JPH112234A - Dynamic pressure gas bearing structure - Google Patents

Abstract

PROBLEM TO BE SOLVED: To increase rotation precision in high speed rotation and a bearing life by providing a cylindrical shaft body and a hollow cylindrical bearing body opposed to the shaft body at an interval in the radius direction and providing at least one clearance variation part with a predetermined clearance changing rate in the clearance between them. SOLUTION: This structure is provided with a cylindrical shaft body 1 and a hollow cylindrical shaft bearing body 2 opposed to the shaft body 1 at a certain interval in the radius direction. A clearance 3 between the shaft body 1 and the shaft bearing body 2 is provided with at least one clearance variation part 31. In the clearance variation part 31, thickness of the clearance 3 is varied with respect to the center angle corresponding to the circumference along the outer circumference of the shaft body 1. When a diameter of the shaft body 1 is represented by D1 , a thickness changing quantity of the clearance 3 is represented by Δh, and a changing quantity of the center angle is represented by Δθ, a clearance changing rate α is represented by an equation: α=(Δh/D1 )/Δθ[/ deg.]. The clearance changing rate αis within a range of 1.0×10<-4> <=α<=10.0×10<-4> .

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention generally relates to a dynamic pressure gas bearing structure, and more particularly to a dynamic pressure gas bearing structure for supporting a rotating body rotating at a high speed.

[0002]

2. Description of the Related Art In recent years, with the increase in storage capacity and the shortening of access time of a rotation drive unit of a magnetic recording apparatus, for example, a hard disk driver (hereinafter, referred to as an "HDD"), an HDD has been required. The driving spindle motor is required to have a correspondingly high rotation speed and high rotation accuracy. It has been proposed to use an air bearing (dynamic pressure gas bearing) for the rotation drive unit in order to rotate a precision motor requiring such high rotation speed and high rotation accuracy at higher speed. In the rotary drive unit using the air bearing, when the rotating body rotates, air is forcibly introduced into at least a gap between the radial gas bearing body and the rotating body. This allows
The air pressure in the gap is increased, and the rotating body rotates at a high speed via the air bearing. In this way, the use of the air bearing is expected to maintain the rotation accuracy even during high-speed rotation.

[0003] In the radial type gas bearing described above, for example, Shinichi Togo, "Gas Bearing", Kyoritsu Shuppan (1)
984), a wedge-shaped gap is formed by the eccentricity of the shaft body in the bearing body. When the air passes through the wedge-shaped gap, the pressure is generated because the air is compressed. This makes it possible to support the shaft body and the bearing body in a non-contact manner.

However, Atsushi Mori, "On Whirl of Gas Bearings", pp. 481-488, "Lubrication", No. 20
According to Vol. 7 (1975), when a cylindrical journal bearing is placed in a no-load state such as when supporting a vertical shaft, an unstable phenomenon called “whirl” is observed. This phenomenon causes the shaft to oscillate inside the bearing such that the shaft is pressed against the bearing surface by centrifugal force at any rotational speed. In a cylindrical journal bearing, a pressure is generated at one place due to a deviation of the center of rotation from the center of rotation due to a static load, thereby providing stable rotation. However, when a cylindrical journal bearing is used in a vertical structure, that is, a structure that supports a vertical shaft, the bearing is placed in a no-load state, so that the pressure generation point changes due to disturbance, and rotation becomes unstable. .

When the above-described dynamic gas bearing is applied to the HDD, the positional accuracy of the rotating body is regarded as important.
It is necessary to eliminate the above instability factors.

Therefore, a dynamic pressure gas bearing structure capable of maintaining high rotational accuracy at high speed rotation is disclosed in Japanese Patent Laid-Open No. Hei 8-3.
12639.

However, in the proposed dynamic pressure gas bearing structure, at high speed rotation, the shaft body and the bearing body can rotate stably without contact due to the generated dynamic pressure. The number of rotations when rotating in a state where the body and the bearing body are in contact with each other is relatively large. For this reason, there is a problem that the shaft body and the bearing body are likely to come into contact with each other at a high rotation speed at the time of starting and stopping the rotation to cause damage or seizure. Therefore, in the dynamic pressure gas bearing structure proposed in the above publication, the service life may be shortened.

An object of the present invention is to provide a dynamic pressure gas bearing structure that can maintain high rotation accuracy at high speed rotation and can increase the life of the bearing.

