Sintered oil-impregnated bearing and fluid dynamic pressure bearing device

Abstract

The present invention comprises a radial dynamic pressure generation part in which a step surface having axial-direction grooves which have rectangular cross sections and are disposed at predetermined intervals along the circumferential direction is formed on the inner diameter surface of a bearing hole. The rectangularity of the axial-direction grooves calculated from a shape obtained by linearly expanding the roundness of the step surface measured by least squares centering is at most 0.025, and the flatness of hill parts between the axial-direction grooves adjacent to each other in the circumferential direction is at least 0.80. Rectangularity is an index that approaches 0 as the similarity of a shape to a rectangle increases, and flatness is an index that approaches 1 as the number of flat parts increases.

Description

Sintered oil-impregnated bearing and fluid dynamic bearing device

[0001] The present invention relates to a sintered oil-impregnated bearing and a fluid dynamic bearing device.

[0002] Oil-impregnated sintered bearings are generally used for bearings that support the rotation of small motors such as spindle motors and fan motors for magnetic disk drives (HDDs) used in polygon mirrors of laser beam printers (LBPs).

[0003] This type of sintered oil-impregnated bearing can be used in a fluid dynamic bearing device in which herringbone or spiral type dynamic pressure grooves are provided on the bearing surface, and a dynamic pressure oil film is generated in the bearing gap by the action of the dynamic pressure grooves as the shaft rotates, thereby floating and supporting the shaft.

[0004] Conventionally, there is a sintered oil-impregnated bearing in which the inner diameter surface of the bearing bore is provided with substantially rectangular stepped portions in the circumferential direction, each having a sliding surface concentric with the rotating shaft, and the gap between adjacent stepped portions and the inner diameter surface of the bearing bore and the rotating shaft is substantially rectangular (Patent Document 1). Specifically, in the sintered oil-impregnated bearing described in Patent Document 1, the inner diameter surface of the bearing bore is provided with a stepped surface having axial grooves with rectangular cross sections arranged at a predetermined pitch along the circumferential direction. Therefore, a sintered oil-impregnated bearing with a stepped surface can be called a stepped hydrodynamic bearing. Sintered oil-impregnated bearings are bearings in which a porous sintered material made by compressing and heating metal powder is impregnated with lubricating oil, while hydrodynamic bearings are bearings in which hydrodynamic grooves are provided in the inner diameter surface of a sintered oil-impregnated bearing, and an oil film is formed between the bearing and the shaft to rotate the bearing.

[0005] A fluid dynamic bearing device for a fan motor equipped with blower blades includes a cylindrical bearing member that supports the motor's main shaft in a non-contact manner so that it can rotate freely in the radial direction. This bearing member can be a true circular bearing, in which the radial bearing surface that forms a radial bearing gap between the outer circumferential surface of the shaft is a smooth cylindrical surface without any irregularities, or a dynamic pressure bearing (sintered oil-impregnated bearing) in which a dynamic pressure generator is provided on the radial bearing surface. When a dynamic pressure bearing is used as the bearing member, as the shaft rotates relative to the dynamic pressure bearing, the dynamic pressure action of the dynamic pressure generator increases the pressure (rigidity) of the fluid film that forms in the radial bearing gap (strictly speaking, the narrowest gap width within the radial bearing gap). For this reason, dynamic pressure bearings are considered to have superior bearing rigidity to true circular bearings.

[0006] Known hydrodynamic bearings include: (1) those having a hydrodynamic pressure generating portion composed of a plurality of hydrodynamic grooves inclined with respect to the axial direction and spaced apart in the circumferential direction, and convex ridges that define the hydrodynamic grooves, the ridges forming a herringbone pattern; and (2) those having a hydrodynamic pressure generating portion with an uneven surface, composed of a plurality of linear hydrodynamic grooves extending in the axial direction and convex ridges alternately arranged in the circumferential direction. When a radial bearing gap is formed between the radial bearing surface and the outer peripheral surface of the shaft as the shaft rotates relative to these hydrodynamic bearings (1) and (2), the opposing region of the convex ridges becomes the "minimum width portion of the radial bearing gap." Note that, hereinafter, when describing the hydrodynamic bearings (1) and (2) separately, the hydrodynamic bearing (1) will also be referred to as a "herringbone-type hydrodynamic bearing," and the hydrodynamic bearing (2) will also be referred to as a "step-type hydrodynamic bearing."

[0007] In order to ensure the required bearing rigidity while reducing torque during relative rotation with the shaft, a hydrodynamic bearing often has two radial bearing surfaces spaced apart in the axial direction. In this case, a large-diameter cylindrical surface with a larger diameter than the radial bearing surfaces is provided on the inner peripheral surface of the hydrodynamic bearing in the axially intersecting region between the two radial bearing surfaces. This large-diameter cylindrical surface (the portion where it is provided) is also referred to as a "middle relief portion" (see, for example, Patent Document 2).

[0008] Of the hydrodynamic bearings having the above configuration, herringbone hydrodynamic bearings have the advantage of being superior to step-type hydrodynamic bearings in terms of their ability to form a fluid film in the radial bearing gap, and therefore in terms of bearing rigidity, but they are less costly than step-type hydrodynamic bearings for reasons such as the need to change the inclination direction of the hydrodynamic grooves depending on the direction of rotation of the shaft, which necessitates measures to prevent incorrect assembly into the bearing device, and the time and effort required to finish the hydrodynamic pressure generating section to the required precision. For this reason, step-type hydrodynamic bearings are preferably used in applications where cost is a particular concern.

[0009] Japanese Patent Laid-Open No. 5-115146 Japanese Patent Laid-Open No. 2001-275306

[0010] As described in Patent Document 1, in a bearing having a step surface with axial grooves of rectangular cross section arranged at a predetermined pitch along the circumferential direction, if the axial grooves are perfectly rectangular in cross section, the oil film pressure will be high, but if they are not perfectly rectangular, the oil film pressure will be high or unstable. Also, the flatter the ridges between the axial grooves are, the larger the contact area will be, and the lower the contact surface pressure will be.

[0011] However, the bearing described in Patent Document 1 does not disclose any specification for the degree of rectangularity of the cross-sectional shape of the axial grooves (specification of rectangularity). Furthermore, the description only states that the ridges between the axial grooves are concentric with the rotation axis, but does not disclose the degree of flatness (flatness) of these ridges. Therefore, the bearing described in Patent Document 1 may not have a perfect rectangular shape, which may result in high or unstable oil film pressure. Furthermore, if the ridges have a low flatness, the contact area cannot be increased, potentially making it difficult to reduce the contact surface pressure. Therefore, it was unclear whether the bearing described in Patent Document 1 could effectively exert a dynamic pressure effect.

[0012] Step-type hydrodynamic bearings are superior to herringbone-type hydrodynamic bearings in terms of cost, but are said to be somewhat inferior in terms of performance (especially bearing rigidity). Therefore, if the bearing rigidity of step-type hydrodynamic bearings could be increased without compromising their cost advantage, it would be possible to adopt step-type hydrodynamic bearings in applications requiring high bearing rigidity, which would contribute to reducing the cost of motors.

[0013] In view of the above situation, the present invention aims to provide a sintered oil-impregnated bearing in which a step surface having axial grooves with rectangular cross sections arranged at a predetermined pitch along the circumferential direction is formed on the inner diameter surface of the bearing hole, and which is capable of effectively exerting the dynamic pressure effect.

[0014] Incidentally, step-type hydrodynamic bearings are superior to herringbone-type hydrodynamic bearings in terms of cost, but are said to be somewhat inferior in terms of performance (especially bearing rigidity). Therefore, if the bearing rigidity of step-type hydrodynamic bearings could be increased without compromising their cost advantage, it would be possible to adopt step-type hydrodynamic bearings in applications requiring high bearing rigidity, which would contribute to reducing the cost of motors.

[0015] Therefore, another object of the present invention is to provide a hydrodynamic bearing having excellent bearing rigidity, in which the hydrodynamic pressure generating portion provided on the radial bearing surface is made of an uneven surface with a plurality of hydrodynamic grooves extending in the axial direction and convex portions arranged alternately in the circumferential direction, and can be molded inexpensively.

