JP2007333004A - Hydrodynamic fluid bearing apparatus, spindle motor, and recording disk driving device equipped with this spindle motor - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a low-cost hydrodynamic fluid bearing apparatus which has a low friction loss and wear at the time of starting/stopping, and a high stability. <P>SOLUTION: A hydrodynamic fluid bearing apparatus, which comprises: a sleeve 2; a shaft 1 which rotates relatively with respect to the sleeve 2; and a lubrication fluid 5 filling a gap between the shaft 1 and the sleeve 2, characterized in that in either the inner peripheral surface 2a of the sleeve 2 or the outer peripheral surface of the shaft 1, a herringbone dynamic pressure groove constituted by a groove portion 10 great in gap with respect to facing periphery and a hill portion 12 having a small gap with respect to the facing periphery is formed, and in that the groove portion 10 is formed in 45°≤β≤75° of slope angle β with respect to rotational axis direction. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a hydrodynamic bearing device that is particularly small and thin, a spindle motor and a recording disk drive device that are configured using the hydrodynamic bearing device.

  In recent years, recording disk drive devices such as hard disks have been further reduced in size and can be attached to portable audio-visual devices and mobile phones such as mobile phones. As a result, the spindle motor built in the recording disk drive is also directly connected to the ease of use of the device, such as portability, freedom of direction of use, impact resistance, repeated start / stop, quick startup, and low power consumption. Performance that has been achieved has become more demanding than ever. Therefore, there is a need for a compact and low-cost hydrodynamic bearing device that is superior in quietness, impact resistance, repeated start-up life, bearing loss, reliability, etc., so that even better performance can be achieved in a small space. It has been.

  As a hydrodynamic bearing device, a herringbone type hydrodynamic bearing is generally used in which grooves arranged in the shape of a fish bone with grooves inclined with respect to the rotation axis face each other are provided on the bearing surface. (Patent Document 1). The dynamic pressure generation principle of this bearing type is to push the lubricant from the groove with a wide bearing gap into the hill with a narrow bearing gap by using the viscous characteristics of the lubricating fluid and the viscous force due to the relative speed between the shaft and the bearing. The squeeze effect generated by the above and the pumping effect of pushing into the bearing along the inclined groove portion pressurize the lubricating fluid filled in the bearing gap, and the rotating body is floated and stably rotated and held. The development of this bearing has a long history, and many studies have been conducted on the characteristics of the bearing at the center position in a high-speed rotation state, and an optimum specification is also required. However, due to the principle of dynamic pressure generation, during low-speed rotation with a small relative speed difference between the shaft and the bearing, the force for flowing the lubricating fluid is small, so that it is difficult to form a normal lubrication interface and the bearing holding capacity is low. For this reason, when starting and stopping, there was a weak point that the friction loss was large and it was easy to wear, but at the time of low speed rotation, the friction and wear characteristics under the condition that it was in contact with the bearing wall surface, and the optimum specification that also considered it The research was difficult and no improvement was found.

  In addition, the small and thin rotary machine supported by the fluid dynamic pressure bearing has an overhang structure due to its structural limitations, and the center of gravity of the rotating body cannot be placed in the middle part of the bearing in the axial direction. In many cases, the load is disconnected to the output end side where the load is mounted. In particular, since a thin bearing cannot provide a sufficient bearing span, a large load is applied to the bearing on the output end side, and the output end side portion of the bearing tends to be easily damaged.

JP 2004-56963 A

  The present invention has been made in view of the above-described current problems, and an object of the present invention is to provide a low-cost, small-sized hydrodynamic fluid bearing that has low friction torque and wear during start and stop and is stable even at high speed rotation. An apparatus, a spindle motor, and a recording disk driving device including the spindle motor are provided.

The invention according to claim 1 includes a substantially cylindrical shaft, a substantially cylindrical sleeve through which the shaft is inserted, and a lubricating fluid filled in a gap between the shaft and the sleeve, The shaft is held by the sleeve so as to be rotatable relative to the sleeve around the rotation axis, and the dynamic pressure at which a load is connected with one end of the shaft or the sleeve in the rotation axis direction as an output end In the hydrodynamic bearing device, one of the opposed outer peripheral surface of the shaft and the inner peripheral surface of the sleeve is defined as a G surface, and the other surface is defined as an F surface, which is generated by relative rotation between the shaft and the sleeve. In the rotational flow motion of the lubricating fluid, the relative rotational direction of the lubricating fluid when viewed with the G plane as a reference is defined as the ROT direction. The groove portion having an open end is formed with a dynamic pressure groove array alternately arranged in the circumferential direction in an inclined state with respect to the rotating shaft, and the G surface has a substantially reverse direction with respect to the rotating shaft. The two rows of the dynamic pressure grooves inclined to each other are arranged in a pair of herringbone-like shapes in which the groove opening ends are arranged on the side away from each other and faced substantially symmetrically with respect to the plane orthogonal to the rotation axis. At least one hydrodynamic bearing groove is formed in the rotational axis direction, and the groove width ratio indicates a ratio of the circumferential width angle (θ _p ) of the groove portion and the circumferential width angle (θ _l ) of the hill portion. θ ^* _p (= θ _p / (θ ₁ + θ _p )) is 0.4 ≦ θ ^* _p ≦ 0.6, and an acute angle side formed by the pair of dynamic pressure groove rows and the rotating shaft The angles β ₁ and β ₂ are 45 ° ≦ β ₁ ≦ 75 ° and 45 ° ≦ β ₂ ≦ 75 °.

The invention according to claim 2 is the hydrodynamic bearing device according to claim 1, wherein, in the pair of dynamic pressure groove rows, a boundary portion from the hill portion to the groove portion viewed in the ROT direction is formed. Angles β _1a , β _2a formed by the wall surface a to be formed and the rotation axis, and a wall surface b forming the boundary portion from the groove portion to the hill portion viewed in the ROT direction and the rotation shaft are formed. The acute angles β _1b and β _2b are 45 ° ≦ β _1b <β _1a ≦ 75 °, or 45 ° ≦ β _2b <β _2a ≦ 75 °.

  The invention according to claim 3 is the hydrodynamic bearing device according to claim 1 or 2, wherein, in the pair of dynamic pressure groove rows, the groove portion has a gap dimension with the F surface. It is formed in a tapered shape that gradually decreases in the ROT direction.

  Invention of Claim 4 is a hydrodynamic bearing apparatus in any one of Claim 1 thru | or 3, Comprising: In said pair of dynamic pressure groove | channel row | line | columns, the said dynamic pressure arrange | positioned at the said output end side at least The groove portion on the groove row side is formed in a tapered shape so that the gap dimension with the F surface gradually decreases toward the side where the pair of dynamic pressure groove rows face each other.

  A fifth aspect of the present invention is the hydrodynamic bearing device according to any one of the first to fourth aspects, wherein in the pair of dynamic pressure groove rows, the dynamic pressure disposed at least on the output end side. The groove width ratio on the groove row side is formed so as to gradually decrease toward the side where the pair of dynamic pressure groove rows face each other.

A sixth aspect of the present invention is the hydrodynamic bearing device according to any one of the first to fifth aspects, wherein the dynamic pressure is disposed at least on the output end side in the pair of dynamic pressure groove rows. The number N _{g of} the groove portions constituting the dynamic pressure groove array on the groove array side is 6 ≦ N _g ≦ 10, and preferably 7 ≦ N _g ≦ 9.

A seventh aspect of the present invention is the hydrodynamic bearing device according to any one of the first to sixth aspects, wherein the dynamic pressure is arranged at least on the output end side in the pair of dynamic pressure groove rows. the ratio of the bearing gap is the difference between the sleeve within a radius r of the groove array side to the shaft outer radius R Δ (= R-r) , the average depth h _p from the lands to the bottom of the groove A certain groove depth ratio h ^* _p (= h _p / Δ) is 0.5 ≦ h ^* _p ≦ 3.0.