[0009]

SUMMARY OF THE INVENTION A dynamic pressure gas bearing structure according to the present invention comprises a cylindrical shaft and a hollow cylindrical bearing opposed to the shaft with a gap in the radial direction. .
When the shaft and the bearing are arranged such that their respective central axes coincide, a substantially cylindrical gap is formed by the outer peripheral surface of the shaft and the inner peripheral surface of the bearing. In the cross-sectional shape perpendicular to the axis of the shaft body and the bearing body, a point where a radial straight line passing through the center axis intersects an outer shape line corresponding to the outer circumferential surface of the shaft body, and an outer shape corresponding to the inner circumferential surface of the bearing body. The distance (h) of the air gap is defined by the distance between the point at which the line intersects.

[0010] The gap defined in this way is at least one.
It has two void changing portions. In the gap change part,
The thickness of the air gap changes with respect to the center angle corresponding to the circumference along the outer peripheral surface of the shaft. Assuming that the diameter of the shaft body is D ₁ , the amount of change in the thickness of the gap is Δh, and the amount of change in the central angle is Δθ, the rate of change of the gap α is represented by the following equation.

Α = (Δh / D ₁ ) / Δθ [/ °] The gap change portion has a gap change rate in the following range.

1.0 × 10 ⁻⁴ ≦ α <10.0 × 10 ⁻⁴ The length of the gap change portion along the longitudinal direction of the shaft is the length of the cylindrical gap along the longitudinal direction of the shaft. 20% or more of 90
% Or less.

[0013] Preferably, the gap changing portion is formed in the cylindrical gap other than at least one end in the longitudinal direction of the shaft. In other words, the gap change portion does not exist at at least one end in the cylindrical gap.

[0014] Preferably, the gap has a constant gap portion having a substantially constant thickness, and a gap expansion portion having a thickness larger than the constant thickness. The gap enlargement portion includes a gap change portion.

Preferably, the difference (h _max −) between the thickness (h ₀ ) of the fixed gap portion and the maximum thickness (h _max ) of the gap enlarged portion expressed as a ratio to the diameter (D ₁ ) of the shaft body. h
₀ ) / D ₁ is 0.007 or less.

More preferably, the enlarged gap portion has a width (W) corresponding to a central angle of 5 ° or more.

The diameter of the shaft (D₁ )
Thickness (h₀ / D ₁ ) Is 0.001
It is preferably 25 or less.

By making the cross-sectional shape of at least one of the shaft body and the bearing body non-circular, an enlarged gap portion is formed.

It is preferable that three or more gap enlarged portions are arranged along the outer peripheral surface of the shaft body.

In the present invention, when the shaft body and the bearing body are arranged so that their respective central axes coincide with each other, a gap change portion having a predetermined gap change rate is formed by an outer circumferential surface of the shaft body and an inner circumferential surface of the bearing body. It is formed between the surface. Therefore, the wedge-shaped gap can be formed without the shaft body being eccentric inside the bearing body. Therefore, when air, lubricating oil, or the like flows through the gap formed by the shaft and the bearing, a dynamic pressure is generated by a wedge effect.

Specifically, as shown in detail in part A of FIG. 3, the flow path of the fluid expands in the expanded portion of the gap change portion, so that the density of stream lines per unit sectional area is reduced. descend. As a result, a negative pressure portion is formed. On the other hand, in the reduced portion of the gap change portion, the streamline density per unit cross-sectional area increases, so that the pressure increases.
The pressure generated in this way supports the radial load.

In order to efficiently perform such a pressure change, the air gap change rate α is 1.0 × 10 ⁻⁴ / ° or more and 10.0 or more.
It must be within the range of less than × 10 ⁻⁴ / °. This is because if the rate of change of the gap is out of the above range, the effect of the shape of the gap change portion cannot be sufficiently obtained, and the pressure rise due to entrainment of the viscous fluid becomes insufficient.

With respect to the stability in a state where the shaft body and the bearing body are rotating while maintaining a non-contact state due to the generated dynamic pressure,
A gap changing portion having the above-described shape is required. In the present invention, the bearing structure having a longer life can be obtained as a result of examining and paying attention to the process from rotation to the rotation in the non-contact state (referred to as a floating state) from the stopped state. it can.

When the rotation of the shaft or the bearing starts, the shaft and the bearing rotate relatively while in contact with each other.
When the rotation speed increases, the generated dynamic pressure increases, and when the rotation speed reaches a certain rotation speed, the shaft body and the bearing body are brought into a non-contact state, that is, a floating state. The rotation speed at this time is called a floating rotation speed. This means that the shaft or the bearing is rotating while the shaft and the bearing are in contact with each other until the floating speed is reached.

In consideration of the influence of the contact on the shaft and the bearing, it is preferable that the rotational speed in the contact state is small. In other words, the smaller the floating rotation speed is, the less the damage caused by the start and stop of the rotation is.

