JP2001248632A - Fluid bearing device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a dynamic pressure type fluid bearing device having high bearing rigidity in the radial direction and excellent resistance against impact and vibration. SOLUTION: When an upper radial bearing part 6 and a lower radial bearing part 7 are formed on an outer peripheral face of a shaft member 10 and a groove shape of the radial bearing part is formed like a circular arc on a cross section vertical for an axial line 44 of the radial bearing part, a groove having a groove inclination angle of 45 or 70 degrees on one side, a symmetrical shape, and a dog leg shape is formed by forming a line forming a right angle for the axial line 44 symmetrical in a plurality of grooves having substantially dog leg shape along the direction of rotary shaft centered on the line forming a right angle for the axial line 44 of the shaft member 10.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION The present invention relates to a lubricating oil filled between a shaft member and a sleeve member fitted thereto.
The present invention relates to a hydrodynamic bearing device that rotatably supports a shaft member or a sleeve member, and more particularly to a device that improves the bearing rigidity of a radial dynamic pressure bearing.

[0002]

2. Description of the Related Art Heretofore, as a bearing device for rotatably supporting a sleeve member or a shaft member rotated in a predetermined direction, a shaft member and a sleeve member rotatably supported relative to the shaft member have been known. And, a hydrodynamic bearing provided with lubricating oil filled in a gap between the sleeve member and the shaft member has been put to practical use. A typical example is a bearing device called a herringbone type, in which an apparently U-shaped groove is formed on either the inner peripheral surface of the sleeve member or the outer peripheral surface of the shaft member. Things.

FIG. 15 is a sectional view of a motor equipped with the above-mentioned conventional dynamic pressure type fluid bearing device. As shown in the figure, a fixed shaft 10 is press-fitted into a chassis 14, and a core around which a coil 13 is wound is also fixed to an upper portion. Radial dynamic pressure generating grooves (herringbone grooves) 6, 7 are formed on the outer peripheral surface of the fixed shaft 10 at predetermined intervals in the axial direction. Hereinafter, the radial dynamic pressure generating groove is referred to as a radial bearing, the radial bearing 6 is referred to as an upper radial bearing, and the radial bearing 7 is referred to as a lower radial bearing. Usually, an oil reservoir 5 is formed between the upper radial bearing and the lower radial bearing,
The pressure propagation of the upper and lower radial bearings is buffered.
The fixed shaft 10 has a thrust disk 2 protruding in the radial direction.
And a sleeve 4 rotatable with respect to the fixed shaft 10.
Is provided with a thrust cover 3 for covering the thrust disk 2. The lubricating oil 1 is filled in a gap between the fixed shaft 10 and the sleeve 4 facing the thrust disk 2. The hub 11 including the magnet 12 and the back yoke 15 is connected to the sleeve 4.

[0004] In the radial bearing in the dynamic pressure type fluid bearing having such a configuration, for example, as described in Japanese Patent Application Laid-Open No. 9-35404, the axis 44 of the fixed shaft 10 is used.
The groove inclination angles 8, 9 (θ ₁ , θ ₂ ) of the herringbone grooves 6, 7 of 10 to 30 degrees are formed with respect to a line perpendicular to the above. In a cross section perpendicular to the axis of the shaft member, the shape of the groove is often carved in an arc shape. For example, as described in Japanese Patent Publication No. 4-78364 (method of forming a groove in a shaft member) and Japanese Patent Publication No. 3-49648 (method of forming a groove in a sleeve member), the groove shape of the bearing is ball. This is because it is manufactured by rolling using, and its shape becomes an arc shape.

[0005] FIG.
FIG. 4 is a diagram illustrating a general method of forming a herringbone groove in Step 2, wherein a hard ball 32 is pressed into a guide pin 31 and advances while rotating along the axis of the guide pin 31. By manipulating (changing) the amount of advance in the axial direction when the hard ball 32 makes one rotation around the axis, the groove inclination angle θ34 of the herringbone can be set freely. Further, by adjusting the ball diameter 35 and the protrusion amount 36 of the tool having the above-described configuration, the shape (groove depth, groove width ratio) of the arc groove can also be operated (changed).
If the tool is adjusted in this way, a groove having an arbitrary shape can be formed, and research on a groove shape that maximizes the bearing characteristics has been actively conducted.

Incidentally, a hard disk drive is indispensable as an external recording device of a computer.
The recording capacity is increasing year by year, and the hard disk is becoming smaller with the spread of notebook computers. For this reason, motor bearings mounted on hard disks are required to rotate at even higher speeds and have a smaller NRRO (non-reproducible runout), and it is said that conventional ball bearings can no longer satisfy the demand. The development of high-performance motors on board is urgent.

Therefore, the bearing characteristics have been improved by contriving the groove depth, groove inclination angle, groove width ratio, etc., which constitute the groove shape of the fluid bearing. In particular, bearing stiffness is an important factor in bearing characteristics. The bearing rigidity indicates the degree to which the shaft member can rotate without contact between the shaft member and the sleeve member when a shock or vibration is transmitted to the hydrodynamic bearing device in the rotating motor. Its unit is gf / μ
m, which is defined as the ratio of the load gf required to decenter the shaft member rotating by 1 μm with respect to the sleeve member.