[0016] The sintered oil-impregnated bearing of the present invention is a sintered oil-impregnated bearing in which a shaft member is inserted into a bearing bore, and is provided with a radial dynamic pressure generating portion formed by a step surface formed on the inner diameter surface of the bearing bore, the step surface having axial grooves with a rectangular cross section arranged at a predetermined pitch along the circumferential direction, and the rectangularity of the axial groove calculated from the linearly developed shape of the circularity shape of the step surface measured using the least squares central method is 0.025 or less, and the flatness of the hill portions between circumferentially adjacent axial grooves is 0.80 or more, the rectangularity being calculated based on the inclination of the groove side surface between a first portion that is 20% of the groove depth from the circumscribed circle and a second portion that is 20% of the groove depth from the inscribed circle, and the rectangularity is an index of the degree of rectangularity being closer to 0, and the flatness is calculated based on the ratio of the total width dimension of the hill portions at the second portion to the total length of the linearly developed shape, and the flatness is an index of the degree of flatness being closer to 1. In other words, flatness is an index where 1 is the maximum, and the closer it is to 1, the larger the contact area becomes, allowing the contact area to be reduced and the more effectively the dynamic pressure effect can be exerted. Here, the least squares central method finds the reference circle where the square of the error is the smallest, and then finds the radii of the circumscribed and inscribed circles that are concentric with this reference circle. In this case, the roundness is calculated as (radius of the circumscribed circle) - (radius of the inscribed circle). Here, the circumscribed circle is the maximum groove depth (maximum diameter), and the inscribed circle is the minimum groove depth (minimum diameter). Linear expansion is the expansion of polar coordinates obtained from roundness data into Cartesian coordinates, and the Cartesian coordinates are obtained by analyzing the roundness data using flatness.

[0017] In the sintered oil-impregnated bearing of the present invention, the cross-sectional shape of the axial groove is close to rectangular, which increases the oil film pressure, and the flatness of the ridges is high and the flat portions of the ridges are wide, which increases the contact area and reduces the contact surface pressure, thereby effectively exerting the dynamic pressure effect.

[0018] The rectangularity is calculated based on the inclination of the groove side surface between the first portion at 20% of the groove depth from the maximum groove depth and the second portion at 20% of the groove depth from the minimum groove depth, so that the rectangularity can be calculated stably.

[0019] In particular, when the total length of the linearly developed shape of the step surface roundness profile measured by the least squares central method is defined as W1, the total groove width of all grooves in the first region is defined as W2, and the total width of all hills in the second region is defined as W3, W1 - (W2 + W3) is defined as the inclined portion data, the ratio of the inclined portion data to the total length is calculated, and the calculated value is divided by the number of grooves to obtain the rectangularity. By setting it in this way, the reliability of the calculated rectangularity is improved.

[0020] The total width of all the hills in the second region is defined as W3, and the ratio (W1 / W3) of this W3 to the total length W1 is calculated. The calculated value is divided by the value obtained by hill-groove ratio / (hill-groove ratio+1) to obtain the flatness. By setting the flatness in this manner, the reliability of the calculated flatness is improved.

[0021] The groove depth of the axial groove is preferably 2 μm or more. In this case, the groove depth is the groove depth at the axial middle position of the axial groove. By setting it in this manner, dynamic pressure can be generated stably.

[0022] The hill-groove ratio is the hill-groove ratio on the squared centerline (on the squared centerline) of the axial groove, and this hill-groove ratio is preferably 0.1 to 1.0. In this case, when the width dimension of the hill is W10 and the width dimension of the groove is W11, the hill-groove ratio is W10 / W11. By setting it in this manner, it is possible to generate dynamic pressure and ensure a sufficient sliding area, thereby obtaining stable rotational motion. Here, the hill-groove ratio on the squared centerline of the axial groove refers to the root mean square height of the axial groove.

[0023] The bearing rigidity of a step-type hydrodynamic bearing can be increased, for example, by increasing the area of ​​the ridges of the hydrodynamic pressure generating portion, which form the narrowest portion of the radial bearing gap during relative rotation with the shaft. The area of ​​the ridges can be increased by elongating the ridges in the axial and / or circumferential directions. However, elongating the ridges in the axial direction counteracts this by shortening the central relief portion formed between the two radial bearing surfaces in the axial direction, potentially reducing the torque reduction effect. Furthermore, elongating the ridges in the circumferential direction counteracts this by shortening the hydrodynamic grooves in the circumferential direction. In this case, the distance over which the fluid can flow within the radial bearing gap is shortened, weakening the fluid's pushing force toward the ridges, potentially reducing the ability to form a fluid film in the region facing the ridges and, ultimately, the bearing rigidity. Therefore, we have proposed a configuration that allows the hydrodynamic grooves to be elongated in the axial direction without changing the area (axial and / or circumferential dimensions) of the ridges that make up the hydrodynamic pressure generating portion, thereby improving the ability to form a fluid film in the region facing the ridges (bearing rigidity).

[0024] That is, the sintered oil-impregnated bearing has two radial bearing surfaces spaced apart in the axial direction that form a radial bearing gap between itself and the outer peripheral surface of the shaft to be supported, and an inner peripheral surface on which a large-diameter cylindrical surface having a diameter larger than that of the radial bearing surfaces is provided between the two radial bearing surfaces, and each radial bearing surface is provided with a radial dynamic pressure generating portion for generating a dynamic pressure action on the fluid present in the radial bearing gap, and this radial dynamic pressure generating portion is made up of dynamic pressure grooves that are a plurality of axial grooves extending in the axial direction and an uneven surface on which convex ridges are arranged alternately in the circumferential direction, and it is preferable that one of the radial bearing surfaces and the large-diameter cylindrical surface are connected via a steeply inclined surface that is inclined with respect to the axial direction, and the other radial bearing surface and the large-diameter cylindrical surface are connected via a gently inclined surface whose inclination angle with respect to the axial direction is smaller than that of the steeply inclined surface, and that one axial end of the dynamic pressure grooves provided in the other radial bearing surface is open to the gently inclined surface.

[0025] As described above, by making one axial end of the hydrodynamic groove provided on the other radial bearing surface open to a gentle slope, the hydrodynamic groove can be elongated in the axial direction without elongating the ridge portion that constitutes the hydrodynamic pressure generating portion in the axial and / or circumferential directions. By elongating the hydrodynamic groove in the axial direction, the amount of fluid flowing within the radial bearing gap during relative rotation between the shaft and the hydrodynamic bearing, and ultimately the amount of fluid pushed toward the ridge portion, can be increased accordingly, thereby improving the ability to form a fluid film in (the narrowest portion of) the radial bearing gap and increasing bearing rigidity.

[0026] In the above configuration, when the dynamic pressure generating portion is developed on a plane, the ratio (= Z2 / Z1) of the circumferential dimension Z2 of the ridge portion to the circumferential dimension Z1 of the dynamic pressure groove (groove width) is set to 1 or less. In other words, the circumferential dimension of the dynamic pressure is set to be equal to or greater than the circumferential dimension of the ridge portion. In this way, when the shaft and dynamic pressure bearing rotate relative to each other, the fluid in the radial bearing gap can be caused to move significantly in the circumferential direction, thereby improving the ability to form a fluid film in the narrowest part of the radial bearing gap.

[0027] In the above configuration, the difference in height between the inner diameter end face of the ridge portion and the bottom face of the dynamic pressure groove is preferably 2 μm or more.

[0028] The hydrodynamic bearing according to the present invention may be made of a porous sintered metal. A hydrodynamic bearing made of sintered metal allows the hydrodynamic pressure generating portion of the radial bearing surface to be formed easily and accurately. In this case, the other radial bearing surface may be a formed surface formed by subjecting the sintered metal to plastic processing.

[0029] A fluid dynamic bearing device comprising a bearing member comprising the dynamic pressure bearing according to the present invention described above, a bottomed cylindrical housing that houses this bearing member on its inner periphery, lubricating oil as a fluid present in the radial bearing gap, and a radial bearing portion that supports the shaft in a non-contact manner in the radial direction by an oil film of lubricating oil formed in the radial bearing gap, can exhibit high bearing performance at low cost because the dynamic pressure bearing according to the present invention has the above-mentioned characteristics.

[0030] It is now possible to numerically determine the shape of the ridges of a bearing (sintered oil-impregnated bearing) that excels in oil film formation. This makes it possible to provide a sintered oil-impregnated bearing that effectively exerts hydrodynamic effects as a finished product. Furthermore, it is possible to provide a hydrodynamic bearing that excels in the formation of a fluid film in the radial bearing gap and has excellent bearing rigidity, while the hydrodynamic pressure generating portion provided on the radial bearing surface can be molded inexpensively.

[0031] FIG. 1 is a diagram showing a linear development of the roundness profile of a step surface measured by the least-squares central method on the inner diameter surface of a bearing hole of the present invention. FIG. 2 is a cross-sectional view of a spindle motor using a sintered oil-impregnated bearing. FIG. 3 is a transverse cross-sectional view of a sintered oil-impregnated bearing. FIG. 4 is a longitudinal cross-sectional view of a sintered oil-impregnated bearing. FIG. 5 is a block diagram showing steps in a method for manufacturing a sintered oil-impregnated bearing. FIG. 6 is an enlarged view of a main part of FIG. 1. FIG. 7 is a simplified cross-sectional view of a non-defective product showing axial grooves. FIG. 8 is a simplified cross-sectional view of a defective product showing axial grooves. FIG. 9 is a simplified diagram showing the roundness profile measured by the least-squares central method. FIG. 10 is a cross-sectional view conceptually showing an example of a fan motor. FIG. 11 is a longitudinal cross-sectional view conceptually showing a fluid dynamic bearing device that can incorporate a dynamic bearing (sintered oil-impregnated bearing) according to an embodiment of the present invention. FIG. 12 is a longitudinal cross-sectional view of a dynamic bearing (sintered oil-impregnated bearing) according to an embodiment of the present invention. FIG. 1 is a schematic vertical cross-sectional view showing a state in which raw material powder filled in a cavity of a compression molding die is being compressed. FIG. 1 is a schematic vertical cross-sectional view showing a state in which a green compact has been discharged outside the compression molding die. FIG. 1 is a schematic vertical cross-sectional view showing a state in which a sintered body has been placed in a sizing die. FIG. 2 is a schematic vertical cross-sectional view showing a state in which a sintered body placed in the sizing die is being sized. FIG. 3 is a schematic vertical cross-sectional view showing a sizing die used to obtain a dynamic pressure bearing (sintered oil-impregnated bearing) according to another embodiment of the present invention, showing a state in which a sintered body has been placed in the sizing die. FIG. 4 is a schematic vertical cross-sectional view showing a sizing die used to obtain a dynamic pressure bearing (sintered oil-impregnated bearing) according to another embodiment of the present invention, showing a state in which a sintered body placed in the sizing die is being sized.