The floating speed decreases as the dynamic pressure required for floating increases. The dynamic pressure required for floating, apart from the wedge-shaped effect, becomes more pronounced as the gap is smaller. Therefore, the gap changing portion is a portion in which the gap is widened, and acts in a direction to reduce the dynamic pressure.

In order to effectively generate the dynamic pressure for the floating, the length of the gap change portion along the longitudinal direction of the shaft body is determined by adjusting the length of the gap between the shaft body and the bearing body. If it is 90% or less of the length, dynamic pressure is effectively generated in a portion where the gap change portion does not exist, and the floating rotation speed is reduced. The length of the gap change portion is 20% of the length of the cylindrical gap.
When the value is less than the above, the wedge-shaped effect becomes small, and the above-mentioned shape effect of the gap changing portion cannot be sufficiently obtained, so that the stability in the floating rotation state cannot be obtained.

The cylindrical gap means a cylindrical gap 3 formed by the outer peripheral surface of the shaft 1 and the inner peripheral surface of the bearing 2 shown in FIG. Also,
As shown in FIG. 2, the length of the cylindrical gap 3 means a length along a longitudinal direction of the shaft body 1 at a portion where the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 face each other. I do. The gap changing portion 31 is formed in a part of the cylindrical gap 3. Void change part 3
The length of 1 is defined as the length along the longitudinal direction of the shaft body 1. The ends of the cylindrical gap 3 mean one end and the other end along the longitudinal direction of the shaft 1 at a portion where the outer peripheral surface of the shaft 1 and the inner peripheral surface of the bearing 2 oppose each other.

Preferably, the gap changing portion 31 is formed in the cylindrical gap 3 at a position other than at least one end in the longitudinal direction of the shaft 1. In other words, the gap changing portion 31
Is a cylindrical gap 3 formed between the shaft 1 and the bearing 2.
Is not preferably present at at least one end.
The end where the gap change portion 31 does not exist may be both ends of the cylindrical gap or any one of the ends.

As described above, the gap changing portion tends to reduce the dynamic pressure for floating, but when the gap changing portion exists at the end of the cylindrical gap, the dynamic pressure is reduced more remarkably. When a gap change portion exists at the end of the cylindrical gap, the fluid inside the gap change portion and the fluid outside the bearing portion communicate. That is, the fluid inside the gap change portion can enter and exit the space outside the bearing structure. This is due to the pressure (dynamic pressure) generated in the gap.
Is escaped to the external space through the gap change portion. As a result, the generated dynamic pressure tends to decrease, and the floating speed tends to increase.

The thickness (h₀ ) And gap enlargement
Min thickness (h_max) And the difference (h _max-H₀ ),sand
That is, the maximum depth (d) of the enlarged space portion is determined by the diameter (D
₁ ) To 0.00025 or more and 0.007 or less
Preferably. The maximum depth of the enlarged space exceeds the upper limit
The effect of the shape of the wall surface of the enlarged space does not work.
The dynamic pressure effect by the rust effect cannot be obtained. Also, the gap expansion
The lower limit of most of the maximum depth (d) is the required accuracy in implementation
Depends on.

It is preferable that the enlarged space portion has a width (W) corresponding to a central angle of 5 ° or more. This is the width (W)
If the angle is less than 5 °, it is considered that the flow of the fluid does not sufficiently develop with respect to the change in the shape of the gap, and the desired effect cannot be obtained. The upper limit of the width (W) is determined by the number of the gap-enlarging portions arranged on the outer peripheral surface of the shaft, and the ratio of the gap-enlarging portion to the circumference of the cross-sectional shape of the shaft is two minutes. Is preferably 1 or less. This is thought to be due to the fact that, when the gap-enlarged portion having a large thickness increases with respect to the fixed gap portion, the volume of the fluid flowing between the shaft body and the bearing body increases, so that the efficiency of dynamic pressure generation decreases. Can be

The thickness (h ₀ ) of the constant gap portion is preferably 0.00125 or less as a ratio to the diameter (D ₁ ) of the shaft body. In other words, the difference between the diameter of the shaft and (D ₁₎ to the inner diameter of the bearing member _{(D 2) (D 2 -D} 1) is in the range 0.0025 or less with respect to the axis body diameter (D ₁₎ preferable. The reason for this is that, like the effect of the maximum depth of the gap-enlarging portion, if the thickness of the constant-gap portion increases, the shape effect of the gap does not work sufficiently and the dynamic pressure generating effect of the wedge-shaped gap does not function. Can be

The enlarged gap portion is obtained by forming at least one of the outer peripheral surface of the shaft body and the inner peripheral surface of the bearing body into a non-circular cross-sectional shape. Note that the cross-sectional shapes of the shaft body and the bearing body can be determined by the shape of the prescribed gap described above.