  The invention according to claim 8 is the hydrodynamic bearing device according to any one of claims 1 to 7, wherein the dynamic pressure disposed at least on the output end side in the pair of dynamic pressure groove rows. The hill portion on the groove row side is formed in a taper shape such that the gap dimension with the F surface gradually decreases in the ROT direction.

The invention according to claim 9 is the hydrodynamic bearing device according to claim 8, wherein, in the pair of dynamic pressure groove rows, the dynamic pressure groove row side disposed at least on the output end side. the ratio of the amount of taper [delta] _l and bearing clearance delta of the lands toward the ROT direction ^{_{δ * l (= δ l /}} Δ) _{is, 1-cos (θ l /} 2) <δ * l <1-cos ( θ _l + θ _p ).

  A tenth aspect of the present invention is the hydrodynamic bearing device according to any one of the first to ninth aspects, wherein the dynamic pressure is arranged at least on the output end side in the pair of dynamic pressure groove rows. The corner of the wall surface b that forms a boundary portion from the groove portion to the hill portion as viewed in the ROT direction of the groove row is chamfered.

  The invention according to claim 11 is the hydrodynamic bearing device according to any one of claims 1 to 10, wherein the dynamic pressure disposed at least on the output end side in the pair of dynamic pressure groove rows. The bearing gap Δ on the groove row side is formed so as to gradually decrease toward the side where the pair of dynamic pressure groove rows face each other.

A twelfth aspect of the present invention is the hydrodynamic bearing device according to the eleventh aspect, wherein in the pair of dynamic pressure groove rows, the dynamic pressure groove row side disposed at least on the output end side. bearing clearance delta is, the pair of dynamic pressure groove array the gradually becomes smaller as the bearing gap formed variation [delta] _delta and towards the opposite side an average depth h _p and the ratio [delta] ^* _delta (= δ _Δ / h _p ) is 1/8 <δ ^* _Δ <1/2, preferably δ ^* _Δ≈1 / 3.

  A thirteenth aspect of the present invention is the hydrodynamic bearing device according to any one of the first to twelfth aspects, wherein the sleeve is made of a molded body of sintered porous metal or resin.

  The invention according to claim 14 is the hydrodynamic bearing device according to any one of claims 1 to 13, wherein the groove and the hill are formed in the sleeve.

  The invention according to claim 15 is an electric spindle motor, wherein the rotary shaft is centered between the rotor magnet having a field magnet arranged around the rotary shaft and the field magnet. And a stator part that rotatably supports the rotor part about the rotation shaft, and any one of the rotor part and the stator part comprises: The sleeve which comprises the hydrodynamic bearing apparatus in any one of thru | or 14 is attached, and the other of the said rotor part and the said stator part is provided with the shaft inserted in the said sleeve.

  The invention described in claim 16 is a recording disk drive device, wherein a head for reading and writing data on the recording disk in the housing, an actuator for moving the head on the disk surface, and A spindle motor.

  The hydrodynamic bearing device described in claim 1 of the present invention is composed of a plurality of hills and grooves on at least one of the outer peripheral surface of the shaft and the inner peripheral surface of the sleeve constituting the radial hydrodynamic bearing device. At least one pair of herringbone-shaped hydrodynamic bearing grooves is provided. The ratio between the hill portion and the groove portion is set to an optimum ratio that provides a large rigidity even during high-speed rotation used in the central portion of the bearing. When the lubricating fluid flows from the groove into the bearing gap of the hill, a large dynamic pressure is generated by the squeeze effect, but when the shaft eccentricity is large, the bearing gap with the hill is small, so it is difficult to flow in, The lubricating fluid tends to flow mainly in the groove. In particular, when the inclination angle of the groove is large, that is, when the groove is directed in the circumferential direction, the tendency becomes remarkable, so that it is difficult for pressure to be generated and the rotating body is difficult to float. As a result, wear during start-up becomes severe. In the invention of claim 1, when the eccentricity ratio is large, it is clarified that there is a groove inclination angle at which the amount of wear increases rapidly, and the groove angle is kept below that inclination and an appropriate angle is maintained. Therefore, stable rotation characteristics during high-speed rotation can be maintained.

  The hydrodynamic bearing device according to claim 2 of the present invention is optimal in the hydrodynamic bearing device according to claim 1 by setting the inclination angles of the two wall surfaces forming the groove with respect to the rotation axis to different angles. It has become. As described above, the smaller the angle of the wall surface from the groove to the hill, the more stable the positive dynamic pressure is, and the more stable, but conversely, if the angle of the wall surface from the hill to the groove is small, negative dynamic pressure is generated and cavitation occurs. Is likely to cause instability. Therefore, it is preferable that the angle of the wall surface from the groove to the hill is appropriately reduced, and conversely, the angle of the wall surface from the hill to the groove is appropriately increased.

  The hydrodynamic bearing device according to claim 3 of the present invention is such that, in the hydrodynamic bearing device according to claim 1 or 2, the bottom of the groove is tapered in the direction in which the lubricating fluid flows. Even if the bearing gap at the hill portion is narrow, the lubricating fluid is easily poured smoothly into the hill portion, so that the squeeze effect at the hill portion is likely to occur, and the hill is likely to rise even at a low speed. As a result, the amount of wear can be reduced. It should be noted that the influence of the cross-sectional shape of the groove itself is relatively small for the characteristics during high-speed rotation. Furthermore, in the overhang structure, it is easy to apply a load to the output end side, so the bearing specification on the output end side is the optimum specification at startup, and the opposite side of the output end is the specification at high speed rotation. By doing so, it is possible to achieve further optimization that it is easy to float at the time of start-up and can obtain a stable rotation at the time of high-speed rotation.

  A hydrodynamic bearing device according to a fourth aspect of the present invention is the hydrodynamic bearing device according to any one of the first to third aspects, wherein the herringbone-like hydrodynamic bearing groove is arranged from the end side to the center portion. By gradually decreasing the depth of the groove toward the surface, the flow path area is narrowed, thereby increasing the pressure increasing effect due to the flow accompanying the pumping effect in the groove. When combined with the structure of claim 2, the pressure increasing effect is further enhanced, so that it becomes easy to float and wear at the time of activation can be reduced.

  A hydrodynamic bearing device according to a fifth aspect of the present invention is the hydrodynamic bearing device according to any one of the first to fourth aspects, wherein the center portion is located from the end side of the herringbone-like hydrodynamic bearing groove. By increasing the groove width ratio and reducing the flow path area toward the surface, the pressure increasing effect due to the flow accompanying the pumping effect in the groove is enhanced. When combined with the structure of claim 4, the pressure increasing effect is further enhanced, so that it becomes easy to float and wear at the time of activation can be reduced.

  A hydrodynamic bearing device according to a sixth aspect of the present invention is the hydrodynamic bearing device according to any one of the first to fifth aspects, wherein the wear is achieved by setting the number of grooves to an optimum number. The amount can be reduced.

  A hydrodynamic bearing device according to a seventh aspect of the present invention is the hydrodynamic bearing device according to any one of the first to sixth aspects, wherein the groove depth ratio is set to an appropriate value. Reduce wear during startup. The optimum groove depth ratio for high-speed rotation at the center of the bearing is generally said to be the same as the bearing gap and groove depth, that is, about 1.0. In the starting state in contact with the bearing wall surface, the amount of wear increases as the groove depth ratio increases. Therefore, by suppressing the wear amount to a groove depth ratio or less where the amount of wear increases rapidly, and to a groove depth ratio or more that does not deteriorate the high-speed rotation characteristics, it is possible to maintain good characteristics at both startup and high-speed rotation.

