 6.
It is.

From this, it can be considered that most of the bearing load capacity is generated by the perfect circular dynamic air bearing in the flat portion 19 at the center of the bearing. That is, in the case of this one bearing embodiment, the effective bearing span is 3.3 mm of the flat portion 19 at the center of the bearing.

Next, the load capacity in each case will be described with reference to FIG. Bearing length L in the case of a pair of two bearings L = (LT−0.4) /2=0.5
mm (0.4 mm is the distance between the two bearings) and L / D =
0.5. Herringbone grooves 1a ', 1
b ′, 1a ″, 1b ″ is the axial length of the groove; Lg = L / 3
When provided at both ends of the bearing length; L = 5.0 mm, a flat portion 19 of about 1.6 mm is formed at the center of the bearing length; L = 5.0 mm. The two flat portions 19 serve as perfect circular dynamic pressure air bearings. As a result, in the case of the embodiment of the pair of two bearings, about 6.6 m including the flat portion 19 at the center of the two bearings.
m can be said to be the effective bearing span.

Therefore, when the same whirling load of the rotating body is received within a 3.3 mm span of one bearing (FIG. 3).
(Corresponding to (a))) and the case where two bearings receive a pair of bearings with a span of 6.6 mm (corresponding to FIG. 3 (b)). principle)
Therefore, the load capacity in the case of one bearing in this embodiment is 3 mm /
6 mm = 0.5, that is, half is sufficient. If this is taken as one bearing, it will be further reduced by half.As a result, the load capacity per bearing in the case of two pairs of bearings will be less than the load capacity per bearing in the case of one bearing. A quarter would be fine. "

Next, with reference to FIG. 2, which one of the case of one bearing and the case of two pairs of bearings is more optimal (larger load capacity) will be described. (Note that the distance between the two bearings in the case of a pair of two bearings has no effect and will be described below as not present.) Similarly to the above description, the rotation speed of the motor will be described at 12,000 rpm. Also, in consideration of the allowable bearing space of this motor, the shaft diameter D is set to φ10 mm, and the allowable bearing length (total bearing length L) is set to 10 mm.

Now, the number of bearings shown as the horizontal axis in the graph of FIG.

(Equation 1) However,; mu is viscosity counts air, omega is angular velocity, p _a is the air density (atmospheric pressure, room temperature), R is the radius of the shaft 1, Cr is the radius clearance.

In equation (1), the viscosity coefficient of air; μ = 1.83
3 × 10^-10kg. sec / cm ², Angular velocity; ω =
1,257 rad / sec, specific gravity of air (atmospheric pressure, normal
Temperature); p _a= 1.0197kg / cm², Radius of axis 1;
R = 0.5 cm, radius gap; Cr = 0.0005 cm
Substituting and calculating the number of bearings Λ, we get Λ = 1.35.
You.

When one bearing is considered within the above-described allowable bearing space, a shaft diameter D = 1.0 cm and a bearing length L = 1.
0 cm, and L / D = 1.0.

On the other hand, considering two pairs of bearings within the allowable bearing space, a shaft diameter D = 1.0 cm and a bearing length L =
0.5 cm, and L / D = 0.5.

In FIG. 2, the number of bearings Λ = 1.35 and L /
From the intersection of the curves of D = 1.0, the dimensionless load capacity = 0.3
Find one. Similarly, the number of bearings Λ = 1.35 and L / D = 0.35.
From the intersection of the curves of No. 5, a dimensionless load capacity = 0.11 is obtained.

The equation of the load capacity W in which the obtained value is shown on the vertical axis in the graph of FIG.

(Equation 2) However, W is the load capacity (dimensionless), p _a is the air density (atmospheric pressure, room temperature), D is the axial diameter, L is the bearing length, epsilon is the eccentricity.

When the bearing load capacity W is obtained by substituting into Equation 2, the bearing load capacity W = 0.19 when L / D = 10
85 kg, and the bearing load capacity W when L / D = 0.50 is 0.0352 kg.

As a result, the load capacity per bearing in the case of a pair of two bearings may be 1/4 of the load capacity per bearing in the case of one bearing in this embodiment. In the light of the above, the load capacity per bearing in the case of one bearing is the load capacity per bearing W = 0.0352 kg in the case of two pairs of bearings. There are 64 times, and it is understood that the case of one bearing is more optimal (having a larger load capacity).

Assuming that the number of bearings Λ = 8.00, the number of bearings Λ = 8.00 and L / D = 10 in FIG.
The dimensionless load capability = 0.78 is obtained from the intersection of the curves. Similarly, the dimensionless load capacity = 0.50 is obtained from the intersection of the curve of the number of bearings Λ = 8.00 and L / D = 0.5.

By substituting the obtained value into Equation 2, the bearing load capacity W is obtained. When L / D = 1.0, the bearing load capacity W becomes 0.4995 Kg, and L / D = 0.50 kg. The bearing load capacity W at this time is 0.1601 kg.

As a result, the load capacity per bearing in the case of two pairs of bearings may be 1/4 of the load capacity per bearing in the case of one bearing in this embodiment. In the light of the above, the load capacity per bearing in the case of one bearing is the load capacity per bearing W in the case of a pair of two bearings, W = 0.1601 Kg. It is only 12 times, and in this case, it can be seen that the case of two bearings and a pair of bearings is more optimal (large load capacity).

The number of bearings is also called a compressibility number.
It is used as an important index that regulates the operating conditions and performance of gas bearings.

As described above, in consideration of the allowable bearing space of the motor, the bearing diameter is set to 10 mm, and the allowable bearing length (total bearing length; L _T ) is set to 10 mm. L = 10 mm, L / D = 1.0, or a pair of dynamic pressure bearings with L / D = 0.5, L = 5 mm and L / D = 0.5. Whether a bearing is composed of bearings is determined by the value of the number of bearings の described above. Composed of bearings,
The bifurcation point of the value of the number of bearings か の of whether two are constituted by a pair of dynamic pressure bearings is approximately 4.8.
In the case of ≦ 4.8, it is better to configure the bearing with one dynamic pressure bearing. On the other hand, in the case of Λ ≧ 4.8, it is better to configure two and a pair of dynamic pressure bearings.

As in the above example, the bearing dimensions (D = 10.
0 mm, Cr = 0.005 mm), the rotation speed is 4
If it is 2,700 rpm or more, it is better to form two dynamic pressure bearings. If it is less than this (for example, 12,000 rpm) as in the present embodiment, it is better to form one dynamic pressure bearing. Gorgeous.

The rotation speed of the rotating body is 12,000 rpm.
m, the bearing dimensions are D ≧ 18.8 mm and Cr =
In the case of 0.005 mm, L ≧ 9.4 and L / D = 0.
5 is better to be composed of a pair of dynamic pressure bearings, but this is the case where the bearing allowable space is secured.
As in the above example, when the distance is smaller (D = 10.0 mm, Cr
= 0.005 mm), L = 10.0 mm and L
It is better to use one dynamic pressure bearing of /D=1.0.