When the present invention is applied to an actual bearing structure, it is necessary to support a load in the radial direction, and therefore it is necessary to balance mechanically at least two places. For this reason, specifically, it is necessary to arrange the above-mentioned gap enlarged portion at two places along the outer peripheral surface of the shaft body. However, if only two gap-enlarging portions are provided, there is a possibility that dynamic fluctuation easily occurs with respect to disturbance in a direction perpendicular to the balancing direction. Therefore, it is more preferable that the mechanical balance is achieved at three or more locations, that is, three or more locations are provided along the outer peripheral surface of the shaft body.

In order to obtain a pressure rise more efficiently,
It is preferable that the rate of change in the gap between the expanded portion and the reduced portion is different in the gap change portion. Further, it is preferable that the pressure is gradually reduced in the expansion portion accompanied by the pressure drop, so that the flow loss is suppressed as much as possible. In the reduced part,
It is considered effective to suppress the loss of fluid energy due to friction on the wall surface by rapidly increasing the pressure. Even when the rate of change of the gap between the expanded portion and the reduced portion is made different, the rate of change of the gap must be within the above range.

Also, the dynamic pressure gas bearing structure of the present invention can obtain excellent rotational accuracy even when applied to a horizontal structure, that is, a bearing structure for supporting a horizontal shaft. Even in a state where the shaft body is eccentric in the bearing body due to the static load in the horizontal structure, pressure is generated by the gap changing portion of the present invention in addition to the wedge-shaped gap formed between the shaft body and the bearing body due to the eccentricity. . Therefore, in the horizontal structure, it is considered that rotation stability is increased by the same operation and effect as in the case of the vertical structure.

[0038]

【Example】

Example 1 First, the inventors logically simulated the influence of the shape of the gap formed between the shaft body and the bearing body on the generated dynamic pressure by numerical analysis.

FIG. 3 is a cross-sectional view showing the outer peripheral surface of the shaft and the inner peripheral surface of the bearing used to define the shape of the gap in the dynamic pressure gas bearing structure of the present invention. As shown in FIG. 3, a gap 3 is formed between the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2. Shaft 1 has a diameter D _1. Bearing body 2
It has an inner diameter D _2. The thickness of the gap 3 is represented by h.

As shown in detail in part A of FIG. 3, an enlarged gap portion and a fixed gap portion are defined. The gap enlargement portion includes a gap change portion. The gap enlarged portion has a width W (represented by a central angle corresponding to a circumference which is a cross-sectional shape of the shaft).
The gap enlarged portion has a base width Wb. The gap changing portion has an expanded portion and a reduced portion. The thickness of the constant gap is h
Represented by ₀ . The maximum thickness of the air gap expanding portions is represented by h _max. The thickness of the air gap constant part h ₀ is one-half of the difference between the inner diameter D ₂ of diameter D ₁ and the bearing of the shaft (diameter difference). The maximum depth d of the gap enlarged portion is the maximum thickness h of the gap enlarged portion.
It is represented by the difference between _max and the thickness h ₀ of the constant gap portion. The void change rate is represented by a slope α. The shape of the gap defined as described above is determined when the shaft body 1 and the bearing body 2 are arranged so that the central axes 50 coincide with each other.

FIG. 4 is a cross-sectional view showing the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 shown for defining the shape function and the air gap function of the radial bearing structure as shown in FIG. is there. The gap 3 shown in FIG. 4 is defined by a gap function h (θ). The shape of the outer peripheral surface of the shaft body 1 is defined by a shape function g (θ). In this way, the dynamic pressure generated during rotation was obtained by numerical calculation by changing the shape of the gap defined by the function.

The numerical calculation was performed under the following assumptions. (I) The fluid (here, air) is an incompressible fluid, and the flow is laminar.

(Ii) The fluid satisfies the following equation with respect to the gap function h (θ) in the circumferential direction.

[0044]

(Equation 1)

As shown in FIG. 4, by substituting the air gap function h (θ) into the equation (1) corresponding to each circumferential position on the outer peripheral surface of the shaft body 1 defined by the central angle θ, The pressure generated at each circumferential position was determined.