When the rigidity of the bearing is reduced, the rotating shaft member and the fixed sleeve member come into direct contact with each other, causing a so-called burn-in phenomenon, so that the motor cannot rotate and the basic function as the motor disappears. become.
In addition, when a hydrodynamic bearing is mounted on the motor of a hard disk drive, even a very small operation failure of NRRO in the radial direction as compared with a large accident such as seizure loses its commercial value as a motor. Requires large bearing stiffness. Further, since the space given to the bearing device is reduced due to the downsizing of the hard disk, it is a problem how to design a bearing device having a large bearing rigidity.

At present, given the dimensions of the bearing shape and the viscosity of the lubricating oil, the bearing stiffness can be calculated by substituting these values into a design theory based on fluid dynamics.
The shape is determined by repeating trial calculations several times to obtain the required bearing rigidity. However, there are many issues that need to be improved in the design theory as the premise. As one of them, according to the present invention, as described above, the actual groove shape is formed into an arc shape by ball rolling, but the groove shape is processed into a rectangular shape due to the theoretical shape. Take up the issues.

For example, “Finite Element Analysis of H
erringbone Groove Journal Bearings: A Parametric S
tudy ”(author Nicole Zirkelback, ref. Page.234-, Vo
The simulation model described in l.120, APRIL 1998, Transactions of the ASME) has a rectangular shape. Also, in the “Comparative Study on Characteristics of Various Dynamic Pressure Bearings for HDD Spindle Motors” (author Keisuke Ono, the subject of the '98 Motor Technology Symposium organized by the Japan Management Association), the simulation model of the herringbone type fluid bearing is rectangular. It has a shape. It should be noted that Keio Ono has also applied for a patent (Japanese Patent Laid-Open No. 10-267029) in accordance with the above description, in which divergence formulating is used as a method for calculating the characteristics of a hydrodynamic bearing. There is a statement that the difference method was used together.

FIG. 4 shows a case where the groove shape of the fluid bearing is rectangular. This is a type in which a groove is dug in the sleeve member, and the inside of the bearing is represented in an image-friendly manner in a cross section perpendicular to the axis 27 of the shaft member. In the case of an actual bearing, for example, a hard disk drive for 3.5 inches, the sleeve radius 24 is 1 mm to 2 mm,
The radial gap Cr22 is 2 μm to 5 μm, and the groove depth 2
1 is also 2 μm to 5 μm, the groove width ratio defined by the following equation 1 is 0.4 to 0.6, and the bearing length is often designed to be 2 mm to 5 mm.

(Equation 1) However, as described above, the actual groove shape is often an arc (see FIG. 5), and even if the result of the calculated value obtained by approximating the rectangular shape is applied to the arc shape, whether or not the result is always the optimal design. The question remains. That is, the relation between the groove width ratio and the groove inclination angle is particularly problematic. In the design theory utilized in the above-mentioned documents and the like, when calculating the relationship between the groove inclination angle and the bearing rigidity, the groove width ratio is fixed.
There is no problem if the groove has a rectangular shape, but when the groove is formed by the ball rolling method shown in FIG. 12, the groove width ratio increases as the groove inclination angle decreases, as described in Equation 2 below. Although there is a characteristic, the conventional calculation method has a fixed groove width ratio, and thus the conventional calculation clearly includes unnecessary calculation errors. FIG. 7 shows the initial condition of the ball diameter at which the groove inclination angle is θ = 90 degrees and the groove width ratio is 0.12.
The relationship between the groove inclination angle θ and the groove width ratio is obtained from the following equation (2). In a radial bearing for a hard disk, the groove width ratio is 0.4 when the groove inclination angle is in the range of 10 to 20 degrees.
Often designed to about 0.6.

Further, only for the shaded area 29 shown in FIG.
There is also a difference in the groove area, and the internal pressure of the lubricating oil calculated by the conventional design theory is different from the actual one.

(Equation 2) grv ₉₀ : grv land ₉₀ when θ = 90 degrees: land when θ = 90 degrees

In summary, the groove shape (groove depth, radial gap, groove width ratio, groove inclination angle) for many radial bearings
While the relationship between the bearing rigidity is studied, there is a problem as to what is the shape that maximizes the bearing rigidity under limited design conditions. In addition, the design theory that has been cited for obtaining the bearing rigidity from the bearing shape has a problem that the groove shape is limited to a rectangle, and some calculation errors must be ignored in order to apply the groove shape to an arc shape. is there.

[0014]

The conventional fluid bearing device is configured as described above, and the theoretical groove shape (groove depth, radial gap, groove width ratio, groove inclination angle) calculated by calculation. Of bearings with reference to the relationship between bearings and bearing stiffness does not match the actual device, and it is not always possible to obtain bearings with sufficient rigidity and low friction torque There was a point. The present invention has been made to solve the above problems, and an object of the present invention is to provide a fluid bearing device having high rigidity and low friction.

[0015]

According to a first aspect of the present invention, there is provided a fluid bearing device which is in contact with a sleeve member via lubricating oil and is rotatable relative to the sleeve member. A fluid bearing device, comprising: a radial bearing portion formed on one of the inner and outer peripheral surfaces of the sleeve member and the outer peripheral surface of the shaft member. It consists of a plurality of substantially U-shaped grooves along the rotation axis direction about a line perpendicular to the axis, and the plurality of grooves are symmetrical with respect to a line perpendicular to the axis of the shaft member and 45 degrees on one side. A symmetrical U-shaped groove having an inclination angle of 70 to 70 degrees, and wherein the groove shape is engraved in an arc shape in a cross section perpendicular to the axis of the shaft member.

