JP2007177807A - Hydrodynamic bearing unit - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To sufficiently exhibit a hydrodynamic effect without being largely affected by the viscous friction of a lubricating oil due to a variation in use environment temperature and to secure the smooth startability of a rotating shaft, by specifying the clearance between the rotating shaft and a bearing in a proper area. <P>SOLUTION: The maximum clearance 42 formed between the flange 32 of the rotating shaft 30 and the upper end face of the bearing 20 facing the flange 32 and brought into partial contact with the flange 32 when the rotating shaft 30 is not rotated, is set at 2 to 6 μm. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  In the present invention, the lubricating oil supplied to the gap between the bearing and the rotating shaft generates a high dynamic pressure during rotation of the rotating shaft, and supports the load in both the radial and thrust directions of the rotating shaft in a non-contact state. The present invention relates to a hydrodynamic bearing unit. The hydrodynamic bearing unit of the present invention is suitable as a bearing unit of a spindle motor used as a drive motor of various information devices such as a drive source of a disk drive device that reads and writes magnetic disks and optical disks and a polygon motor of a laser printer. is there.

  In this kind of spindle motor for information equipment, high-speed and high-precision rotation performance is required in order to achieve high recording density and high data transfer speed. In recent years, non-contact type hydrodynamic bearings have been widely adopted as bearings for supporting a rotating shaft. In a hydrodynamic bearing, lubricating oil is supplied to a minute gap between a rotating shaft and a bearing to form an oil film, and the oil film is increased in pressure by rotating the rotating shaft to support the rotating shaft with high rigidity. The dynamic pressure is effectively obtained by forming a dynamic pressure generating groove (dynamic pressure groove) on one of the sliding surfaces of the rotating shaft and the bearing.

  The spindle motor bearing is configured to support a thrust load in addition to a normal radial load support mainly for the purpose of preventing the shaft from rotating. The thrust load is configured such that a flange provided integrally with the rotating shaft is opposed to the bearing end surface so as to be slidable, and the thrust load of the rotating shaft is received by the end surface of the bearing through the flange. The dynamic pressure is generated on both the thrust side and the radial side, that is, a thrust dynamic pressure groove is formed on one of the flange of the rotating shaft and the bearing end surface facing each other, and the inner circumferential surface of the bearing A radial dynamic pressure groove is formed on one of the outer peripheral surfaces of the rotating shaft facing the inner peripheral surface. These dynamic pressure grooves are devised in shape and depth so that the oil film becomes high pressure with the rotation of the rotating shaft, for example, a wedge shape whose width and depth become narrower toward the rotating direction of the rotating shaft. It is common.

  By the way, it is desirable that the gap between the rotating shaft on which the oil film is formed and the bearing is as small as possible from the viewpoint that the hydraulic pressure tends to be high and the bearing rigidity is improved. Control affects bearing performance. For example, according to Patent Document 1, the perpendicularity and flatness of the bearing end surface are defined to be equal to or less than a predetermined value, and thereby the bearing clearances on the radial side and the thrust side are optimized.

JP 2005-188753 A

  When the bearing gap increases, the oil film pressure decreases, the dynamic pressure does not increase sufficiently, and the bearing rigidity tends to decrease. The high temperature (for example, around 60 ° C.) accompanying the continuous operation of the spindle motor causes the viscosity of the lubricating oil to increase. The decrease in the bearing rigidity becomes remarkable by promoting the decrease. Therefore, it is effective to design the bearing gap to be as small as possible as described in the above document, but on the other hand, if the bearing gap is too small, the viscous friction of the lubricating oil increases, so that the motor It is difficult to start up, that is, it is difficult to start rotation from a stationary state. In particular, when the rotating shaft is stopped, such a starting failure is likely to occur on the thrust side where the flange tends to come into close contact with the bearing end surface by gravity, and when the operating environment temperature is low, for example, 0 ° C. or less, the viscosity of the lubricating oil There is also concern that it may become unable to start due to increased friction. In addition, since a spindle motor of a portable device of a type that uses a battery power source such as a notebook PC (personal computer) has a small starting torque, there is a limit to miniaturization of the bearing gap from the viewpoint of starting performance.