[0046] (1) Thickness h ₀ Impact gap constant portion void rate of change has on the dynamic pressure in proportion to the diameter D ₁ 0.
000625 (diameter D _{1 of} shaft 1 and inner diameter D _{2 of} bearing ₂₎
Difference of a ratio to the diameter D ₁ 0.00125), 0.0 maximum depth d of the air gap expanding portions a ratio to the diameter D ₁
0125. The width W of the enlarged space portion was set to 60 ° at the central angle. At this time, the maximum pressure generated when the shaft body 1 and the bearing body 2 are arranged concentrically and the bearing body 2 is rotated at 20,000 rpm is obtained by numerical calculation, and the influence of the air gap change rate on the dynamic pressure is determined. investigated. The shape function g (θ) was defined as shown in FIG. 5, and only the void change rate α was changed.

As a result, as shown in Table 1, when the shaft body is a perfect circle, the pressure difference from the atmospheric pressure is 0, while the gap change rate α is 1.0 × 10 ⁻⁴ / ° or more. 10.0 × 10 ^-4 /
It was found that a pressure rise of 0.1 × 10 ⁵ Pa or more was obtained within a range of less than °. Therefore, it has been found that in the radial bearing structure of the present invention, a dynamic pressure is generated even if the shaft body and the bearing body are arranged concentrically, so that it is possible to provide a bearing structure with high rotation accuracy even at high speed rotation.

[0048]

[Table 1]

(2) Influence of Maximum Depth of Enlarged Gap on Dynamic Pressure Regarding a radial bearing structure having a gap change portion having a gap change rate in which good dynamic pressure is obtained in the above (1), The generated pressure was calculated by changing only the maximum depth d of the gap expanding portion. That is, the shape function g (θ) is shown in FIG.
And only the maximum depth d was changed.

As a result, the change in the generated dynamic pressure with respect to the maximum thickness d of the void enlarged portion expressed as a ratio to the diameter D ₁ is shown in Table 2.

[0051]

[Table 2]

When the ratio d / D ₁ of the maximum depth of the gap expanding portion to the diameter of the shaft body is 0.007 or less, at least the dynamic pressure necessary to support the high-speed rotating bearing body can be obtained. Therefore, it can be inferred that a dynamic pressure gas bearing structure with small fluctuations due to rotation can be obtained at high speed rotation.

(3) Influence of the width of the enlarged gap on the dynamic pressure In the radial bearing structure having the shape function g (θ) shown in FIG. 7, the pressure generated when only the width W of the enlarged gap is changed. Was calculated. The results are shown in Table 3.
Here, the case where the gap enlarged portion includes only the gap change portion (Wb = 0) and the case where the gap enlarged portion has a bottom portion (Wb>)
The calculation was performed for each of 0).

[0054]

[Table 3]

As shown in Table 3, when the width W of the enlarged gap portion increases, the pressure difference from the atmospheric pressure increases.
It can be inferred that a dynamic pressure gas bearing structure with smaller fluctuations can be obtained at high speed rotation.

(4) Influence of thickness of fixed gap portion on dynamic pressure In a radial bearing structure having a shape function g (θ) shown in FIG. 8, thickness h _{0 of} fixed gap portion (diameter D of shaft body)
The generated pressure when only the difference 2h ₀ ) between ₁ and the inner diameter D ₂ of the bearing body was changed was calculated. The results are shown in Table 4.

[0057]

[Table 4]

As shown in Table 4, if the diameter difference 2h ₀ is 0.0025 or less as a ratio to the diameter D ₁ , dynamic pressure is generated. Therefore, if the diameter difference is within the range, it can be inferred that a dynamic pressure gas bearing structure with little fluctuation during high-speed rotation and high rotation accuracy can be obtained.

Embodiment 2 On the basis of the above simulation results, a radial bearing structure was actually manufactured for a good dynamic pressure generation result that was stable during high-speed rotation, and the floating rotation speed due to the generated dynamic pressure was measured. Was measured. The void change rate α is 3.0 ×
10 ⁻⁴ / deg, the number of the void enlarged portions (the number of processed portions) was 3, the depth d / D ₁ was 0.00125, and the width W of the void enlarged portion (the processed portion) was 10 °. The shape function g (θ) shown in FIG. 10 was employed.

Using a shaft having an outer diameter of 10 mm, a diameter difference of 2 h
₀ / D ₁ is 0.000625, and the length (length of the cylindrical gap) along the longitudinal direction of the shaft body 1 where the outer circumferential surface of the shaft body 1 and the inner circumferential surface of the bearing body 2 face each other is 25 mm. The bearing structure was configured such that

As shown in FIG. 12, a method for processing a shaft having the shape function shown in FIG. 10 is as follows. The grinding wheel 6 is indicated by an arrow P using the inner peripheral surface of the annular grinding wheel 6 as a grinding surface 61. The outer peripheral surface of the shaft body 1 was ground as shown by a broken line while being rotated in the direction shown in FIG.