[0032] An embodiment of the present invention will now be described with reference to FIGS. 1 to 17B. FIG. 2 shows a spindle motor used in a disk drive device for an HDD. This spindle motor includes a fluid dynamic bearing device 1, a disk hub 3 fixed to a shaft member 2 of the fluid dynamic bearing device 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 disk 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 of disks 10 (two in the illustrated example) are held on the disk hub 3. When current is applied to the stator coil 4, the rotor magnet 5 rotates, causing the disks 10 held on the disk hub 3 to rotate integrally with the shaft member 2.

[0033] The fluid dynamic bearing device 1 comprises a sintered oil-impregnated bearing 8 according to an embodiment of the present invention, a shaft member 2 inserted into the inner periphery of the sintered oil-impregnated bearing 8, a cylindrical housing 7 with a bottom and the sintered oil-impregnated bearing 8 fixed to the inner diameter surface, and a seal member 9 disposed at the opening of the housing 7. For convenience in the following description of the fluid dynamic bearing device 1, the open side of the housing 7 in the axial direction will be referred to as the upper side, and the opposite side as the lower side.

[0034] A flange portion 2 b provided at the lower end of the shaft member 2 is housed between the lower end surface 8 e of the sintered oil-impregnated bearing 8 and the upper end surface 7 b 1 of the opposing bottom portion 7 b of the housing 7 .

[0035] The sintered oil-impregnated bearing 8 has a bearing hole 8A into which a shaft member is inserted. A first radial bearing surface and a second radial bearing surface are formed on the inner diameter surface 8A1 of the bearing hole 8A. This results in a first radial dynamic pressure generating portion 8a and a second radial dynamic pressure generating portion 8b being formed at two locations axially spaced apart. As shown in FIG. 3 , the first radial dynamic pressure generating portion 8a is configured as a step surface having a plurality (four in the illustrated example) of axial grooves G (G1) with a rectangular cross section, arranged at a predetermined pitch (90-degree pitch in the illustrated example) along the circumferential direction. A ridge portion H (H1) is provided between adjacent axial grooves G1, G1 in the circumferential direction. The second radial dynamic pressure generating portion 8b is also configured as a step surface having a plurality (four in the illustrated example) of axial grooves G (G2) with a rectangular cross section, arranged at a predetermined pitch (90-degree pitch in the illustrated example) along the circumferential direction. A ridge portion H (H2) is provided between adjacent axial grooves G2, G2 in the circumferential direction.

[0036] In a hydrodynamic bearing device configured in this manner, when the shaft member 2 rotates, the region where the axial grooves G formed on the inner diameter surface of the sintered oil-impregnated bearing 8 are formed forms a radial bearing gap with the opposing outer peripheral surface 2a of the shaft member 2. As a result, the dynamic pressure action of the lubricating oil generated by the axial grooves G forms a first radial bearing portion R1 and a second radial bearing portion R2 that support the shaft member 2 in a non-contact manner in the radial direction, as shown in Figure 2.

[0037] At the same time, the pressure of the lubricating oil film formed in the thrust bearing gap between the lower end surface 8e (dynamic pressure groove forming region) of the sintered oil-impregnated bearing 8 and the opposing upper end surface of the flange portion 2b, and in the thrust bearing gap between the upper end surface 7b1 (dynamic pressure groove forming region) of the bottom portion 7b of the housing 7 and the opposing lower end surface of the flange portion 2b, is increased by the dynamic pressure action of the dynamic pressure grooves.The pressure of these oil films then forms a first thrust bearing portion T1 and a second thrust bearing portion T2 that support the flange portion 2b (shaft member 2) in the thrust direction without contact.

[0038] An annular groove (not shown) and a plurality of radial grooves (not shown) provided on the inner diameter side of the annular groove are formed in the upper end surface 8d of the sintered oil-impregnated bearing 8. A plurality of axial grooves 8c1 are provided at equal intervals in the circumferential direction on the outer peripheral surface 8c of the sintered oil-impregnated bearing 8. The space on the outer diameter side of the flange portion 2b of the shaft member 2 communicates with the seal space via these axial grooves 8c1, the annular groove, the radial grooves, etc., thereby preventing the generation of negative pressure in this space.

[0039] The sintered oil-impregnated bearing 8 is manufactured through the steps shown in Figure 5. That is, this manufacturing process involves, in order, a powder compacting step S1, a sintering step S2, and a dynamic pressure groove forming step S3. The powder compacting step S1 is a step in which metal powder, such as Cu powder, Cu alloy powder, Fe powder, or Cu-coated Fe powder, is compressed into a cylindrical shape. The sintering step S2 is a step in which the powder compact obtained in the powder compacting step is sintered at a predetermined sintering temperature. The dynamic pressure groove forming step S3 is a step in which dynamic pressure grooves are formed on the inner diameter surface after the inner diameter surface has been sealed.

[0040] The powder compacting step S1 is preceded by a raw material powder mixing step in which raw material powder for the sintered oil-impregnated bearing 8 is prepared by mixing a plurality of types of powder.

[0041] If the grooves formed on the inner diameter surface 8A1 of the bearing hole 8A are not axial grooves but are herringbone-shaped or the like, when the core rod having a forming die formed on its outer peripheral surface is pulled out from the hole of the sintered compact in the dynamic pressure groove forming step, the hole of the sintered compact will be peeled off from the forming die of the core rod whose diameter has been expanded by spring back. Therefore, when the core rod is pulled out from the sintered compact, the ridges between the grooves will be scraped off to some extent.

[0042] However, when the sintered body has axial grooves G, the core rod can be pulled out of the sintered body in the axial direction without using springback. Moreover, the risk of the ridges H between the grooves being scraped off is reduced.

[0043] If the axial grooves G have a perfectly rectangular cross section, the oil film pressure will be high. Also, if the ridges H between the axial grooves are large, the contact area can be increased, and the contact surface pressure can be reduced.

[0044] For this reason, in this bearing, the rectangularity of the axial grooves calculated from the linearly developed shape of the circularity profile of the step surface measured using the least squares center method is 0.025 or less, and the flatness of the hills between circumferentially adjacent axial grooves is 0.80 or more. Here, the least squares center method involves finding a reference circle where the square of the error is smallest, and then finding the radii of the circumscribed and inscribed circles that are concentric with this reference circle. In this case, the circularity is calculated as (radius of the circumscribed circle) - (radius of the inscribed circle). Figure 8 shows the circularity profile measured using the least squares center method.

[0045] In this case, the rectangularity is calculated based on the slope of the groove side surfaces Ga and Gb (see FIG. 6) between a first portion (portion D) at 20% of the groove depth (dimension C) from the maximum groove depth (value at portion A) and a second portion (portion E) at 20% of the groove depth (dimension C) from the minimum groove depth (value at portion B), as shown in FIG. 1. Furthermore, the flatness is calculated based on the width dimension of the hill portion H at the second portion (portion E). The maximum groove depth (value at portion A) represents the circumscribed circle shown in FIG. 8, and the minimum groove depth (value at portion B) represents the inscribed circle.

[0046] Specifically, the rectangularity is calculated by taking W1 as the total length of the shape obtained by linearly expanding the circularity shape of the step surface measured using the least squares central method, W2 as the total groove width at the first portion of all grooves (portion D), and W3 as the total hill width of all hill portions at the second portion, and then calculating W1-(W2+W3) as the inclined portion data, and dividing the ratio of the inclined portion data to the total length by the number of grooves (four in this embodiment).

[0047] If the groove width of one groove G in the first region (region D) is w2, the total groove width W2 is (w2 + w2 + w2 + w2 = 4w2) since there are four grooves G in this case. Also, if the width of one hill portion H in the second region (region E) is w3, the total hill width W3 is (w3 + w3 + w3 + w3 = 4w3) since there are four hill portions H in this case.