  A hydrodynamic bearing device according to an eighth aspect of the present invention is the hydrodynamic bearing device according to any one of the first to sixth aspects, wherein the shape of the hill portion is tapered in the circumferential direction. The squeeze effect is enhanced by narrowing the bearing gap at the portion in the rotational direction. As described above, the optimum groove depth ratio when rotating at the center of the bearing at a high speed is generally said to be about the same as the bearing gap and the groove depth, that is, about 1.0. In the vicinity of the place where the rotating body is in contact with the bearing wall surface, the local bearing gap is substantially zero, and the local groove depth ratio is a very large value. Tapering the shape of the hill in the circumferential direction also corresponds to increasing the local bearing gap and reducing the local groove depth ratio.

  A hydrodynamic bearing device according to a ninth aspect of the present invention is a hydrodynamic bearing device according to the eighth aspect, in which a specific taper amount of a hill portion is numerically shown. The curvature of the smooth surface and the curvature at the contact portion of the surface provided with the grooves and hills are substantially the same, so that a smooth lubricating fluid can be introduced and an appropriate squeeze effect can be obtained. As a result, it is possible to reduce wear during startup.

  The hydrodynamic bearing device according to claim 10 of the present invention is the hydrodynamic bearing device according to any one of claims 1 to 9, wherein the corner of the wall surface extending from the groove on the downstream side of the lubricating fluid to the hill is provided. By chamfering, the lubricating fluid can easily flow into the hill and enhance the squeeze effect. As a result, it is possible to obtain a bearing that easily floats and has little wear.

  A hydrodynamic bearing device according to an eleventh aspect of the present invention is the hydrodynamic bearing device according to any one of the first to tenth aspects, wherein the bearing gap on the shaft end side of the herringbone-shaped hydrodynamic groove bearing is provided. By gradually increasing the width, the local groove depth ratio can be reduced, the flying characteristics can be improved, and the wear characteristics can be improved.

  A hydrodynamic bearing device according to a twelfth aspect of the present invention is the hydrodynamic bearing device according to the eleventh aspect, in which the value of the axial change in the bearing gap is optimized, Variation in the degree of improvement can be suppressed and quality can be improved.

  A hydrodynamic bearing device according to claim 13 of the present invention is the hydrodynamic bearing device according to any one of claims 1 to 12, wherein the sleeve is molded with a sintered porous metal or resin. Since sintered porous metals and resins are excellent in mass productivity, the cost can be reduced. Furthermore, it is very effective in terms of seizure resistance.

  A hydrodynamic bearing device according to claim 14 of the present invention is the hydrodynamic bearing device according to any one of claims 1 to 13, wherein grooves and hills are formed in the sleeve. If manufactured with a soft material such as sintered porous metal, the recesses can be easily formed by a process such as coining, so that the cost can be reduced.

  According to a fifteenth aspect of the present invention, there is provided an electric spindle motor having a torque centered on the rotating shaft between the rotor portion having the field magnet arranged around the rotating shaft and the field magnet. And a stator part that rotatably supports the rotor part about a rotation axis, and any one of the rotor part and the stator part is according to any one of claims 1 to 14. A spindle motor with a sleeve that constitutes a hydrodynamic bearing device attached, and the other of the rotor part and the stator part having a shaft inserted into the sleeve, and friction loss and wear during low-speed rotation such as starting and stopping Since the amount is small and stable high-speed rotation performance can be obtained, the reliability can be improved.

  The invention described in claim 16 is a recording disk drive device, wherein a head for reading and writing data on the recording disk in the housing, an actuator for moving the head on the disk surface, and It has a spindle motor and is reliable and inexpensive.

  The best mode for carrying out the present invention will be described with reference to the drawings. In the description of the present embodiment, expressions relating to directions such as up, down, left, and right indicate directions on the drawing, unless otherwise specified. Therefore, it does not limit the actual direction of implementation.

(First embodiment)
FIG. 1 is a sectional view of a hydrodynamic bearing device 6 and a spindle motor 30 including the hydrodynamic bearing device 6 according to the first embodiment of the present invention. The members to be omitted are omitted. In addition, the display of parallel diagonal lines in the cross section is omitted.

  The hydrodynamic bearing device 6 of the present invention includes a substantially cylindrical shaft 1, a substantially cylindrical sleeve 2 through which the shaft 1 is inserted, and a lubricating fluid 5 filled between the shaft 1 and the sleeve 2. The outer diameter side surface of the sleeve 2 is attached to the inner diameter side surface of the substantially cup-shaped bearing housing 4, and the upper end portion of the shaft 1 is attached to the rotor hub 3 having a top plate portion 3b extending in the radial direction. ing. The space surrounded by them is filled with the lubricating fluid 5. The sleeve 2 is a molded body made of sintered porous metal. The bearing housing 4 is made of aluminum, stainless steel, brass, resin, or the like. The shaft 1 is formed integrally with the rotor hub 3 by pressing or cutting.

  The sleeve 2 and the bearing housing 4 may be integrally formed. For example, the outer peripheral surface may be sealed after being integrally formed of aluminum, stainless steel, brass, or the like, or integrally formed of a sintered porous body. The sleeve 2 may be insert-molded with resin.

  The outer peripheral surface 1a of the shaft 1 and the inner peripheral surface 2a of the sleeve 2 are opposed to each other via a minute gap in the radial direction. Oil is held as a lubricating fluid 5 in the minute gap in the radial direction. Further, the upper end surface of the sleeve 2 and the upper end surface 4a of the bearing housing 4 face the lower surface 3a of the top plate portion 3b of the rotor hub 3 with a slight gap in the axial direction. In the minute gap in the axial direction, the same oil as that held in the minute radial gap is continuously held in the minute radial gap. The oil is continuously held between the lubricating fluid holding portion 16 and the lower end portion of the shaft 1 formed on the lower side of the sleeve 2 and the bottom surface of the bearing housing 4, and thus the oil is contained in the bearing housing 4. It is filled without interruption. This state is called a full fill state. A circulation groove 15 is formed on the outer peripheral surface of the sleeve 2, and the minute gap formed on the upper end surface side of the sleeve 2 and the lubricating fluid holding portion 16 formed on the lower side of the sleeve 2 are provided. By connecting, it is possible to circulate oil and discharge bubbles generated in the oil.

  The rotor hub 3 has a hanging wall 3 c on the outer periphery of the outermost peripheral edge of the bearing housing 4. Between the inner circumferential surface of the retaining member 21 attached to the hanging wall 3 c and the outer circumferential surface of the bearing housing 4. Only one interface 17 between the oil and air is formed. The space between the inner peripheral surface of the retaining member 21 and the outer peripheral surface of the bearing housing 4 is tapered so as to expand toward the outside of the bearing, and oil is prevented from leaking out of the bearing. Further, a step is provided in the circumferential direction on the outer peripheral surface of the bearing housing 4, and a retaining member 21 provided on the lower peripheral wall 3 c of the rotor hub 3 faces the step so that the shaft 1 and the sleeve 2 have a certain axial direction or more. Movement is restricted.

  When the shaft 1 is rotated with respect to the sleeve 2, the oil held in the minute radial gap generates dynamic pressure between the outer peripheral surface of the shaft 1 and the inner peripheral surface of the sleeve 2, and radial (radial direction) ) Is generated. Thus, the radial dynamic pressure fluid bearing device 6 is formed. Similarly, when the shaft 1 is rotated with respect to the sleeve 2, the oil held in the minute axial gap generates dynamic pressure between the upper end surface of the bearing housing 4 and the lower surface of the top plate portion 3b, Thrust (axial direction) load support pressure is generated. Thus, the thrust hydrodynamic bearing device 13 is formed.