As shown in FIG. 13, the length of the portion where the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 face each other, that is, the length of the gap 3 (25 mm) is the processed portion (gap enlargement). part)
, That is, the length of the gap changing portion 31 is 12.5 m
m (50% of the length of the gap). As shown in FIGS. 13A to 13E, the number and positions of the gap changing portions 31 were five types.

The shaft manufactured as described above was used as a fixed shaft, and the bearing was used as a movable body. As shown in FIG. 16, the motor 100 was manufactured by incorporating the shaft 1 and the bearing 2. A lid 7 was provided on the upper part of the bearing body 2. Bearing body 2
The magnet 8 was provided at the lower part of. A similar magnet 8 is provided on the outer peripheral surface of the shaft body 1 so that a repulsive force acts on the magnet 8 provided below the bearing body 2. Further, an annular disk 9 was provided on the upper outer peripheral surface of the bearing body 2.

By rotating the motor 100 thus constructed at high speed, that is,
After rotating at a rotation speed of 0 rpm, the driving source for the rotation was stopped and the bearing body 2 was left until stopped. At this time, the shaft 1
The number of rotations at which the bearing and the bearing body 2 started to contact was measured, and this was taken as the floating rotation number.

Table 5 shows the measurement results of the floating rotation speed.

[0066]

[Table 5]

When the machining position of the gap changing portion 31 with respect to the shaft body 1 is compared with the machining position at both ends of the gap 3 (D in FIG. 13), the machining position at one end of the gap 3 (A and B in FIG. 13)
Those not provided at both ends (C and E in FIG. 13) showed that the dynamic pressure generated for floating was high and the floating rotation speed was low.

Embodiment 3 On the basis of the simulation result of Embodiment 1, for which good dynamic pressure was obtained, several radial bearing structures were manufactured, and the number of floating rotations was measured by the generated dynamic pressure. Was performed. The void change rate α is 3.0 × 10 ⁻⁴ /
The number of degs, the number of portions with enlarged voids (the number of processed portions) was 3, and the depth d / D ₁ was 0.00125.

As the shape function g (θ), those shown in FIGS. 9 and 10 were employed. In actual production of the shaft, a shaft having a perfect circular cross section was machined to obtain a desired shape function.

Specifically, the shaft body was processed as follows. The shape function shown in FIG.
While rotating the grinding wheel 4 in the direction indicated by the arrow P on the V block 5 as shown in FIG.
Is moved in the direction indicated by the arrow Q to perform grinding.

As shown in FIG. 12, a method for processing a shaft body having the shape function shown in FIG. 10 uses the inner peripheral surface of an annular grindstone 6 as a grinding surface 61 and indicates the grindstone 6 by an arrow P. The outer peripheral surface of the shaft body 1 was ground as shown by a broken line while being rotated in the direction shown in FIG.

The processing of the shaft having the shape function shown in FIG. 10 was performed by a method different from the above two methods.
That is, after finishing the grinding shown in FIG. 11, the shaft body 1 was slightly rotated, the grinding shown in FIG. 11 was performed again, and the processing was repeated until the processed portion became a predetermined width. . According to this processing method, in the actual shape function, as shown in FIG. 14, irregularities are formed in the bottom portion that defines the maximum thickness of the enlarged-gap portion. In this case, FIG.
As shown in 4, by employing an average depth d _m as maximum depth of the air gap expanding portions, it believed shaft having the same shape function and FIG.

As a comparative example, as shown in FIG. 15, the shape function (α =
A radial bearing structure was also manufactured for a shaft using a shaft having 20 × 10 ⁻⁴ / deg), and the floating speed was measured.

When machining the enlarged gap portion (gap changing portion) on the shaft body 1, machining is performed from the end face of the shaft body 1 and the length of the enlarged gap portion (machining length) is changed to produce several types of shafts. I got a body.

The shaft manufactured as described above was used as a fixed shaft as in Example 2, and the bearing was used as a movable body. FIG.
As shown in (1), the motor 100 was manufactured by incorporating the shaft body 1 and the bearing body 2. A lid 7 was provided on the upper part of the bearing body 2. A magnet 8 was provided below the bearing body 2. A similar magnet 8 is provided on the outer peripheral surface of the shaft body 1 so that a repulsive force acts on the magnet 8 provided below the bearing body 2. Further, an annular disk 9 was provided on the upper outer peripheral surface of the bearing body 2.

In the manufacture of the above radial bearing structure, a shaft having an outer diameter of 10 mm was used, and the length of the shaft facing the bearing, that is, the length of the cylindrical gap was set to 25 mm. And the bearing body were assembled.