[0048] Therefore, the inclined portion data, W1-(W2+W3), becomes W1-(4w2+4w3). If the groove width in the second portion (portion E) is w4, then the total groove W4, since there are four grooves G in this case, becomes (w4+w4+w4+w4=4w4). The groove width w4 in the second portion (portion E) becomes w5+w2+w6.

[0049] Therefore, w4 - w2 = w5 + w6. Here, w5 is the amount of inclination of one groove side surface Ga of groove G, and w6 is the amount of inclination of the other groove side surface Gb. As shown in FIG. 6 , the amount of inclination of one groove side surface Ga is w5 = L tan α, where α is the inclination angle of this groove side surface Ga (the inclination angle relative to the groove depth direction) and L is the dimension between the first portion (portion D) and the second portion (portion E). The amount of inclination of the other groove side surface Gb is w6 = L tan β, where β is the inclination angle of this groove side surface Gb (the inclination angle relative to the groove depth direction) and L is the dimension between the first portion (portion D) and the second portion (portion E).

[0050] The inclined portion data is W1-(4w2+4w3). Therefore, the degree of rectangularity is (W1-(4w2+4w3)) / W1. In this case, if it is a perfect rectangle, W1=(4w2+4w3), and the degree of rectangularity is 0. That is, in the present invention, the degree of rectangularity is a value (index) that approaches 0, becoming closer to a rectangle. Therefore, by making the degree of rectangularity 0.025 or less, it becomes closer to a perfect rectangle.

[0051] Furthermore, the total length of the linearly developed shape of the step surface roundness profile measured using the least squares center method is defined as W1, and the total width of all hill portions in the second region is defined as W3 (w3 + w3 + w3 + w3). The ratio of W3 to the total length W1 (W1 / W3) is calculated, and this calculated value is divided by the value expressed as hill-groove ratio / (hill-groove ratio + 1) to determine the flatness. This is because the hill-groove ratio affects this calculated value. The flatness is an index in which the closer the value is to 1, the more flat portions there are.

[0052] Here, the hill-groove ratio is the ratio of the hill width to the groove width on the center line (root mean square height Ra (see FIG. 6)). In this case, as shown in FIG. 6, when the width dimension of the hill is W10 and the width dimension of the groove is W11, the hill-groove ratio is W10 / W11.

[0053] For example, if the hill-groove ratio is 2, then as shown in Table 1 below, the hill-groove ratio + 1 results in a hill ratio of 2 (hill-groove ratio) and a groove ratio of 1, so hill-groove ratio + 1 = 3. Therefore, hill-groove ratio / (hill-groove ratio + 1) is 0.67. Furthermore, the calculated value (W1 / W3) (value assuming a perfect rectangle: 0.667) is calculated from the linearly developed coordinate data, and the ratio at this time is 1.0, and this 1.0 means that the flat portion is wider.

[0054] Table 1 also shows cases where the flatness is 1.0 when the hill groove ratio is 0.2, 0.5, 1.0, and 3.0, in addition to 2.0. When the hill groove ratio is 0.2, the hill groove ratio + 1 is 1.2, the hill groove ratio / (hill groove ratio + 1) is 0.17, the value assuming a perfect rectangle is 0.167, and the flatness is 1.0. When the hill groove ratio is 0.5, the hill groove ratio + 1 is 1.5, the hill groove ratio / (hill groove ratio + 1) is 0.33, the value assuming a perfect rectangle is 0.333, and the flatness is 1.0. When the hill groove ratio is 1.0, the hill groove ratio + 1 is 2, the hill groove ratio / (hill groove ratio + 1) is 0.50, the value assuming a perfect rectangle is 0.500, and the flatness is 1.0. When the hill-groove ratio is 3.0, the hill-groove ratio + 1 is 4.0, the hill-groove ratio / (hill-groove ratio + 1) is 0.75, the value assuming a perfect rectangle is 0.750, and the flatness is 1.0.

[0055] 7A and 7B show the linear development of the circularity profile of the step surface measured using the least squares central method. In FIG. 7A, the rectangularity is 0.019 and the flatness is 0.805. In FIG. 7B, the rectangularity is 0.041 and the flatness is 0.706. That is, in FIG. 7A, the rectangularity is less than 0.025 and the flatness is greater than 0.80, while in FIG. 7B, the rectangularity is greater than 0.025 and the flatness is less than 0.80. Therefore, the product in FIG. 7A is non-defective, while the product in FIG. 7B is defective.

[0056] It is preferable to set this hill-groove ratio to 0.1 to 1.0. If it is less than 0.1, the hill portion H is too small, making it difficult to obtain a stable rotational force, and conversely, if it exceeds 1.0, the groove G is too small, making it difficult to obtain a stable dynamic pressure. For this reason, in terms of flatness, the closer the flatness value is to 1, the more flat portions there are.

[0057] In the sintered oil-impregnated bearing of the present invention, the cross-sectional shape of the axial groove G approaches a rectangle, increasing the oil film pressure, and the flatness of the ridge portion H is high, resulting in a wide flat portion of the ridge portion, which increases the contact area and reduces contact. This allows for effective dynamic pressure effect. This makes it possible to numerically determine the ridge shape of a bearing (sintered oil-impregnated bearing) that excels in oil film formation. This makes it possible to provide a sintered oil-impregnated bearing that can effectively exert dynamic pressure effect as a product.

[0058] The rectangularity is calculated based on the inclination of the groove side surface between the first portion at 20% of the groove depth from the maximum groove depth and the second portion at 20% of the groove depth from the minimum groove depth, so that the rectangularity can be calculated stably.

[0059] The groove depth of the axial groove G is preferably 2 μm or more. By setting it in this way, dynamic pressure can be generated stably.

[0060] The hill-groove ratio is preferably 0.1 to 1.0. By setting it in this way, it is possible to generate dynamic pressure and ensure a sufficient sliding area, thereby obtaining stable rotational motion.

[0061] In particular, when the total length of the linearly developed shape of the step surface roundness profile measured by the least squares central method is defined as W1, the total groove width of all grooves in the first region is defined as W2, and the total width of all hills in the second region is defined as W3, W1 - (W2 + W3) is defined as the inclined portion data, the ratio of the inclined portion data to the total length is calculated, and the calculated value is divided by the number of grooves to obtain the rectangularity. By setting it in this way, the reliability of the calculated rectangularity is improved.

[0062] The total width of all the hillocks in the second region is defined as W3, and the ratio (W1 / W3) of this W3 to the total length W1 is calculated, and the value obtained by dividing this calculated value by the following formula can be used as the flatness. By setting the flatness in this way, the reliability of the calculated flatness is improved. The flatness is an index where 1 is the maximum, and the closer to 1 the contact area is, the larger the contact area becomes, which can be reduced and the more the dynamic pressure effect can be exerted.

[0063] Below (Figure 9 onwards), we will first explain an example of a motor (here, a fan motor) that uses a fluid dynamic bearing device as a component, and then explain a fluid dynamic bearing device that incorporates a dynamic bearing (a sintered oil-impregnated bearing) as a bearing member.

[0064] 9 is a longitudinal cross-sectional view conceptually illustrating an example of a small fan motor that sends cooling air (airflow) toward an object to be cooled. This fan motor includes a fluid dynamic bearing device 51, a motor base 55 that constitutes the stationary side of the motor, a rotor 53 fixed to a shaft member 52 of the fluid dynamic bearing device 51 that also functions as the motor's main shaft, impellers 4 attached to the rotor 53, and a stator coil 56a and a rotor magnet 76b arranged opposite each other with a radial gap between them. The stator coil 56a is attached to a housing 57 of the fluid dynamic bearing device 51, and the rotor magnet 56b is attached to the rotor 53. In this fan motor configuration, when current is applied to the stator coil 56a, electromagnetic force between the stator coil 56a and the rotor magnet 56b rotates the rotor magnet 56b, causing the shaft member 2 and the rotor 53 to rotate together. As the rotor 53 rotates, impellers 54 attached to the rotor 53 generate an axial airflow.

[0065] FIG. 10 is a longitudinal cross-sectional view conceptually illustrating a fluid dynamic bearing device 1 that can be incorporated into the fan motor shown in FIG. 9 , with components such as the shaft member 52 and bearing member 58 depicted in simplified form. The fluid dynamic bearing device 51 shown in FIG. 10 is a so-called rotating shaft type bearing device, primarily comprising the shaft member 52 constituting the rotating side, the housing 7, bearing member 58, and seal member 59 constituting the stationary side, and lubricating oil (not shown) as a lubricating fluid filled in the internal space of the housing 77. The rotating shaft member 2 is supported in a non-contact manner for radial rotation by two radial bearing portions R1 and R2 spaced apart in the axial direction. A fluid dynamic bearing according to an embodiment of the present invention can be used for the bearing member 58. For ease of explanation, the upper side of FIG. 10 , where the seal member 59 is located, will be referred to as the "upper side," and the lower side of FIG. 10 will be referred to as the "lower side." However, this is not intended to limit the orientation of the fluid dynamic bearing device 1 during use.

