Dynamic pressure bearing, fluid dynamic pressure bearing device, and motor

Abstract

In the present invention, a dynamic pressure generation part has a plurality of dynamic pressure grooves arranged in a herringbone pattern, and the dynamic pressure grooves include first and second groups of dynamic pressure grooves. Each of the first and second groups of dynamic pressure grooves is provided with inclined ridge parts having an inclined shape formed between the dynamic pressure grooves. An annular ridge part connected to each inclined ridge part and extending in the circumferential direction is provided between the first and second groups of dynamic pressure grooves. A recess for generating dynamic pressure is formed in a partial area of the annular ridge part.

Description

Dynamic pressure bearing, fluid dynamic pressure bearing device, and motor

[0001] The present invention relates to a dynamic pressure bearing, a fluid dynamic pressure bearing device, and a motor.

[0002] A fluid dynamic bearing device supports a shaft member in a non-contact manner, allowing relative rotation, by means of pressure generated in a fluid film (e.g., an oil film) in a radial bearing gap between the outer peripheral surface of the shaft member and the inner peripheral surface of the bearing member. Due to their high rotational accuracy and quietness, fluid dynamic bearing devices are suitable for use in small motors such as spindle motors for information devices (e.g., magnetic disk drives such as HDDs, optical disk drives for CD-ROMs, CD-R / RWs, DVD-ROMs / RAMs, and Blu-ray disks, and magneto-optical disk drives for MDs and MOs), polygon scanner motors for laser beam printers (LBPs), color wheels for projectors, and fan motors used in cooling fans for electrical equipment.

[0003] There is a strong demand for thinner portable information devices (so-called mobile devices) such as laptops and tablet PCs. In recent years, information devices have become more sophisticated in order to support 5G (five-generation) technology. This has led to increased heat generation from circuits, which has led to heightened demands for cooling performance. This has resulted in larger impellers attached to rotating shafts. This increases the load on the hydrodynamic bearings that support the rotating shafts of fan motors. 5G (fifth-generation) refers to the "fifth-generation mobile communications system," and its three main features are "high speed, large capacity," "multiple simultaneous connections," and "ultra-low latency."

[0004] A known hydrodynamic bearing for thin fan motors is described in Patent Document 1. The hydrodynamic bearing described in Patent Document 1 aims to increase bearing rigidity against moment loads and suppress shaft whirling without increasing the axial dimension. In this case, hydrodynamic grooves with different groove specifications are arranged in two rows in the axial direction.

[0005] JP 2022-54860 A

[0006] However, in the motor described in Patent Document 1, which uses a hydrodynamic bearing, it is necessary to change the groove specifications depending on the center of gravity position of the rotor, which can result in poor versatility and high costs. Furthermore, in this case, since the groove specifications are different between the top and bottom, if there is a motor with the same motor specifications but a different rotation direction, it cannot be used even if it is turned upside down.

[0007] In view of the above, the present invention provides a hydrodynamic bearing, a fluid dynamic bearing device, and a motor that can improve load capacity without increasing the axial dimension and without compromising versatility.

[0008] The hydrodynamic bearing of the present invention has an inner diameter surface of a bearing that faces the outer diameter surface of a shaft member, and is equipped with a hydrodynamic pressure generating portion on the inner diameter surface of the bearing, the hydrodynamic pressure generating portion having a plurality of hydrodynamic pressure grooves arranged in a herringbone pattern, the hydrodynamic pressure grooves being provided with first and second groups of hydrodynamic pressure grooves that are spaced apart along the axial direction, with inclined sloped hill portions formed between the hydrodynamic pressure grooves of the first and second groups of hydrodynamic pressure grooves, and annular hill portions that are connected to each of the sloped hill portions and extend circumferentially are provided between the first group of hydrodynamic pressure grooves and the second group of hydrodynamic pressure grooves, and a recess is formed in part of the annular hill portion for generating hydrodynamic pressure.

[0009] According to the hydrodynamic bearing of the present invention, a step is formed between the inclined groove (hydrodynamic groove) in each hydrodynamic groove group and the annular hill portion, and pressure (hydrodynamic pressure) is generated at this step, and pressure (hydrodynamic pressure) is generated downstream of the recess in the annular hill portion in the direction of rotation due to the recess. In other words, the annular hill portion has a recess for generating pressure by a wedge effect, thereby improving load capacity.