From these results, the motors used for rotating the disks 5 having a diameter of about 50 mm vertically (horizontal axis of the motor) and arranging several or dozens of the disks at the same time as the motors used for the simultaneous rotation are shown. The bearing length can be set to 15 mm or less, and the total motor length can be set to 20 mm.

Next, with reference to FIG. 4, an embodiment in which the end surface of the bearing bush 2 on the front end side is sealed will be described. FIG.
FIG. 2 is a cross-sectional side view of a motor using a dynamic pressure air bearing in which a shaft 1-1 of a bearing bush 2 has a tip end side end surface in a sealed state and is used horizontally. The same functional portions as those in FIG. 2 are denoted by the same reference numerals, and the description of those portions will be omitted to avoid duplication.

In this configuration, the front end face of the shaft 1-1 is sealed, so that air is taken into the radial gap of the dynamic pressure bearing by the first groove pattern 1a formed on the shaft 1-1. The operation is performed only from a gap of approximately 0.25 mm formed between the opposed coil 11 and the end face of the rotor magnet 10.

At this time, the floating dust having a relatively large particle diameter and the floating dust having a relatively large specific gravity floating in the air taken into the dynamic pressure bearing radial gap from this gap are removed by the rotor McNet 10. Due to the centrifugal force generated by the rotation, the particles are scattered in the outer peripheral direction, and entry into the bearing gap can be prevented as much as possible.

Also, the air taken into the gap can form a high-pressure air layer up to the end of the shaft 1-1 without any escape space, so that the most load at the tip of the shaft 1-1 can be reduced. The load capacity of the circular dynamic pressure radial air bearing in a state close to a nearly complete wall can be obtained at the portion where the bearing is hung.

FIG. 4 is referred to for a description of the perfect circular dynamic pressure radial air bearing, but FIG. 1 and FIGS.
Similar effects can be obtained for FIG.

Further, referring to FIG. 4, a sliding bearing material is provided on the end face sealing portion of the bearing bush 2 on the tip end side of the shaft 1-1, and the thrust load is applied by a sliding bearing (pipot) between the bearing bush 2 and the tip of the shaft 1-1. An embodiment of the present invention will be described.
A disk 5 is fitted into the outer peripheral portion of the disk holder, and the disk 5 is fastened and fixed by the disk retainer 4. A stepped portion is provided inside the disk retainer 4, and the thrust receiver 13 is fitted into the disk retainer 4 by light press-fitting. A thrust load caused by magnetic attraction generated between the rotor magnet 10 and the fixed yoke 14 is received by a sliding bearing (pivot) between the rotor magnet 10 and the tip of the shaft 1-1.

The thrust bearing 13 (sliding bearing material) can be fitted and inserted into the disk holder 4 attached to the end face of the bearing bush 2 from the inside of the bearing bush 2, and can be easily attached and detached. ing.

That is, in assembling, the thrust receiver 13
When only the rotating body including the bearing bush 2 is removed, the bearing inner diameter of the rotating body including the bearing bush 2 is substantially penetrated, and the balance of the rotating body alone is corrected through an ultra-precise arbor (not shown) through the bearing inner diameter. , And the rotary member can be assembled to the shaft 1-1 by inserting the thrust receiver 13 from the inside of the bearing bush 2.

As a result, since the balance is corrected by the rotating body alone, precise balance can be corrected easily and efficiently. Further, after the balance is corrected, the balance can be adjusted even if the thrust receiver 13 is inserted from the inside of the bearing bush 2. Almost no collapse occurs.

That is, in this embodiment, the balance correction of the rotating body is performed with the correction radius = 10 mm, the residual unbalance amount is made up to 2 mg or less, and the total weight of the rotating body is 56 g. , 0.36 μm or less.

Assuming that the eccentricity of the inner stepped hole of the disc holder 4 provided for inserting the thrust receiver 13 with respect to the inner diameter of the bearing bush 20 is 20 μm, the weight of the thrust receiver 13 is 0.1 g. When this amount of unbalance is converted into the amount of residual unbalance at the position of the correction radius = 10 mm, it becomes 0.2 mg, and when this is converted into the eccentricity distance of the total weight of the rotating body, it becomes 0.036 μm.

That is, the eccentricity distance of 0.36 μm or less due to the residual unbalance amount in the balance correction of the rotating body,
The vector sum with the eccentricity distance 0.036 μm due to the eccentricity generated when the thrust receiver 13 is inserted into the thrust receiver 13 is obtained by moving the rotating body to the shaft 1-1.
This is the distance of the eccentricity of the rotating body when assembled. Therefore, in the worst case, the eccentric distance only increases by 0.036 μm, and in some cases, the eccentric distance is 0.03 μm.
It may be reduced to 6 μm.

It should be noted that applying a resin coating to iron, metal, ceramic, or the like as a sliding bearing material is because the mechanical strength of the base material, such as iron, metal, or ceramic, against the load (pressure) applied to the thrust bearing 13 In this case, the sliding friction coefficient is reduced by a resin coating material to improve the life of the bearing.

Further, the first shaft provided on the base side of the shaft 1-1
3, the compressibility number (number of bearings) when the bearing is rotated at 8,000 rpm is 0.9, and the bearing load capacity is 12,000, as shown in FIG.
mm, the problem was caused by insufficient bearing load capacity when starting and stopping the motor.

Therefore, the pair of groove patterns 1 shown in FIG.
a, 1b are obtained by leaving only the first groove pattern 1a provided on the base side of the shaft 1 for air intake and removing the groove pattern 1b provided on the tip side of the shaft 1 to thereby flatten the shaft 1. The bearing load capacity is increased by doubling the portion 19, and at the same time, the effective bearing span is also doubled (6.6 mm), and the bearing becomes more stable when the compressibility number (number of bearings) Λ ≦ 1.0. Is being planned.

In this case, in the groove pattern 1a, the axial length of the groove on one side is 以下 or less and １ ／ or more of the bearing length, and the number of grooves is 10 or more and 15 or less. The inflow angle of the groove is not less than 20 ° and not more than 30 °, and the relation between the groove width and the ridge width is such that the ratio of the groove width to the ridge width is 3: 7 to 4: 6.

Next, referring to FIG. 5 and FIG.
A variation of the groove pattern of the hydrodynamic air bearing constituted by one bearing when the value of changes below 4.8 will be described. The same functional portions as those in FIG. 1 are denoted by the same reference numerals, and the description thereof will be omitted.