[0066] The shaft member 52 is made of a highly rigid metal material such as stainless steel, and its outer peripheral surface 52a is formed into a smooth cylindrical surface without any irregularities, while its lower end surface 52b is formed into a convex spherical surface. A rotor 53, to which blades 54 and a rotor magnet 56b (see FIG. 9) are attached, is fixed to the upper end of the shaft member 52.

[0067] The housing 57 is cylindrical and has a cylindrical portion 57a ​​and a bottom portion 57b that closes the opening at the lower end of the cylindrical portion 57a. In the illustrated example, the cylindrical portion 57a ​​and the bottom portion 57b are integrally formed from resin or metal. The inner peripheral surface 57a1 of the cylindrical portion 57a ​​is formed into a cylindrical surface with a constant diameter, and the outer diameter end of the lower end is connected to the outer peripheral surface 57a2 of the cylindrical portion 57a. The stator coil 56a and the motor base 55, which are components of the fan motor, are fixed to the outer peripheral surface 57a2 of the cylindrical portion 57a.

[0068] In this embodiment, a thrust plate 57c made of a material with better sliding properties than the material of the housing 57 is placed on the inner bottom surface 57b1 of the housing 57, and the upper end surface of the thrust plate 57c contacts and supports the lower end surface 52b of the shaft member 52 (contact supports the shaft member 52 in the thrust direction). However, the thrust plate 57c is not necessarily provided and may be omitted. If the thrust plate 57c is omitted, the lower end surface 52b of the shaft member 52 is contact-supported by the inner bottom surface 57b1 of the housing 57.

[0069] The seal member 59 is formed in an annular shape from a resin or metal material, and its inner peripheral surface 59a forms an annular seal space S with the outer peripheral surface 52a of the opposing shaft member 52. This seal space S prevents as much as possible the lubricating oil filled in the housing 57 from leaking out of the device.

[0070] The fluid dynamic bearing device 51 may be used with the entire interior space of the housing 57 filled with lubricating oil, or with lubricating oil present in a portion of the interior space of the housing 57 (lubricating oil and air mixed in the interior space of the housing 57). In the former case, the volume of the seal space S is determined so that the lubricating oil level will always be within the axial range of the seal space S, even if the lubricating oil level position fluctuates in the axial direction due to temperature changes.

[0071] The bearing member 58 is fixed to the inner periphery of the cylindrical portion 57a ​​of the housing 57 with its lower end surface abutting against the shoulder surface 57b2 of the housing 57. The bearing member 58 can be fixed to the inner periphery 57a1 of the cylindrical portion 57a ​​by press-fitting, adhesive bonding, or press-fit adhesive bonding (a combination of press-fitting and adhesive bonding), or it can also be fixed to the inner periphery of the cylindrical portion 57a ​​by a clearance fit (see JIS B 0401-1) to the inner periphery of the cylindrical portion 57a ​​and then sandwiched from both axial sides between the seal member 59 and the shoulder surface 57b2 of the housing 57. In particular, the latter fixing method allows the bearing member 58 to be fixed to the housing 57 at the same time as the seal member 559 is fixed to the housing 57, thereby reducing the effort required to assemble the components. Furthermore, for example, if bearing member 58 is press-fitted into the inner periphery of cylindrical portion 57a ​​with a large interference, deformation of bearing member 58 caused by the press-fitting will spread to inner periphery 58a of bearing member 58, which may have an adverse effect on the accuracy of the radial bearing gap and, ultimately, the bearing performance of radial bearing portions R1, R2. In contrast, the method of sandwiching bearing member 58 between seal member 59 and housing 57 from both axial sides can prevent such problems from occurring.

[0072] The bearing member 58 is formed into a cylindrical shape from a porous sintered metal body whose main components are, for example, copper and iron, and is fixed to the inner periphery of the cylindrical portion 57a ​​of the housing 57 in an oil-impregnated state with its internal pores impregnated with lubricating oil. The size of the bearing member 58 is, for example, an inner diameter of 4 mm or less, an outer diameter of 10 mm or less, and an axial dimension of 15 mm or less.

[0073] Fig. 11 is a longitudinal cross-sectional view of a hydrodynamic bearing according to one embodiment of the present invention, which is used as the bearing member 58 in the fluid dynamic bearing device 51 shown in Fig. 10. The drawing on the left side of Fig. 12 is an expanded plan view of the inner circumferential surface of the bearing member 58 shown in Fig. 11, and the drawing on the right side of Fig. 12 is a cross-sectional view taken along line X-X in the drawing on the left side. As shown in Figs. 11 and 12, the inner circumferential surface 58a of the bearing member 58 is provided with two radial bearing surfaces A1 and A2 spaced apart in the axial direction, which form a radial bearing gap between the inner circumferential surface 52a of the shaft member 52 and the outer circumferential surface 52a of the shaft member 52. In the illustrated bearing member 88, the radial bearing surface A1 is provided at the upper end of the inner circumferential surface 58a, and the radial bearing surface A2 is provided at the lower end of the inner circumferential surface 58a.

[0074] Between the two radial bearing surfaces A1, A2, on the inner peripheral surface 58a of the bearing member 58, there is provided a large-diameter cylindrical surface 58b that is larger in diameter than the radial bearing surfaces A1, A2 and has a cylindrical shape with a constant diameter and no irregularities. By providing such a large-diameter cylindrical surface 58b between the two radial bearing surfaces A1, A2, it is possible to reduce the torque when the shaft member 2 rotates while ensuring the bearing rigidity required for the entire bearing member 58. In short, the large-diameter cylindrical surface 58b forms a so-called "middle relief portion."

[0075] The upper (one) radial bearing surface A1 provided on the inner peripheral surface 58a of the bearing member 58 and the large-diameter cylindrical surface 58b are connected via a tapered surface 58c inclined at an angle θ1 with respect to the axial direction, and the lower (other) radial bearing surface A2 and the large-diameter cylindrical surface 58b are connected via a tapered surface 58d inclined at an angle θ2 with respect to the axial direction. There is a magnitude relationship between the above angles θ1 and θ2, θ1 > θ2. In other words, a steep slope is a slope with a larger inclination angle than a gentle slope, and the tapered surfaces 58c and 58d respectively constitute the "steep slope" and "gentle slope" as defined in the present invention.

[0076] Each of the radial bearing surfaces A1, A2 is formed with a dynamic pressure generating portion P for generating a dynamic pressure action on the lubricating oil in the radial bearing gap. Each of the dynamic pressure generating portions P has an uneven surface with a plurality of axially extending dynamic pressure grooves 60 and convex ridge portions (regions indicated by cross-hatching in FIGS. 11 and 12) 61 alternately arranged in the circumferential direction. Therefore, the bearing member 58 of this embodiment is a type of the "step-type dynamic pressure bearing" described above. The height difference h between the inner diameter surface (inner diameter end surface) 61a of the ridge portion 61 and the groove bottom surface 60a of the dynamic pressure groove 60 (groove depth of the dynamic pressure groove 60) is set to 2 μm or more.

[0077] In the dynamic pressure generating portion P provided in the upper radial bearing surface A1, each dynamic pressure groove 60 is formed so that its lower end (the other axial side) opens onto the tapered surface 58c, which serves as a "steep slope," while each hill portion 61 is formed so that its lower end terminates at the upper end of the tapered surface 58c (the boundary between the radial bearing surface A1 and the tapered surface 58c). In the dynamic pressure generating portion P provided in the lower radial bearing surface A2, each dynamic pressure groove 60 is formed so that its upper end (one axial side) opens onto the tapered surface 58d, which serves as a "gentle slope," while each hill portion 61's upper end terminates at the lower end of the tapered surface 58d (the boundary between the radial bearing surface A2 and the tapered surface 58d). Therefore, in both the upper and lower dynamic pressure generating portions P, the dynamic pressure grooves 60 are longer in the axial direction than the hill portions 61.

[0078] In the fluid dynamic bearing device 51 having the above configuration, when the shaft member 52 rotates about its central axis, a radial bearing gap, with lubricating oil present, is formed between the opposing outer peripheral surface 52a of the shaft member 52 and the radial bearing surfaces A1, A2 of the bearing member 58. Furthermore, due to the generation of pressure accompanying the rotation of the shaft member 52 and the thermal expansion of the lubricating oil, the lubricating oil impregnated in the internal pores of the bearing member 58 gradually seeps out through the surface perforations of the bearing member 58 to the outside of the bearing member 58 and is drawn into the radial bearing gap.

[0079] In this embodiment, the dynamic pressure generating portions P (first radial dynamic pressure generating portion P1, second radial dynamic pressure generating portion P2) provided on the radial bearing surfaces A1, A2 are formed as uneven surfaces with a plurality of axially extending dynamic pressure grooves 60 and convex ridges 61 alternately arranged in the circumferential direction, and as the shaft member 52 rotates, the lubricating oil flowing circumferentially within the radial bearing gap (strictly speaking, the arc-shaped gap within the radial bearing gap defined by the dynamic pressure grooves 60) collides with the ridges 61 and is forced into the opposing regions of the inner diameter end faces 61 a of the ridges 61 (the narrowest portions of the radial bearing gap), thereby increasing the pressure (rigidity) of the oil film within the radial bearing gap. The dynamic pressure action of these dynamic pressure generating portions P forms two radial bearing portions R1, R2 spaced axially apart that support the shaft member 52 in a non-contact manner for radial rotatability.

