Sintered oil-retaining bearing, bearing unit, and motor

Abstract

Provided is a sintered oil-retaining bearing in which five or more axial grooves are formed at predetermined intervals in the circumferential direction on the inner diameter surface of the bearing, the sintered oil-retaining bearing being formed by the grooves and mounds formed between the grooves adjacent in the circumferential direction. The surface aperture ratio of the groove portion is set to be larger than the surface aperture ratio of the mound portion. The hill-groove ratio is set to 1 or less, and the circumferential length of the hill part is set to 0.1 mm or more. The depth of the groove portion is set to be less than eight times the gap size between the inner diameter surface and the outer diameter surface of the shaft member. The interval angle of the circumferential interval between the mountains adjacent in the circumferential direction is 72 DEG or less.

Description

Technical Field

[0001] The present invention relates to a sintered oil-impregnated bearing, a bearing unit, and a motor. Background Art

[0002] Sintered oil-impregnated bearings are commonly used for rotating small motors such as spindle motors and fan motors used in hard disk drives (HDDs) for polygon mirrors in laser beam printers (LBPs).

[0003] This type of oil-containing sintered bearing can use, for example, a fluid dynamic pressure bearing device, which has herringbone or spiral dynamic pressure grooves on the bearing surface. The dynamic pressure grooves generated as the shaft rotates generate a dynamic pressure oil film in the bearing gap to support the shaft.

[0004] Among conventional oil-impregnated sintered bearings, one described in Patent Document 1 includes a plurality of rectangular steps, each with a concentric sliding surface with the rotating shaft, arranged along the circumference of the bearing hole inner diameter surface facing the rotating shaft (shaft member). These steps form a rectangular gap surrounded by adjacent steps, the bearing hole inner diameter surface between the steps, and the rotating shaft. This gap generates oil film pressure.

[0005] Patent Document 1 states that by configuring as described above, "a sufficiently large oil film pressure can be obtained, thereby being able to support a large bearing load. Furthermore, a certain oil intake and discharge can be ensured, and the occurrence of seizure can be prevented."

[0006] In Patent Document 2, a plurality of axially extending grooves are formed on the inner surface of a cylindrical bearing. The grooves include downward grooves connected only to one (lower) end surface and upward grooves connected only to the other (upper) end surface.

[0007] Patent Document 2 states that by configuring as described above, “even if the motor receives an impact, the characteristics of the motor can be maintained to the greatest extent possible.”

[0008] Prior art literature

[0009] Patent Literature

[0010] Patent Document 1: Japanese Patent Application Laid-Open No. 5-115146

[0011] Patent Document 2: Japanese Patent Application Laid-Open No. 10-68418 Summary of the Invention

[0012] Problems to be solved by the invention

[0013] However, in the cited document 1, noise reduction is listed as a subject, butFigure 6 In the examples shown in [1], [2], [3], [4], the radial clearance (the gap between the groove and the shaft member) is set to 5 μm, and the eccentricity (the value obtained by dividing the distance from the center of the bearing to the center of the rotating shaft by the radial clearance) is set to 0.9. This results in an extremely narrow gap between the shaft member and the inner diameter surface of the bearing. Consequently, contact between the shaft member and the bearing is unavoidable in terms of the roundness of the shaft member outer diameter and the bearing inner diameter (roundness that takes into account productivity and accuracy), making it difficult to function as a bearing.

[0014] Patent Document 2 does not consider use in a high-temperature environment. That is, when used in a high-temperature environment, oil leakage may occur due to expansion of the lubricating oil.

[0015] Specifically, in bearing units equipped with such oil-impregnated sintered bearings, the temperature rise within the bearing gap (the gap between the rotating shaft and the bearing) caused by the rotation of the shaft member (rotating shaft) and the surrounding environment (high temperature) can cause the lubricating oil to expand, causing the oil level to rise and leaking out of the bearing gap. If the lubricating oil leaks out of the bearing gap, it cannot stably lubricate the bearing and cannot suppress whirling vibration of the shaft. It should be noted that the lubricating oil is initially injected (during assembly) at room temperature (approximately 23°C) to fill the internal volume of the bearing. Therefore, if the oil level rises due to temperature increases, as described above, the lubricating oil can leak out of the bearing gap.

[0016] Therefore, in view of the above-mentioned problems, the present invention provides a sintered oil-impregnated bearing that can suppress the whirling vibration of the shaft and ensure quietness, and provides a sintered oil-impregnated bearing, a bearing unit, and a motor that can return the overflowed oil to the bearing gap even if the oil leaks out of the bearing gap due to oil expansion.

[0017] Means for solving problems

[0018] A sintered oil-impregnated bearing according to the present invention includes a bearing inner diameter surface opposing the outer diameter surface of a shaft member, and five or more axial grooves are formed on the bearing inner diameter surface at predetermined intervals along the circumferential direction. The sintered oil-impregnated bearing is formed by the grooves and hillocks formed between circumferentially adjacent grooves. The surface opening ratio of the grooves is greater than the surface opening ratio of the hillocks, and the hillock-to-groove ratio, which is the ratio of the hillocks to the grooves, is set to 1 or less. The circumferential length of the hillocks is set to 0.1 mm or greater, and the depth of the grooves is set to 8 times or less the gap dimension between the inner diameter surface and the outer diameter surface of the shaft member. Furthermore, the circumferential spacing angle between circumferentially adjacent hillocks is set to 72° or less. Here, when the circumferential length of the hillocks is H2 and the circumferential length of the grooves is H1, the hillock-to-groove ratio is H2 / H1.