[0010] It is preferable that a recess for generating the dynamic pressure is provided at the confluence of the annular hill portion and the inclined hill portion, the circumferential width of the recess is smaller than the circumferential width of the confluence, the number of dynamic pressure grooves in the first and second dynamic pressure groove groups is the same, the confluence of the inclined hill portion and the annular hill portion of the first dynamic pressure group coincides with the confluence of the inclined hill portion and the annular hill portion of the second dynamic pressure group, and the recess is provided at all confluences.

[0011] By providing recesses at all confluences between the annular hill and the inclined hill, the load capacity can be further increased.Furthermore, by making the circumferential width of the recesses smaller than the circumferential width of the confluences, the recesses can be set larger within the dimensions that do not contact the contours of the annular hill and the inclined hill, thereby further increasing the load capacity.

[0012] A pair of dynamic pressure generating portions having an upside-down shape and spaced apart in the axial direction may be formed on the inner diameter surface of the bearing. By configuring it in this way, it can be adapted to motors with the same specifications but different rotation directions by simply reversing the bearing upside down.

[0013] The fluid dynamic bearing device of the present invention comprises the dynamic bearing, a shaft member inserted into the inner circumference of the dynamic bearing, and a radial bearing portion that supports the relative rotation of the shaft member by the dynamic pressure action of the lubricating fluid in the radial bearing gap formed between the inner surface of the dynamic bearing and the outer surface of the shaft member.

[0014] The fluid dynamic bearing device of the present invention uses a dynamic bearing that can improve load capacity, so it is possible to support the load on the bearing even if the axial dimension of the bearing is reduced to make it thinner.

[0015] The motor according to the present invention comprises the fluid dynamic bearing device, a rotor that rotates integrally with the shaft member or the dynamic bearing, and a drive unit that drives the rotor to rotate, the rotor having an impeller unit.

[0016] Even if the impeller size is increased, the load on the bearing can be sufficiently supported, and cooling performance can be improved.

[0017] The present invention improves load capacity, allowing the bearing to function satisfactorily even when miniaturized. In other words, it is possible to provide a bearing that improves load capacity without increasing the axial dimension. Furthermore, in motors using this hydrodynamic bearing, there is no need to change the groove specifications depending on the center of gravity of the rotor, and there is no risk of the bearing's versatility being compromised.

[0018] FIG. 1 is a cross-sectional view of a dynamic pressure bearing according to the present invention. FIG. 2 is an enlarged simplified view of a main part of FIG. 1. FIG. 3 is a cross-sectional view of a dynamic pressure bearing device using the dynamic pressure bearing according to the present invention. FIG. 4 is a cross-sectional view of a motor equipped with the dynamic pressure bearing device shown in FIG. 3. FIG. 5 is a simplified view of a dynamic pressure groove. FIG. 6 is a groove specification diagram of an embodiment 1. FIG. 7 is a pressure distribution diagram of an embodiment 1. FIG. 8 is a groove specification diagram of an embodiment 1. FIG. 9 is a pressure distribution diagram of an embodiment 1. FIG. 10 is a groove specification diagram of an embodiment 1. FIG. 11 is a pressure distribution diagram of an embodiment 1. FIG. 12 is a groove specification diagram of an embodiment 1. FIG. 13 is a pressure distribution diagram of an embodiment 1.

[0019] An embodiment of the present invention will now be described with reference to Figures 1 to 4. Figure 1 shows a dynamic pressure bearing according to this embodiment, Figure 2 shows an enlarged simplified view of the main parts of the dynamic pressure bearing, Figure 3 shows a fluid dynamic pressure bearing device using the dynamic pressure bearing according to the present invention, and Figure 4 shows a cooling fan motor using this fluid dynamic pressure bearing device. This fan motor is incorporated into, for example, information devices, particularly mobile devices such as mobile phones and tablet terminals.

[0020] This fan motor comprises a fluid dynamic bearing device 1 according to one embodiment of the present invention, a rotor 3 mounted on a shaft member 2 of the fluid dynamic bearing device 1, an impeller (vanes) 4 attached to the outer diameter end of the rotor 3, a stator coil 6a and a rotor magnet 6b opposed to each other across a radial gap, and a casing 5 that houses these. The stator coil 6a is attached to the outer periphery of the fluid dynamic bearing device 1, and the rotor magnet 6b is attached to the inner periphery of the rotor 3. When current is applied to the stator coil 6a, the rotor 3, impeller 4, and shaft member 2 rotate together, thereby generating an airflow in the axial or outer diameter direction.