[0080] Furthermore, when the shaft member 52 rotates, the lower end surface 52b of the shaft member 52 is contact-supported by the upper end surface of a thrust plate 57c placed on the bottom portion 57b of the housing 57. This forms a thrust bearing portion T that supports (contact-supports) the shaft member 52 in the thrust direction.

[0081] In the bearing member 58 of the present embodiment described above, one (upper) radial bearing surface A1 and the large-diameter cylindrical surface 58b are connected via a tapered surface 58c that is a steeply inclined surface inclined with respect to the axial direction, and the other (lower) radial bearing surface A2 and the large-diameter cylindrical surface 58b are connected via a tapered surface 58d that is a gently inclined surface whose inclination angle with respect to the axial direction is smaller than that of the steeply inclined surface, and the dynamic pressure grooves 60 of the dynamic pressure generating portion P provided in the lower radial bearing surface A2 have their upper ends open to the tapered surface 58d that is a gently inclined surface. With this configuration, in the radial bearing surface A2, the dynamic pressure grooves 60 can be made longer in the axial direction than the convex hill portions 61 that constitute the dynamic pressure generating portion P provided therein, without making the convex hill portions 61 longer in the axial and / or circumferential directions. In this case, the amount of lubricating oil flowing within the radial bearing gap when the shaft member 52 rotates, and therefore the amount of lubricating oil pushed toward the land portion 61, can be increased by the amount that the dynamic pressure grooves 60 are elongated in the axial direction. Therefore, the oil film formation in the radial bearing gap of the radial bearing portion R2 can be improved, and the bearing rigidity of the radial bearing portion R2 can be increased.

[0082] In the bearing member 58 of this embodiment, the lower ends of the dynamic pressure grooves 60 provided in the upper radial bearing surface A1 open to the tapered surface 58c as a steeply inclined surface, and therefore, in the radial bearing surface A1, the dynamic pressure grooves 60 are (slightly) longer in the axial direction than the land portions 61. In this case, even in the radial bearing gap of the radial bearing portion R1, although not as large as in the radial bearing gap of the radial bearing portion R2, the flow rate of the lubricating oil, and therefore the amount of lubricating oil pushed toward the land portions 61, can be increased, and therefore the oil film formation in the narrowest portion of the radial bearing gap of the radial bearing portion R1 can be improved, and the bearing rigidity of the radial bearing portion R1 can also be increased.

[0083] Incidentally, the deeper the groove depth h of the dynamic pressure grooves 60 that constitute the dynamic pressure generating portion P, the greater the amount of lubricating oil that can be present in the radial bearing gap, which is advantageous in preventing oil film breakdown in the radial bearing gap, but the greater the amount of lubricating oil that is present in the radial bearing gap, the greater the viscous resistance of the lubricating oil, which reduces the flow rate of the lubricating oil when it flows circumferentially in the radial bearing gap as the shaft member 52 rotates, and this may make it difficult to improve oil film formation in the narrowest part of the radial bearing gap. Therefore, from the perspective of increasing the bearing rigidity of the radial bearing portions R1, R2, it is preferable to set the groove depth h of the dynamic pressure grooves 60 as small as possible within the range that prevents oil film breakdown in the radial bearing gap, and specifically to set it to 20 μm or less, more preferably 10 μm or less, and even more preferably 5 μm or less.

[0084] As shown in FIG. 12 , when the groove width (circumferential dimension) of the dynamic pressure groove 60 is Z1 and the circumferential dimension of the ridge portion 61 is Z2, the ratio of Z2 to Z1 (= Z2 / Z1) is 1 or less, preferably in the range of 0.3 to 0.7. In other words, the circumferential dimension Z2 of the ridge portion 61 is preferably set to be smaller than the circumferential dimension Z1 of the dynamic pressure groove 60 by a predetermined amount. The dynamic pressure generating portion P of this embodiment has five dynamic pressure grooves 60 and five ridge portions 61, and the ratio (= Z2 / Z1) is 0.5. This allows the lubricating oil in the radial bearing gap (within the arc-shaped gap of the radial bearing gap defined by the dynamic pressure groove 60) to move significantly in the circumferential direction as the shaft member 52 rotates. This is therefore advantageous for improving oil film formation in the radial bearing gap and increasing the bearing rigidity of the radial bearing portions R1 and R2. The number of the dynamic pressure grooves 60 and the ratio described above can be changed as desired depending on the required characteristics, etc.

[0085] The bearing member 8 made of the sintered metal body having the above-mentioned structure is obtained by undergoing a compression molding process, a sintering process, a sizing process, and an oil impregnation process in this order. Each of the above processes will be described below.

[0086] [Compression Molding Step] In this step, raw material powder filled in the cavity of a compression molding die (details of which will be described later) is compression molded to obtain a green compact 68 as shown in FIGS. 14 and 15, which is the base material of a bearing member 58 as a hydrodynamic bearing.

[0087] 13 and 14 are schematic vertical cross-sectional views of a compression molding die 70. The compression molding die 70 has a cylindrical die 71 that shapes the outer peripheral surface of the powder compact 18, a core pin 72 that shapes the inner peripheral surface of the powder compact 68, a lower punch 73 that shapes the lower end surface of the powder compact 68, including the lower outer peripheral chamfer and lower inner peripheral chamfer of the powder compact 68, and an upper punch 74 that shapes the upper end surface of the powder compact 68, including the upper outer peripheral chamfer and upper inner peripheral chamfer of the powder compact 68, and the core pin 72, lower punch 23, and upper punch 74 are provided so as to be able to move up and down relative to the die 71.

[0088] The die 71 has a small-diameter inner circumferential surface 71a, a large-diameter inner circumferential surface 71b provided above the small-diameter inner circumferential surface 71a, and a tapered inner circumferential surface 71c connecting the inner circumferential surfaces 71a, 71b. The core pin 72 has a stepped shaft shape, with a large-diameter outer circumferential surface 72a, a small-diameter outer circumferential surface 72b provided above the large-diameter outer circumferential surface 72a, and a tapered outer circumferential surface 72c connecting the outer circumferential surfaces 72a, 72b. The small-diameter outer circumferential surface 72b is provided with a molding die portion 72d corresponding to the shape of the dynamic pressure generating portion P having an uneven surface to be provided on the radial bearing surface A1 of the bearing member 58. The molding die portion 72d has a convex shape 72e that molds the dynamic pressure grooves 10.

[0089] 13, in the compression molding die 70 having the above-described configuration, a core pin 72 and a lower punch 73 are first arranged on the inner periphery of a die 71 to form a cavity 75 to be filled with raw material powder, and raw material powder M is then filled into this cavity 75. The raw material powder M is, for example, a metal powder mainly composed of copper or iron, to which graphite, solid lubricant powder such as molybdenum disulfide, low-melting-point metal powder such as tin powder or zinc powder, and the like are appropriately added or mixed. When forming the cavity 75, the core pin 72 is arranged so that its small-diameter outer peripheral surface 72b and tapered outer peripheral surface 72c are located radially inside the large-diameter inner periphery 71b of the die 71.

[0090] As shown in Fig. 14 , after the raw material powder M is filled into the cavity 75, the upper punch 24 is lowered to compress the raw material powder M in the axial direction and form a cylindrical green compact 18. Thereafter, as shown in Fig. 15 , the core pin 72, the lower punch 73, and the upper punch 74 are raised together to eject the green compact 68 to the outside (upper side) of the die 71, and then the upper punch 74 is raised (releasing the axial restraining force exerted by the lower punch 73 and the upper punch 74), and the core pin 72 is removed from the inner periphery of the green compact 68. The inner and outer peripheral shapes of the green compact 68 obtained in this manner follow the outer peripheral shapes of the core pin 72 and the inner peripheral shapes of the die 71, respectively.

[0091] That is, the inner peripheral surface of the powder compact 68 is provided with a small-diameter inner peripheral surface 68a, a large-diameter inner peripheral surface 68b, and a tapered inner peripheral surface 68c that respectively follow the shapes of the small-diameter outer peripheral surface 72b, the large-diameter outer peripheral surface 72b, and the tapered outer peripheral surface 72c of the core pin 72, and the outer peripheral surface of the powder compact 68 is provided with a large-diameter outer peripheral surface 68e1, a small-diameter outer peripheral surface 68e2, and a tapered surface connecting both outer peripheral surfaces 68e1, 68e2 that respectively follow the shapes of the large-diameter inner peripheral surface 71b, the small-diameter inner peripheral surface 71a, and the tapered inner peripheral surface 71c of the die 71. The small-diameter inner peripheral surface 68a of the powder compact 68 is provided with a dynamic pressure generating portion P (with irregularities corresponding to the shape of) molded by the molding die portion 71d of the core pin 71.