[0019] The oil-impregnated sintered bearing of the present invention can reduce the eccentric angle by setting the dimensions and shape as described above. Specifically, by providing axial grooves, the eccentric angle is reduced compared to a case without grooves (a true circle), and the eccentric angle further decreases when the hill-to-groove ratio is less than 1. Reducing the hill-to-groove ratio can reduce the eccentric angle. However, if the eccentric angle is too small, the hill width becomes narrow, which may reduce the load capacity and also reduce workability and wear resistance. Therefore, it is preferable to set the hill-to-groove ratio so that the circumferential length of the hill is greater than 0.1 mm. Furthermore, if the groove depth is large (deep), it is difficult to adjust the eccentric angle. Therefore, it is preferable to set the groove depth to no more than 8 times the gap dimension between the inner diameter surface of the bearing and the outer diameter surface of the shaft member. Furthermore, in bearings where the circumferential spacing between adjacent hills is widely spaced, if a load is applied to the shaft member, the bearing may be unable to support the load. However, if the interval angle of the circumferential interval between adjacent hill portions is 72° or less, the load can be sufficiently supported.

[0020] Preferably, an oil buffer portion is provided on at least one side of the bearing end face for preventing leakage of the lubricating oil and for circulating the lubricating oil. The oil buffer portion is composed of a groove connected to the bearing gap between the inner diameter surface of the bearing and the outer diameter surface of the shaft member, and an axial groove is provided on the outer diameter surface of the bearing to connect the two end surfaces of the bearing, so that the volume of the groove is larger than the volume of the oil in the groove that expands as the temperature rises.

[0021] By providing an oil buffer portion consisting of a groove connected to the bearing gap between the inner diameter surface of the bearing and the outer diameter surface of the shaft member, oil leakage (scattering) from the bearing can be effectively prevented, and the volume of the groove is larger than the volume of the oil in the groove that expands as the temperature rises, so that the oil that enters the groove when the temperature rises can be effectively prevented from leaking from the groove.

[0022] In this case, it is preferable that the axial groove portion of the bearing inner diameter surface, the concave groove of the bearing end surface, and the axial groove of the bearing outer diameter surface are connected. If so connected, lubricating oil flowing into the concave groove can return to the bearing interior through the axial groove.

[0023] Preferably, the shaft end portion of the bearing inner diameter surface has a chamfered portion, with the chamfer angle of the chamfer being set to 60° or greater. If the chamfer angle is less than 60°, the capillary force in the space (tapered space) formed by the shaft outer diameter surface and the chamfered portion makes it difficult for oil to flow toward the bearing outer diameter surface. Conversely, if the chamfer angle is greater than 60°, the capillary force is less effective, and oil is more likely to flow toward the bearing outer diameter surface.

[0024] A bearing unit of the present invention comprises the sintered oil-impregnated bearing, a shaft member inserted into the axial center hole of the sintered oil-impregnated bearing, and a housing accommodating the sintered oil-impregnated bearing on the inner diameter surface thereof, wherein the bearing unit is assembled in such a manner that the groove on the end surface of the bearing faces the side open to the atmosphere, thereby forming the following circulation path: due to a temperature rise, a flow of lubricating oil toward the side open to the atmosphere is generated in the bearing gap (radial gap) between the outer diameter surface of the shaft member and the opposing inner diameter surface of the bearing, the lubricating oil overflowing from the bearing gap enters the groove on the end surface of the bearing, and the lubricating oil returns to the gap from the groove via the axial groove on the outer diameter surface of the bearing.

[0025] A bearing unit according to the present invention has a circulation path where lubricating oil that overflows from the bearing gap (radial gap) due to rising temperature enters a groove on the bearing end surface. From there, the lubricating oil returns to the bearing gap via an axial groove on the bearing outer diameter surface. Consequently, the bearing unit according to the present invention forms an oil film immediately after operation begins, preventing the oil film from breaking down in the bearing gap during operation even at high temperatures. This reduces bearing surface wear and extends the bearing life.

[0026] Another bearing unit of the present invention comprises a sintered oil-impregnated bearing having no axial grooves formed on the inner diameter surface of the bearing but having other dynamic pressure generating grooves formed thereon, a shaft member inserted into the axial center hole of the sintered oil-impregnated bearing, and a housing for accommodating the sintered oil-impregnated bearing on the inner diameter surface, wherein an oil buffer portion is provided on at least one side of the end surface of the bearing to prevent leakage of the lubricating oil and to circulate the lubricating oil, the oil buffer portion being composed of a groove communicating with the bearing gap between the inner diameter surface of the bearing and the outer diameter surface of the shaft member, and a groove is provided on the outer diameter surface of the bearing to connect the two ends of the bearing The axial grooves connected to each other on the surface make the volume of the groove larger than the volume of the oil in the groove that expands as the temperature rises. The other bearing unit is assembled in a manner that the groove of the bearing end face is facing the atmosphere-opening side, and forms the following circulation path: due to the temperature rise, the bearing gap between the outer diameter surface of the shaft component and the opposite bearing inner diameter surface generates a flow of lubricating oil toward the atmosphere-opening side, the lubricating oil overflowing from the bearing gap enters the groove of the bearing end face, and the lubricating oil returns to the bearing gap from the groove via the axial groove on the bearing outer diameter surface.

[0027] Another bearing unit according to the present invention also features a circulation path. Lubricating oil that overflows from the bearing gap due to temperature rise enters a groove in the bearing end surface, and the lubricating oil then returns to the bearing gap via an axial groove in the bearing outer diameter surface. Consequently, the bearing unit according to the present invention forms an oil film immediately after operation begins, preventing the oil film from breaking down in the bearing gap during operation even at high temperatures. This reduces bearing surface wear and extends the bearing life.