That is, in FIG. 5, the same pair of groove patterns 1a and 1c that are inclined with respect to the axis are provided on the air intake side of the shaft 1-2 and the tip side of the shaft 1-2. In FIG. 6, a pair of groove patterns 1a and 1d are engraved on the shaft 1-3, but only the air intake side is a first groove pattern 1a inclined with respect to the axis, and On the tip side, a second groove pattern 1d parallel to the axis is formed.

In this case, since a pair of groove patterns 1a and 1d are provided, the axial length of each groove is 1/3 or less and 1/4 or more of the bearing length, and the number of grooves is 10 or less. Not less than 15 and not more than 15, and the inflow angle of the groove is 20
The angle between the groove width and the ridge width is
The ratio of the groove width to the ridge width is 3: 7 to 4: 6.

When the bearing bush 2 is made of an Invar alloy (commonly referred to as Invar, which generates magnetic anomaly accompanied by volume distortion = spontaneous magnetization distortion = at a temperature lower than its inherent magnetic transformation point), it is contained therein. Particulate graphite having a particle size of about 1 μm falls off and separates during rotation, performs solid lubrication within the bearing radial gap, and particularly starts the motor. Axis 1, 1 when stopped
The effect becomes remarkable when the bearing bush 2 is rotated in contact with -1, 1-2, 1-3 and the seizure of the bearing at this time can be prevented.

Each of the above-described embodiments can be applied to any use state (posture) of the motor regardless of whether the axes 1, 1-1, 1-2, 1-3 are vertical or horizontal.

Next, in the embodiment described below, the use state (posture) of the motor is unique to the axis horizontal, which can be applied only to the axis horizontal.

Referring to FIG. 1, the self-weight of the rotating body is lifted by the magnetic attraction force of a magnet provided on the rotating body, and the rotating body is inserted into the shaft 1 in a cantilevered state in a horizontal posture. A ring-shaped magnetic body 6 made of iron is provided on a part of the outer peripheral portion of the rotating body including the bearing bush 2 so as to be integral with the rotating body, and a magnet 7 is disposed above the magnetic body 6 through an appropriate gap. It is fixedly installed by the net holder 8. The magnet 7 has a magnetic attraction force applied to the link-shaped magnetic body 6 of the magnet 7 so as to be substantially equal to the total weight of the rotating body.
, The magnetic flux density, the gap with the ring-shaped magnetic body 6, and the like are adjusted in advance. As a result, the bearing load is reduced by the amount reduced by the magnetic attraction force of the total weight of the rotating body, and the dynamic pressure air levitation limit rotation speed decreases, and the motor starts (rotation speed increase process) and stops accordingly. The chance of contact rotation (sliding bearing state) between the shaft 1 and the bearing bush 22 during (rotational speed lowering process) is reduced, and the life of the bearing can be extended.

Further, by the action of the link-shaped magnetic body 6 and the magnet 7 fixedly provided, the link body is provided at a position passing through at least the center of gravity of the rotating body and including a plane perpendicular to the rotating shaft 1. And the downward moment due to the weight of the rotating body, the rotation center of the rotating body can minimize the generation of a moment that causes the rotation center to tilt with respect to the shaft 1. . In this case, ideally, a plane perpendicular to the axis 1 passing through the center of gravity of the rotating body and the ring-shaped magnetic body 6
It is desirable that a plane perpendicular to the axis 1 that passes through the midpoint of the respective lengths of the magnet and the magnet 7 coincides with each other.

Next, a configuration for reducing the influence of the rotating body in the direction of its own weight will be described with reference to FIGS. FIG. 7A is a cross-sectional side view of a thin hydrodynamic bearing motor using a shaft horizontal attitude, and FIG. 7B is a perspective view of a ring-shaped magnet and a magnet (notched) used in the motor. FIG. 8 is a graph of the amount of displacement between the ring-shaped magnet and the magnet (notched) and the thrust repulsion. FIG. 9 is a graph showing notch angles of a ring-shaped magnet and a magnet (notched) and a thrust repulsion force or a radial suction force. The same functional portions as those in FIG. 6 are denoted by the same reference numerals, and the description of those portions will be omitted to avoid duplication.

As shown in FIG. 7 (a), one end face of the rotating body including the bearing bush 2 inserted into the shaft 1-4 in a cantilevered state with the shaft horizontal is attached to one end face. 23. Outer diameter 23. Single pole magnetized so that the end face on the opposite side is an S pole.
A ring-shaped magnet 6a having a thickness of 5mm and a thickness of 6mm is provided so as to be integral with the rotating body. Further, the outer diameter of this magnet is 24.0mm in outer diameter, 24.0mm in inner diameter, and a notch angle of 38 °. And the notched magnet 7a having a single pole magnetized so that one end face is an N pole and the opposite end face is an S pole, is not in contact with the ring-shaped magnet 6a (0. (A gap of 25 mm), and the magnetic poles on the same side end surfaces of the two magnets of the notched magnet 7a and the ring-shaped magnet 6a are different from each other, and the notched magnet 7a is provided in a part of the rotating body. It is fixed and installed using a magnet holder 8a at a position shifted from the ring-shaped magnet 6a by 2.0 mm toward the opposite side (the disk 5 side) of the fixed yoke 14.

FIG. 8 shows the relationship between the amount of misalignment of the notched magnet 7a and the ring-shaped magnet 6a in the direction of the end face (thickness) and the thrust repulsion force by this experiment. .

The notched magnet 7a and the link-shaped magnet 6a used in this experiment are made of a plastic magnet and have the same thickness of 6 m as that actually used.
m. However, as for the notched magnet 7a, an experiment was performed using a magnet having no notch.

The amount of deviation between the notched magnet 7a and the ring-shaped magnet 6a in the end face (thickness) direction is from 1 mm to 6 mm.
mm was performed every 1 mm. The result is shown in FIG.

Based on the results of this experiment, the amount of deviation between the notched magnet 7a and the ring-shaped magnet 6a in the direction of the end face (thickness) was determined to be 2 mm, and then the notched angle of the notched magnet 7a was determined. At this time, the relationship between the thrust repulsion force and the radial magnetic attraction force (the magnetic attraction force floating of the rotating body's own weight) was determined by experiments. Macnet 7 with notch
The notch angle of a is 9 from 0 ° (none) to 180 °.
° was changed and implemented. The result is shown in FIG.

From these results, the weight of the rotating body was 58 g.
Therefore, when the intersection of this and the radial suction force curve is obtained, 38 ° at this time is the optimum notch angle for magnetic levitation of the weight of the rotating body. Further, the intersection of this 38 ° line and the thrust repulsion line is found to be 244 g. At this time, the repulsive force acts in a direction opposite to the magnetic attraction force of the rotor magnet 10 and the fixed yoke 14, and the load applied to the thrust receiver 13 at this time is 290 g (actual measurement) of the magnetic attraction force with the fixed yoke 14-repulsion. Force 244g = 46
g, which is reduced to about 1/6 or less.