[0092] [Sintering Process] In this sintering process, the green compact 68 obtained in the compression molding process is heated for a predetermined time, for example, at a temperature equal to or higher than the sintering temperature of the main component metal, thereby obtaining a sintered body 68′ (see FIGS. 16A and 16B ) in which the metal powders constituting the green compact 68 are neck-bonded together.

[0093] [Sizing Step] In this sizing step, a portion of the sintered body 68' is plastically deformed using a sizing die described below, thereby finishing the sintered body 68' into its final shape (the shape of the bearing member 58 shown in FIG. 11).

[0094] 16A and 16B show schematic vertical cross-sectional views of the sizing die 30. The sizing die 80 includes a cylindrical die 81, a core pin 82 that is provided so as to be able to move up and down relative to the die 81, a lower punch 83, and an upper punch 84.

[0095] A forming die 82a having an uneven surface corresponding to the shape of the dynamic pressure generating portion P to be provided on the radial bearing surface A2 of the bearing member 58 is provided at a predetermined axial position on the outer peripheral surface of the core pin 82. That is, the forming die 82a is configured by circumferentially arranging convex portions 32b extending in the axial direction and spaced apart to form the dynamic pressure grooves 60 of the dynamic pressure generating portion P. The outer peripheral surface of the core pin 82, except for the forming die 82a, is a smooth cylindrical surface without any unevenness, and is formed with a diameter smaller than the small-diameter inner peripheral surface 68a of the sintered body 68′ by a predetermined amount so that, when the sintered body 68′ is subjected to a sizing process, the inner peripheral surface of the sintered body 68′, except for the region to be formed of the radial bearing surface A2 (a part of the large-diameter inner peripheral surface 68b), is in non-contact or light contact to the extent that it does not deform.

[0096] The inner surface of the die 81 is provided with a large-diameter inner surface 81a formed into a cylindrical surface of a constant diameter, a cylindrical constriction surface 81b formed with a smaller diameter than the large-diameter inner surface 81a and constricting the lower end of the small-diameter outer surface 68e2 of the sintered body 68', and a tapered inner surface 81c connecting the large-diameter inner surface 81a and the constriction surface 81b.

[0097] 16A, in the sizing die 80 having the above-described configuration, the sintered body 68' is first placed between the die 81 and the core pin 82. At this time, the sintered body 68' is positioned so that the lower region of the large-diameter inner circumferential surface 68b, which is the region intended to form the radial bearing surface A2, faces the forming die 82a of the core pin 82.

[0098] Then, while maintaining the relative axial positional relationship between the sintered body 68' and the core pin 82, the core pin 82, the lower punch 83, and the upper punch 84 are moved downward together, and (the lower region of) the small-diameter outer peripheral surface 68e2 of the sintered body 68' is guided by the tapered inner peripheral surface 81c of the die 81 and gradually pressed into the drawing surface 81b of the die 81. As a result, the lower region of the small-diameter outer peripheral surface 68e2 of the sintered body 68' is drawn by the drawing surface 81b of the die 81 (receives a radially inward compressive force), so that the vicinity of the lower end of the sintered body 68' is entirely reduced in diameter and deformed, and the lower region of the large-diameter inner peripheral surface 68b of the sintered body 68' (the region to be molded for the radial bearing surface A2) is pressed against the molding die 82a of the core pin 82 [see FIG. 16B ]. As a result, a radial bearing surface A2 having substantially the same diameter as the radial bearing surface A1 and having a dynamic pressure generating portion P is formed in the lower region of the inner peripheral surface of the sintered body 68'.

[0099] The dynamic pressure generating portion P provided on the radial bearing surfaces A1, A2 has a simple shape consisting of an uneven surface with multiple axially extending dynamic pressure grooves 60 and convex ridges 61 alternately arranged in the circumferential direction, so the required precision can be ensured simply by molding using the sizing die 80 (and the aforementioned compression molding die 70). In contrast, with a "herringbone-type dynamic pressure bearing" in which the ridges that make up the dynamic pressure generating portion have a herringbone shape, the required precision cannot be ensured by molding alone, and separate finishing processing (such as rotational sizing) is essential after molding (after the sizing process). For this reason, the bearing member 8 of this embodiment, which is a "step-type dynamic pressure bearing," requires few man-hours and processes until completion, and can be manufactured at low cost.

[0100] At the same time that the radial bearing surface A2 is formed on the inner peripheral surface of the sintered body 68', a gently sloping tapered surface 58d (see Figure 12) is formed on the portion of the inner peripheral surface of the sintered body 68' adjacent to the upper side of the radial bearing surface A2, and the lower region of the small-diameter outer peripheral surface 68e2 of the sintered body 68', which is narrowed by the narrowing surface 81b of the die 81, is formed into the small-diameter outer peripheral surface 58e3 of the bearing member 58.

[0101] By performing the sizing process on (a part of) the sintered compact 68' as described above, the radial bearing surface A2 having the dynamic pressure generating portion P and the tapered surface 58d are formed on the large-diameter inner peripheral surface 68b of the sintered compact 68'. During this process, the axial region of the outer peripheral surface of the sintered compact 68', excluding the lower region of the small-diameter outer peripheral surface 68e2, is not subjected to a compressive force in the diameter-reducing direction from the die 81 and is therefore not deformed. Therefore, the axial central region of the large-diameter inner peripheral surface 68b of the sintered compact 68' becomes the large-diameter cylindrical surface 58b of the bearing member 58. Furthermore, the upper region of the small-diameter outer peripheral surface 68e2 of the sintered compact 68' that is not drawn down by the drawing surface 81b of the die 81 becomes the medium-diameter outer peripheral surface 58e2 provided on the outer peripheral surface 58e of the bearing member 58.

[0102] In this step, lubricating oil is impregnated into the internal pores of the sintered body 68′ (bearing member 8) that has been finished to its final shape in the sizing step by a known method such as vacuum impregnation, thereby obtaining an oil-impregnated bearing member 58 (dynamic bearing) that can be incorporated into a fluid dynamic bearing device 51.

[0103] As in the embodiment described above, the surface open area ratio of the radial bearing surface A1 (radial bearing surface A1 having the dynamic pressure generating portion P) formed by compression molding rather than sizing is likely to be greater than the surface open area ratio of the radial bearing surface A2 formed by sizing, and as a result, of the two radial bearing portions R1, R2, the bearing rigidity of radial bearing portion R1 may be lower than the bearing rigidity of radial bearing portion R2. Therefore, although the bearing performance of the bearing member 8 of this embodiment may be somewhat inferior to that in a case where both radial bearing surfaces A1, A2 are formed by sizing, when supporting the main shaft of a fan motor (shaft member 52), it does not matter if the required bearing performance is somewhat lower than when supporting the main shaft of a spindle motor for an HDD, for example. If there is a difference in bearing rigidity between the radial bearing portions R1 and R2, it is preferable to position the radial bearing portion with superior bearing rigidity (here, radial bearing portion R2) near the center of gravity of the motor's rotating body (in this embodiment, the entire component that rotates integrally with the rotating shaft member 52, such as the shaft member 52 and rotor 53).

[0104] The dynamic pressure bearing and its manufacturing method according to the embodiment of the present invention, as well as the fluid dynamic pressure bearing device 1 that uses this dynamic pressure bearing as the bearing member 8, have been described above, but the embodiments of the present invention are not limited to these.

[0105] 11 shows a bearing member 58 as a dynamic pressure bearing, in which the radial bearing surface A1 of the two radial bearing surfaces A1 and A2 having dynamic pressure generating portions P is formed by compression molding and the other radial bearing surface A2 is formed by sizing, but both radial bearing surfaces A1 and A2 may be formed by sizing. In this case, the compression molding step uses a compression molding die (not shown) having a core pin 22 that does not have a molding die portion 72d corresponding to the shape of the dynamic pressure generating portion P on its outer surface, while the sizing step uses a sizing die 80' having a core pin 82 with a molding die 82a on its outer surface that can mold the dynamic pressure generating portions P on both radial bearing surfaces A1 and A2, as shown schematically in FIGS. In the sizing die 80′ shown in FIGS. 17A and 17B, the convex portions 82b for forming the dynamic pressure grooves 60 are longer in the axial direction than the convex portions 82b provided in the sizing die 80 shown in FIGS. 16A and 16B.

[0106] 10, the thrust bearing portion T that supports the shaft member 52 in the thrust direction is configured as a pivot bearing, but the thrust bearing portion T can also be configured as a so-called dynamic pressure bearing. Although not shown in the figures, such a configuration can be realized, for example, by forming the lower end surface of the shaft member 2 into a flat surface that is perpendicular to the axial direction, and providing a dynamic pressure generating portion such as a dynamic pressure groove in the lower end surface of the shaft member 52 or in the inner bottom surface 57b1 of the housing 57 that faces this.