[0028] The motor may include the above-mentioned sintered oil-impregnated bearing or may include the above-mentioned bearing unit.

[0029] Effects of the Invention

[0030] The present invention reduces the eccentric angle of the shaft during operation, resulting in a bearing that suppresses shaft vibration and achieves excellent quietness. Furthermore, an oil film forms immediately after operation begins, preventing the oil film from breaking up within the bearing gap during operation even at high temperatures. This results in a bearing unit that suppresses wear on the bearing surface and achieves a long lifespan. BRIEF DESCRIPTION OF THE DRAWINGS

[0031] Figure 1 This is a cross-sectional view of a spindle motor using the fluid dynamic bearing device (bearing unit) of the present invention.

[0032] Figure 2 It is a top view of the first oil-impregnated sintered bearing of the present invention.

[0033] Figure 3 It is a cross-sectional view of a first oil-impregnated sintered bearing according to the present invention.

[0034] Figure 4A This is a simplified diagram showing a chamfered portion formed on a bearing end surface, where the chamfer angle is less than 60°.

[0035] Figure 4B This is a simplified diagram showing a chamfered portion formed on a bearing end surface, where the chamfer angle is 60° or greater.

[0036] Figure 5 It is a top view of a second oil-impregnated sintered bearing according to the present invention.

[0037] Figure 6 It is a cross-sectional view of a second oil-impregnated sintered bearing according to the present invention.

[0038] Figure 7 It is a schematic top view of a third oil-impregnated sintered bearing according to the present invention.

[0039] Figure 8 It is a schematic cross-sectional view of a third oil-impregnated sintered bearing according to the present invention.

[0040] Figure 9 It is a schematic top view of a fourth oil-impregnated sintered bearing according to the present invention.

[0041] Figure 10 It is a schematic cross-sectional view of a fourth oil-impregnated sintered bearing according to the present invention.

[0042] Figure 11 This is a schematic diagram of a fifth oil-impregnated sintered bearing according to the present invention.

[0043] Figure 12It is an enlarged view of the main part of the fifth oil-impregnated sintered bearing of the present invention.

[0044] Figure 13A This is a simplified diagram showing the shapes of the dynamic pressure generating grooves and dynamic pressure generating hills, which is a basic shape.

[0045] Figure 13B This is a schematic diagram showing the shapes of the dynamic pressure generating grooves and the dynamic pressure generating hills, and is a first modified example.

[0046] Figure 13C This is a schematic diagram showing the shapes of the dynamic pressure generating grooves and the dynamic pressure generating hills, and is a second modified example.

[0047] Figure 14 This is a schematic cross-sectional view of another oil-impregnated sintered bearing used in a spindle motor using another fluid dynamic bearing device according to the present invention.

[0048] Figure 15 This is a simplified diagram showing the relationship between the dynamic pressure generating grooves and the dynamic pressure generating hills.

[0049] Figure 16 Graph showing the relationship between the hill-to-groove ratio and the eccentric angle.

[0050] Figure 17 This is a graph showing the results of measuring the oil film forming properties and vibration values ​​in a bearing provided with axial grooves having a hill-to-groove ratio of 0.7.

[0051] Figure 18 This is a graph showing the results of measuring the oil film formation and vibration value in a bearing provided with axial grooves having a hill-to-groove ratio of 2.5.

[0052] Figure 19 This is a graph showing the results of measuring the oil film formation properties and vibration values ​​in a true circular bearing.

[0053] Figure 20 This is a graph showing the relationship between groove depth and radial clearance.

[0054] Figure 21A This is a simplified diagram showing the relationship between the shaft member and the bearing, when the number of grooves is three.

[0055] Figure 21B This is a simplified diagram showing the relationship between the shaft member and the bearing when the number of grooves is four.

[0056] Figure 21C This is a simplified diagram showing the relationship between the shaft member and the bearing, when the number of grooves is five. DETAILED DESCRIPTION

[0057] exist Figure 1Figure 2 shows a spindle motor used in a disk drive device for an HDD. This spindle motor includes a fluid dynamic bearing device (bearing unit) 1, a hub 3 fixed to a shaft member 2 of the fluid dynamic bearing device (bearing unit) 1, a stator coil 4 and a rotor magnet 5 facing each other with a radial gap therebetween, and a bracket 6. The stator coil 4 is fixed to the bracket 6, and the rotor magnet 5 is fixed to the hub 3. A housing 7 of the fluid dynamic bearing device 1 is fixed to the inner diameter surface of the bracket 6. A predetermined number (two in the illustrated example) of magnetic disks D are held on the hub 3. When current is applied to the stator coil 4, the rotor magnet 5 rotates, causing the magnetic disks 10 held on the hub 3 to rotate integrally with the shaft member 2.

[0058] A fluid dynamic bearing device 1 includes a sintered oil-impregnated bearing 8 according to one embodiment of the present invention, a shaft member 2 inserted into the inner circumference of the sintered oil-impregnated bearing 8, a bottomed cylindrical housing 7 that secures the sintered oil-impregnated bearing 8 to its inner diameter surface, and a sealing member 9 disposed at the opening of the housing 7. In the following description of the fluid dynamic bearing device 1, for convenience, the opening side of the housing 7 in the axial direction is referred to as the upper side, and the opposite side is referred to as the lower side.