  FIG. 2 is an enlarged cross-sectional view showing the first embodiment of the radial hydrodynamic bearing device 6 according to the present invention in an enlarged manner, and FIG. 3 omits the description of the shaft 1 from FIG. It is the figure which made it easy to see the inner surface structure of. FIG. 4 is a schematic view in which the sleeve 2 is cut open and developed. FIG. 5 is a simplified partial sectional view of the sleeve 2 in the direction perpendicular to the rotation axis. In these drawings, the display of parallel oblique lines in the cross section is omitted. In this embodiment, as shown in FIG. 2, the shaft 1 has a cylindrical outer shape, is formed in substantially the same shape from one end to the other end, and rotates in the direction of the arrow 14. With this rotation, the lubricating fluid 5 also rotates and flows in the same direction, so the relative flow rotation direction (ROT direction) of the lubricating fluid 5 with respect to the inner peripheral surface 2a (G surface) of the stationary sleeve 2 is the shaft. The direction of the arrow 14 is the same as the direction of rotation 1. As shown in FIG. 3, on the inner circumferential surface 2a (G surface) of the sleeve 2, a herringbone-shaped hydrodynamic bearing in which a groove for generating dynamic pressure is folded back at an intermediate portion to form a "V" shape. Grooves are provided. As shown in FIG. 4, a groove portion 10 (colored and displayed) having a large radial clearance with the outer peripheral surface 1a (F surface) of the shaft 1 and a hill portion 12 having a small radial clearance with the shaft 1 are circumferentially arranged. The wall surface 11a is formed at the boundary portion between the downstream end of the hill portion 12 and the upstream end of the groove portion 10 when viewed from the relative rotation direction (ROT direction) 14 of the lubricating fluid 5 alternately. A wall surface 11b is formed at the boundary between the downstream side of the hill and the upstream side of the hill portion 12. By the rotation of the shaft 1, the lubricating fluid 5 is configured to gather at the center portion of the dynamic pressure bearing groove facing the “V” shape from the lubricating fluid holding portion 16 side on both sides of the bearing portion.

As shown in FIG. 4, the angle formed by the groove 10 and the rotation axis is β on average, but the details differ depending on the location. The angle between the wall surface 11a of the upper groove and the rotation axis is β _1a , the angle between the wall surface 11b of the upper groove and β _1b is the angle between the wall surface 11a of the lower groove and the rotation axis is β _2a , the lower side The angle formed with the wall surface 11b of the groove is β _2b so that 45 ° ≦ β _1b <β _1a ≦ 75 ° or 45 ° ≦ β _2b <β _2a ≦ 75 ° is satisfied. , Β _1a = β _2a = 62 ° and β _1b = β _2b = 50 °. By setting in this way, the squeeze effect in the hill and the pumping effect in the groove can be enhanced.

Further, the groove width ratio θ ^* _p (= θ _p / (θ _l + θ _p )) indicating the ratio of the circumferential width angle θ _{p of} the groove portion and the circumferential width angle θ _l of the hill portion is 0.4 as a whole. ≦ θ ^* _p ≦ 0.6 and average 0.5, but with the difference in inclination angle of the wall surfaces 11a and 11b of the groove, the entrance portion is wide and the back side is narrow, that is, the entrance side The groove width ratio is large, and the groove width ratio gradually decreases toward the back side. As shown in FIG. 6, the wall surfaces 11a and 11b of the grooves may have the same inclination angle, the groove width ratio may be constant, and a symmetric simple shape centered on the V-shaped folded portion may be used. An asymmetrical shape in which the groove angle and the groove width ratio are changed at the folded portion may be used.

  In the present embodiment, since the shaft 1 having a smooth outer peripheral surface rotates, the actual rotation direction of the shaft 1 and the lubricating fluid 5 and the relative rotation of the lubricating fluid 5 with respect to the stationary sleeve 2 are used. The directions 14 are all directed in the same direction. However, for example, when the sleeve 2 in which the groove is formed rotates, the relative flow rotation direction 14 of the lubricating fluid 5 with respect to the sleeve 2 is different. So be careful. In this embodiment, the number of grooves is six, but the number of grooves is not limited to this, and may be seven, eight, nine, or ten. In particular, 7, 8, and 9 are good. In the present embodiment, the pair of dynamic pressure groove rows has a structure in which the grooves are connected to each other at the center. However, a circumferential hill is arranged in the middle portion to separate the two dynamic pressure groove rows. May be. FIG. 7 shows an example of a structure in which the number of grooves is 8, and a circumferential hill is arranged in the middle of a pair of dynamic pressure groove rows to separate the two dynamic pressure groove rows.

  Next, the shape of the groove part 10 and the hill part 12 is demonstrated with reference to FIG.8, FIG.9 and FIG.10. 8, 9, and 10 are cross-sectional views taken along a plane perpendicular to the rotation axis at a height position passing through the groove 10 and the hill portion 12, and the groove portion 10, the hill portion 12, and a part of the shaft 1 are illustrated. It is a model enlarged view shown. The radial rigidity of the hydrodynamic bearing device 6 when rotating at a high speed at the approximate center position of the sleeve 2 in a normal operation state is isotropic, and the shaft 1 can be reliably supported.

  In the present embodiment, the hill portion 12 is the same surface as the inner peripheral surface 2 a of the sleeve 2, and constitutes an arcuate surface in which the gap with the outer peripheral surface of the shaft 1 is uniform. The interval Δ between the peripheral surface of the hill portion 12 and the outer peripheral surface of the shaft 1 is small, for example, set to about 0.003 mm. The peripheral surface of the hill portion 12 cooperates with each other to define cylindrical support surfaces, and these support surfaces support the shaft 1 via the lubricating fluid 5 so as to be rotatable.

Further, the groove portion 10 has a shape recessed in the radial direction from the inner peripheral surface 2 a of the sleeve 2. Distance _{h p} of the upper surface of the bottom and the lands 12 of the groove 10, a fraction of several times the size of the delta, is set to, for example, about 0.0015Mm～0.009Mm. The circumferential surface of the groove portion 10 is tapered so as to gradually approach the shaft 1 from the downstream end of the groove portion 10 to the upstream end of the hill portion 12 as viewed in the relative rotation direction 14 of the lubricating fluid 5. (Fig. 8), or concentric circles with a uniform spacing from the shaft 1 to facilitate processing, and may be connected to the hills and steps (Fig. 8). 9). Furthermore, it is even better if the corner of the wall surface 11b of the groove is chamfered (FIG. 10). Even in this case, the groove depth ratio h ^* _p (= h _p / Δ) to the average groove depth is set so as to satisfy 0.5 ≦ h ^* _p ≦ 3.0. .

  A pressure generation state in the bearing will be described along the flow of the lubricating fluid 5. As described above, when the shaft 1 relatively rotates in the direction indicated by the arrow 14, the lubricating fluid 5 interposed between the sleeve 2 and the shaft 1 also rotates and flows in the direction indicated by the arrow 14. In this embodiment, the pressure of the lubricating fluid 5 is increased by the squeeze effect by flowing from the groove portion having a wide interval with the outer peripheral surface of the shaft 1 toward the hill portion having a small interval, and the shaft 1 is rotatably supported. . On the other hand, when the lubricating fluid 5 next flows from the hill portion to the groove portion due to the relative rotation of the shaft 1, the pressure of the lubricating fluid 5 raised in the hill portion 12 is released, and the boundary portion between the hill portion 12 and the groove portion 10. The pressure of the lubricating fluid 5 is significantly reduced. When negative pressure is generated, the lubricating film of the lubricating fluid 5 is broken, and the film state becomes unstable, causing non-reproducible run-out and reduction in radial rigidity.