According to a second aspect of the present invention, there is provided a fluid bearing device comprising a shaft member which is in contact with a sleeve member via lubricating oil and is rotatable relative to the sleeve member. Having a radial bearing portion formed on one of the rotation sliding surfaces of the inner peripheral surface of the sleeve member and the outer peripheral surface of the shaft member, wherein the radial bearing portion is perpendicular to the axis of the shaft member. The groove comprises a plurality of substantially U-shaped grooves extending along the direction of the rotation axis about the line formed, and the plurality of grooves are formed by the inclination angle θ _{1 of} one of the grooves forming the U shape.
Is symmetrical with respect to a line perpendicular to the axis of the shaft member and is in the range of 45 to 70 degrees, the inclination angle θ ₂ of the other groove is
Is an angle range in which the ratio of θ ₂ / θ ₁ has a relationship of 0.6 to 1.1, and the groove shape is carved in an arc shape in a cross section perpendicular to the axis of the shaft member. is there.

The fluid bearing device according to a third aspect of the present invention is a fluid bearing device comprising a shaft member which is in contact with a sleeve member and is rotatable relative to the sleeve member. A radial bearing portion is formed on one of the rotational sliding surfaces of the inner peripheral surface of the sleeve member and the outer peripheral surface of the shaft member, and the radial bearing portion is centered on a line perpendicular to the axis of the shaft member. It consists of a plurality of substantially U-shaped grooves along the rotation axis direction, and the plurality of grooves have an inclination angle of 45 degrees to 70 degrees on one side by symmetrically forming a line perpendicular to the axis of the shaft member. The groove is a symmetrical U-shaped groove, and the groove is carved in an arc shape in a cross section perpendicular to the axis of the shaft member, and the gap between the shaft member and the sleeve member is filled with air as a lubricating fluid. Is what it is.

According to a fourth aspect of the present invention, there is provided a fluid bearing device comprising a shaft member which is in contact with a sleeve member via lubricating oil and is rotatable relative to the sleeve member. Having a radial bearing portion formed on one of the rotation sliding surfaces of the inner peripheral surface of the sleeve member and the outer peripheral surface of the shaft member, wherein the radial bearing portion is perpendicular to the axis of the shaft member. It consists of a plurality of substantially U-shaped grooves along the rotation axis direction with the line formed as the center, and the plurality of grooves are symmetrical with respect to a line perpendicular to the axis of the shaft member, and each side has an angle of 45 to 70 degrees. The groove is a symmetrical C-shaped groove having an inclination angle, and the groove is formed in an arc shape in a cross section perpendicular to the axis of the shaft member.

[0019]

DESCRIPTION OF THE PREFERRED EMBODIMENTS In order to solve such a conventional problem, the inventors of the present invention have made intensive studies and as a result, have developed an analysis system capable of calculating bearing stiffness by using the finite element method in both rectangular and circular groove shapes. Through the development, the relationship between the groove shape (groove depth, radial gap, groove angle, groove width ratio) of the arc groove and bearing rigidity was found. That is, as a result, the present invention has an arc-shaped groove formed by the ball rolling method and has a groove inclination angle θ.
Of the conventional radial bearing (with a groove inclination angle of 10 °) is formed in the range of 45 ° to 70 °.
(35 ° to 35 °), the vibration resistance and shock resistance of the motor rotatably supported by the hydrodynamic bearing device can be improved.

The effect of the groove inclination angle θ of 45 ° to 70 ° will be described below in connection with a radial bearing having an arcuate groove shape, particularly a size mounted on a motor used in a 3.5-inch hard disk drive. This will be described with reference to the analysis results of the radial bearing. FIG. 13 shows various dimensions required for analyzing the bearing rigidity. The bearing length 41 is 3 mm,
The apex of the groove-shaped square is located at the center 42 of the axial length.
That is, the length 42 is half of the length 41. The sleeve 4 member is eccentrically rotated with respect to the shaft member 10, and the amount of eccentricity ΔE is the distance between the axis of the shaft member 10 at the two-dot chain line 44 and the axis of the sleeve 4 member at the two-dot line 43. 45. However, in general, the eccentric state is often represented by an eccentricity ε defined by the following equation (3). Here, the bearing rigidity is analyzed under the condition of ε = 0.1. Note that Cr in Equation 3 is the radial gap 22 shown in FIG. 5 and is 2 μm to 10 μm.
In the range.

(Equation 3) The groove inclination angles θ ₁ and θ ₂ forming a square shape are vertically symmetrical at θ ₁ = θ ₂ , and the relationship between the groove inclination angle θ and the groove width ratio refers to the data in FIG. In this case, the label θ on the horizontal axis is θ = θ ₁
= Θ ₂ . The sleeve diameter was 3.5 mm, the groove depth was 5 μm, the number of grooves was 8, and the number of rotations of the shaft member was 720.
0 rpm, and the viscosity of the lubricating oil is 20 × 10 ⁻³ Pa · sec.

FIG. 8 shows a result of analyzing a rectangular groove instead of an arc by the new analysis system described above. The radial bearing was configured with the above dimensions, the groove shape was rectangular, and the groove rigidity was determined under analysis conditions in which the groove width ratio was fixed at 0.6. This is to verify whether the same results as the conventional design theory can be obtained under the same conditions in the new analysis system. As a result, it was confirmed that the bearing stiffness was maximum when the groove inclination angle was around 28 degrees in both cases of the radial gaps of 5 μm and 7 μm, which agreed with the conventional design theory.