As described above, the notched macnet 7a and the ring-shaped magnet 6a used in this embodiment are both made of the same material and have the same thickness of 6 mm, as described above. The notch angle is 38 °, and other specific dimensions are shown in FIG.

The ratio between the thrust repulsion force and the radial suction force can be freely selected by selecting the notch angle in FIG. 9, and the thrust repulsion force relative to the notch angle can be selected. The value (curve) of the radial attractive force is not limited to this, and can be variously changed according to changes in the size, magnetic flux density, magnetic gap, etc. of the notched magnet 7a and the ring-shaped magnet 6a.

As described above, according to the present invention, the relationship between the bearing length L and the shaft diameter D is L / D ≧ 1, and the air dynamic pressure groove provided on the outer peripheral surface of the shaft 1 Provided symmetrically with respect to the center of the length, the axial length of one side groove is 1/3 or less and 1/4 or more of the bearing length, and the number of grooves is 10 or more and 15 or less. , Groove inflow angle is more than 20 ° and 30 °
The relationship between the groove width and the ridge width is as follows: the ratio of the groove width to the ridge width is 3: 7 to 4: 6, and the dynamic pressure bearing length is 1
A load capacity that can support the rotating body in one cantilever with one dynamic pressure air bearing of 5 mm or less is obtained, and the total motor length is 2 mm.
A thin dynamic pressure bearing motor used for long-time continuous rotation of a disk or the like having a diameter of 0 mm or less was obtained.

Further, by closing the end face of the shaft of the bearing bush on the tip end side, the air pressure in the bearing gap was increased, and a thin dynamic bearing motor having an increased load capacity was obtained.

Further, a sliding bearing material is provided in a sealed portion of the end surface on the tip end side of the bearing bush shaft, and a thrust load is received by a sliding bearing (pipot) between the bearing bush shaft and the tip of the shaft, whereby the thickness can be further reduced.

Further, the bearing bush 2 has a structure in which the sliding bearing material provided in the sealing portion at the tip end side of the shaft is press-fitted from the inside to the end face of the bearing bush 2 so that it can be easily attached and detached. The inner diameter becomes almost penetrating, and the bearing bush changes into a substantially pipe-shaped bearing bush. As a result, precise balance correction can be easily performed with a single rotating body, and an improved rotational accuracy and an inexpensive thin dynamic pressure bearing motor can be obtained.

The thrust slide bearing material is made of resin,
By using a resin coating of iron, metal, ceramic, or the like, a thin dynamic pressure bearing motor with a longer life was obtained.

Further, a combination of a radial dynamic pressure air bearing with a herringbone groove provided on the base side of the shaft from the center of the bearing length and a perfect circular dynamic pressure air bearing provided on the tip side of the shaft. Thus, the load capacity at the tip of the shaft in the cantilevered state was increased, and a more stable thin dynamic pressure bearing motor was obtained.

Further, the air pressure in the radial gap of the bearing was increased by using the irregular ring for air dynamic pressure provided on the outer peripheral surface of the shaft, and a thin dynamic pressure bearing motor having a high load capacity was obtained.

Further, by using an Invar-type alloy for one or both of the shaft and the bearing bush and simultaneously using the solid lubrication effect of the granular graphite contained therein, the motor is started. The bearing damage in the sliding bearing state at the time of stoppage was reduced, and a long-life thin dynamic pressure bearing motor was obtained.

Further, since the magnet is fixed and installed at a distance where the magnetic attraction force is substantially equal to the total weight of the rotating body, the radial and thrust load loads are reduced, and the life is prolonged. A hydrodynamic bearing motor was obtained.

Next, as a second embodiment of the present invention,
A description will be given of a dynamic pressure radial air bearing motor for continuous operation for a long time.

FIG. 10 is a sectional side view of a thin dynamic bearing motor for a long time continuous operation used in a disk drive, showing a second embodiment of the present invention.

The shaft 31 having the irregular herringbone groove pattern 31a for radial dynamic pressure provided on the outer peripheral surface thereof is
Bearing bush 3 inserted into 1 and rotatably fitted and supported
2 and the outer periphery of the bearing bush 32 is
2a, and is integrated by bonding and fixing. Also,
The shaft 31 is supported and fixed by a shaft support 43 integrated with a cover 45. A substrate 42 and a coil 41 adhered and fixed to the substrate 42 are fixed to a shaft-side end surface of the cover 45 by a fastening screw 54. Furthermore, they are fixed by the housing 48 by means of the fastening screws 53, forming a fixed part of the motor.

On the other hand, a bearing bush 32 is fitted on the inner periphery of the bearing main body 32a, concentrically with the bearing main body 32a on the side of the rotating body, and is bonded and fixed to be integrated. Bearing body 32
A magnetic ring 36 is fitted around the outer periphery of a, and is bonded and fixed to be integrated.

Further, a rotor magnet 40 integrated with a rotor yoke 39 is provided at one end of the bearing main body 32a on the shaft root side, and the rotor magnet 40 is assembled to the shaft 1 so that the end face of the rotor magnet 40 and the coil 4
1 is located at a position where the end faces face each other. The end face of the coil is formed as a flat surface by a resin molding or the like. Also,
The cylindrical portion of the bearing body 32a protruding from the end face of the rotor magnet 40 penetrates the center hole of the
2 stick out behind.

The magnetic converging yoke 4 is fixed to the foremost cylindrical portion of the protruding bearing body 32a, and rotates together with the bearing body 32a.

At this time, the magnetic focusing yoke 44 is
And a cover 45 to form a labyrinth structure through a small gap.

With these structures, the end surface of the bearing bush 32 on the shaft tip side is sealed, so that air is taken into the radial gap of the dynamic pressure bearing by the groove pattern 31a and the end surface of the opposing coil 41 and the rotor magnet. This is performed from a gap of about 0.25 mm formed by the end face of the forty. At this time, the floating dust having a relatively large particle diameter and the floating dust having a relatively large specific gravity floating in the air taken in from the gap are generated in the outer circumferential direction by centrifugal force generated by rotation of the rotor magnet 40. It can be scattered and prevented from entering the gap.

Further, the air is substantially 0.25 mm between the lower surface of the substrate 42 and the upper surface of the magnetic focusing yoke (rotating yoke) 44.
Pass through the gap, and the magnetic focusing yoke (rotating yoke)
44 lower surface and cover 45 upper surface, approximately 0.50 mm
Then, this air is taken into the dynamic bearing radial gap.

At this time, the floating dust having a relatively large particle diameter and the floating dust having a relatively large specific gravity, which float in the air taken into the radial gap of the dynamic pressure bearing from this gap, are removed by the magnetic focusing yoke ( Due to the centrifugal force generated by the rotation of the rotating yoke () 44, the particles are scattered in the outer peripheral direction and can be prevented from entering the gap in the bearing radial direction. This is the dustproof effect of the bearing due to the labyrinth structure of the air inflow passage.