Further, the groove portion 10 and the hill portion 12 are arranged so that the lubricating fluid 5 is pushed from the lubricating fluid holding portion 16 side into the central portion side of the herringbone-shaped dynamic pressure bearing groove by the rotation of the shaft 1 with respect to the rotation axis direction. Te beta _1, so inclined by beta _2, the pressure on the center side of the hydrodynamic bearing groove than more by the pumping effect is increased. As described above, in this embodiment, the groove width ratio is gradually reduced toward the center of the hydrodynamic bearing groove to improve the pressure in the groove.

  In addition, in this embodiment, the pressure in the groove is increased by gradually decreasing the groove depth toward the center of the hydrodynamic bearing groove. FIG. 11 is a longitudinal sectional view in which the central portion of the groove portion is cut along the inclination of the groove, and is a schematic enlarged view showing the groove portion 10, the hill portion 12, and a part of the shaft 1. As shown in FIG. 11, the groove depth is tapered such that the groove depth becomes tapered from both sides so that the V-shaped turn-back point is the shallowest. The groove depth taper may be provided only in the upper groove row portion close to the interface 17 of the lubricating fluid 5 in order to enhance the function of holding the lubricating fluid 5 in the bearing portion.

  Here, the characteristics of the dynamic pressure type bearing at the time of starting and stopping, which is one of the main points of the present invention, will be examined. In the hydrodynamic bearing represented by the herringbone type and the multi-arc type, the bearing gap in the circumferential direction is discontinuous in order to generate dynamic pressure. For example, as in this embodiment, the sleeve 2 is discontinuous. Recesses are formed. In a normal use state, the shaft 1 floats due to the dynamic pressure generated by the high-speed rotation and rotates while being held at a substantially central position of the sleeve 2. Therefore, the directionality of the circumferential bearing characteristics by these concave portions is ignored. In particular, when starting from a contact state in which the rotating body is placed horizontally and the gravity of the shaft 1 is applied in the direction of the inner wall surface of the sleeve 2, the start-up characteristics are large depending on the contact state with the recess at the start of start-up. You can expect it to change. Similarly, for the characteristics at the time of stopping, it is necessary to consider the influence of the shape and phase of the recess and the direction of gravity.

  Next, the results of confirming the effects of the groove portion 10 and the hill portion 12 in the hydrodynamic bearing device 6 of the first embodiment described above by numerical analysis using an analysis model will be described in detail. The hydrodynamic bearing device 6 is characterized by a finite element method so that transient behavior such as solid contact between the shaft 1 and the sleeve 2 at the start and stop and the boundary lubrication state in which the lubricating fluid 5 is interposed can be analyzed in detail. Thus, the average flow Reynolds equation and the pressure equation taking into account the elasto-plastic deformation of the fine irregularities on the bearing surface were established, and were obtained by simultaneous equations with the equation of motion of the rotating body. The numerical analysis was performed in detail by dividing it into several tens in the circumferential direction and the axial direction.

The numerical analysis model and symbols will be described with reference to FIGS. In the analysis model, all the V-shaped grooves are the same, and are symmetrical with respect to the vertex position of the V-shape. The circumferential angle of the groove 10 was θ _p , the circumferential angle of the hill portion 12 was θ _l , and the circumferential angle for one pitch was 2θ (= θ _p + θ _l ). Further, the acute angle side of the inclination angle of the groove with respect to the rotation axis is β (= β ₁ = β ₂ = β _1a = β _1b = β _2a = β _2b ). In this numerical analysis, the diameter of the shaft 1 is 2.5 mm, the width in the axial direction of the hydrodynamic bearing device is 2.5 mm, and the distance Δ between the inner peripheral surface of the hill portion 12 and the outer peripheral surface of the shaft 1. the go 0.0025 mm, 0.0050Mm spacing _{h p} of the circumferential surface and the outer circumferential surface of the shaft 1 of the groove 10, as a reference value the combination of the six grooves number Ng, based on those values parameters survey It was. The oil viscosity coefficient of the lubricating fluid 5 was 0.013 Pa · s.

  FIG. 12 shows the result of analysis of dimensionless friction loss in the low-speed rotation region after startup. It can be seen that the herringbone bearing (hereinafter referred to as “HB”) cannot easily rise due to the low dynamic pressure in the low-speed rotation region, and finally starts to rise at about 200 rpm, and the friction loss is reduced. For this reason, the friction loss in a low speed region becomes large and the amount of wear also increases.

FIG. 13 shows the result of calculating the dimensionless wear amount V ^* _W of the sleeve at the time of start and stop, and shows the change with respect to the circumferential angle ratio θ ^* _p of the groove. The dimensionless wear amount V ^* _W tends to increase as the circumferential angle ratio θ ^* _p of the groove portion increases. Considering that the optimum θ ^* _p for high-speed stable rotation is about 0.5, it can be said that 0.4 ≦ θ ^* _p ≦ 0.6 is preferable.

FIG. 14 shows the calculation result of the relationship between the inclination angle β with respect to the rotation axis of the groove and the dimensionless wear amount V ^* _W of the sleeve. When the inclination angle β exceeds 70 °, the dimensionless wear amount V ^* _W increases rapidly. This means that the direction of the groove and the hill almost coincides with the flow direction (ROT direction) of the lubricating fluid, so that the squeeze effect due to the lubricating fluid flowing into the hill from the groove can hardly be obtained. ing. In addition, the groove depth ratio h ^* _p (= h _p / Δ), which is the ratio of the bearing clearance to the groove depth, is used as a parameter, but basically the same tendency is found, and the groove depth ratio h ^* _p The larger the is, the larger the base wear amount and the increase rate of the wear amount. FIG. 15 shows the calculation result of the relationship between the groove inclination angle β and the bearing stiffness K _xx during high-speed rotation. It can be seen that the inclination angle β that maximizes the bearing rigidity is in the vicinity of 40 °, and that if the inclination angle β is too small, the bearing rigidity decreases. Further, when the inclination angle β of the groove portion decreases, the wall between the groove and the hill blocks the flow of the lubricating fluid, so that the axial loss during high-speed rotation increases. Therefore, it is not preferable to make the inclination angle β of the groove portion too small, and it can be said that the groove angle β should be 45 ° ≦ β ≦ 75 ° in consideration of characteristics during high-speed rotation.

FIG. 16 shows the result of calculating the relationship between the groove depth ratio h ^* _p and the dimensionless wear amount V ^* _W of the sleeve with the groove angle β fixed at 70 °. The groove depth ratio was analyzed by changing the groove depth while keeping the bearing gap constant. The shaft is in contact with the upper surface of the hill portion, and the lubricating fluid does not easily flow into the gap from the groove portion, and it is difficult to obtain a squeeze effect. Therefore, in order to reduce the amount of wear, it is better that the groove is shallow, and it is substantially the same as the dynamic bearing clearance at the time of contact determined by the curvature of the shaft outer diameter and the sleeve inner diameter, and the profile of the hill and groove. It can be seen that it is better to make the groove depth to a certain extent. FIG. 17 shows the result of analyzing the relationship between the groove depth h ^* _p and the axial loss in the high-speed rotation state. In this way, when the groove depth is too shallow, 0.5 ≦ h ^* _p ≦≦ in consideration of an increase in axial loss during high-speed rotation and a tendency to generate a whirling vibration of half-speed whirl. It can be said that 3.0 is good.

  This calculation result suggests more specifically. Considering the case where the rising part from the groove to the hill is directly below and the shaft is in contact at that point, the groove side, which is the first region, has a wedge-shaped sleeve inner wall surface and shaft outer wall surface However, since there is a groove, the gap is large and the squeeze effect of the lubricating fluid does not occur. Next, on the hill side, which is the second region, the gap formed by the hill surface and the outer surface of the shaft expands in the circumferential direction, so that negative pressure is generated by the squeeze effect, and conversely, the shaft is drawn toward the sleeve side. Therefore, in order to obtain levitation force, the outer surface of the shaft and the hill part which is the second region are passed from the groove part which is the first region through the place where the outer surface of the shaft is in contact with the hill part. This means that it is better to activate the function of injecting the lubricating fluid into the gap or prevent the generation of negative pressure in the second area, and conversely generate levitation force in this area .