  Therefore, in the present invention, by defining the clearance between the rotating shaft and the bearing within an appropriate range, the dynamic pressure effect is sufficiently exerted without being greatly affected by the viscous friction of the lubricating oil accompanying the fluctuation of the operating environment temperature. In addition, an object of the present invention is to provide a hydrodynamic bearing unit that can ensure smooth startability of a rotating shaft.

The inventor has intensively studied the form and amount of the bearing clearance on the side that supports the thrust load of the hydrodynamic bearing unit used in the magnetic recording disk device or the like. The thrust side bearing clearance here is a minute gap between the flange of the rotating shaft and the bearing end face facing this flange, and the rotating shaft is not rotating and the flange partially contacts the bearing end face. It is a minute gap generated between the opposing surfaces in contact with each other. The gap is a perpendicularity of one or both of the opposing surfaces (degree of inclination of the surface that should be perpendicular to the axial direction, zero and perfect right angle) or flatness (flatness caused by undulation or surface roughness). When the squareness is zero and the flatness is zero, there is no gap.

  According to the research of the present inventor, when the thrust side bearing gap is configured with an accuracy of 2 to 6 μm at the maximum, a hydrodynamic bearing unit satisfying both sufficient hydrodynamic effect and startability can be obtained. I found out. The maximum in this case is the maximum gap generated between the two rotating shafts with respect to the bearing, and a gap larger than this does not occur unless they are separated.

  The reason for this numerical range is that when the gap is less than 2 μm, the viscous friction generated by the lubricating oil present in the gap is remarkably increased, and it is necessary to increase the starting torque for rotating the rotating shaft. This means that the startability is deteriorated, and particularly when the use environment temperature is 0 ° C. or less. If the startability is low, it is not suitable for a spindle motor for a portable device such as a notebook PC having a small start torque, and may not be operable in a cold use environment. On the other hand, if the bearing clearance on the thrust side exceeds 6 μm, the oil film pressure becomes too low to obtain a sufficient dynamic pressure effect, and the flange remains in contact with the bearing end surface, so that the dynamic pressure bearing itself functions. It becomes difficult to be demonstrated.

  The hydrodynamic bearing unit of the present invention is made on the basis of the above knowledge, and a rotating shaft in which a flange is provided on a cylindrical shaft main body, and the shaft main body of the rotating shaft are rotatably inserted. A cylindrical bearing having an end face opposed to the flange in the state, and a thrust dynamic pressure groove is formed on one of the opposing faces of the flange and the bearing end face, and the inner peripheral face of the bearing, and the inner peripheral face A radial dynamic pressure groove is formed on one of the outer peripheral surfaces of the shaft body opposite to the shaft, and lubricating oil is supplied to a minute gap formed between the rotating shaft and the bearing. In a dynamic pressure bearing unit in which dynamic pressure is generated in the lubricating oil supplied to the groove and the radial dynamic pressure groove, and the rotary shaft is supported by the bearing in a non-contact state in the thrust direction and the radial direction, the flange and the bearing end surface Max formed between Gap, is characterized in that it is 2-6 [mu] m.

  In the present invention, the bearing material is preferably a sintered metal obtained by sintering a molded body formed by compression molding a raw metal powder because the thrust dynamic pressure groove and the radial dynamic pressure groove are easily formed. These dynamic pressure grooves can be easily formed by plastic working such as coining.

  According to the present invention, the thrust side bearing clearance between the flange of the rotating shaft and the bearing end surface that supports the rotating shaft and faces the flange is defined as an optimum condition. In this case, it is possible to provide a hydrodynamic bearing unit that exhibits both sufficient hydrodynamic effect and startability.