Further, suppose that the coil 4 facing the ultrafine particles is
1 and the end face of the rotor magnet 40, even if they enter the gap made by the magnetic converging yoke 44, they still enter the pool 44 a at the tip of the magnetic converging yoke 44 due to the centrifugal force of the magnetic converging yoke 44. The accumulated pressure is prevented from entering the dynamic pressure bearing radial gap.

Further, a disc 35 is fitted into the protruding portion of the bearing bush 32 on the end face of the bearing main body 32a on the shaft tip side, and a disc holder 34a composed of two parts is further fitted. When the tightening screw 51 is tightened while pressing these against the bearing body 32a, the inner diameter side of the disk holder 34a that is in close contact with the disk is deformed, and the outer periphery of the bearing bush 32 is tightened and fixed at this position.

When the tightening screw 51 is further tightened, the disk holder 34 composed of the two parts on the side opposite to the disk is next.
While a is tightened on the outer periphery of the bearing bush 32 and is pulled into the other disk holder 34a fixed at this position, it simultaneously presses the outer periphery thereof. The disc holder 34a, which is in close contact with the disc whose outer periphery is pressed, presses and fixes the disc against the bearing body 32a while elastically deforming.

Further, the inner diameter of the protruding portion of the bearing bush 32 on the shaft tip side of the bearing bush 32 is substantially the same as the diameter of the shaft 31, and is provided with a projection having a height of about 0.5 mm to 40 mm at the center. Similarly, a thrust receiving holder 34b and a thrust receiving 33 are provided, which are composed of two parts of a disk made of an elastic material (brass or the like) having substantially the same diameter as the shaft diameter, and a thrust receiving holder 34b composed of two parts is provided. A thrust receiver 33 made of a polyimide resin is bonded and fixed to one surface of the disc having no projection, and a disc having another projection is combined with the projection inside, and a fastening screw 52 is attached. And the bearing bush 3
The disk provided with the raised portion is elastically deformed by being slidably mounted in the inside 2 and tightening the tightening screw 52, and is fixed to the inner diameter of the bearing bush 32 by crimping.

In other words, a disk composed of two layers can be fixed at an arbitrary position in the bearing bush 32, and at the same time, a thrust receiving holder 34 having a bearing structure for sealing the bearing on this side.
b. As an actual assembling and adjusting method, a two-part ring-shaped spacer (thickness equivalent to an air gap) is inserted between an end face of the rotor magnet 40 and an end face of the coil 41, which face each other, and rotated. When the tightening screw 52 is tightened while holding down the entire body and the thrust holder 34b, the thrust receiver can be positioned at this position (arbitrary). Thereafter, if the ring-shaped spacer divided into two parts is removed, assembly and adjustment are completed.

Next, a combination of a perfect circular dynamic pressure radial air bearing and an irregular dynamic pressure radial air bearing with a ring bone groove was used.
An embodiment in which a cup-shaped (one of the bearings is sealed) bearing and a shaft groove pattern are used will be described with reference to FIGS. 3 (a) and (b) to FIG.

The motor shown in the embodiment is 8,000 to 3
Used at 000 rpm. In consideration of the allowable bearing space of this motor, the shaft diameter is set to 40 mm and the bearing length is set to 10 mm.

First, FIG. 11A is a sectional view of a general perfect circular dynamic pressure radial air bearing, and FIG. 11B is a side view thereof. A cylindrical shaft 3 is attached to a cylindrical bearing bush 32a.
1a are fitted, and neither the bearing bush 32a nor the shaft 31a has a groove pattern formed at all. Note that R is the radius of the shaft 31a, W is the load on the shaft 31a, Wn and Wt are its components, h is the amount of eccentricity, and L is the length of the bearing bush 32a.

In this case, the bearing bush 32a and the shaft 31 are caused by the eccentric rotation of the shaft 31a with respect to the bearing bush 32a.
The dynamic pressure is generated by the wedge effect due to the viscous resistance ζ of the air within the deviation of the gap a. The load capacity of the dynamic pressure bearing at this time is the ambient pressure (in the normal environment, the specific gravity of air at normal temperature and atmospheric pressure;
1.0197 Kg / cm ² ), which is proportional to the diameter of the shaft 31a and the length of the bearing bush 32a.

FIG. 12 is a side view of a general dynamic pressure radial air bearing with a herringbone groove. A cylindrical shaft 31b is fitted on a cylindrical bearing bush 32b, and a groove pattern 31ba formed by a herringbone groove is provided on either the shaft 31b or the bearing bush 32b.

In this case, the air in the atmosphere is taken in from both ends of the bearing (the bearing bush 32b) by the action of the herringbone groove, and a high-pressure air layer is formed in the gap between the bearings. There are two dynamic pressure generating mechanisms of the dynamic pressure bearing with a herringbone groove.

One is a flat portion 50 between the herringbone groove (groove pattern 31ba) and the herringbone groove (groove pattern 31ba) in the length in the shaft 31b direction, which is caused by the eccentric rotation of the shaft 31b with respect to the bearing bush 32b.
The dynamic pressure is generated by the wedge effect due to the viscous resistance of the air within the deviation of the gap between the bearing bush 32b and the shaft 31b. At this time, the ambient pressure of the dynamic pressure bearing Specific gravity of air at normal temperature and atmospheric pressure; 1.0197 Kg / cm
² ) is high pressure due to the action of the herringbone groove,
The bearing load capacity at this point is proportional to this high pressure.

The second is a herringbone groove, which is a herringbone groove and a ridge 50 which is higher than the herringbone groove in the circumferential direction. Due to the viscous resistance of the air accompanying the rotation of the shaft, a dynamic pressure is generated when the air is drawn into the narrow gap at the ridge 50.

The sum of the above two dynamic pressure generating mechanisms is the bearing load capacity of the dynamic pressure bearing with a herringbone groove. Table 1 shows the calculation results of the load capacities of these two dynamic pressure bearings, and FIG. 13 is a graph thereof.

[Table 1] The load capacity in each case will be described below with reference to Table 1 and FIG. Each load capacity differs depending on the eccentricity ε at that time. (Ε = e / Cr, where e is the amount of eccentricity, and Cr is the radial gap) FIG. 13 shows the eccentricity ε of a general circular dynamic pressure radial air bearing and a dynamic pressure radial air bearing with a herringbone groove. 0.
The graph is obtained by calculating bearing load capacity with respect to the number of rotations at 2, 0.4, 0.6, and 0.7.