[0021] 3, the fluid dynamic bearing device 1 comprises a shaft member 2, a housing 7, a bearing sleeve 8 as a dynamic bearing according to the present invention, a seal member 9, and a thrust bearing 10. In the following, the open side of the housing 7 will be referred to as the upper side in the axial direction (the vertical direction in FIG. 2), and the side of the bottom 7b of the housing 7 will be referred to as the lower side.

[0022] The shaft member 2 is formed in a cylindrical shape from a metal material such as stainless steel, etc. The shaft member 2 has a cylindrical outer circumferential surface 2a and a spherical protrusion 2b provided at the lower end.

[0023] The housing 7 has a substantially cylindrical side portion 7a and a bottom portion 7b that closes the opening below the side portion 7a. In the illustrated example, the side portion 7a and the bottom portion 7b are integrally injection-molded from resin. The casing 5 and the stator coil 6a are fixed to the outer peripheral surface 7a2 of the side portion 7a. The bearing sleeve 8 is fixed to the inner peripheral surface 7a1 of the side portion 7a. A shoulder surface 7b2 is provided at the outer diameter end of the upper end surface 7b1 of the bottom portion 7b, located above the inner diameter portion, and the lower end surface 8c of the bearing sleeve 8 abuts against this shoulder surface 7b2. A circumferential notch 7b4 is formed in the outer peripheral surface 7a2 of the side portion 7a, and the stator coil 6a is fitted into this circumferential notch 7b4. A resin thrust receiver 10 is disposed in the center of the upper end surface 7b1 of the bottom portion 7b. Instead of (or in addition to) providing the radial groove 7b3 on the shoulder surface 7b2 of the housing 7, a radial groove may be formed on the lower end surface 8c of the bearing sleeve 8.

[0024] The bearing sleeve 8 is cylindrical and fixed to the inner peripheral surface 7a1 of the side portion 7a of the housing 7 by an appropriate means such as gap bonding, press fitting, or press fitting adhesion (press fitting with an adhesive interposed therebetween). In this embodiment, the bearing sleeve 8 can be made of sintered metal processed by pressing, ingot material such as brass or stainless steel processed by cutting, or resin processed by injection molding.

[0025] A pair of dynamic pressure generating portions 20 (20A, 20B) are provided axially spaced apart on the inner peripheral surface 8a of the bearing sleeve 8, which serves as the radial bearing surface. Each dynamic pressure generating portion 20A, 20B has a plurality of dynamic pressure grooves 11a, 11b arranged in a herringbone pattern. The dynamic pressure grooves 11a, 11a on the axially outer side of the dynamic pressure generating portions 20A, 20B form a first dynamic pressure groove group 11A, 11A, and the dynamic pressure grooves 11b, 11b on the axially inner side of the dynamic pressure generating portion 20 form a second dynamic pressure groove group 11B, 11B. The pair of dynamic pressure generating portions 20A, 20B have a vertically inverted shape.

[0026] Furthermore, in each dynamic pressure generating portion 20A, 20B, the dynamic pressure grooves 11a, 11a on the axially outer side and the dynamic pressure grooves 11b, 11b on the axially inner side have different inclination directions. In the illustrated example, the dynamic pressure grooves 11a, 11a on the axially outer side are inclined from the axially outer side to the axially inner side along the rotation direction of the shaft member 2, while the dynamic pressure grooves 11b, 11b on the axially inner side are inclined from the axially inner side to the axially outward side along the rotation direction of the shaft member 2. The bottom surfaces of the dynamic pressure grooves 11a, 11b are provided on the same cylindrical surface. The bottom surface of the dynamic pressure groove 11b on the axially inner side is continuous with a cylindrical surface 13 provided axially between both dynamic pressure generating portions 20 (20A, 20B).

[0027] Inclined hills 11c and 11d are provided between the dynamic pressure grooves 11a and 11b, respectively. Furthermore, in each dynamic pressure generating portion 20 (20A, 20B), annular hills 11e and 11e are provided between the first dynamic pressure groove group 11A and the second dynamic pressure groove group 11B. The inclined hills 11c and 11d and the annular hills 11e and 11e are indicated by cross-hatching. The annular hills 11e and 11e and the inclined hills 11c and 11d rise radially inward from the bottom surfaces of the dynamic pressure grooves 11a and 11b. The inner diameter surfaces of the annular hills 11e and 11e and the inclined hills 11c and 11d are provided on the same cylindrical surface. The annular hills 11e and 11e and all of the inclined hills 11c and 11d are provided continuously.