The motor 100 thus configured is rotated at high speed, that is, the bearing 2 is
After rotating at a rotation speed of 0 rpm, the driving source for the rotation was stopped and the bearing 2 was left until the rotation of the bearing 2 stopped. Then, the rotation speed at which the shaft body 1 and the bearing body 2 started to contact was measured, and this was defined as the floating rotation speed.

Table 6 shows the results obtained by using the shaft having the shape function shown in FIG. The result of using the shaft body provided with the shape function shown in FIG. 10 and the shaft body processed by the method shown in FIG. 12 was obtained by using the shaft body manufactured by the processing method shown in Table 7 and FIG. The results are shown in Table 8. As a comparative example, the results obtained by using the shaft body having the shape function shown in FIG. 15 are shown in Table 9.

In Tables 6 to 8, comparative examples indicate those in which the processing length (length of the gap change portion) is out of the range of the present invention.

[0080]

[Table 6]

[0081]

[Table 7]

[0082]

[Table 8]

[0083]

[Table 9]

Embodiment 4 A shaft body is machined by further changing the condition of the shape function.
In the same manner as in the above, a radial bearing structure was prepared, and the floating rotation speed was measured.

A shaft body and a bearing body having the same dimensions as those of the third embodiment are used, the gap changing portion α is 3.0 × 10 ⁻⁴ / deg, and the processing length from the end face of the shaft body (length of the gap changing portion) Was 7 mm.

FIG. 9, FIG. 10, and FIG.
Was adopted. A shaft body was manufactured by the processing method described in Example 3.

As a comparative example, a radial bearing structure was manufactured for a shaft body and a bearing body having a perfect circular cross section, and the number of floating revolutions was measured.

The assembling of the motor and the measurement of the floating speed were performed in the same manner as in Example 3.

Table 10 shows the results obtained by using the shaft having the shape function shown in FIG. The results obtained by using the shaft body having the shape function shown in FIG. 10 and the shaft body processed by the method shown in FIG. 12 were obtained by using the shaft body manufactured by the processing method shown in Table 11 and FIG. The results are shown in Table 12. Also, as a comparative example, the result of using a perfect circular structure is shown in Table 1.
It is shown in FIG.

In Tables 10 to 12, the comparative examples include the depth d / D ₁ , the width W, and the diameter difference 2h ₀ / D.
Any number of ₁ and the gap enlarged portion (the number of processed points) indicates those outside the scope of the present invention.

[0091]

[Table 10]

[0092]

[Table 11]

[0093]

[Table 12]

[0094]

[Table 13]

In the comparative examples having a processing length of 4 mm in Tables 6 to 8 and the comparative examples in Tables 9 to 13, the floating rotation speed was 2000 rpm.
It is considered that the value of m or more is due to the influence of instability in the floating state.

As is clear from the above results, in the example of the present invention, it is recognized that the rotational speed at which the shaft body and the bearing body are not in contact, that is, the floating rotational speed is extremely low, and the dynamic pressure generated by the rotation is large. Can be As a result, contact between the shaft body and the bearing body occurs in an extremely low rotation speed region, the impact due to the contact is small, and breakage and seizure are extremely unlikely to occur. In addition, since the generated dynamic pressure is high, high stability is exhibited even with a load such as vibration from the outside while rotating at a high speed in a non-contact manner, and high rotation accuracy can be obtained.

On the other hand, in the comparative example, generation of sufficient dynamic pressure was not recognized, and it was found that the shaft body and the bearing body were in contact with each other at a high rotation speed, so that there was a high possibility of causing damage and seizure. In addition, since the generated dynamic pressure is low even during high-speed rotation, it is unstable, and high rotation accuracy cannot be obtained.

The embodiments disclosed above are to be considered in all respects as illustrative and not restrictive. The scope of the present invention is defined by the terms of the claims, rather than the examples above, and should be construed as including any modifications and alterations within the scope and meaning equivalent to the terms of the claims.

[0099]

As described above, by providing the gap and the gap change portion defined according to the present invention, a dynamic pressure gas bearing structure having high rotational accuracy under high-speed rotation and having a longer life is provided. be able to.

[Brief description of the drawings]

FIG. 1 is a diagram showing a cylindrical gap provided in a region where an outer peripheral surface of a shaft body and an inner peripheral surface of a bearing body face each other.

FIG. 2 is a diagram showing an example of a gap changing portion formed according to the present invention.

FIG. 3 is a cross-sectional view showing an outer peripheral surface of a shaft body and an inner peripheral surface of a bearing body used for defining a gap according to the present invention.