[0059] The flange portion 2b provided at the lower end of the shaft member 2 is accommodated between the lower end surface 8b of the oil-impregnated sintered bearing 8 and the upper end surface 7b1 of the bottom portion 7b of the housing 7 facing the lower end surface 8b.

[0060] like Figure 2 as well as Figure 3 As shown, the bearing inner diameter surface 8a of the sintered oil-impregnated bearing 8 (hereinafter sometimes referred to simply as the inner diameter surface) has a plurality (five in the example) of axial grooves 11 formed at predetermined intervals along the circumferential direction. Specifically, the inner diameter surface 8a of the sintered oil-impregnated bearing 8 includes the axial grooves 11 and hillocks 12 formed between adjacent grooves along the circumferential direction. Note that the hillocks 12 are formed from the remaining portion of the inner diameter surface resulting from the formation of the grooves 11.

[0061] However, the depth d of the groove 11 (see Figure 15 ) is set as the gap dimension c1 between the outer diameter surface 2a of the shaft member 2 and the inner diameter surface of the sintered oil-impregnated bearing 8 (ie, the hill portion 12) (refer to Figure 15 )) is less than 8 times. In this case, it is preferable to set the maximum value of the gap size c1 to 10 μm. In addition, the hill-to-groove ratio is set to 1 or less. Here, the hill-to-groove ratio of 1 means that the circumferential length H2 of the hill portion 12 (refer to Figure 15 ) and the circumferential length H1 of the groove portion 11 (refer to Figure 15) ratio (H2 / H1). Therefore, in this embodiment, the circumferential length H2 of the hill portion 12 is set to be the same as the circumferential length H1 of the groove portion 11, or the circumferential length H2 of the hill portion 12 is set to be shorter than the circumferential length H1 of the groove portion 11. However, the circumferential length H2 of the hill portion 12 is preferably 0.1 mm or more. Moreover, the interval (circumferential interval) B between the hill portions 12 (see Figure 2 ) is preferably set to 72° or less. Note that in this specification, the gap c1 between the outer diameter surface of the shaft member 2 and the inner diameter surface of the oil-impregnated sintered bearing 8 (ie, the hillock 12) is sometimes referred to as a bearing gap or a radial gap.

[0062] Furthermore, the surface opening ratio of the groove portion 11 is preferably larger than the surface opening ratio of the hill portion 12. The surface opening ratio refers to the ratio of the total area of ​​the openings (total area) per unit area.

[0063] However, this bearing 8 has a radial groove 15 formed on one (upper) bearing end surface, connecting the bearing inner diameter surface 8a with the bearing outer diameter surface (hereinafter sometimes referred to as the inner diameter surface) 8c. This groove 15 serves as a groove (oil buffer) for accumulating oil that overflows onto this bearing end surface. In this case, the volume of the groove 15 is preferably larger than the volume of the oil within the groove 15, which expands with rising temperature. In this embodiment, the groove 15 is located at a position corresponding to the groove portion 11 on the inner diameter surface. Specifically, in this embodiment, the grooves 15 are arranged at 72° intervals along the circumference. These grooves 15 are also arranged at 72° intervals along the circumference and are also set to be aligned in phase.

[0064] In addition, an axial groove 16 connecting the two end faces of the bearing is provided on the outer diameter surface of the bearing 8. The axial groove 16 may be an axial groove that is connected to the groove 15 or an axial groove that is not connected to the groove 15. Figure 2 as well as Figure 5 In the bearing 8 shown, the axial groove 16 communicates with the recessed groove 15. Therefore, the axial groove 16 is provided at a position corresponding to the recessed groove 15. Therefore, in this case, five axial grooves 16 are also provided at 72° intervals along the circumferential direction.

[0065] Furthermore, chamfered portions 17a, 17b, 18a, 18b are provided on the inner diameter side and the outer diameter side of both bearing end surfaces of the bearing 8. In this case, the chamfered portions 17a, 17b on the inner diameter side are chamfered at an angle θ (see Figure 4A as well as Figure 4B ) is set to 60° or above.

[0066] like Figure 4AAs shown in FIG. 1 , if the chamfer angle θ of the chamfered portion 17a (17b) on the inner diameter side is smaller than 60°, the oil will not easily flow to the bearing outer diameter surface 8c due to the capillary force of the space (tapered space) formed by the shaft outer diameter surface 2a and the chamfered portion 17a (17b). On the contrary, Figure 4B As shown in FIG. 1 , if the chamfer angle θ of the chamfered portion 17 a is as large as 60° or more, the capillary force is less likely to act, and the oil easily flows toward the bearing outer diameter surface 8 c.

[0067] In a spindle motor using a bearing 8 constructed in this manner, the following circulation path is formed: when the temperature of the bearing gap increases due to the rotation of the shaft member 2 and the surrounding environment is high temperature (for example, 150°C), causing the lubricating oil to expand and the oil level to rise, if the lubricating oil leaks out of the bearing gap, the lubricating oil accumulates in the groove 15 on the bearing end face, and flows downward from the groove 15 to the axial groove 16 on the outer diameter surface and returns to the inside of the bearing.

[0068] In this case, although not shown in the figure, a region in which a plurality of dynamic pressure generating grooves are arranged in a spiral shape is formed as a thrust dynamic pressure generating portion on the entire surface or a partial annular region of the lower end surface 8b of the oil-impregnated sintered bearing 8 .