[0028] Let A and B be the axial lengths (width dimensions) of the first dynamic pressure groove group 11A and the second dynamic pressure groove group 11B, and add them to the axial length (width dimension) C of the annular hill portions 11e, 11e, so that A = B > C. In each of the first dynamic pressure generating portions 20A and second dynamic pressure generating portions 20B, the inclination angle θa (see FIG. 2) of the dynamic pressure grooves 11a, 11a of the first dynamic pressure groove group 11A, 11A relative to the circumferential direction is set equal to the inclination angle θb (see FIG. 2) of the dynamic pressure grooves 11b, 11b of the second dynamic pressure groove group 11B, 11B relative to the circumferential direction.

[0029] Each annular hill 11e has a recess 11e1. In this case, the recess 11e1 is formed at the confluence U between the inclined hills 11c and 11d and the annular hill 11e. In the illustrated example, the recess 11e1 is rectangular, and its depth is the same as the height of the annular hill 11e. That is, the bottom surface of the recess 11e1 is aligned with the cylindrical surface 13. The width (axial length) of the recess 11e1 is approximately the same as the width (axial length) of the annular hill 11e. In this case, the circumferential length of the recess 11e1 is such that it is not exposed from the confluence U. That is, when the circumferential length of the confluence U is L and the circumferential length of the recess 11e1 is L1, L > L1. When the axial length of the confluence U is H and the axial length of the recess 11e1 is H1, H ≥ H1.

[0030] Specifically, the bearing sleeve 8 is formed from a sintered metal, for example, a sintered metal containing 35 wt. % or more copper, particularly a sintered metal containing 35 wt. % or more each of copper and iron. The bearing sleeve 8 is manufactured by the following method. First, raw material powder is compressed to form a green compact (compacting process). The raw material powder contains, as the main component metal powder, either or both of a copper-based powder (copper powder or copper alloy powder) and an iron-based powder (iron powder or iron alloy powder). The raw material powder may also contain a high-hardness powder such as stainless steel powder. In this embodiment, the raw material powder contains, as the main component metal powder, pure iron powder and pure copper powder. In addition to the main component metal powder, the raw material powder may also contain a low-melting-point metal powder such as tin powder, a carbon powder such as graphite powder, or a molding lubricant. The green compact is sintered at a predetermined sintering temperature to obtain a sintered body (sintering process). The sintered body is then sized to form dynamic pressure grooves 11, 11 on the inner circumferential surface (sizing process). In this embodiment, the inner peripheral surface of the sintered body is subjected to a sealing treatment such as rotary sizing, etc. The bearing sleeve 8 is completed by impregnating the internal pores of this sintered body with lubricating oil.

[0031] The bearing sleeve 8 has a density ratio of 80 to 95%. The bearing sleeve 8 has communicating pores that connect the interior and the surface; specifically, communicating pores with an oil content of 4% or more are formed. That is, the molding conditions of the bearing sleeve 8 (e.g., the compression ratio in the compacting process and sizing process) are set so that communicating pores with an oil content of 4% or more are formed. The sizing of the bearing sleeve 8 sets the surface opening ratio of the inner circumferential surface 8a (radial bearing surface) to the value of the porosity of the bearing sleeve 8 (= 100% - density ratio) or less; specifically, it is set to 10% or less, preferably 8% or less, and more preferably 5% or less. By setting the opening ratio to 10% or less, it is possible to effectively prevent dynamic pressure loss from the inner diameter surface 8a.

[0032] Axial grooves 8d1 are formed on the outer peripheral surface of the bearing sleeve 8. The number of axial grooves 8d1 is arbitrary, and for example, they are formed at three locations equidistantly spaced in the circumferential direction.

[0033] The seal member 9 is formed in an annular shape from resin or metal and is fixed to the upper end of the inner circumferential surface 7a1 of the side portion 7a of the housing 7. The seal member 9 abuts against the upper end face 8b of the bearing sleeve 8. The inner circumferential surface 9a of the seal member 9 faces the outer circumferential surface 2a of the shaft member 2 in the radial direction, forming a seal space S therebetween. When the shaft member 2 rotates, the seal space S prevents the lubricating oil inside the bearing from leaking to the outside. A radial groove 9b1 is formed in the lower end face 9b of the seal member 9. Note that instead of (or in addition to) forming the radial groove 9b1 in the lower end face 9b of the seal member 9, a radial groove may be formed in the upper end face 8b of the bearing sleeve 8.