FIG. 9 shows an analysis result when the rectangular groove is changed to an arc groove under the above conditions. In other words, under the analysis conditions that the groove shape is an arc, but the groove width ratio is fixed at 0.6 regardless of the groove inclination angle, the analysis contents ignore the fact that the groove width ratio fluctuates with the groove inclination angle in the actual ball rolling method It is. Also in this case, as in FIG. 8, the bearing rigidity is maximum when the groove inclination angle is around 28 degrees. From the results of FIGS. 8 and 9, it was found that the bearing stiffness was maximized at a groove inclination angle of about 28 degrees regardless of the groove shape under the condition that the groove width ratio was fixed.

FIG. 10 shows the results of analysis conditions (FIG. 7) in which the groove width ratio changes in conjunction with the groove inclination angle in the form of an arc-shaped groove. Under these analysis conditions, it was found that the bearing rigidity was large when the groove inclination angle was around 45 to 70 degrees, and the bearing rigidity became maximum especially when the groove inclination angle was around 60 degrees in any radial gap Cr. The bearing stiffness was maximized at a groove inclination angle that was significantly different from the above two analysis results (conditions in which the groove width ratio was fixed).

FIG. 11 shows a comparison of the absolute values of the bearing stiffness when the radial gap is 5 μm under the above-mentioned three types of analysis conditions. It can be seen that there is a remarkable difference in the change characteristic of the bearing rigidity with respect to the groove inclination angle between the condition and other analysis conditions.

FIG. 14 shows a groove inclination forming a U-shape.
Angle θ₁And θ_TwoIs θ₁≠ θ_TwoStiffness in case of vertical and asymmetric
FIG. The horizontal axis is θ_TwoAnd θ₁Ratio, the vertical axis is the bearing
When the maximum value of rigidity is 100%, θ₁Is 45 degrees,
Each stiffness at 60 degrees and 70 degrees is shown. With the wavy line in the figure
Enclosed area (θ_Two/ Θ₁If the ratio is 0.6-1.1)
If the bearing stiffness is greater than 95% of the maximum value
, We have θ ₁And θ_TwoAxis even if it varies slightly
It can be seen that the receiving rigidity does not decrease rapidly.

As described above, conventionally, the bearing rigidity is said to be maximum when the groove inclination angle is around 30 degrees. However, if the groove shape is an arc shape, by configuring the groove inclination angle at around 45 degrees to 70 degrees, greater bearing rigidity can be obtained, and a motor excellent in shock resistance and vibration resistance can be provided. It can be said.

According to FIG. 10, when the groove inclination angle is in the range of 10 ° to 35 °, the groove inclination angle and the bearing rigidity have a linear relationship. When the groove inclination angle is designed within this range, if the groove inclination angle varies in the machining process, it is expected that this directly affects the bearing rigidity. For example, when the groove inclination angle is 20
The bearing rigidity differs by 1% between the degrees and 21 degrees. At the current level of processing technology, it is unlikely that the groove inclination angle will vary by ± 2 degrees. However, when the groove inclination angle is in the range of 45 to 70 degrees, the bearing rigidity is flat with respect to the groove inclination angle. Even if the angle varies, the fluctuation of the bearing rigidity is negligible, and has an advantage that the variation of the groove inclination angle is less sensitive than the conventional range of 10 to 35 degrees. Therefore, the conventional groove machining accuracy is not required, and as a result, the cost of the fluid bearing device can be reduced.

In addition, since the groove inclination angle was previously considered to be 10 degrees to 35 degrees, if the design conditions were not satisfied with the bearing stiffness obtained at this groove inclination angle, a small groove depth was obtained by making full use of precision machining technology. In addition, the rigidity of the bearing has been increased by forming a radial gap. Therefore, there is a problem that the processing cost is necessarily high. However, when the groove inclination angle is set in the range of 45 degrees to 70 degrees, the bearing rigidity increases, so that a high precision machining technique is not required for machining the groove depth and the radial gap. It is considered that the cost of the device can be reduced. Based on the above results, an embodiment suitable for manufacturing a fluid bearing device will be described below.

(Embodiment 1) A fluid bearing device according to Embodiment 1 of the present invention will be described below. FIG.
FIG. 2 is a partial sectional view showing a half of a symmetrical structure about a fixed shaft of a motor of the dynamic pressure type fluid bearing device according to the first embodiment. In the figure, a fixed shaft 10 is press-fitted into a chassis 14, and a core around which a coil 13 is wound is also fixed to an upper portion. An upper radial bearing 6 and a lower radial bearing 7 are formed on the outer peripheral surface of the fixed shaft 10 at predetermined intervals in the axial direction. The method of forming this radial bearing is described in, for example, Japanese Patent Publication No. 4-78364.
As described in the publication, a hard ball coaxially fitted to the inner surface of the guide pipe is pressed against the shaft member 10 and rotated forward to engrave an arc-shaped groove in the shaft member 10.

The fixed shaft 10 is provided with a radially protruding thrust disk 2, and a sleeve 4 rotatable with respect to the fixed shaft 10 is provided with a thrust cover 3 for covering the thrust disk 2. The lubricating oil 1 is filled in a gap between the fixed shaft 10 and the sleeve 4 facing the thrust disk 2. The hub 11 including the magnet 12 and the back yoke 15 is connected to the sleeve 4.