FIG. 4 is a diagram showing an outer peripheral surface of a shaft body and an inner peripheral surface of a bearing body as a model for a simulation calculation of the dynamic pressure gas bearing structure of the present invention.

FIG. 5 is a diagram showing a shape function used for examining the effect of a gap change rate on dynamic pressure.

FIG. 6 is a diagram showing a shape function used for examining the influence of the maximum depth of the gap enlarged portion on dynamic pressure.

FIG. 7 is a diagram showing a shape function used for examining the effect of the width of the gap enlarged portion on dynamic pressure.

FIG. 8 is a diagram showing a shape function used for examining the effect of the thickness (diameter difference) of a fixed gap portion on dynamic pressure.

FIG. 9 is a diagram showing a shape function used in Examples 3 and 4.

FIG. 10 is a diagram showing a shape function used in Examples 2 to 4.

FIG. 11 is a perspective view illustrating an example of a method of processing a shaft body.

FIG. 12 is a cross-sectional view showing another example of a method of processing a shaft body.

FIG. 13 is a longitudinal sectional view showing a schematic cross section of a shaft body and a bearing body showing various positions A to E of a gap change portion employed in the second embodiment.

FIG. 14 is a diagram illustrating a shape function used in Examples 3 and 4.

FIG. 15 is a diagram showing a shape function used in a comparative example.

FIG. 16 is a diagram showing a schematic configuration of a motor to which the dynamic pressure gas bearing structure of the present invention is applied.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Shaft body 2 Bearing body 3 Air gap 31 Air gap change part 50 Center axis

Claims (8)

[Claims] 1. A cylindrical shaft body, and a hollow cylindrical bearing body opposed to the shaft body while keeping a gap in a radial direction, wherein the shaft body and the bearing are aligned so that their respective central axes coincide with each other. When the body is arranged, a substantially cylindrical gap is formed by the outer peripheral surface of the shaft body and the inner peripheral surface of the bearing body, and in a cross-sectional shape perpendicular to the axis of the shaft body and the bearing body, The gap is determined by a distance between a point at which a radial straight line passing through a central axis intersects an outer shape line corresponding to an outer peripheral surface of the shaft body and a point intersecting an outer shape line corresponding to an inner peripheral surface of the bearing body. The gap has at least one gap change portion in which the thickness of the gap changes with respect to a central angle corresponding to a circumference along the outer peripheral surface of the shaft body. The diameter of the shaft is D ₁ , and the change in the thickness of the gap is
h, when the amount of change in the central angle is Δθ, the air gap change rate α is represented by α = (Δh / D ₁ ) / Δθ [/ °], and the air gap change portion is 1.0 × 10 ^-4 ≦ α <10.0 ×
A gap change rate in the range of 10 ^-4 , wherein the length of the gap change portion along the longitudinal direction of the shaft is the length of the cylindrical gap along the longitudinal direction of the shaft. A dynamic pressure gas bearing structure that is not less than 20% and not more than 90%. 2. The dynamic pressure gas bearing structure according to claim 1, wherein the gap changing portion is formed in the cylindrical gap other than at least one end in the longitudinal direction of the shaft. 3. The air gap according to claim 1, wherein the air gap includes a constant air gap portion having a substantially constant thickness, and an air gap expansion portion having a thickness greater than the predetermined thickness, wherein the air gap expansion portion includes the air gap change portion. Item 3. A dynamic pressure gas bearing structure according to item 1 or 2. 4. A difference between the shaft diameter (D ₁₎ of the gap constant portion represented by a ratio to the thickness (h ₀₎ and the maximum thickness of the air gap expanding portions (h _max) (h _{ma x} -h ₀₎
/ D ₁ is 0.007 or less, the dynamic pressure gas bearing structure according to claim 3. 5. The air gap enlargement portion has a central angle of 5 °.
The dynamic pressure gas bearing structure according to claim 3 or 4, having a width (W) corresponding to the above. 6. The thickness (h ₀ / D ₁ ) of the constant gap portion expressed as a ratio to the diameter (D ₁ ) of the shaft body is:
The dynamic pressure gas bearing structure according to any one of claims 3 to 5, which is 0.00125 or less. 7. The air gap enlargement portion according to claim 3, wherein the cross-sectional shape of at least one of the shaft body and the bearing body is formed as a non-circular shape to form the enlarged-gap portion. Dynamic pressure gas bearing structure. 8. The hydrodynamic gas bearing structure according to claim 3, wherein the enlarged gap portion is disposed at three or more locations along the outer peripheral surface of the shaft.