[0034] The hydrodynamic bearing device 1 is assembled as follows. First, the thrust receiver 10 is fixed to the upper end surface 7b1 of the bottom portion 7b of the housing 7. Next, the bearing sleeve 8, whose internal pores have been pre-impregnated with lubricating oil, is inserted into the inner periphery of the side portion 7a of the housing 7. With the lower end surface 8c of the bearing sleeve 8 abutting against the shoulder surface 7b2 of the bottom portion 7b, the outer periphery 8d of the bearing sleeve 8 is fixed to the inner periphery 7a1 of the side portion 7a. Next, the seal member 9 is fixed to the upper end of the inner periphery 7a1 of the side portion 7a of the housing 7. At this time, the seal member 9 is press-fitted into the side portion 7a of the housing 7, and the bearing sleeve 8 is sandwiched between the seal member 9 and the shoulder surface 7b2 of the bottom portion 7b of the housing 7 from both axial sides, thereby restraining the bearing sleeve 8 in the axial direction. Next, lubricating oil is dripped onto the inner periphery of the bearing sleeve 8, and the shaft member 2 is inserted, completing the assembly of the hydrodynamic bearing device 1. At this time, the internal space of the housing 7 sealed by the seal member 9 (including the internal cavity of the bearing sleeve 8) is filled with lubricating oil, and the oil level is maintained within the range of the seal space S.

[0035] In the fluid dynamic bearing device 1 configured as described above, when the shaft member 2 rotates, a radial bearing gap is formed between the inner circumferential surface 8a of the bearing sleeve 8 and the outer circumferential surface 2a of the shaft member 2. The dynamic pressure generating portions 20 (20A, 20B) formed on the inner circumferential surface 8a of the bearing sleeve 8 generate dynamic pressure in the lubricating oil in the radial bearing gap. More specifically, the lubricating oil in the radial bearing gap is collected toward the axial center of each dynamic pressure generating portion 20 (20A, 20B) along the dynamic pressure grooves 11a, 11b, increasing the fluid pressure in this area. This forms a radial bearing portion R (R1, R1) that supports the shaft member 2 in a non-contact manner in the radial direction. Furthermore, the protrusion 2b at the lower end of the shaft member 2 comes into contact with and slides against the thrust receiver 10, thereby forming a thrust bearing portion T that supports the shaft member 2 in the thrust direction.

[0036] In the hydrodynamic bearing of the present invention, a step is formed between the inclined grooves (hydrodynamic grooves) 11a, 11b in each hydrodynamic groove group 20A, 20B and the annular hill portion 11e. This step generates pressure (dynamic pressure). Furthermore, the recess 11e1 of the annular hill portion 11e generates pressure (dynamic pressure) downstream of the recess 11e in the direction of rotation. This increases the pressure (dynamic pressure) generated on the annular hill portion 11e, improving the load capacity. In other words, the improved load capacity allows the bearing to function satisfactorily even when miniaturized. This means that a bearing with improved load capacity can be provided without increasing the axial dimension. Furthermore, in a motor using this hydrodynamic bearing, there is no need to change the groove specifications depending on the center of gravity of the rotor, and there is no risk of the bearing's versatility being compromised.

[0037] In addition, by providing recesses 11e1 at all confluences U between the annular hill portion 11e and the inclined hill portions 11c and 11d, the load capacity can be further increased.Furthermore, by making the circumferential width of the recesses 11e1 smaller than the circumferential width of the confluences U, the recesses 11e1 can be set larger within the dimensions that do not contact the contours of the annular hill portion 11e and the inclined hill portions 11c and 11d, thereby further increasing the load capacity.

[0038] Furthermore, the fluid dynamic bearing device of the present invention uses a dynamic bearing that can improve load capacity, so it is possible to support the load on the bearing even if the axial dimension of the bearing is reduced to make it thinner.

[0039] The motor according to the present invention can sufficiently support the load on the bearing even if the size of the impeller 4 is increased, and can improve the cooling performance.

[0040] The above describes an embodiment of the present invention, but the present invention is not limited to the above embodiment and various modifications are possible. In the above embodiment, the sintered oil-impregnated bearing 8 is fixed and the shaft member 2 rotates, but the present invention 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.