When the motor starts rotating, that is, when the sleeve 4 starts rotating with respect to the fixed shaft 10, the lubricating oil 1 between the upper and lower radial bearings 6, 7 and the sleeve 4 also starts to rotate together with the sleeve 4. On the surface of the fixed shaft 10,
With respect to a line perpendicular to the axis 44 of the fixed shaft 10, a plurality of U-shaped grooves formed at an inclination angle 8 (θ ₁ ) and an inclination angle 9 (θ ₂ ) are formed in the axial direction. I have. The lubricating oil 1 filled between the groove and the sleeve 4 is brought to the center of the groove along the dogleg-shaped groove by the rotation of the sleeve 4 and generates a large oil pressure. This oil pressure allows the sleeve to rotate without contact with the fixed shaft 10.

In the radial bearing having this configuration, the groove shape is an arc shape in a cross section perpendicular to the axis of the shaft member 10. And the relationship between the two groove inclination angles is
45 degrees ≦ θ ₁ ≦ 70 degrees, 45 degrees ≦ θ ₂ ≦ 70 degrees, θ ₁ = θ ₂
The structure of the symmetrical herringbone groove formed by the formula (1) is a characteristic structure of the present invention.

That is, in the present embodiment, by adopting the above-described configuration, the bearing rigidity in which the groove inclination angles (θ ₁ , θ ₂ ) are in the range of 45 to 70 degrees in FIG. Comparing with the conventional one in the range of 10 degrees to 35 degrees, it is understood that there is an advantage that the bearing rigidity of the radial bearing is increased by about 1.3 times. For example, when the groove inclination angle is 20 degrees, the bearing rigidity is 75%, whereas when the groove inclination angle is 55 degrees, the bearing rigidity is 100%.
100% ÷ 75% = 1.33 times.

By obtaining a large bearing rigidity with such a radial bearing, a motor equipped with a dynamic pressure type fluid bearing device has improved shock resistance and vibration resistance. As a result, a hard disk drive, an optical disk or the like employing this motor has been developed. The shock resistance and vibration resistance of the information recording / reproducing apparatus are improved.

In this embodiment, the case where the grooves of the radial bearings 6 and 7 are formed on the shaft member 10 is shown.
As shown in FIG. 5, the same effect as described above can be obtained by forming a groove serving as a radial bearing on the side of the sleeve 4 member.

(Embodiment 2) Next, a fluid bearing device according to Embodiment 2 of the present invention will be described. "Dynamic pressure type grooved bearing" is a document that describes the basic formula for calculating the bearing stiffness of a herringbone type radial bearing (author:
Iba Koji, literature name Fujikoshi technique VOL45No. 1
(1989), Vol. 101). According to this, the load capacity W which is the basis of the bearing rigidity is represented by the following equation (4). Here, μ is the viscosity coefficient of the fluid, R is the radius of the shaft member,
ω is the angular velocity, Cr is the above-mentioned radial clearance, ε is the above-mentioned eccentricity, and W ″ is the dimensionless load capacity. W ″ is the bearing length, groove inclination angle, groove width ratio as described in the article. , Groove depth and radius gap.

(Equation 4) The above equation (4) is satisfied when the Reynolds number Re is much smaller than 1.

Using the lubricating oil and air presented in the means for solving the problem, the Reynolds number is determined using the following equation (5). This equation was quoted from "Mechanics of viscous fluid" (Takefumi Ikui and Masahiro Inoue: Rigakusha: page 61). Here, ν is the kinematic viscosity coefficient, and U is the tangential velocity of the shaft member.

(Equation 5)The viscosity coefficient of the lubricating oil at room temperature is 20 × 10^-3P
a · sec and the specific gravity is 0.9, the kinematic viscosity coefficient
Is 22.2 × 10^-6m^Two/ Sec. In the case of air
The viscosity coefficient is 1.884 × 10^-6The specific gravity is Pa · sec
1.226 × 10 ^-3So the kinematic viscosity coefficient is 15.012 ×
10^-6m^Two/ Sec.

Now, R = 1.5 × 10 ⁻³ m, rotation speed 720
At 0 rpm, U = Rω = 1.13 m / s, and Cr
= 5 μm, the Reynolds number of the lubricating oil is 1.35.
An apparent viscous flow that is sufficiently smaller than 1 at × 10 ^-4 . The Reynolds number in the case of air is also 1.99 × 10 ^−4, which is the same viscous flow as lubricating oil. Therefore, when the lubricating fluid is either oil or air, the above equation (4) is satisfied.

In the above equation (4), W ″ is a factor caused only by the shape and is not affected by the viscosity coefficient.
Therefore, in the case of the radial bearing having the V-shape described in the first embodiment, when the lubricating fluid is changed from oil to air, the absolute value of the rigidity becomes small due to the difference in the viscosity coefficient. The tendency that the bearing rigidity is large is maintained when the inclination angle of the U-shaped groove is in the range of 45 to 70 degrees. It should be noted that the bearing stiffness reduced by the difference in the viscosity coefficient is often compensated by increasing the number of rotations of the motor, and the stiffness can be similarly compensated in the present embodiment.