Hereinafter, an embodiment of the present invention will be described with reference to the drawings.
FIG. 1 shows a longitudinal section of a fluid dynamic bearing unit 1 according to an embodiment, FIG. 2 is a plan view of the bearing, and FIG. 3 is a transverse sectional view of the bearing. This dynamic pressure bearing unit 1 is a bearing unit suitable for a spindle motor of a magnetic recording disk drive device. A cylindrical bearing indicated by reference numeral 20 has a small outer diameter of about 6 mm and is inserted into a shaft hole 21. A rotating shaft 2 having a diameter of about 3 mm is supported rotatably.

  As shown in FIG. 1, the hydrodynamic bearing unit 1 includes a bottomed cylindrical housing 10 that opens upward in the drawing, a bearing 20 accommodated in the housing 10, and a rotary shaft 30 that is rotatably supported by the bearing 20. It consists of. The housing 10 includes a cylindrical housing main body 11 and a disk-shaped bottom plate 12 that is fixed to the inner peripheral edge of the opening on the lower side of the housing main body 11 by welding means such as an electron beam or a laser and hermetically closes the opening. Has been. The bearing 20 is fixed in the housing 10 by means of press fitting or fitting into the housing main body 11 by means such as welding or adhesion.

  The rotary shaft 30 is formed by integrally fixing an annular flange 32 to a shaft main body 31, and is formed from a melted material such as brass or stainless steel. The flange 32 is formed on the outer peripheral surface of the shaft main body 31 so that the axial direction is positioned and the surface direction is perpendicular to the axial direction of the shaft main body 31, that is, the perpendicularity is zero. Locked to the step 31a. The shaft body 31 is divided into a large-diameter portion 31b and a small-diameter portion 31c with the stepped portion 31a as a boundary. The large-diameter portion 31b is inserted into the shaft hole 21 of the bearing 20 from above in the figure, and the flange 32 is located above the bearing 20. It is made to oppose the end surface 22. The radial load of the rotating shaft 30 is supported by the inner peripheral surface 23 of the bearing 20, and the thrust load of the rotating shaft 30 is supported by the upper end surface 22 of the bearing 20.

  Between the inner peripheral surface 23 of the bearing 20 and the outer peripheral surface of the shaft body 31, and between the upper end surface 22 of the bearing 20 and the lower surface 32a of the flange 32, a radial-side bearing gap 41 and a thrust-side bearing gap 42, respectively. The lubricating oil is supplied to the gaps 41 and 42. The radial-side bearing gap 41 is set to about 1 to 3 μm, and the thrust-side bearing gap 42 is set to 2 to 6 μm as will be described in detail later.

  A cover member 13 made of an annular plate material is fixed to the opening end of the housing 10. The cover member 13 suppresses the scattering of the lubricating oil, and the flange 32 of the rotary shaft 30 that floats due to the thrust dynamic pressure contacts the cover member 13 to prevent the rotary shaft 30 from coming off.

  As shown in FIG. 2, the upper end surface 22 of the bearing 20 has a plurality of spiral grooves 24 that are curved toward the inner peripheral side toward the rotational direction R of the rotary shaft 30 and are equally spaced in the circumferential direction. It is formed (in FIG. 2, a hatched line is drawn to distinguish it from the upper end surface 22). The outer peripheral ends of these spiral grooves 24 are open to the outer peripheral surface, but the inner peripheral end is not opened to the inner peripheral surface 23 and is closed. About 10 spiral grooves 24 are formed (12 in FIG. 2), and the maximum depth is about 10 μm.

  Further, as shown in FIG. 3, the inner peripheral surface 23 of the shaft hole 21 of the bearing 20 has a plurality of separation grooves 25 having a semicircular cross section and extending straight along the axial direction between both end surfaces. It is formed at equal intervals in the direction. And between each separation groove 25 of the inner peripheral surface 23, it decenters with respect to the shaft center P of the outer diameter of the bearing 20, and it shrinks | reduces to an inner peripheral side as it goes to the rotation direction of the rotating shaft 30 shown by arrow R. An eccentric groove 26 having a cross-sectional shape that increases in diameter is formed. In this case, in the illustrated example, five separation grooves 25 and five eccentric grooves 26 are formed, but 3 to 6 is a suitable number.