[0041] A fluid dynamic bearing device incorporating an oil-impregnated sintered bearing 8 according to the present invention can be used not only in spindle motors used in HDD disk drive devices, but also in a wide range of other small motors, such as spindle motors incorporated in other information devices, polygon scanner motors in laser beam printers, color wheels in projectors, or cooling fan motors.

[0042] In the embodiment, a pair of dynamic pressure generating portions 20A, 20B are provided, but a single dynamic pressure generating portion 20 may be provided. Furthermore, the inclination angles θa, θb of the dynamic pressure generating grooves 11a, 11b are not limited to those in the embodiment. The number, arrangement pitch, groove width, etc. of the dynamic pressure generating grooves 11a, 11b can be set arbitrarily, and the depth of the dynamic pressure generating grooves can also be changed as needed. In the embodiment, the depth of the dynamic pressure generating grooves (inclined grooves) and the depth of the recess 11e1 provided in the annular hill portion 11e are set to be the same, but they may be different. Note that making the depths the same can improve productivity.

[0043] In the embodiment, when the axial lengths (width dimensions) of the first dynamic pressure groove group 11A and the second dynamic pressure groove group 11B are A and B, and the axial length (width dimension) of the annular hill portions 11e, 11e is C, A = B > C, but it may also be A = B < C, A = B = C, or A ≠ B. Furthermore, although the shape of the recess 11e1 is rectangular in the embodiment, it may also be a square, a polygonal shape with pentagons or more sides, or the like.

[0044] The bearing sleeve 8 is a cylindrical body having a porous structure, and is formed of sintered metal as in the embodiment, for example, but it can also be formed of a porous body made of a non-metallic material such as resin or ceramic. In addition to a porous body, it can also be formed of a structure that does not have internal pores or has pores of a size that do not allow lubricating oil to enter or exit.

[0045] For a bearing (Product 1) with the specifications shown in Table 1, the pressure distribution on the outer diameter surface of the shaft member and the forces acting on the outer diameter surface were calculated using thermal fluid analysis software (SIEMENS STAR-CCM+). STAR-CCM+ is integrated CAE software for multi-domain problems. Its core function is fluid analysis using the finite volume method. It also includes fluid physics functions, structural analysis using the finite element method, and particle analysis using the DEM method. STAR-CCM+ handles everything from CAD geometry creation to evaluation of calculation results in a single package, facilitating automation of the workflow. It integrates the workflow required for simulation setup, calculation execution, and result evaluation into a single GUI, and is fully equipped with all the necessary automation functions. STAR-CCM+'s extensive physical models and functions enable the complex analysis of a wide range of phenomena. In addition to basic functions such as 2D / 3D, laminar / turbulent / inviscid, compressible / incompressible, buoyancy, translation / rotation, and porous regions, it also includes various turbulence models used in fluid analysis.

[0046] In this case, to shorten the analysis time, the dynamic pressure generating portions 20 were arranged in a single row, as shown in Figure 5. The bearing specifications of Example 1 are shown in Table 1. The shaft (shaft member 2) used had an outer diameter (diameter) of 1.99 mm.

[0047] The inner diameter (diameter) of the bearing was 2 mm, the width (axial length of the bearing) was 1.80 mm, the groove depth was 10 μm, the hill-groove ratio was 1, the groove angle (θa, θb) was 20 degrees, the inclined groove width (axial length A of the hydrodynamic groove groups 11A, 11B) was 0.7 mm, the annular hill portion width (axial length C of the annular hill portion 11e) was 0.4 mm, the radial gap was 10 μm, and the eccentricity was 0. Here, the hill-groove ratio is H2 / H1, where H2 is the circumferential length of the hill portion and H1 is the circumferential length of the groove.

[0048] In this case, as shown in Figure 5 for Example 1, the axial length (width dimension) H of the recess 11e1 was 0.4 mm, the circumferential length L1 of the recess 11e1 was 0.32 mm, and the recess 11e1 was arranged in a position not offset from the confluence part U, which was used as Example 1. Calculations were also carried out for two comparative examples 1 and 2, which had different positions and sizes of the recess 11e1, and a conventional product without a recess. Therefore, the circumferential length of the ridge part of Comparative Example 1 was half that of Example 1, and the axial length of the ridge part of Comparative Example 2 was approximately 0.1 mm shorter than that of Example 1, and the ridge part was offset approximately 0.22 mm in the direction of rotation. The rotation speed of the shaft member 2 was set to 5,500 rpm.