As described above, according to the second embodiment in which the lubricating fluid is changed from oil to air in the first embodiment,
The motor equipped with the dynamic pressure type fluid bearing device is excellent in shock resistance and vibration resistance similarly to the first embodiment, and as a result, the resistance of an information recording / reproducing device such as a hard disk drive or an optical disk using this motor is improved. The shock characteristics and the vibration resistance characteristics are improved, and the configuration can be simplified, and the manufacturing process can be simplified.

(Embodiment 3) Hereinafter, a fluid bearing device according to Embodiment 3 of the present invention will be described with reference to FIGS. FIG. 2 is a partial cross-sectional view showing a half of a symmetrical structure about a fixed shaft of a motor of the dynamic pressure type fluid bearing device according to the third embodiment. As shown in the figure, the basic configuration is the same as that of the first embodiment,
In the third embodiment, the inclination angles θ ₂ /
theta ₁ ratio is characterized in herringbone grooves asymmetrical type with the inclination angle of theta ₂ in a relationship of from 0.6 1.1.

By forming such a bearing shape,
As in the first embodiment, a greater bearing rigidity than that of the conventional product can be obtained, and at the same time, the lubricating oil can be prevented from flowing out of the bearing opening on the lower radial bearing side, and a sealing effect can be expected.

FIG. 14 shows a groove inclination angle θ ₁ forming a dogleg.
FIG. 6 is a diagram showing bearing rigidity when θ ₂ is θ ₁ ≠ θ ₂ and is vertically asymmetric. The horizontal axis theta ₂ and theta ₁ ratio, and the vertical axis is obtained by the maximum value of the bearing rigidity and 100%, theta ₁ is 45 degrees,
Each stiffness at 60 degrees and 70 degrees is shown. In the region surrounded by the dashed line (the ratio θ ₂ / θ ₁ is 0.6 to 1.1), the bearing stiffness has a large stiffness of 95% or more of the maximum value. From this, even if the angle is not a symmetrical V-shape of θ ₁ = θ ₂ , a large bearing stiffness can be obtained if θ ₁ is 45 to 70 degrees and the θ ₂ / θ ₁ ratio is 0.6 to 1.1. You can see that. It can also be seen that the bearing stiffness does not drop sharply even if θ ₁ and θ ₂ vary somewhat.

In addition, by setting θ ₁ ≠ θ _{2, there} is an effect that a slight pressure difference is provided inside the radial bearing to generate a force for moving the lubricating oil 1 in the axial direction.
More specifically, a pump-in operation is performed when the lubricating oil 1 moves toward the inside of the bearings 6a and 7a due to the above effect, and a pump-in operation is performed when the lubricating oil 1 flows out toward the outside of the bearing.
Called out action. For example, in FIG. 2, if θ ₂ > θ _{1 in} a range of a small groove inclination angle (about 55 degrees or less) until the bearing rigidity reaches the maximum value, the bearing rigidity on the θ ₂ side is larger than θ _1. Become. The bearing stiffness is the product of the internal pressure of the bearing and the area.In other words, if the bearing stiffness is large, the pressure of the lubricating oil will increase, so the lubricating oil will be affected by the pump-in flowing from the bottom up on the paper. become. By this operation, the lubricating oil 1 can be prevented from flowing out of the bearing opening on the lower radial bearing 7a side.

On the other hand, when the groove inclination angle is 60 degrees or more (an angle region where the bearing rigidity is larger than the inclination angle at which the bearing rigidity shows the maximum value), θ ₂ <
At θ ₁ , the lubricating oil 1
In. However, if the action of the pump-in is too great, lubricating oil may flow out from the bearing opening on the thrust bearing side. In this regard, the pumping-in or pump-out action is created by the herringbone of the thrust bearing. It is necessary to balance the lubricating oil pressure in the entire device.

As described above, according to the third embodiment, θ ₂ > θ ₁ is satisfied in a plurality of U-shaped grooves having the inclination angles θ ₁ and θ ₂ constituting the radial bearings 6a and 7a. By doing so, the pump-in action is generated, so that the lubricating oil 1 is prevented from flowing out of the bearing opening on the lower radial bearing 7a side while securing high bearing rigidity, and an oil seal effect is obtained. Can be expected.

(Fourth Embodiment) Next, a fluid bearing device according to a fourth embodiment of the present invention will be described with reference to FIG. FIG. 3 is a partial cross-sectional view showing a half of a symmetrical structure about a fixed shaft of a motor of a dynamic pressure type fluid bearing device according to a fourth embodiment. As shown in the drawing, the basic configuration is the same as that of the first embodiment. However, in the second embodiment, in the bearings 6b and 7b, the respective groove inclination angles are 45 degrees ≦ θ ₁ ≦ 70 degrees, and It is characterized by having a C-shaped groove shape having a relationship of 45 degrees ≦ θ ₂ ≦ 70 degrees and θ ₁ = θ ₂ .

Normally, as described in the first embodiment, the radial bearing portion is constituted by two upper and lower U-shaped grooves (herringbone grooves) in the axial direction. However, when there is little space in the radial bearing portion of the motor, "Study on unbalanced vibration of herringbone grooved hydrodynamic gas bearing" (Junichi Ichihara, Transactions of the Japan Society of Mechanical Engineers, C,
As described in Vol. 53, No. 495, Showa 62-11), it may be constituted by a single U-shaped radial bearing.

In the fourth embodiment, by setting the angle of the groove in the shape of a square to be 45 degrees to 70 degrees, the angle of the groove in the angle of the square is 10 to 35 degrees. It has an effect of obtaining a large bearing rigidity.