[0069] The pressure of the lubricating oil film formed in the thrust bearing gap between the lower end surface 8b of the sintered oil-impregnated bearing 8 (where the dynamic pressure grooves are formed) and the upper end surface of the flange portion 2b facing it, and in the thrust bearing gap between the upper end surface 7b1 of the bottom 7b of the housing 7 (where the dynamic pressure grooves are formed) and the lower end surface of the flange portion 2b facing it, is increased by the dynamic pressure of the dynamic pressure grooves. Furthermore, under the pressure of these oil films, as shown in FIG. Figure 1 As shown, a first thrust bearing portion T1 and a second thrust bearing portion T2 are respectively configured to support the flange portion 2 b (shaft member 2 ) in a thrust direction in a non-contact manner.

[0070] The sintered oil-impregnated bearing of the present invention can reduce the eccentric angle by setting the dimensions and shape as described above. Specifically, by providing the axial groove 11, the eccentric angle is reduced compared to a case without the groove 11 (a perfect circle). Moreover, the eccentricity is further reduced when the hill-to-groove ratio is less than 1. Reducing the hill-to-groove ratio can reduce the eccentric angle. However, if the eccentric angle is too small, the hill width becomes narrow, which may lead to a reduction in load capacity, and thus, workability and wear resistance. Therefore, it is preferable to set the hill-to-groove ratio so that the circumferential length of the hill 12 is greater than 0.1 mm.

[0071] Here, when the shaft eccentricity is e and the radial clearance is c1, the eccentricity is e / c1 (refer to Figure 15). That is, relative to the state in which the shaft member 2 is not eccentric (indicated by the imaginary line), the actual shaft member 2 is eccentric as shown by the solid line. Therefore, the angle formed by the eccentric direction with respect to the reference line L0 is the eccentric angle. That is, the angle α formed by the line L3 connecting the bearing center O and the rotating shaft center O1 with respect to the reference line L0 is the eccentric angle, and the eccentricity refers to the value obtained by dividing the distance e between the bearing center O and the rotating shaft center O1 by the radial clearance c1. Therefore, when the shaft member 2 is in contact with the bearing 8, the eccentricity is 1.0, and when the shaft member 2 is located at the center position under the action of dynamic pressure, the eccentricity becomes 0.0.

[0072] Furthermore, if the groove depth d is large (deep), it becomes difficult to adjust the eccentric angle. Therefore, it is preferable to set the groove depth d to no more than eight times the gap dimension c1 between the bearing inner diameter surface 8a and the outer diameter surface 2a of the shaft member 2. Furthermore, if the circumferential spacing angle between adjacent hillocks is wide, and a load is applied to the shaft member 2, the load may not be supported. Therefore, if the circumferential spacing angle B between adjacent hillocks is no more than 72°, the load can be adequately supported.

[0073] By making the surface opening ratio of the groove 11 larger than that of the hill 12, it is possible to prevent pressure leakage on the bearing surface (hill 12) and prevent the generation of negative pressure on the bottom surface of the groove 11. Furthermore, sintered oil-impregnated bearings undergo sizing. Therefore, sizing is not necessary for the groove 11. Sizing refers to a light cold pressing process used to compress the sintered product to flatten warpage, correct dimensions, or improve surface conditions.

[0074] By having an oil buffer portion consisting of a groove 15 connected to the radial gap c1 between the outer diameter surface 2a of the shaft member 2, it is possible to effectively prevent oil leakage (scattering) from the bearing 8, and the volume of the groove 15 is larger than the volume of the oil in the groove 15 that expands as the temperature rises, so it is possible to effectively prevent the leakage of oil entering the groove 15 when the temperature rises.

[0075] In this case, it is preferable that the axial groove 11 on the bearing inner diameter surface 8a, the concave groove 15 on the bearing end surface 8d, and the axial groove 16 on the bearing outer diameter surface 8c are connected. This connection allows lubricating oil that flows into the concave groove 15 to return to the bearing interior via the axial groove 16. In other words, by aligning (connecting) the axial groove 11 on the bearing inner diameter surface 8a, the concave groove 15 on the bearing end surface, and the axial groove 16 on the bearing outer diameter surface 8c in phase, oil leakage can be effectively prevented. However, they may not be connected (connected) and not in phase.

[0076] Preferably, the shaft end portion of the bearing inner diameter surface 8a includes a chamfered portion 17a, and the chamfer angle θ of the chamfered portion 17a is 60° or greater. When the chamfer angle of the chamfered portion 17a is smaller than 60°, the capillary force in the space (tapered space) formed by the shaft outer diameter surface 2a and the chamfered portion 17a makes it difficult for oil to flow toward the bearing outer diameter surface 8c. Conversely, if the chamfer angle θ of the chamfered portion 17a is larger than 60°, the capillary force is less effective, and oil is more likely to flow toward the bearing outer diameter surface 8c.

[0077] A motor using the oil-impregnated sintered bearing of the present invention forms a circulation path: Lubricating oil that overflows from the bearing gap due to rising temperature enters grooves 15 in bearing end surface 8d, and from there, returns to bearing gap c1 via axial grooves 16 in bearing outer diameter surface 8c. Consequently, a motor using the oil-impregnated sintered bearing 8 of the present invention forms an oil film immediately after operation begins. This prevents the oil film from breaking down in the bearing gap during operation, even at high temperatures. This reduces wear on the bearing surface and extends its life.

[0078] Next, in Figure 5 as well as Figure 6 shows a second oil-impregnated sintered bearing 8. In this case, a circular countersink 20 is provided on the bearing end surface on the side where the groove 15 is formed. In this case, chamfered portions 17a and 17b are provided below the countersink 20. Therefore, the groove 15 provided on the bearing end surface communicates with the countersink 20.