  As the former method, there are a method of taking the edge of the hill, a method of tapering the bottom of the groove, and a method of eliminating three-dimensionally by appropriately selecting the inclination angle of the groove.

As the latter method, there is a method in which a hill portion is tapered in the circumferential direction and a gap shape that expands in the circumferential direction is not formed. That is, since the interval between the outer surface of the shaft and the hill portion at the corner portion that moves from the hill portion to the groove portion is Δ (1-cos θ _l ), if such a taper is given in the circumferential direction of the hill portion, the outer surface of the shaft Since the hill surface is concentric, a gap shape that expands in the circumferential direction is not formed. Actually, since the contact position between the outer surface of the shaft and the inner surface of the sleeve is changed by this taper application, the effect can be obtained even with a smaller taper amount. In addition, an excessively large taper impairs the characteristics of the original herringbone bearing. Accordingly, the following range can be selected as an appropriate taper range. That is, from the lower limit value that causes the above effect to be generated from at least half the width of the hill portion and the upper limit value that the influence on the bearing characteristics during high-speed rotation increases when one entire pitch of the hill groove is tapered, the ROT direction the ratio of the circumferential taper amount [delta] _l and bearing clearance delta of lands ^{_{δ * l (= δ l /}} Δ) towards,
1-cos (θ _l / 2) <δ ^* _l <1-cos (θ _l + θ _p )
It can be said that it is good.

FIG. 18 shows the result of calculating the relationship between the number of grooves and the dimensionless wear amount V ^* _W of the sleeve. Groove depth ratio h ^{* p} ₌ 2.0, since the groove angle is an analysis of beta = 70 ° just before the amount of wear increases rapidly, but not in a very distinctive differences, the amount of wear by the number of grooves N _g Fluctuates, and the wear amount is minimum when the number is eight. Considering characteristics during high-speed rotation and mass production processability, it can be said that the number of grooves N _g is preferably 6 ≦ N _g ≦ 10, preferably 7 ≦ N _g ≦ 9.

FIG. 19 shows the calculation result of the relationship between the taper amount of the sleeve inner diameter in the rotation axis direction and the dimensionless wear amount V ^* _W of the sleeve. The model of this analysis is different from the previous models. The shaft diameter is 2.5 mm, the sleeve axial width is as long as 4.8 mm, and there are herringbone type radial dynamic pressure bearings on the top and bottom. interval Δ is 0.00275mm the inner periphery and the outer periphery of the shaft, the distance _{h p} of the circumferential surface and the outer circumferential surface of the shaft 1 of the groove 10 shallow as 0.0015 mm, that is, the groove depth ratio is 0.55 It is small. The number of grooves Ng is 6 for both the upper and lower bearings. In this analysis, for the upper bearing on the output end side, the bearing clearance was increased by gradually expanding the sleeve inner diameter while keeping the groove depth constant from the herringbone groove turn-up to the upper opening. This is the result of calculating the relationship between the taper variation δ _{Δ of} the bearing gap and the dimensionless wear amount V ^* _W of the sleeve, which is particularly affected in the case of a structure constituted by two bearings described later in Example 3. Become prominent. Since the bearing is supported by other bearings in the stationary state, the shaft and sleeve are not in contact with each other on the opening side of the upper bearing, and locally between the shaft outer diameter surface and the hill surface. The gap can be held. By increasing the taper amount, the lubricating fluid flows into the gap portion and it becomes easy to exert a squeeze effect, and the amount of wear is reduced because it becomes easier to float. However, if the taper amount becomes too large, the bearing characteristics as a whole deteriorate, and the wear amount increases. Under the conditions of this analysis, it can be seen that if the taper is 0.0004 mm to 0.0005 mm, which is about 1/3 of the groove depth h _p = 0.0015 mm, the amount of wear of the sleeve is minimized. From this analysis result, the groove depth _{h p} ratio of the tapered variation [delta] _delta of bearing clearance delta against _{^{δ * Δ (= δ Δ /}} h p = (δ Δ / Δ) / (h p / Δ)) is 1 It can be seen that / 8 <δ ^* _Δ <1/2, preferably about 1/3. Since this is a phenomenon caused by the relative speed, it is clear that the same effect can be obtained even if the bearing clearance Δ is changed by changing the diameter on the shaft side.

The above analysis result diagram is the result of extracting and displaying the result of combining several parameters mainly based on the analysis reference parameters shown above, but changing the parameters within a practical range. However, there is no significant difference in trend except as noted. Therefore, from the above analysis results, it is possible to stably keep the friction loss and the wear amount at the start-up while maintaining the rotation rigidity such as the support rigidity K _xx at the high speed rotation, the whirl characteristic, or the bearing loss. the optimum Introduction specifications circumferential angle ratio theta ^* _p of the groove is 0.40 to 0.60, the inclination angle β is 45 ° to 75 ° of the groove, the groove number _{N g} is present 6 to 10 (more preferably 7 to 9), and the groove depth ratio h ^* _p is preferably 0.5 to 3.0. Further, as the circumferential taper ratio δ ^* _l of the hill portion, 1−cos (θ _l / 2) <δ ^* _l <1−cos (θ _l + θ _p ), axial variation ratio δ ^* _{Δ of the} bearing clearance 1/8 <δ ^* _Δ <1/2 is good.

(Second Embodiment)
FIG. 20 shows a second embodiment according to the present invention. Items that are not specially described are the same as in the first embodiment, and the display of parallel diagonal lines in the cross section is omitted.

  In the hydrodynamic bearing device of FIG. 23, a sleeve 2 formed by sintering and molding metal powder is used. On the inner peripheral surface 2 a of the sleeve 2, a hill portion 12 and a groove portion 10 that are inclined by β with respect to the rotation axis are formed. A resin-molded thrust bush 8 is fixed to the lower end of the sleeve 2 with an adhesive 9 to form a bottomed container, and the lubricating fluid 5 is held therein. A thrust bearing 13 is formed between the thrust plate 7 formed at the lower end of the shaft 1 manufactured by forging and grinding the SUS material, and the sleeve 2 and the thrust bush 8 to support the axial load. Yes.

  In this embodiment, when the powder metal sleeve 2 is pressure-molded, the hill portion 12 is raised on the inner diameter side of the sleeve 2 by compressing and pressing in the direction of the rotation axis, thereby forming a herringbone groove. Further, the top portion of the hill portion 12 formed into a gentle convex shape toward the inner diameter side is flattened when sizing the inner diameter of the sleeve 2 to form a hill portion having a taper in the circumferential direction. ing.

The shape of the hill portion is shown in FIG. 21 as a schematic diagram, and shown in FIG. 22 as a further enlarged schematic diagram. Even when the outer surface 1a of the shaft 1 is in contact with the inner wall surface 2a of the sleeve 2, the circumferential taper amount δ _l of the hill portion so that the lubricating fluid 5 can easily flow into the gap portion and exert a squeeze effect. and the ratio of the bearing gap ^{_{Δ δ * l (= δ l}} / Δ) _{is, 1-cos (θ l /} 2) <δ * l <1-cos (θ l + θ p) are set to satisfy the Yes.

  Further, the sleeve 2 is formed in a taper shape in which the inner diameter is gradually increased toward the upper side to which a load is applied when the rotor hub 3 is attached during the sizing process. Specifically, the radius R in the sleeve is expanded by about 1/3 of the bearing gap Δ. The shape of the sleeve is shown in FIG. 23 as an enlarged schematic diagram. As a result, the output end side has a shape in which the bearing gap is large and the groove depth is small, that is, the groove depth ratio is small. For this reason, at the time of start-up, the contact pressure between the shaft 1 and the sleeve wall surface of the output end side portion of the bearing is smaller than the opposite side of the output end, and the gap between the hill portion 2 and the shaft 1 is relatively large. Can easily flow in, and the rotating body can easily float.