That is, even if the bearing length is the same in the U-shape and the U-shape, the net bearing length is shorter in the U-shape. Therefore, if the bearing length is the same, the absolute value of the bearing stiffness will be larger in the V-shaped groove. However, regarding the bearing stiffness, the relationship between the groove inclination angle and the bearing length is independent, and is 45 to 7 degrees in the fourth embodiment from the conventional groove inclination angle (10 to 35 degrees).
It can be said that 0 degrees is better.

[0051]

As described above, according to the fluid bearing device according to the first aspect of the present invention, in the radial bearing formed by the arc-shaped groove, the groove inclination angle is set to 10 degrees to 35 degrees in the related art.
By changing the range from 45 degrees to 70 degrees, a large bearing rigidity can be obtained, and there is an effect that a motor resistant to vibration and impact can be provided. In addition, in order to increase the bearing stiffness, a fine groove depth and a radial gap were conventionally formed by making full use of precision processing technology, but in the present invention, it is possible to respond by a simple operation of changing the groove inclination angle. The effect of cost reduction can also be obtained. Further, when the groove inclination angle is in the range of 45 ° to 70 °, even if the groove inclination angle varies in processing, the bearing rigidity is hardly affected. Therefore, it is not necessary to use a nerve for the groove forming operation, and the cost can be reduced. This has the effect of increasing the effect and suppressing the variation in the performance of the hydrodynamic bearing device.

A fluid bearing according to claim 2 of the present invention.
According to the shaking device, one of the inclination angles θ forming the dogleg₁But
When in the range of 45 to 70 degrees, the other inclination angle θ
_TwoIs θ _Two/ Θ₁With a ratio of 0.6 to 1.1
To form a U-shaped groove
Large bearing stiffness even if the two corners are not the same
And pump-in or pump-
Out action can be generated,
Lubricating oil can be opened without the need for ring members
The effect of being able to prevent leakage from the mouth
is there.

Further, according to the fluid bearing device of the third aspect of the present invention, in the radial bearing portion of the fluid bearing using air instead of the lubricating oil, the V-shaped groove forming the radial bearing portion is formed. By setting the cross-sectional shape to an arc groove and setting the groove inclination angle in the range of 45 to 70 degrees, a large bearing rigidity can be obtained, a motor resistant to vibration and impact can be provided, and the configuration is simplified to reduce costs. There is an effect that can be.

Further, according to the fluid bearing device of the fourth aspect of the present invention, the angle of inclination of the C-shaped groove is set to 45 degrees to 7 degrees.
By setting it to the range of 0 degrees, a large bearing rigidity is provided,
The effect is obtained that a radial bearing having practical bearing rigidity can be provided for a small motor in which two herringbone type radial bearings cannot be installed in terms of space.

[Brief description of the drawings]

FIG. 1 is a partial cross-sectional view showing a half of a symmetrical structure about a fixed shaft of a motor of a fluid bearing device according to Embodiment 1 of the present invention.

FIG. 2 is a partial cross-sectional view showing a half of a left-right symmetric structure centered on a fixed shaft of a motor of a fluid bearing device according to a second embodiment of the present invention.

FIG. 3 is a partial sectional view showing a half of a left-right symmetric structure centered on a fixed shaft of a motor of a fluid bearing device according to a third embodiment of the present invention.

FIG. 4 shows an exaggerated drawing of a rectangular shape of the groove shape of the radial bearing portion in a cross section perpendicular to the axis of the shaft member, which is conventionally used when calculating the bearing rigidity of a radial bearing. It is a figure showing a model

FIG. 5 is an exaggerated drawing of the groove shape of the radial bearing portion formed in an arc shape in a cross section perpendicular to the axis of the shaft member, the groove shape when the groove is formed by ball rolling; FIG.

FIG. 6 is a diagram for explaining that when a radial bearing is projected on a cross section perpendicular to the axis of a shaft member, a difference occurs only in the hatched region between the arc and the rectangle in the area of the groove.

FIG. 7 is a view for explaining that a groove width ratio is linked with a groove inclination angle when a groove of a radial bearing is formed by a ball rolling method.

FIG. 8 is a radial bearing having a rectangular groove shape, a fixed groove width ratio of 0.6, a sleeve diameter of 3.5 mm, and a bearing length of 3 mm.
FIG. 9 is a relative comparison diagram showing the result of analyzing the relationship between the groove inclination angle and the bearing rigidity when the radial gap Cr is 5 μm and 7 μm, with the maximum value of the bearing rigidity being 100%.

FIG. 9 is a radial bearing having an arc-shaped groove shape, a fixed groove width ratio of 0.6, a sleeve diameter of 3.5 mm, and a bearing length of 3 mm.
FIG. 9 is a relative comparison diagram showing the result of analyzing the relationship between the groove inclination angle and the bearing rigidity when the radial gap Cr is 5 μm, 6 μm, and 7 μm, with the maximum value of the bearing rigidity being 100%.

FIG. 10 is an arc-shaped groove shape having a sleeve diameter of 3.5 mm,
A radial bearing having a bearing length of 3 mm and a radial gap Cr of 3.25 μm under the analysis conditions of the groove inclination angle and the groove width ratio shown in FIG.
The results of analyzing the relationship between the groove inclination angle and the bearing rigidity at m, 5 μm, 6 μm, and 7 μm indicate that the maximum value of the bearing rigidity is 1
It is a relative comparison figure shown as 00%.