  The minute gap between the inner surface of the eccentric groove 26 and the outer peripheral surface of the rotating shaft 30 has a wedge shape in a cross section that gradually narrows toward the rotating direction of the rotating shaft 30. The width of the separation groove 25 is a length corresponding to 8 to 20 degrees in the circumferential angle θ around the axis P of the dynamic pressure bearing 20 shown in FIG. It is about 10 mm.

  The bearing 20 of the present embodiment is a sintered bearing obtained by sintering a molded body obtained by compression molding a raw metal powder. By being a sintered bearing, the spiral groove 24, the separation groove 25, and the eccentric groove 26 can be easily formed by plastic working. For example, when the inner peripheral surface 23 of the bearing 20 is formed, the male pin that can form the separation groove 25 and the eccentric groove 26 is press-fitted into the shaft hole of the sintered bearing material. it can. Further, the spiral groove 24 can be formed by stamping a punch having a plurality of convex portions that can form the spiral groove 24 on the end face of the material of the sintered bearing 20. Since the sintered bearing is porous, the spring back amount is small, and the spiral groove 24, the separation groove 25, and the eccentric groove 26 can be formed with high dimensional accuracy by plastic working.

  As the raw material powder of the sintered bearing, for example, iron powder and copper powder are almost the same amount, such as alloy powder of iron: 40-60 wt%, copper: 40-60 wt%, tin: 1-5 wt%, In addition, those containing several wt% tin powder are preferably used. According to such a composition, in addition to the properties of a sintered material whose main component is copper with good workability, the strength is improved by containing a large amount of iron, and further, by containing tin, The conformability to the rotating shaft 30 and the plastic workability are further improved. For this reason, it becomes easy to form the spiral groove 24, the separation groove 25, and the eccentric groove 26 by plastic working as described above, and the friction coefficient is reduced and the wear resistance is improved.

  The bearing 20 is impregnated with lubricating oil to form an oil-impregnated bearing. When the rotary shaft 30 inserted into the shaft hole 21 rotates in the direction of the arrow R shown in FIGS. 2 and 3, the lubricating oil that leaks into and stores in each separation groove 25 of the inner peripheral surface 23 rotates efficiently. It is wound around the shaft 30 and enters the wedge-shaped minute gap between the eccentric groove 26 and the rotating shaft 30 to form an oil film. The lubricating oil entering the minute gap flows toward the narrow side of the minute gap, thereby generating a wedge effect and a high pressure, and a high radial dynamic pressure is generated. The portion where the oil film is increased in pressure is generated at equal intervals in the circumferential direction according to the eccentric groove 26, whereby the radial load of the rotating shaft 30 is supported in a balanced manner with high rigidity.

  On the other hand, the lubricating oil oozes out and is stored in the spiral groove 24 formed in the upper end surface 22 of the bearing 20, and a part of this lubricating oil comes out of the spiral groove 24 by the rotation of the rotating shaft 30, An oil film is formed between the upper end surface 22 and the flange 32. Further, the lubricating oil retained in the spiral groove 24 flows from the outer peripheral side to the inner peripheral side in the spiral groove 24, and a thrust dynamic pressure that generates the highest pressure is generated at the end on the inner peripheral side. Then, when the thrust dynamic pressure is received by the flange 32, the rotating shaft 30 is slightly lifted as shown in FIG. 4, whereby the thrust load is supported in a balanced manner and with high rigidity.

  When the magnetic recording disk is rotationally driven, the flying height of the rotary shaft 30 is required to be 6 μm or more at a temperature of about 60 ° C. The radial bearing clearance 41 is about 1 to 3 μm, but this value is for suppressing the shaft runout of the rotating shaft 30 to μm or sub μm or less. In particular, in order to guarantee a high recording density on the magnetic recording disk, it is indispensable to suppress the NRRO value (non-repetitive vibration value) to 0.05 μm or less. Bearing rigidity in both radial and thrust is essential.