[0079] Other structures and Figure 2 、 Figure 4A as well as Figure 4B The first sintered oil-impregnated bearing 8 is the same as that in the first embodiment, so the same parts and structures are marked as in the first embodiment. Figure 2 、 Figure 4A as well as Figure 4B The same reference numerals are used to denote the same reference numerals as those in the figure and their descriptions are omitted. Figure 5 as well as Figure 6 The second sintered oil-impregnated bearing 8 plays the role of Figure 2 、 Figure 4A as well as Figure 4B The same effect as the first oil-impregnated sintered bearing 8 is achieved. In this case, a circulation path is also formed: when the oil level rises and the lubricating oil leaks out of the bearing gap, the lubricating oil accumulates in the groove 15 on the bearing end surface, flows downward from the groove 15 to the axial groove 16 on the outer diameter surface, and returns to the inside of the bearing.

[0080] In addition, the shape of the hill 12 may be as follows: Figure 7 as well as Figure 8As shown, the inner diameter of the hill 12 is set to be smaller on the axial end side than on the axial center. Here, the inner diameter of the hill 12 refers to the inner diameter of the cylindrical surface formed by connecting the inner diameter end faces of the hill 12. When the inner diameter size on the axial end side is set to D1 and the inner diameter size on the axial center side is set to D2, it is set to D1<D2. By setting it in this way, the shaft member 2 can be supported on both end sides, and the effect of improving the supporting rigidity for moment loads and reducing torque can be expected. It should be noted that, in Figure 8 In the figure, the groove portion 11 of the inner diameter surface 8a is omitted for simplification of the drawing.

[0081] like Figure 9 as well as Figure 10 As shown, the circumferential length H2 of the hill 12 is shorter at the axial center than at the axial end. That is, if the circumferential length H1 at the axial end is A1 and the circumferential length H2 at the axial center is A2, then A1 > A2. This setting can achieve a torque reduction effect.

[0082] like Figure 11 as well as Figure 12 As shown, the lines L2 and L3 connecting the hill 12 and the groove 11 are formed parallel to the line L1 connecting the center of the hill 12 and the bearing center O. This shape can improve the workability of the mold.

[0083] In addition, in each of the above embodiments, the shape of the groove portion 11 is as follows: Figure 13A As shown, it is rectangular in shape when unfolded, but it can also be Figure 13B as well as Figure 13C That is, in the expanded view Figure 13A In the embodiment, the bottom surface 11a of the groove portion 11 is set as a flat surface with a constant depth dimension, but in the development view Figure 13B In the embodiment, the bottom surface 11a is set as an inclined surface that becomes deeper from the upstream side to the downstream side in the rotation direction of the shaft member. Figure 13C In the figure, the bottom surface is set to an arc shape surface.

[0084] So, even if Figure 13B as well as Figure 13C The shape shown also plays a role with Figure 13A The same shape as shown has the same effect.

[0085] Figure 14 The illustrated sintered oil-impregnated bearing 8 does not have axial grooves formed on its inner diameter surface. Instead, a region comprising multiple dynamic pressure generating grooves is formed as a radial dynamic pressure generating portion, either entirely or in a portion of the cylindrical region of the inner diameter surface 8a of the sintered oil-impregnated bearing 8. In this embodiment, two regions comprising multiple dynamic pressure generating grooves 8a1 and 8a2 arranged in a herringbone pattern are formed, separated in the axial direction.

[0086] In a fluid dynamic bearing device using the sintered oil-impregnated bearing 8 with the above-described structure, when the shaft member 2 rotates, the areas where the dynamic pressure generating grooves 8a1 and 8a2 are formed on the inner diameter surface 8a of the sintered oil-impregnated bearing 8 form a radial bearing gap with the opposing outer diameter surface 2a of the shaft member 2. Furthermore, as the shaft member 2 rotates, the lubricating oil in this radial bearing gap is forced toward the axial center of the dynamic pressure generating grooves 8a1 and 8a2, causing its pressure to increase. This creates a radial bearing portion that provides non-contact support for the shaft member 2 in the radial direction due to the dynamic pressure of the lubricating oil generated by the dynamic pressure generating grooves 8a1 and 8a2.

[0087] In this case, the sintered oil-impregnated bearing 8 is also provided with a groove 15 on the bearing end surface and an axial groove 16 on the outer diameter surface, thereby forming the following circulation path: when the oil level rises and the lubricating oil leaks out of the bearing gap, the lubricating oil accumulates in the groove 15 on the bearing end surface and flows downward from the groove 15 to the axial groove 16 on the outer diameter surface 8c and returns to the inside of the bearing. It should be noted that the dynamic pressure groove is provided as the radial dynamic pressure generating portion. Figure 4A as well as Figure 4B The threads are arranged in a herringbone shape, but can also be in other shapes such as a spiral.

[0088] Thus, a motor using the oil-impregnated sintered bearings of the above-described embodiments forms an oil film immediately after operation begins, preventing the oil film from breaking in the bearing gap during operation even at high temperatures, thereby suppressing wear on the bearing surface and extending its life.

[0089] The embodiments of the present invention have been described above, but the present invention is not limited to the embodiments described above and can be modified in various ways. For example, the embodiments described above show a case where the sintered oil-impregnated bearing 8 is fixed and the shaft member 2 rotates, but the present invention is not limited to this. A structure in which the shaft member 2 is fixed and the sintered oil-impregnated bearing 8 is rotated, or a structure in which both the shaft member 2 and the sintered oil-impregnated bearing 8 are rotated can also be adopted.