FIG. 11 is a diagram in which bearing stiffness is plotted as an absolute value with respect to data of a radial gap Cr = 5 μm in the results calculated in FIGS. 8, 9 and 10;

FIG. 12 is a diagram for explaining a method of forming a groove in a sleeve member by a ball rolling method.

FIG. 13 is a diagram showing dimensions to be substituted into an analysis of bearing stiffness.

FIG. 14 is a radial bearing having an arc-shaped groove shape, a groove width ratio of 0.6 fixed, a sleeve diameter of 3.5 mm, and a bearing length of 3 mm, and a letter C in a radial gap Cr of 5 μm, 6 μm and 7 μm. FIG. 10 is a relative comparison diagram showing a result of analyzing a relationship between a ratio θ ₂ / θ _{1 of} two groove inclination angles (θ ₁ , θ ₂ ) to be formed and a bearing rigidity with a maximum value of the bearing rigidity being 100%.

FIG. 15 is a partial cross-sectional view showing a half of a left-right symmetric structure centered on a fixed shaft of a motor of a fluid bearing device in a conventional fluid bearing device.

[Explanation of symbols]

REFERENCE SIGNS LIST 1 lubricating oil 2 thrust disk 3 thrust cover 4 sleeve 5 lubricating oil reservoir 6 upper radial bearing 7 lower radial bearing 8 groove inclination angle θ ₁ 9 groove inclination angle θ ₂ 10 shaft member 11 hub 12 magnet 13 coil 14 Chassis 15 Back yoke 21 Groove depth 22 Radial gap Cr 23 Radius of shaft member 24 Radius of sleeve member 25 Angle of groove portion 26 Angle of non-groove portion 27 Axis of shaft member 28 Radius of curvature of arc portion 29 At the same groove width ratio Difference between arc-shaped groove and rectangular groove 31 Guide pin 32 Hard ball 33 Sleeve member 34 Groove inclination angle θ1 35 Ball diameter 36 Ball protrusion amount 41 Bearing length 42 Length from the lower end of bearing to the end of the letter 43 Sleeve member Axis 44 Axis of shaft member 45 Eccentricity ΔE

Claims (4)

[Claims] Claims: 1. Contact with a sleeve member via lubricating oil,
A fluid bearing device comprising a shaft member that is rotatable relative to the sleeve member, wherein the fluid bearing device is formed on one of a rotation sliding surface of an inner peripheral surface of the sleeve member and an outer peripheral surface of the shaft member. A radial bearing portion, the radial bearing portion comprises a plurality of substantially U-shaped grooves along the rotation axis direction about a line perpendicular to the axis of the shaft member, the plurality of grooves, A symmetrical U-shaped groove having a tilt angle of 45 to 70 degrees on one side by symmetrically forming a line perpendicular to the axis of the shaft member, and a cross section perpendicular to the axis of the shaft member. 3. The fluid bearing device according to claim 1, wherein the groove shape is carved in an arc shape. 2. Contacting a sleeve member via a lubricating oil,
A fluid bearing device comprising a shaft member that is rotatable relative to the sleeve member, wherein the fluid bearing device is formed on one of a rotation sliding surface of an inner peripheral surface of the sleeve member and an outer peripheral surface of the shaft member. A radial bearing portion, the radial bearing portion comprises a plurality of substantially U-shaped grooves along the rotation axis direction about a line perpendicular to the axis of the shaft member, the plurality of grooves, The inclination angle θ of one groove forming the U-shape
₁ is a symmetrical line perpendicular to the axis of the shaft member, and 45
When the angle is in the range of 70 to 70 degrees, the inclination angle θ of the other groove is
₂ is an angle range in which a ratio of θ ₂ / θ ₁ has a relationship of 0.6 to 1.1, and a groove shape is engraved in a cross section perpendicular to the axis of the shaft member, A fluid bearing device characterized by the above-mentioned. 3. A fluid bearing device comprising: a shaft member that is in contact with a sleeve member and is rotatable relative to the sleeve member, wherein the inner peripheral surface of the sleeve member and the outer peripheral surface of the shaft member A radial bearing portion formed on one of the rotary sliding surfaces, wherein the radial bearing portion has a plurality of substantially rectangular shapes along the rotation axis direction about a line perpendicular to the axis of the shaft member; The plurality of grooves are symmetrical U-shaped grooves having a tilt angle of 45 to 70 degrees on one side by symmetrically forming a line perpendicular to the axis of the shaft member, and In a cross section perpendicular to the axis of the shaft member, the groove shape is carved in an arc shape,
A gas bearing device, wherein a gap between the shaft member and the sleeve member is filled with air as a lubricating fluid. 4. Contacting the sleeve member via lubricating oil,
A fluid bearing device comprising a shaft member that is rotatable relative to the sleeve member, wherein the fluid bearing device is formed on one of a rotation sliding surface of an inner peripheral surface of the sleeve member and an outer peripheral surface of the shaft member. A radial bearing portion, the radial bearing portion comprises a plurality of substantially C-shaped grooves along a rotational axis direction around a line perpendicular to the axis of the shaft member, the plurality of grooves, A symmetrical C-shaped groove having an inclination angle of 45 to 70 degrees on one side with respect to a line perpendicular to the axis of the shaft member, and a cross section perpendicular to the axis of the shaft member 3. The fluid bearing device according to claim 1, wherein the groove shape is carved in an arc shape.