[0049] Figures 6A and 6B show Example Product 1, with Figure 6A being a groove specification diagram and Figure 6B being a pressure distribution diagram, Figures 7A and 7B show Comparison Product 1, with Figure 7A being a groove specification diagram and Figure 7B being a pressure distribution diagram, Figures 8A and 8B show Comparison Product 2, with Figure 8A being a groove specification diagram and Figure 8B being a pressure distribution diagram, and Figures 9A and 9B show Conventional Product 1, with Figure 9A being a groove specification diagram and Figure 9B being a pressure distribution diagram.

[0050] As can be seen from each pressure distribution diagram, an increase in pressure was observed on the downstream side of the groove in the rotation direction in Example 1, Comparative Example 1, and Comparative Example 2, which were provided with recess 11e1. That is, the pressure on the outer diameter surface of the shaft member was 0.0049 N in Example 1, 0.0042 N in Comparative Example 1, 0.0042 N in Comparative Example 2, and 0.0039 N in Conventional Example 1. In this case, Example 1 had the highest pressure, 1.3 times that of the Conventional Example.

[0051] Next, a functional evaluation was carried out using Example Product 2 having the bearing specifications shown in Table 2. The shaft (shaft member 2) used had an outer diameter (diameter) of 1.99 mm.

[0052] The inner diameter (diameter) of the bearing (bearing sleeve 8) was 2 mm, the width (length in the bearing axial direction) was 2.35 mm, the groove depth was 3 μm, the hill-groove ratio was 1, the groove angle (θ1, θ2) was 20 degrees, the inclined groove width (length of the dynamic pressure groove in the bearing axial direction) was 0.25 mm, the annular hill width (length of the annular hill portion 11e in the bearing axial direction) was 0.2 mm, and the radial gap was 8 μm.

[0053] In this case, three examples, No. 1 to No. 3, were produced as Example Product 2, and three examples, No. 1 to No. 3, were produced as Conventional Product 2. Example Product 2 No. 1 to No. 3 were each as shown in FIG. 10 . The axial length (width) of the recess 11e1 was 0.2 mm, the circumferential length of the recess 11e1 was 0.25 mm, and the recess 11e1 was positioned so as not to be offset from the confluence portion U. The dimension L2 from the axial outer edge of the annular hill portion 11e of one dynamic pressure generating portion 20A to the other end face of the bearing (bearing sleeve 8) was 1.9 mm, and the dimension L3 from the axial outer edge of the annular hill portion of the other dynamic pressure generating portion 20B to the other end face of the bearing was 0.45 mm. In this case, Example Product 2 No. 1 to No. 3 were identical in design.

[0054] Furthermore, as shown in Figure 11, Conventional Product 2 No. 1 to No. 3 are the same as those shown in Figure 10 except that they do not have recesses. In this case, Conventional Product 2 No. 1 to No. 3 were identical in design.

[0055] 11 , a pair of dynamic pressure generating portions 120 (120A, 120B) are provided axially spaced apart on the inner circumferential surface 108a of a bearing sleeve 108, which serves as a radial bearing surface. Each of the dynamic pressure generating portions 120A, 120B has a plurality of dynamic pressure grooves 111a, 111b arranged in a herringbone pattern. The dynamic pressure grooves 111a, 111a on the axially outer sides of the dynamic pressure generating portions 120A, 120B form a first dynamic pressure groove group 111A, 111A, and the dynamic pressure grooves 111b, 111b on the axially inner sides of the dynamic pressure generating portions 120 form a second dynamic pressure groove group 111B, 111B. The pair of dynamic pressure generating portions 120A, 120B have an upside-down shape.

[0056] In each dynamic pressure generating portion 120A, 120B, the dynamic pressure grooves 111a, 111a on the axially outer side are inclined in a different direction from the dynamic pressure grooves 111b, 111b on the axially inner side. Sloping hill portions 111c, 111d are provided between the dynamic pressure grooves 111a, 111b, respectively. The sloping hill portions 11c, 11d and the annular hill portions 11e, 11e are indicated by cross-hatching.

[0057] The following Table 3 shows the results of the functional evaluation. In Table 3, a load (compression load) of 0.1 N was applied to the shaft (shaft member 2), and the shaft member 2 was rotated at a rotation speed of 100 to 1000 rpm for 60 seconds. The presence or absence of contact between the rotating shaft member 2 and the bearing (bearing sleeve 8) was measured. This measurement was performed using a publicly known electrical contact method.