[0090] The fluid dynamic pressure bearing device assembled with the sintered oil-containing bearing 8 of the present invention is not limited to the spindle motor used in the disk drive device of the HDD, but can also be widely used in other small motors such as the spindle motor assembled in other information equipment, the polygon scanner motor of the laser printer, the color wheel of the projector, or the fan motor for cooling.

[0091] Example 1

[0092] Experiments (tests) were conducted to confirm the effects of the present invention.

[0093] First, the relationship between the hill-to-groove ratio and the eccentricity angle was investigated. Figure 15 As shown, a test product was used in which grooves 11 extending axially on the inner diameter surface 8a were formed with six grooves at 60° intervals along the circumferential direction. Specifically, the bearing dimensions were set to an inner diameter of 4 mm, a width (wall thickness) of 2.7 mm, a radial clearance of 6.5 μm, a groove depth d of 35 μm, a lubricating oil viscosity of 0.0601 (Pa·s), and an eccentricity of 0.3.

[0094] A true circular bearing without grooves on its inner diameter surface was designated as Test 1, a bearing with a hill-to-groove ratio of 0.5 was designated as Test 2, a bearing with a hill-to-groove ratio of 1.0 was designated as Test 3, a bearing with a hill-to-groove ratio of 1.5 was designated as Test 4, a bearing with a hill-to-groove ratio of 2.0 was designated as Test 5, and a bearing with a hill-to-groove ratio of 2.5 was designated as Test 6. The rotational speed was set at 3900 rpm.

[0095] exist Figure 16 The relationship between the hill-to-groove ratio and the eccentric angle is shown in FIG. Figure 16 As can be seen, the eccentric angle of a bearing having grooves 11 is smaller than that of a true circular bearing. In particular, the eccentric angle decreases further when the hill-to-groove ratio is 1.0 or less. However, decreasing the hill-to-groove ratio reduces the circumferential width of the hill 12 (narrower), resulting in a reduction in load capacity and, in addition, a decrease in workability and wear resistance. Therefore, it is preferable to set the hill-to-groove ratio so that the circumferential length of the hill 12 is 0.1 mm.

[0096] Next, the oil film formation and vibration values ​​of the bearings of Test 7, which had a hill-to-groove ratio of 0.7, Test 8, which had a hill-to-groove ratio of 2.5, and Test 9, a true circular bearing without grooves on the inner diameter surface, were examined. The bearing dimensions for Test 7, 8, and 9 were: an inner diameter of 4 mm, a width (wall thickness) of 2.7 mm, a radial clearance of 6.5 μm, a lubricating oil viscosity of 0.0601 (Pa·s), a rotational speed of 3900 rpm, and a load of 5 N. Test 7 and Test 8 had six grooves, each with a groove depth of 35 μm.

[0097] In the next Figures 17 to 19 And the test results are shown in Table 1. Figure 17 This is the measurement result of test product 7. Figure 18 This is the measurement result of test product 8. Figure 19 This is the measurement result of Test Sample 9. The vibration values ​​of Test Samples 7 to 9 are shown in Table 1. The vibration values ​​were obtained using a vibration sensor (PV-90B manufactured by RION Co., Ltd.).

[0098] [Table 1]

[0099]

[0100] according to Figures 17 to 19 It can be seen that the provision of the groove portion 11 on the inner diameter surface 8a improves the oil film formation. In addition, in the bearing with a hill-to-groove ratio of 0.7, the oil film is stably formed immediately after startup, and the vibration value is also increased as shown in the following order: Test piece 7 < Test piece 8 < Test piece 9. That is, as shown in Table 1, the vibration value of Test piece 7 with a hill-to-groove ratio of 0.7 is 1.79 mm / s. 2 , in the test piece 8 with a hill-to-groove ratio of 2.50, it was 1.88 mm / s 2 , and 2.00 mm / s in the case of the test piece 9 which is a true round bearing without a groove. 2 .

[0101] Next, we investigated the relationship between groove depth and radial clearance. In this case, the bearing dimensions were set to an inner diameter of 4 mm, a width (wall thickness) of 2.7 mm, six grooves, a lubricating oil viscosity of 0.0601 (Pa·s), a rotational speed of 3900 rpm, and an eccentricity of 0.3. In this case, a bearing with a radial clearance of 2.0 μm was set as test product 10, a bearing with a radial clearance of 4.0 μm was set as test product 11, a bearing with a radial clearance of 6.0 μm was set as test product 12, a bearing with a radial clearance of 8.0 μm was set as test product 13, and a bearing with a radial clearance of 10.0 μm was set as test product 14. For each test product, tests were conducted on test products with groove depths of 5.0 μm, 10.0 μm, 15.0 μm, 20.0 μm, 25.0 μm, 30.0 μm, 35.0 μm, 40.0 μm, 45.0 μm, and 50.0 μm. Figure 20 The results are shown in .

[0102] according to Figure 20 It can be seen that groove depth and radial clearance interact (a synergistic effect occurs when the two factors are combined), resulting in a combination that reduces the eccentric angle. For example, when the groove depth is 10 μm, a radial clearance of 8 to 10 μm can reduce the eccentric angle. Furthermore, increasing the groove depth reduces the difference in eccentric angle based on the radial clearance. Therefore, it can be said that it is best to set the groove depth to no more than 8 times the radial clearance.