[0107] Furthermore, in the fluid dynamic bearing device 51 described above, the shaft member 2 constitutes the rotating side and the bearing member 58 constitutes the stationary side, but conversely, there are also cases where the shaft member 52 constitutes the stationary side and the bearing member 58 constitutes the rotating side. In other words, the bearing member 58 made of a dynamic bearing according to an embodiment of the present invention can be incorporated and used not only in a so-called rotating-shaft type fluid dynamic bearing device 51, but also in a so-called fixed-shaft type fluid dynamic bearing device 51.

[0108] Furthermore, the fluid dynamic bearing device 51 can be used not only for fan motors such as the one shown in Figure 9, but also as a bearing device for other small motors such as spindle motors for HDDs and polygon scanner motors for LBPs. When the fluid dynamic bearing device 51 is used in a spindle motor, a rotor holding one or more recording disks is provided so as to rotate integrally with the rotating member of the fluid dynamic bearing device 51 (e.g., shaft member 52), and when the fluid dynamic bearing device 1 is used in a polygon scanner motor, a rotor holding a polygon mirror is provided so as to rotate integrally with the rotating member of the fluid dynamic bearing device 1.

[0109] 9 to 17B, the rectangularity of the axial grooves G calculated from the linearly developed shape of the step surface circularity profile measured using the least squares central method is 0.025 or less, and the flatness of the hills H between circumferentially adjacent axial grooves is 0.80 or more, and the rectangularity is calculated based on the inclination of the groove side surface between a first portion D at 20% of the groove depth from the circumscribing circle and a second portion E at 20% of the groove depth from the inscribing circle. Furthermore, the rectangularity is an index indicating that the closer the value is to 0, the closer the shape is to a rectangle. The flatness is calculated based on the ratio of the total width of the hills at the second portion E to the overall length of the linearly developed shape, and the flatness is an index indicating that the closer the value is to 1, the more flat the grooves.

[0110] Therefore, even with the sintered oil-impregnated bearings shown in Figures 9 to 17B, the following effect can be achieved: "It is possible to numerically determine the hill shape of a bearing (sintered oil-impregnated bearing) that is excellent in oil film formation.As a result, it is possible to provide a sintered oil-impregnated bearing that can effectively exert dynamic pressure effects as a product."

[0111] Although the embodiments of the present invention have been described above, the present invention is not limited to the above embodiments and various modifications are possible. In the embodiments, the first radial dynamic pressure generating portion 8a and the second radial dynamic pressure generating portion 8b are formed at two locations axially separated on the inner diameter surface 8A1 of the bearing hole 8A of the sintered oil-impregnated bearing 8, but the inner diameter surface of the sintered oil-impregnated bearing 8 may be provided with a single radial dynamic pressure generating portion. That is, in the sintered oil-impregnated bearing 8 shown in FIG. 4, the axial groove G1 and the axial groove G2 are separated, but the axial groove G1 and the axial groove G2 may be continuous, as indicated by the two-dot chain line in FIG. 4. Furthermore, although the number of axial grooves in one radial dynamic pressure generating portion is four in the embodiments, this is not limited to four and the number of axial grooves may be increased or decreased as desired. In the above embodiment, the sintered oil-impregnated bearing 8 is fixed and the shaft member 2 rotates, but this is not limited to this. It is also possible to adopt a configuration in which the shaft member 2 is fixed and the sintered oil-impregnated bearing 8 rotates, or a configuration in which both the shaft member 2 and the sintered oil-impregnated bearing 8 rotate.

[0112] The rectangularity and flatness are averages of the multiple axial grooves and the multiple hills, and therefore, it is not necessary for all of the multiple grooves to have a rectangularity of 0.025 or less and a flatness of 0.80 or more. Even if one or more grooves fall outside these specifications, it is sufficient that the average rectangularity is 0.025 or less and the average flatness is 0.80 or more.

[0113] The present invention is not limited to spindle motors used in HDD disk drive devices, but can also be widely used in other small motors such as spindle motors incorporated in other information devices, polygon scanner motors in laser beam printers, color wheels in projectors, and cooling fan motors.

[0114] DESCRIPTION OF SYMBOLS 2 Shaft member 8a Radial dynamic pressure generating portion 8b Radial dynamic pressure generating portion G Axial groove H Hill portion D First portion E Second portion 51 Fluid dynamic bearing device 52 Shaft member 58 Bearing member (dynamic pressure bearing) 58a Inner peripheral surface 58b Large diameter cylindrical surface 58c Tapered surface (steep slope) 58d Tapered surface (gentle slope) 58e Outer peripheral surface 60 Dynamic pressure groove 61 Hill portion 68 Powder compact 68' Sintered body 70 Compression molding die 80 Sizing die A1, A2 Radial bearing surface P Dynamic pressure generating portion R1, R2 Radial bearing portion T Thrust bearing portion Z1 Circumferential dimension of dynamic pressure groove Z2 Circumferential dimension of hill portion

Claims

1. A sintered oil-impregnated bearing in which a shaft member is inserted into a bearing hole, comprising a radial dynamic pressure generating section formed on the inner diameter surface of the bearing hole by forming a step surface having axial grooves with rectangular cross sections arranged at a predetermined pitch along the circumferential direction, wherein the rectangularity of the axial grooves calculated from the linearly expanded shape of the circularity shape of the step surface measured using the least squares central method is 0.025 or less, and the flatness of the hills between circumferentially adjacent axial grooves is 0.80 or more, wherein the rectangularity is calculated based on the inclination of the groove side surface between a first portion that is 20% of the groove depth from the circumscribing circle and a second portion that is 20% of the groove depth from the inscribing circle, and the rectangularity is an index of the closer the value is to a rectangle, and the flatness is calculated based on the ratio of the total width dimension of the hills at the second portion to the total length of the linearly expanded shape, and the flatness is an index of the number of flat portions increasing as the value approaches 1.

2. A sintered oil-impregnated bearing as described in claim 1, characterized in that when the total length of the shape of the step surface measured by the least squares central method when linearly expanded is defined as W1, the total groove width of all grooves in the first portion is defined as W2, and the total width of all hill portions in the second portion is defined as W3, W1 - (W2 + W3) is defined as the inclined portion data, the ratio of the inclined portion data to the total length is calculated, and the calculated value divided by the number of grooves is the rectangularity.

3. The sintered oil-impregnated bearing of claim 1, characterized in that the total length of the shape of the step surface measured by the least squares central method when linearly expanded is defined as W1, the total width of all hill portions in the second portion is defined as W3, the ratio (W1 / W3) of this W3 to the total length W1 is calculated, and the calculated value is divided by the value obtained by hill-groove ratio / (hill-groove ratio + 1) to obtain the flatness.

4. A sintered oil-impregnated bearing according to any one of claims 1 to 3, characterized in that the groove depth of the axial grooves is 2 μm or more.

5. A sintered oil-impregnated bearing according to claim 3, wherein the hill-groove ratio is a hill-groove ratio on the squared center line of the axial grooves, and this hill-groove ratio is 0.1 to 1.

0.

6. A sintered oil-impregnated bearing according to claim 1, characterized in that two radial bearing surfaces are provided axially spaced apart to form a radial bearing gap between them and the outer peripheral surface of the shaft to be supported, and that an inner peripheral surface is provided between the two radial bearing surfaces and has a large-diameter cylindrical surface with a diameter larger than that of the radial bearing surfaces, and each radial bearing surface is provided with a radial dynamic pressure generating section for generating a dynamic pressure action on the fluid present in the radial bearing gap, and this radial dynamic pressure generating section consists of a plurality of axial grooves extending in the axial direction and a concave-convex surface with dynamic pressure grooves and convex hills arranged alternately in the circumferential direction, one of the radial bearing surfaces and the large-diameter cylindrical surface is connected via a steeply inclined surface inclined with respect to the axial direction, and the other radial bearing surface and the large-diameter cylindrical surface are connected via a gently inclined surface whose inclination angle with respect to the axial direction is smaller than that of the steeply inclined surface, and one axial end of the dynamic pressure groove provided in the other radial bearing surface opens onto the gently inclined surface.

7. A sintered oil-impregnated bearing as described in claim 6, wherein when the radial dynamic pressure generating portion is developed on a plane, the ratio (= Z2 / Z1) of the circumferential dimension Z2 of the hill portion of the step surface to the circumferential dimension Z1 of the dynamic pressure groove is 1 or less.

8. A sintered oil-impregnated bearing according to claim 6 or 7, wherein the difference in height between the inner diameter end face of the hill portion of said step surface and the groove bottom face of said hydrodynamic groove is 2 μm or more.

9. The sintered oil-impregnated bearing according to claim 6, which is made of a porous sintered metal body, and the other radial dynamic pressure generating portion is a formed surface formed by subjecting the sintered metal body to plastic working.

10. A sintered oil-impregnated bearing according to claim 6, comprising: a bearing member made of a porous metal sintered body; a housing that houses the bearing member on its inner periphery; lubricating oil as a fluid present in the radial bearing gap between the bearing member and the housing; and a radial bearing portion that supports a shaft member inserted into the bearing member in a non-contact manner in the radial direction by an oil film of the lubricating oil formed in the radial bearing gap.

Images