  Now, in the hydrodynamic bearing unit 1 of the present embodiment, when the rotary shaft 30 is not rotating, the thrust 32 does not occur and the flange 32 hits the upper end surface 22 of the bearing 20 as shown in FIG. The rotating shaft 30 is lowered to the point. At this time, there is a thrust-side bearing gap 42 between the opposing faces of the two (the lower face 32a of the flange 32 and the upper end face 22 of the bearing 20. The bearing gap 42 is either one of the opposing faces or Due to both squareness (degree of inclination of the surface that should be perpendicular to the axial direction, zero and perfect right angle) and flatness (degree of flatness caused by waviness and surface roughness, zero and perfect plane) It occurs as a result.

  FIG. 5 shows a case where the squareness is not zero because the flange 32 is inclined with respect to the axial direction of the shaft main body 31, and the outer peripheral edge of the lower surface of the flange 32 is partially on the upper end surface 22 of the bearing 20. A gap 42 is formed by contact. 6 and 7 show a case where the perpendicularity is not zero because the upper end surface 22 of the bearing 20 is inclined with respect to the axial direction of the bearing 20, and in FIG. 6, the upper end surface 22 is in a constant direction. In FIG. 7, the upper end surface 22 is a downward slope as it goes to the outer periphery. 6 and 7, the lower surface 32 a of the flange 32 partially contacts the upper end surface 22 of the bearing 20, so that the gap 42 is formed. The gap 42 here is only generated between the lower surface 32 a of the flange 32 and the upper end surface 22 of the bearing 20, and is different from the distance from the bottom surface of the spiral groove 24 to the lower surface 32 a of the flange 32.

  The thrust side bearing gap 42 may be constant or fluctuate even during one rotation of the rotary shaft 30, but in this embodiment, the rotary shaft 30 makes one rotation anyway. Design and processing are performed so that the generated gap 42 is configured with an accuracy of 2 to 6 μm at the maximum. The amount of the gap 42 can be set by adjusting the perpendicularity of the flange 32 or the upper end surface 22 of the bearing 20 or by forming a swell on at least one of the opposing surfaces. The perpendicularity can be controlled to about 1 μm by machining such as cutting. In addition, it is advantageous in manufacturing to process the upper end face 22 of the sintered metal bearing 20 that can be easily formed into a desired size and shape by plastic working.

  By setting the thrust-side bearing gap 42 in the range of 2 to 6 μm in this way, sufficient dynamic pressure effect and startability are exhibited even under conditions of high temperature around 60 ° C. or low temperature below 0 ° C. Is done. The reason for this is that when the thrust side gap 42 is less than 2 μm, the viscous friction generated by the lubricating oil present in the gap 42 increases remarkably, and there is a possibility that it cannot be operated in a cold use environment. On the other hand, if the gap 42 on the thrust side exceeds 6 μm, the oil film pressure becomes too low to obtain a sufficient dynamic pressure effect, and the flange 32 remains in partial contact with the upper end surface 22 of the bearing 20. This is because the function of the hydrodynamic bearing 20 itself is hardly exhibited. By setting the thrust-side bearing gap 42 in the range of 2 to 6 μm as in this embodiment, the occurrence of these problems can be prevented in advance, and the dynamic pressure bearing can achieve both sufficient dynamic pressure effect and startability. You can get units.

  Although the bearing 20 has been described as a sintered bearing, it may be formed of a melted material such as brass or stainless steel. In that case, the spiral groove 24 is formed by means such as chemical etching or electric discharge machining.

Next, examples of the present invention will be described to clarify the effects of the present invention.
The raw material powder having the composition shown in Table 1 is compression-molded, and the molded body is sintered to obtain a cylindrical sintered body having a true density ratio of 6.3 to 7.2%, an outer diameter of 6 mm, an inner diameter of 3 mm, and an axial length of 5 mm. Obtained the required number of materials for the bearing.

  Subsequently, spiral sintered material, separation groove and eccentric groove similar to those shown in FIGS. 2 and 3 are formed on the sintered bearing material by plastic working to produce sintered bearings, and these sintered bearings are made of steel housings. Then, the sintered bearing was impregnated with ester oil as a lubricating oil. Next, a plurality of rotating shafts adjusted in the perpendicularity of the flange are respectively inserted into the sintered bearings, and have the same configuration as that shown in FIG. 1 and are generated on the thrust side between the flange and the upper end surface of the bearing. A plurality of hydrodynamic bearing units having different bearing clearances were obtained.