[0103] Next, the positional relationship between the bearing 8 and the shaft member 2 with three, four, and five grooves was investigated. Figure 21A The figure shows a bearing with 3 grooves. Figure 21B The figure shows a bearing with 4 grooves. Figure 21CThe figure shows a bearing with 5 grooves. A load is applied to the shaft member 2 of each bearing 8 in one direction. When the load is directed toward one groove, pressure is applied to the hillocks at both ends of the groove. Figures 21A to 21C As shown in , the direction of the load does not match the direction of the pressure. In this case, when the direction of the vector is considered in a bearing with three grooves, the difference in the direction of the vector is large, and the difference in the direction of the vector becomes 3 bearings > 4 bearings > 5 bearings. In the case of a bearing with three or four grooves, the load may not be fully supported. However, as Figure 21C As shown, if there are five or more grooves in which the circumferential interval B between circumferentially adjacent hills is 72 degrees, the difference in the direction of the vector is not too great, and even such a load can be sufficiently supported.

[0104] Industrial Applicability

[0105] The sintered oil-impregnated shaft can suppress whirling vibration of the shaft and ensure quietness. A structure in which the shaft member is fixed and the sintered oil-impregnated bearing is rotated, or a structure in which both the shaft member and the sintered oil-impregnated bearing are rotated can also be adopted.

[0106] Description of Reference Numerals

[0107] θ chamfer angle

[0108] 2-axis component

[0109] 2a Outer diameter surface

[0110] 8 Sintered oil-impregnated bearings

[0111] 8a inner diameter surface

[0112] 8a1 dynamic pressure groove

[0113] 11 Groove

[0114] 12 Qiubu

[0115] 15 grooves

[0116] 16 Axial groove

[0117] 17a Chamfered part

[0118] B Circumferential spacing.

Claims

1. A sintered oil-impregnated bearing having an inner diameter surface of the bearing opposed to an outer diameter surface of a shaft member, wherein the inner diameter surface of the bearing has five or more axial grooves formed at predetermined intervals along the circumferential direction, and hills are formed between adjacent grooves along the circumferential direction. The sintered oil-impregnated bearing is characterized in that: The surface opening ratio of the groove portion is made larger than the surface opening ratio of the hill portion, and the hill-to-groove ratio, which is the ratio of the hill portion to the groove portion, is set to less than 1, the circumferential length of the hill portion is set to more than 0.1 mm, and the depth of the groove portion is set to less than 8 times the gap size between the inner diameter surface and the outer diameter surface of the shaft component, and the spacing angle of the circumferential spacing between adjacent hill portions along the circumferential direction is set to less than 72°.

2. The oil-impregnated sintered bearing according to claim 1, characterized in that: The sintered oil-containing bearing has an oil buffer portion on at least one side of the bearing end surface for preventing leakage of lubricating oil and circulating the lubricating oil. The oil buffer portion is composed of a groove that communicates with the bearing gap between the inner diameter surface of the bearing and the outer diameter surface of the shaft member, and an axial groove that connects the two end surfaces of the bearing is provided on the outer diameter surface of the bearing, so that the volume of the groove is larger than the volume of the oil in the groove that expands with the temperature rise.

3. The sintered oil-impregnated bearing according to claim 2, characterized in that: The axial groove portion of the bearing inner diameter surface, the concave groove of the bearing end surface, and the axial groove of the bearing outer diameter surface are in communication with each other.

4. The oil-impregnated sintered bearing according to claim 1, characterized in that: The oil-impregnated sintered bearing has a chamfered portion at an axial end portion of a bearing inner diameter surface, and the chamfered portion has a chamfer angle of 60° or greater.

5. A bearing unit comprising the oil-impregnated sintered bearing according to claim 2, a shaft member inserted into an axial center hole of the oil-impregnated sintered bearing, and a housing accommodating the oil-impregnated sintered bearing on an inner diameter surface thereof. The bearing unit is characterized in that The bearing is assembled in such a manner that the groove on the end face of the bearing faces the side open to the atmosphere, and the following circulation path is formed: due to the temperature rise, the bearing gap between the outer diameter surface of the shaft member and the opposing inner diameter surface of the bearing generates a flow of lubricating oil toward the side open to the atmosphere, the lubricating oil overflowing from the bearing gap enters the groove on the end face of the bearing, and the lubricating oil returns to the bearing gap from the groove via the axial groove on the outer diameter surface of the bearing.

6. A bearing unit comprising: a sintered oil-impregnated bearing having no axial grooves formed on its inner diameter surface but having other dynamic pressure generating grooves formed thereon; a shaft member inserted into an axial center hole of the sintered oil-impregnated bearing; and a housing accommodating the sintered oil-impregnated bearing on its inner diameter surface. The bearing unit is characterized in that The bearing unit has an oil buffer portion on at least one side of the bearing end face to prevent leakage of the lubricating oil and to circulate the lubricating oil. The oil buffer portion is composed of a groove connected to the bearing gap between the inner diameter surface of the bearing and the outer diameter surface of the shaft component, and an axial groove connecting the two end faces of the bearing is provided on the outer diameter surface of the bearing, so that the volume of the groove is larger than the volume of the oil in the groove that expands with the temperature rise, and the bearing unit is assembled in a manner so as to face the atmospheric open side with the groove of the bearing end face facing the atmospheric open side, and the following circulation path is formed: due to the temperature rise, the bearing gap between the outer diameter surface of the shaft component and the opposite inner diameter surface of the bearing generates a flow of lubricating oil toward the atmospheric open side, the lubricating oil overflowing from the bearing gap enters the groove of the bearing end face, and the lubricating oil returns to the bearing gap from the groove via the axial groove of the bearing outer diameter surface.

7. A motor, characterized in that: The motor includes the oil-impregnated sintered bearing according to any one of claims 1 to 4 .

8. A motor, characterized in that: The motor uses the bearing unit according to claim 5 or claim 6.

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