When the eccentricity ε is 0.2, 30,000
If it is less than rpm, the perfect circular dynamic radial air bearing can obtain a larger load capacity. In addition, when the eccentricity ε is 0.4, if the eccentricity is 22,500 rpm or less, a larger load capacity can be obtained with the perfect circular dynamic pressure radial air bearing. A larger load capacity can be obtained.

Further, when the eccentricity ε is 0.6, 10,5
If it is more than 00 rpm, the dynamic pressure radial air bearing with a herringbone groove can obtain a larger load capacity and the eccentricity ε
Is 0.7, a dynamic load radial air bearing with a herringbone groove can obtain a larger load capacity at 5,000 rpm or more.

Generally, the eccentricity ε of these bearings is
Usually designed at 0.4. In addition, this bearing is often used with an eccentricity ε of about 0.2 for those requiring high rotational accuracy, such as a polygon scanner motor.
In the case of the bearing of this embodiment which is used at 8,000 rpm to 30,000 rpm and the eccentricity ε is 0.2 to 0.4, it is possible to use a perfect circular dynamic pressure bearing. If the bearing load increases due to the mass or the like, the eccentricity may become 0.4 or more. Particularly, in the manufacturing process, this possibility is high in the process before the completion of the balance correction of the rotating body.

From the above, the case where the bearing is a herringbone grooved bearing will be described below.

FIG. 14 is a sectional view of a bearing with a herringbone groove.

The bearing load capacity at this time is substantially as shown in FIG. 13 (groove ε = 02 to 0.7).

At this time, most of the load capacity is
4 mm flat portion 50 between the herringbone groove having the length in the 1c direction and the herringbone groove (groove pattern 31ca)
Occurs at c. As a result, the rotating body is held at the span of 4 mm, and the disk mounting side (see FIG. 2) of the rotating body performs a rubbing rotation with the fulcrum as a fulcrum, so that the tip of the shaft 31c and the bearing bush 32c are moved. Contact (air levitation is not possible) and normal bearing function cannot be obtained.

Further, as time passes, minute dust contained in the air taken in by the action of the herringbone grooves on both sides causes a flatness of 4 mm between the herringbone grooves. It accumulates on the parts, and eventually a normal bearing function cannot be obtained.

FIG. 15 is a sectional view of an irregular herringbone grooved bearing. This is one in which the herringbone groove on the shaft tip side of the herringbone groove shown in FIG. 14 is eliminated.

As a result, a perfect circular dynamic bearing is formed on a flat portion approximately 7 mm from the shaft tip. The bearing load capacity at this time is approximately 7 in FIG. 13 (groove ε = 0.2 to 0.7).
/ 10 (proportional to the bearing length). On the other hand, the bearing (bearing bush 3) is closed by the action of the ring bone groove, which seals the bearing 33d on the tip end side of the shaft 31d and leaves it on the root side of the shaft 31d.
2d) forming a high-pressure air layer in the gap, so that the shaft 31
FIG. 13 shows the bearing load capacity of the perfect circular dynamic pressure bearing formed on a flat portion approximately 7 mm from the tip of d (groove ε = 0.2
(Approximately 0.7 / 10) of the value (proportional to the bearing length).

That is, the bearing load capacity at this time can be considered to be substantially equal to the value of (groove ε = 0.2 to 0.7) shown in FIG.

In addition, most of the load capacity at this time occurs in a flat portion of 7 mm between the tip portions of the shaft 31d excluding the herringbone groove having the axial length. As a result, the rotating body is held at the span of 7 mm, and the rotating body disk mounting side (see FIG. 10) at the tip of the cantilever support is firmly supported by the perfect circular dynamic pressure radial air bearing.
The leading end of the bearing d and the bearing bush 32d do not come into contact with each other, enabling complete air levitation, and a normal bearing function can be obtained.

Further, fine dust contained in the air taken in by the action of the herringbone groove also accumulates on a flat portion of 7 mm between the herringbone groove and the tip of the shaft 31d in the bearing gap. Without this, a normal bearing function can be obtained for a long time.

Similarly, an irregular herringbone groove may be provided on the bearing bush 32d side.

Next, these manufacturing methods will be described with reference to FIG.

First, when the motor is used in a state where the axis is horizontal, the coil 41 bonded and fixed to the substrate 42 is molded with glass epoxy resin, and the upper surface of the coil 41 is cut flat by a lathe. When this is assembled, that is, when the thrust bearing 33 of the rotating body hits the protrusion at the tip of the shaft and the thrust bearing by the pivot is formed, the gap between the upper surface of the coil 41 molded with resin and the lower surface of the rotor magnet 40 is zero. .2 mm, the gap between the lower surface of the substrate 42 and the upper surface of the magnetic converging yoke (rotating yoke) 44 is 0.3 mm to 0.2 mm.
4 mm. At this time, if the outer diameters of the rotor magnet 40 and the magnetic converging yoke (rotating yoke) 44 are made substantially the same, the pressure in the gap formed between the upper surface of the coil 41 having the smaller gap and the lower surface of the rotor magnet 40 becomes lower. And the upper surface of the magnetic converging yoke (rotating yoke) 14 has a lower pressure than the larger pressure, and as a result, the entire rotating body is attracted to the coil 41 side, and the thrust receiver 33 is always protruded at the tip of the shaft 1. Part and can be positioned in the thrust direction.

Next, a modification of the manufacturing method will be described with reference to FIG.
This will be described with reference to FIG.

First, the coil 41 bonded and fixed to the substrate 42
Is molded with glass epoxy resin, and the upper surface of the coil 41 is cut flat by a lathe (not shown).

When this is assembled and the motor is used with the shaft 31 tip side up and the shaft vertical, the thrust bearing 33 hits the protrusion at the shaft tip due to the weight of the rotating body, and a thrust bearing by pivot is formed. At this time, if the gap between the upper surface of the coil 41 molded with resin and the lower surface of the rotor magnet 40 is 0.3 mm to 0.4 mm, and the gap between the lower surface of the substrate 42 and the upper surface of the magnetic converging yoke (rotating yoke) 44 is 0.2 mm. , Rotor magnet 40 at this time
When the outer diameters of the magnetic converging yoke (rotating yoke) 44 and the magnetic converging yoke (rotating yoke) 44 are made substantially the same, the pressure in the gap formed between the lower surface of the substrate 42 having the smaller gap and the upper surface of the magnetic converging yoke (rotating yoke) 44 increases. The pressure is lower than the pressure of the wider portion of the gap formed between the lower surface of the rotor magnet 40 and the rotor. As a result, the entire rotating body is sucked upward, and is constantly pressed against the tip of the shaft 31 by its own weight. Yes,
This has the effect of reducing the bearing load of the thrust receiver 33.

Next, a case where the use state (posture) of the motor is axis horizontal will be described with reference to FIG. Each of the examples described below is a specific case of the axis horizontal that can be adopted only for the axis horizontal.