[0058] It can be seen that for the conventional products (No. 1 to No. 3), contact between the shaft member and the bearing begins at 1000 rpm, and the number of contacts increases dramatically below 200 rpm. In contrast, for the example products 2 (No. 1 to No. 3), almost no contact is observed up to 200 rpm, and the number of contacts at 100 rpm is also lower than that of the comparative example. The numerical values ​​in Table 3 indicate the number of contacts, with larger values ​​indicating more contacts. Also, in Table 3, >500 indicates that the shaft member and the bearing came into contact with each other 500 times or more in 60 seconds.

[0059] The presence or absence of contact between the shaft (shaft member 2) and the bearing (bearing sleeve 8) was measured when the load (stress) applied to the shaft (shaft member 2) was gradually increased. In this test, the experimental product 2 (Nos. 1 to 3) and the conventional product 2 (Nos. 1 to 3) were used. The rotation speed was set to 1000 rpm, and the presence or absence of contact between the shaft (shaft member 2) and the bearing (bearing sleeve 8) was measured when the load (stress) applied to the shaft member was gradually increased. The results are shown in Table 4. This measurement was also performed using a known electrical contact method. In Table 4, "-" indicates that no measurement was performed at this load. This is because the number of contacts exceeded 500 at the load immediately prior to this load, and therefore the number of contacts at the subsequent load would also exceed 500.

[0060] As can be seen from Table 4, the load at which frequent contact between the shaft member and the bearing was observed was 1.27 N for the conventional product, while it was 1.57 N (1.2 times) for the implemented product. Therefore, it was confirmed that the applied load of the implemented product 2 was higher than that of the conventional product 2.

[0061] Fluid dynamic bearing devices incorporating oil-impregnated sintered bearings can be widely used in other small motors, such as spindle motors incorporated into information equipment, polygon scanner motors for laser beam printers, color wheels for projectors, and cooling fan motors.

[0062] DESCRIPTION OF SYMBOLS 1 Fluid dynamic bearing device 2 Shaft member 3 Rotor 4 Impeller 5 Casing 8 Sintered oil-impregnated bearing (bearing sleeve) 8a Inner peripheral surface 11A, 11B Dynamic pressure groove group 11a, 11b Dynamic pressure groove 11c Inclined hill portion 11e Annular hill portion 11e1 Recess 20 (20A, 20B) Dynamic pressure generating portion C Confluence portion R1 Radial bearing portion

Claims

1. A hydrodynamic bearing having a bearing inner diameter surface facing the outer diameter surface of a shaft member, the bearing inner diameter surface being provided with a hydrodynamic generating portion, the hydrodynamic generating portion having a plurality of hydrodynamic grooves arranged in a herringbone shape, the hydrodynamic grooves being provided with first and second hydrodynamic groove groups, and inclined inclined hill portions being formed between the hydrodynamic grooves of the first and second hydrodynamic groove groups, respectively, and an annular hill portion connected to each inclined hill portion and extending in the circumferential direction being provided between the first hydrodynamic groove group and the second hydrodynamic groove group, a concave portion for generating hydrodynamic pressure being formed in a part of the annular hill portion.

2. The hydrodynamic bearing according to claim 1, wherein the circumferential width of the concave portion is made smaller than the circumferential width of the confluence portion, the number of hydrodynamic grooves in the first and second hydrodynamic groove groups is made the same, the confluence portion between the inclined hill portion and the annular hill portion of the first hydrodynamic group coincides with the confluence portion between the inclined hill portion and the annular hill portion of the second hydrodynamic group, and the concave portion is provided in all confluence portions.

3. The hydrodynamic bearing according to claim 1, wherein a pair of hydrodynamic generating portions having an upside-down shape separated in the axial direction are formed on the bearing inner diameter surface.

4. A hydrodynamic bearing device comprising the hydrodynamic bearing according to any one of claims 1 to 3, a shaft member inserted into the inner circumference of the hydrodynamic bearing, and a radial bearing portion that supports the relative rotation of the shaft member by the hydrodynamic action of a lubricating fluid in a radial bearing clearance formed between the inner circumference surface of the hydrodynamic bearing and the outer circumference surface of the shaft member.

5. A motor comprising the hydrodynamic bearing device according to claim 4, a rotor that rotates integrally with the shaft member or the hydrodynamic bearing, and a drive portion that rotationally drives the rotor, the rotor having an inverter portion.

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