  The rotary shaft of these dynamic pressure bearing units was rotated at 4200 rpm, and the flying height of the rotary shaft at that time was measured. FIG. 8 is a graph showing the measurement results. According to this graph, it can be seen that when the thrust-side bearing clearance exceeds 6 μm, the flying height of the rotating shaft decreases rapidly. When rotating the magnetic recording disk, the flying height of the rotating shaft is required to be 6 μm or more when the temperature is around 60 ° C. Therefore, the bearing clearance needs to be 6 μm or less in order to satisfy this condition reliably. It turns out that it is.

  Next, FIG. 9 is a graph showing the relationship between the thrust-side bearing clearance and the starting torque. This graph is a trial calculation of viscous friction generated in the thrust-side bearing clearance with the squareness of the flange as the apparent average clearance. 3 shows the friction characteristics when the starting torque for rotating the rotating shaft is 1 when the squareness is 3 μm. As is apparent from FIG. 9, the bearing gap and the starting torque are in an inversely proportional relationship, and when the gap is 1 μm or less, the starting torque increases rapidly. From this, it can be seen that the bearing gap needs to be 2 μm or more.

It is a longitudinal section of a fluid dynamic bearing unit concerning one embodiment of the present invention. It is a top view of the bearing which is a component of the fluid dynamic bearing unit shown in FIG. It is a cross-sectional view of the bearing. It is a longitudinal cross-sectional view of the dynamic pressure bearing unit in a state where the rotating shaft is floating. It is a longitudinal cross-sectional view which shows the case where the bearing clearance by the side of a thrust arises with the inclination of the flange of a rotating shaft. It is a longitudinal cross-sectional view which shows the case where the bearing gap by the side of a thrust arises with the inclination of the upper end surface of a bearing. It is a longitudinal cross-sectional view which shows the case where the bearing clearance by the side of a thrust arises with the inclination of the other form of the upper end surface of a bearing. It is a diagram which shows the relationship between the flying height of the rotating shaft measured in the Example, and a bearing clearance. It is a diagram which shows the relationship between the starting torque of the rotating shaft measured in the Example, and a bearing clearance.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 1 ... Dynamic pressure bearing unit, 20 ... Bearing, 22 ... Upper end surface (bearing end surface, opposing surface) of bearing, 24 ... Spiral groove (thrust dynamic pressure groove), 26 ... Eccentric groove (radial dynamic pressure groove), 30 ... Rotation Shaft, 31 ... shaft body, 32 ... flange, 32a ... lower surface (opposing surface) of the flange, 42 ... bearing clearance on the thrust side.

Claims (3)

A rotating shaft having a cylindrical shaft body provided with a flange;
The shaft body of the rotating shaft is rotatably inserted, and in this state, a cylindrical bearing having an end surface facing the flange,
A thrust dynamic pressure groove is formed on either one of the opposing surfaces of the flange and the bearing end surface,
A radial dynamic pressure groove is formed on either the inner peripheral surface of the bearing or the outer peripheral surface of the shaft main body facing the inner peripheral surface,
Lubricating oil is supplied to a gap formed between the rotating shaft and the bearing,
When the rotary shaft rotates, dynamic pressure is generated in the lubricating oil supplied to the thrust dynamic pressure groove and the radial dynamic pressure groove, and the rotary shaft is supported in a non-contact state in the thrust direction and the radial direction by the bearing. In the hydrodynamic bearing unit
The hydrodynamic bearing unit, wherein a maximum gap formed between the flange and the bearing end surface is 2 to 6 μm.   The hydrodynamic bearing unit according to claim 1, wherein the bearing is made of sintered metal.   The hydrodynamic bearing unit according to claim 1 or 2, wherein the thrust dynamic pressure groove and the radial dynamic pressure groove are formed by plastic working.

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