(B) All but the rotor yoke 39 and the magnetic converging yoke 44 of the cup-shaped rotating body are made of a non-magnetic material, and the disk leaning between the disk 35 and the rotor yoke 39 on the outer periphery of the rotating body. 6mm wide and 1mm thick
A magnetic ring 36 made of the above magnetic material is provided.

(B) A gravity attracting magnet 37 is fixedly provided on the magnetic ring 36 above the magnetic ring 36 via the magnet holder 8 within a range where the magnetic attraction can be exerted.

(C) The self-weight of the rotating body is caused to float by the magnetic attraction force of the self-weight attraction magnet 37, which will be further described below.

A link-shaped magnetic ring 36 made of iron is integrated with the rotating body at a part of the outer peripheral portion of the rotating body including the bearing bush 32 inserted in the shaft 31 in a cantilevered state with the shaft horizontal posture. , A heavy suction magnet 37 is fixedly set by a magnet holder 38 with an appropriate gap therebetween.

The size, magnetic flux density, and link-like magnetic material of the self-weight attracting magnet 37 are set so that the magnetic attractive force exerted on the ring-shaped magnetic ring 36 of the self-weight attracting magnet 37 becomes almost equal to the total weight of the rotating body. The gap with the ring 36 is adjusted.

As a result, the radial bearing load is reduced by an amount reduced by the magnetic attraction force of the total weight of the rotating body (own weight), the dynamic air floating limit speed is reduced, and the motor is started (rotated) accordingly. Number rise process). At the time of stoppage (rotation speed decreasing process), the chance of the contact rotation of the shaft 31 and the bearing bush 32 (sliding bearing state) is reduced, and the life of the bearing can be extended.

Further, the ring-shaped magnetic ring 36 and the fixedly installed gravity attracting magnet 37 are provided at least at a position which can include a plane passing through the center of gravity of the rotating body and perpendicular to the rotation axis. The effect of the upward moment on the rotating body due to the magnetic attractive force and the downward moment due to the weight of the rotating body causes the center of the rotating shaft of the rotating body to minimize the generation of a moment that causes the shaft 31 to tilt with respect to the shaft 31. Ideally, it passes through a plane passing through the center of gravity of the rotating body, perpendicular to the rotation axis, and the midpoint of each of the lengths of the ring-shaped magnetic ring 36 and the gravity attracting magnet 37. It is desirable that the plane perpendicular to the rotation axis coincides with the rotation axis.

(D) Magnetic ring 3 made of a magnetic material
The magnetic center of the self-weight attraction magnet 7 provided above the magnetic ring 6 and within a range where the magnetic attraction action is exerted on the magnetic ring 36 is closer to the base of the axis than the axial position of the magnetic ring 36 in the axial direction. The magnetic attraction force is used to simultaneously obtain the positioning action in the thrust direction.

As described above, according to the present invention, with the shaft fixed,
The axial magnetic gap, the dynamic pressure radial air bearing of the rotating yoke structure, and the disk rotation motor to which this is applied are described above, using a cup-shaped rotating body with one side of the bearing sealed horizontally in a cantilevered support state. With such a configuration, the load applied to the thrust bearing of the rotating body can be significantly reduced (extremely slight), and at the same time, the load capacity of the dynamic pressure radial air bearing can be increased (more than four times the conventional value).

As a result, the diameter was 10 mm and the bearing length was 12 m
m, horizontal axis position, diameter 50 mm x thickness 2 mm. By mounting an aluminum disk having a specific gravity of 2.7 (substantially equivalent to quartz glass), non-contact rotation of 30,000 mm by complete air levitation was realized.

[0168] Further, with respect to the contact (slip) rotation between the shaft and the bearing bush that occurs when the motor is started and stopped, the circular dynamic pressure radial air bearing (groove pattern) is applied to the tip of the shaft where the radial load is most greatly applied. By using (None) together, the mirror surface was smoothly slid and the seizure of the bearing was prevented.

Further, even when the dynamic pressure radial air bearing is operated in an open structure in a normal environment, the floating yoke in the atmosphere is not brought into the bearing by the action of the labyrinth structure of the rotating yoke, and the bearing is seized. No long-term continuous operation is possible.

Further, by covering the rotary yoke in a closed state except for the gap between the substrate and the insertion portion between the rotary member cylindrical portion and the coil fixed to the substrate, the outside air is shut off.
Intrusion of airborne dust into the bearing can be prevented.

Further, by fixing the shaft, the overhang of the disk mounting position can be reduced, and the radial load on the bearing concentrated on the shaft tip can be reduced.

Further, the motor structure is an axial magnetic gap type, a disk-shaped rotor macnet and a rotor yoke are mounted on the opposite side of the disk, and the center of gravity of the whole rotating body is brought closer to the shaft center from the shaft tip to the shaft tip. The radial load on the bearing concentrated on the part can be reduced, and the bearing can be distributed over the entire shaft.

By providing the rotating yoke at the base of the shaft, the effect of dispersing the bearing radial load concentrated at the shaft tip over the entire shaft can be increased, and at the same time the bearing thrust load can be reduced.

In addition, the use of a rotating yoke is effective in reducing power consumption and improving rotation accuracy.

Further, by covering this rotary yoke in a sealed state except for the space between the substrate and the insertion portion between the rotating member cylindrical portion and the coil fixed to the substrate, the outside air is shut off, and Floating dust can be prevented from entering the bearing.

Further, the wear resistance can be improved and the life can be extended by using ceramics for the dynamic pressure radial bearing portion.

The thrust receiving holder is a thrust receiving holder that can be fixed at an arbitrary position (air gap) in the bearing bush, thereby improving the assembling accuracy.

Further, the disc retainer can be indirectly fastened to a disc when a ceramic (extremely difficult-to-work material) bush is used, so that the fastening force can be equalized, the disc can be prevented from being damaged, and the precision can be improved.

In addition, a cup-shaped (one of the bearings is hermetically sealed) using both a perfect circular dynamic pressure radial air bearing and an irregularly shaped dynamic pressure radial air bearing to minimize space and increase load capacity. The bearing can reduce the thickness of the bearing and the motor.

Further, by using a thrust auxiliary air bearing to which the Bernoulli's theorem is applied, it is possible to reduce the thrust load and realize a stable positioning accuracy of the rotating body in the thrust direction when the shaft is used horizontally. .

Further, a high bearing load capacity capable of cantileverly supporting the rotating body with the dynamic pressure air bearing is obtained, and a thin dynamic pressure air bearing motor used for long-time continuous rotation of a disk or the like is obtained.

[0182]

The thin dynamic pressure air bearing according to the present invention and the disk rotation motor to which it is applied have a bearing length of 15
mm or less, non-contact rotation by air levitation becomes possible, and the thickness of the hydrodynamic bearing motor can be made extremely thin.