  In addition, as shown in FIG. 24, you may satisfy the above relationships by changing the outer diameter of the shaft 1 in steps.

  In the present embodiment, the sleeve 2 is made of powder metal, but may be made by injection molding of a resin material.

(Third embodiment)
FIG. 25 shows a third embodiment according to the present invention. Items that are not specially described are the same as in the first embodiment, and the display of parallel diagonal lines in the cross section is omitted.

  In the hydrodynamic bearing device of FIG. 25, two sintered alloy sleeves 102a and 102b are attached to the inside of the housing 4 so that the lubrication holding portion 116 is sandwiched in the axial direction (vertical direction in FIG. 26). A pair of hydrodynamic bearings 106 and 108 are provided in the form. In the sleeves 102a and 102b, the groove portions 110a and 110b and the hill portions 112a and 112b are repeatedly arranged in the circumferential direction, and the inclination angle, the circumferential angle, and the depth of the groove portions 110a and 110b and the hill portions 112a and 112b are arranged. Etc. are respectively set to the proper specifications shown in the first embodiment. On the other hand, the outer shape of the shaft 101 is cylindrical, and is formed in substantially the same shape from one end to the other end.

  In the present embodiment, the sleeve 102b is a herringbone dynamic pressure bearing having a hill at the center of the two rows of dynamic pressure groove rows to increase radial rigidity. However, the specifications of the sleeves 102a and 102b may be selected so as to match the application and needs of the rotating body to be applied, such as making the sleeves 102a and 102b have the same structure and dimensions to facilitate mass production. Of course, it may be combined with other bearing types such as other arc bearings. In this embodiment, the sleeves 102a and 102b are separated from each other, but two dynamic pressure bearing grooves may be provided in one sleeve.

  In this structure, since the distance between the two dynamic pressure bearing portions provided in the sleeves 102a and 102b in the direction of the rotation axis is long, the moment rigidity is increased, and characteristics excellent in rotational stability and impact resistance are obtained. be able to.

(Fourth embodiment)
FIG. 26 is a cross-sectional view of the spindle motor 30 and the recording disk drive device 50 embodying the present invention. In addition, the display of parallel diagonal lines in the cross section is omitted.

  The recording disk drive device 50 includes a recording disk 51, a spindle motor 30 that rotates the recording disk 51, a head 52 that accesses information to the recording disk 51, and a housing 53 that accommodates the whole.

  The spindle motor 30 includes the hydrodynamic bearing device 306 described in the first to third embodiments. The spindle motor 30 uses a part of a housing 53 as a substrate, and a stator 31 and a circuit substrate (not shown) are fixed on the substrate. On the other hand, the rotor magnet 33 fixed to the rotor hub 303 is opposed to the stator 31 in the radial direction, and is supported rotatably by the hydrodynamic bearing device 306 with respect to the substrate member and the stator 31. The stator 31 includes a plurality of coils 32, and energization of the coils 32 is controlled by a control circuit.

  The recording disk 51 is placed on the rotor hub 303 of the spindle motor 30 and is rotated together with the rotor hub 303. When the coil 32 of the stator 31 is energized by the control circuit, the spindle motor 30 starts to rotate. The hydrodynamic bearing device 306 supports the rotating side and the stationary side in a non-contact manner, and the vibration of the spindle motor 30 is suppressed. Thereby, an error in writing to the recording disk 51 is suppressed, and an improvement in reliability and speeding up are achieved. In addition, the recording disk drive device 50 can be made quiet, and even if it is mounted on a portable device or an acoustic device, it is difficult to generate annoying noise.

  In addition, the hydrodynamic bearing device 306 of the present invention can increase the substantial bearing support interval or support the vicinity of the center of gravity of the rotating body, and the friction loss and the amount of wear at the time of starting and stopping are Small and stable high-speed rotation performance can be obtained, so it is effective not only for recording disk drive devices, but also for overhang structures, especially for mobile devices and fans with a high degree of freedom in the direction of use with a large number of start / stop operations. .

  As mentioned above, although embodiment of this invention was described, this invention is not limited to these embodiment, For example, the groove parts 10 and 110 and the hill parts 12 and 112 may be formed in the shafts 1 and 101 side. The shafts 1 and 101 may be fixed and the sleeves 2 and 201 may be rotated. Of the two upper and lower bearings, one bearing is on the sleeve 2 and 201 side, and the other bearing is on the shaft 1 and 101 side. In addition, various modifications can be made without departing from the gist of the present invention, such as a structure in which the hill portions 10 and 110 and the groove portions 12 and 112 are formed.

It is sectional drawing of a spindle motor provided with the hydrodynamic bearing device and hydrodynamic bearing device in a 1st embodiment. It is sectional drawing which expands and shows the hydrodynamic bearing apparatus in 1st Embodiment. It is a schematic diagram of the sleeve inner surface part of the hydrodynamic bearing device in the first embodiment. It is a partial expanded view of the sleeve inner surface part of the hydrodynamic bearing device in the first embodiment. It is a simple fragmentary sectional view of the axial right angle direction of the hydrodynamic bearing apparatus in 1st Embodiment. FIG. 6 is a partial development view of a sleeve inner surface portion of another hydrodynamic bearing device according to the first embodiment. FIG. 6 is a partial development view of a sleeve inner surface portion of another hydrodynamic bearing device according to the first embodiment. It is a partial expansion schematic diagram of the hill groove part by the cross section in the direction perpendicular to the axis of the hydrodynamic bearing device in the first embodiment. It is a partial expanded schematic diagram of the hill groove part by the axial direction orthogonal cross section of the other hydrodynamic bearing apparatus in 1st Embodiment. It is a partial expanded schematic diagram of the hill groove part by the axial direction orthogonal cross section of the other hydrodynamic bearing apparatus in 1st Embodiment. It is the longitudinal cross-sectional schematic diagram cut along the inclination of the groove | channel of the hydrodynamic bearing apparatus in 1st Embodiment. It is an analysis result figure which shows the relationship between the rotation speed in 1st Embodiment, and the power required for rotation. It is an analysis result figure which analyzed the relation between the groove width ratio and the amount of wear in a 1st embodiment on the conditions at the time of starting and stopping. It is an analysis result figure which shows the relationship between the groove | channel angle and wear amount in 1st Embodiment. It is an analysis result figure which shows the relationship between the groove angle in 1st Embodiment, and the rigidity at the time of high speed rotation. It is an analysis result figure which shows the relationship between the groove depth ratio and wear amount in 1st Embodiment. It is an analysis result figure which shows the relationship between the groove depth ratio in 1st Embodiment, and the axial loss at the time of high speed rotation. It is an analysis result figure which shows the relationship of the amount of wear with respect to the groove number in 1st Embodiment. It is an analysis result figure which shows the relationship of the amount of wear with respect to the taper amount of the rotation axis direction of the hill part in 2nd Embodiment. It is sectional drawing which expands and shows the hydrodynamic bearing apparatus in 2nd Embodiment. It is a partial expansion schematic diagram of the hill groove part by the cross section in the direction perpendicular to the axis of the hydrodynamic bearing device in the second embodiment. It is the further partial expansion schematic diagram of the ditch | groove part by the cross section in the direction orthogonal to the axis of the hydrodynamic bearing device in the second embodiment. It is the longitudinal cross-sectional schematic diagram cut along the inclination of the groove | channel of the hydrodynamic bearing apparatus in 2nd Embodiment. It is the longitudinal cross-sectional schematic diagram cut along the inclination of the groove | channel of the other hydrodynamic bearing apparatus in 2nd Embodiment. It is sectional drawing which expands and shows the hydrodynamic bearing apparatus in 3rd Embodiment. It is sectional drawing of the recording disk drive device in 4th Embodiment.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1,101 Shaft 1a Shaft outer surface 2,102,102a, 102b Sleeve 2a Sleeve inner peripheral surface 3,303 Rotor hub 4 Bearing housing 5 Lubricating fluid 6,306 Dynamic pressure fluid bearing device part 7 Thrust plate 8 Thrust bush 9 Adhesive 10 110a, 110b Groove portions 11a, 11b Groove wall surfaces 12, 112a, 112b Hill portions 13 Thrust dynamic pressure fluid bearing device portions 14, 114 Relative rotation direction of lubricating fluid 15 Circulating grooves 16, 116 Lubricating fluid holding portion 17 Interface 21 Disengagement Stop 30 Spindle motor 31 Stator 32 Coil 33 Rotor magnet 50 Recording disk drive 51 Magnetic disk 52 Head 53 Actuator

Claims (16)

A substantially cylindrical shaft, a substantially cylindrical sleeve through which the shaft is inserted, and a lubricating fluid filled in a gap between the shaft and the sleeve,
The shaft is held by the sleeve so as to be rotatable relative to the sleeve around a rotation axis;
One of the outer peripheral surface of the shaft and the inner peripheral surface of the sleeve facing each other in a hydrodynamic bearing device to which a load is connected with one end in the rotation axis direction of the shaft or the sleeve as an output end Is defined as G-plane and the other plane as F-plane,
In the rotational flow motion of the lubricating fluid caused by the relative rotation between the shaft and the sleeve, the relative rotational direction of the lubricating fluid when viewed with respect to the G plane is defined as the ROT direction,
The G surface is formed with a dynamic pressure groove array in which a hill portion and a groove portion having at least one end opened are alternately arranged in the circumferential direction in an inclined state with respect to the rotation axis,
Further, on the G surface, two rows of the dynamic pressure groove rows inclined substantially in the opposite direction with respect to the rotation axis are arranged, and the groove opening end is disposed on the side away from the G surface, so that the surface is perpendicular to the rotation axis. At least one or more herringbone-shaped hydrodynamic bearing grooves configured to face each other substantially symmetrically are formed in the rotational axis direction,
A groove width ratio θ ^* _p (= θ _p / (θ _l + θ _p )) indicating a ratio of the circumferential width angle (θ _p ) of the groove and the circumferential width angle (θ _l ) of the hill portion,
0.4 ≦ θ ^* _p ≦ 0.6
And
Angles β ₁ and β ₂ on the acute angle side formed by the pair of dynamic pressure groove rows and the rotation shaft are:
45 ° ≦ β ₁ ≦ 75 ° and 45 ° ≦ β ₂ ≦ 75 °
A hydrodynamic bearing device characterized by that. In the pair of dynamic pressure groove rows, acute angles β _1a , β _2a formed by the wall surface a forming the boundary portion from the hill portion to the groove portion in the ROT direction and the rotation axis, and the ROT Angles β _1b and β _2b on the acute angle side formed by the wall surface b forming the boundary part from the groove part to the hill part viewed in the direction and the rotation axis,
45 ° ≦ β _1b <β _1a ≦ 75 °, or 45 ° ≦ β _2b <β _2a ≦ 75 °
The hydrodynamic bearing device according to claim 1, wherein   3. The pair of dynamic pressure groove rows, wherein the groove portion is formed in a tapered shape such that a gap dimension with the F surface gradually decreases in the ROT direction. The hydrodynamic bearing device described in 1.   In the pair of dynamic pressure groove rows, at least the groove portion on the dynamic pressure groove row side disposed on the output end side has a gap dimension with the F surface toward the side where the pair of dynamic pressure groove rows face each other. The hydrodynamic bearing device according to any one of claims 1 to 3, wherein the hydrodynamic bearing device is formed in a tapered shape that gradually decreases.   In the pair of dynamic pressure groove rows, at least the groove width ratio on the dynamic pressure groove row side arranged on the output end side is gradually decreased toward the side where the pair of dynamic pressure groove rows face each other. The hydrodynamic bearing device according to any one of claims 1 to 4, wherein the hydrodynamic bearing device is formed. In the pair of dynamic pressure groove rows, at least the number N _{g of} groove portions constituting the dynamic pressure groove row on the dynamic pressure groove row side arranged on the output end side is:
6 ≦ N _g ≦ 10, preferably 7 ≦ N _g ≦ 9
The hydrodynamic bearing device according to any one of claims 1 to 5, wherein In the pair of dynamic pressure groove rows, at least a bearing gap Δ (= R−r) which is a difference between the sleeve inner radius r and the shaft outer radius R on the dynamic pressure groove row side disposed on the output end side. When the hill groove depth ratio is the ratio of the average depth h _p to the bottom of the groove from the unit _{^{h * p (= h p /}} Δ) is,
0.5 ≦ h ^* _p ≦ 3.0
The hydrodynamic bearing device according to claim 1, wherein the hydrodynamic bearing device is a hydrodynamic bearing device.   In the pair of dynamic pressure groove rows, at least the hill portion on the dynamic pressure groove row side arranged on the output end side is such that a gap dimension with the F surface gradually decreases in the ROT direction. 8. The hydrodynamic bearing device according to claim 1, wherein the hydrodynamic bearing device is tapered. In the pair of dynamic pressure groove rows, at least the ratio δ ^* _l of the taper amount δ _{l of the} hill portion in the ROT direction on the dynamic pressure groove row side arranged on the output end side and the bearing gap Δ ( = Δ _l / Δ) is
1-cos (θ _l / 2) <δ ^* _l <1-cos (θ _l + θ _p )
The hydrodynamic bearing device according to claim 8, wherein   In the pair of dynamic pressure groove rows, at least a corner of the wall surface b forming a boundary portion from the groove portion to the hill portion as viewed in the ROT direction of the dynamic pressure groove row disposed on the output end side. 10. The hydrodynamic bearing device according to claim 1, wherein the hydrodynamic bearing device is chamfered.   In the pair of dynamic pressure groove rows, at least the bearing gap Δ on the dynamic pressure groove row side arranged on the output end side is gradually reduced toward the side where the pair of dynamic pressure groove rows face each other. The hydrodynamic bearing device according to claim 1, wherein the hydrodynamic bearing device is formed. In the pair of dynamic pressure groove rows, at least the bearing gap Δ on the dynamic pressure groove row side arranged on the output end side is gradually reduced toward the side where the pair of dynamic pressure groove rows face each other. the ratio of the amount of change in the bearing gap formed [delta] _delta and the average depth _{^{h p δ * Δ (= δ}} Δ / h p) is,
1/8 <δ ^* _Δ <1/2, preferably δ ^* _Δ≈1 / 3
The hydrodynamic bearing device according to claim 11, wherein   The hydrodynamic bearing device according to any one of claims 1 to 12, wherein the sleeve is made of a sintered porous metal or resin mold.   The hydrodynamic bearing device according to claim 1, wherein the groove portion and the hill portion are formed in the sleeve. An electric spindle motor,
A rotor portion having field magnets arranged around the rotation axis;
A stator portion that has an armature that generates torque about the rotation axis with the field magnet, and supports the rotor portion rotatably about the rotation axis;
With
Either one of the rotor part and the stator part is provided with a sleeve constituting the hydrodynamic bearing device according to any one of claims 1 to 14,
A spindle motor, wherein the other of the rotor portion and the stator portion includes a shaft inserted into the sleeve.   A recording disk drive device comprising: a head for reading / writing data from / to a recording disk; an actuator for moving the head on the disk surface; and a spindle motor according to claim 15.

Images