Further, the contact rotation between the shaft and the bearing bush, which occurs when the dynamic pressure bearing motor is started and stopped,
By using solid lubrication together, seizure of the bearing can be prevented, and the service life can be extended.

[Brief description of the drawings]

FIG. 1 is a cross-sectional side view of a hydrodynamic bearing motor according to a first embodiment of the present invention.

FIG. 2 is a graph showing the relationship between the number of bearings and load capacity.

3A is an explanatory diagram of a relationship between a pair of dynamic pressure air bearings and the length of the bearing portion, and FIG. 3B is an explanatory diagram of a relationship between two pairs of dynamic pressure air bearings and the length of the bearing portion. .

FIG. 4 is a cross-sectional side view of a motor using a dynamic pressure air bearing with a horizontal shaft in which a front end surface of a shaft of a bearing bush of the present invention is sealed.

FIG. 5 is a modified example of the groove pattern of the hydrodynamic air bearing constituted by one bearing when the value of the number of bearings changes at 4.8 or less.

FIG. 6 shows another modified example of the groove pattern of the hydrodynamic air bearing constituted by one bearing when the value of the number of bearings changes below 4.8.

FIG. 7 (a) is a cross-sectional side view of a thin dynamic pressure bearing motor according to the present invention which uses a horizontal shaft. (B) is a perspective view of a ring-shaped magnet and a magnet (notched) used therein.

FIG. 8 is a graph showing a displacement amount and a thrust repulsion force between a ring-shaped magnet and a magnet (notched) according to the present invention.

FIG. 9 is a graph showing notch angles of a ring-shaped magnet and a magnet (notched) and thrust repulsion or radial suction of the present invention.

FIG. 10 is a cross-sectional side view of a hydrodynamic bearing motor according to a second embodiment of the present invention.

11A is a cross-sectional view of a general perfect circular dynamic pressure radial air bearing, and FIG. 11B is a side view thereof.

FIG. 12 is a side view of a general dynamic pressure radial air bearing with a herringbone groove.

FIG. 13 is a graph of the load capacity of a dynamic pressure bearing.

FIG. 14 is a sectional view of a bearing with a herringbone groove.

FIG. 15 is a sectional view of an irregular herringbone grooved bearing.

FIG. 16 is a cross-sectional side view of a conventional air dynamic pressure bearing motor.

[Explanation of symbols]

1-1, 1, 2-3, 1-4, 31 ... axis, 1a, 1b,
1c, 1d, 31a: groove pattern, 2, 32: bearing bush, 3, 33: shaft support, 4, 34: disk holder and thrust receiver, 5, 35: disk, 6, 6a, 36 ...
Ring-shaped magnetic material, 7, 7a, 37 ... magnet, 8,
38: magnet holder, 9, 39: rotor yoke, 1
0, 40: rotor magnet, 11, 41: coil, 1
2, 42 ... substrate, 13, 43 ... thrust receiver, 14, 44
... fixed yoke, 15, 45 ... mounting plate, 19 ... flat part

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[Procedure amendment]

[Submission date] August 24, 2000 (2008.2.
4)

[Procedure amendment 1]

[Document name to be amended] Drawing

[Correction target item name] FIG.

[Correction method] Change

[Correction contents]

FIG. 10

[Procedure amendment 2]

[Document name to be amended] Drawing

[Correction target item name] FIG.

[Correction method] Change

[Correction contents]

FIG. 11

[Procedure amendment 3]

[Document name to be amended] Drawing

[Correction target item name] FIG.

[Correction method] Change

[Correction contents]

FIG.

[Procedure amendment 4]

[Document name to be amended] Drawing

[Correction target item name] FIG.

[Correction method] Change

[Correction contents]

FIG. 14

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat ゛ (Reference) G11B 19/20 G11B 19/20 D H02K 5/16 H02K 5/16 Z 7/09 7/09

Claims (9)

[Claims] 1. A thrust bearing is provided at one end, and has a cylindrical bearing bush and a shaft which hermetically seals the thrust receiving side.
The bearing bush is rotatably arranged on the shaft via a predetermined gap, and a thrust bearing is formed by the shaft tip and the thrust receiver. A dynamic pressure bearing motor in which a groove pattern inclined with respect to the axial direction is provided on one of the outer peripheral surface of the shaft and the inner peripheral surface of the bearing bush on the other end side. 2. A second groove pattern is provided on one of the shaft outer peripheral surface and the bearing bush inner peripheral surface on the one end side of the bearing bush in a portion where the shaft and the bearing bush oppose each other. The dynamic pressure bearing motor according to claim 1, wherein: 3. The hydrodynamic bearing motor according to claim 2, wherein the second groove pattern is cut in parallel to an axial direction of the shaft. 4. The dynamic pressure bearing motor according to claim 1, wherein the thrust receiver is a member coated with resin, metal, ceramic, or the like. 5. The hydrodynamic bearing motor according to claim 1, wherein at least one of said shaft and said bearing bush is made of an Invar alloy of an iron-nickel alloy. 6. The shaft is a fixed shaft disposed in a horizontal direction, and a ring-shaped first magnet magnetized so as to have an NS pole in the axial direction is provided on an outer peripheral surface of the bearing bush. A second magnet magnetized so as to have an NS pole in the axial direction is provided at a position on the outer periphery of the ring-shaped magnet except for a predetermined range below the shaft and at a position shifted in the axial direction. The dynamic pressure bearing motor according to claim 1. 7. A bearing bush, one end of which is engaged with a shaft fixed to a shaft support at a predetermined interval, forms a dynamic pressure air bearing with the shaft, and seals a bearing portion on the tip end side of the shaft. A dynamic pressure bearing motor comprising a rotating body fixed to a bearing bush, wherein said rotating body is an axial magnetic gap rotating yoke in which a disk-shaped rotor magnet and a magnetic focusing yoke face each other and both of them rotate the same. Form the structure,
The dynamic pressure air bearing communicates with a gap formed between the rotor magnet and a coil fixed to a substrate facing the rotor magnet, and the outer surface of the coil, the magnetic focusing yoke, and the shaft support A hydrodynamic bearing motor having an air inflow path formed by a gap formed along the gap, and wherein all of the magnetic focusing yoke except for the air inflow path are formed by a closed structure. 8. A gap formed between one end face of the coil or the substrate and an end face of the rotor magnet and a gap formed between the other end face of the coil or the end face of the substrate and an end face of the magnetic focusing yoke. 8. The dynamic pressure bearing motor according to claim 7, wherein said motor is smaller than said motor. 9. The rotating body is made of a non-magnetic material except for the rotor yoke and the magnetic converging yoke, and a ring made of a magnetic material is provided on a part of an outer peripheral portion of the rotating body. The dynamic pressure bearing motor according to claim 7, wherein: