JP7732007B2 - Dynamic pressure bearing, fluid dynamic pressure bearing device, and motor - Google Patents

Description

The present invention relates to a hydrodynamic bearing, a fluid dynamic bearing device, and a motor.

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

There is a strong demand for thinner portable information devices (so-called mobile devices) such as laptops and tablet PCs. Furthermore, in recent years, as information devices have become more sophisticated in order to support 5G (five-generation), the amount of heat generated from circuits has increased, leading to greater demands for cooling performance. As a result, the impellers attached to the rotating shaft tend to be larger. This increases the load on the hydrodynamic bearing device that supports the rotating shaft of the fan motor. 5G (fifth-generation mobile communications system) is characterized by three main features: high speed and large capacity, multiple simultaneous connections, and ultra-low latency.

A conventional hydrodynamic bearing compatible with 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 axial rows.

JP 2022-54860 A

However, in the motor described in Patent Document 1, which uses a hydrodynamic bearing, the groove specifications must be changed depending on the center of gravity of the rotor, which can make it less versatile and more costly. Furthermore, in this case, since the groove specifications are different between the top and bottom, if there is a motor with the same specifications but a different direction of rotation, it cannot be used even if it is turned upside down.

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 axial dimensions or compromising versatility.

The hydrodynamic bearing of the present invention has an inner diameter surface facing 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 has a plurality of hydrodynamic grooves arranged in a herringbone pattern. The hydrodynamic grooves are provided with first and second groups of hydrodynamic grooves spaced apart along the axial direction, with inclined hill portions formed between the hydrodynamic grooves of the first and second groups, and annular hill portions connected to the inclined hill portions and extending circumferentially are provided between the first and second groups of hydrodynamic grooves, with recesses formed in parts of the annular hill portions for generating hydrodynamic pressure.

In the hydrodynamic bearing of the present invention, a step is formed between the inclined grooves (hydrodynamic grooves) in each hydrodynamic groove group and the annular hill portion, generating pressure (hydrodynamic pressure) at this step, and the recess in the annular hill portion generates pressure (hydrodynamic pressure) downstream of the recess in the direction of rotation. In other words, the annular hill portion has a recess that generates pressure through a wedge effect, improving load capacity.

It is preferable that the 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.

By providing recesses at all junctions 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 junctions, the recesses can be set larger within the dimensions that do not contact the contours of the annular hill and the inclined hill, further increasing the load capacity.

A pair of dynamic pressure generating sections, each with an inverted shape and spaced apart in the axial direction, may be formed on the inner diameter surface of the bearing. This configuration makes it possible to accommodate motors with the same specifications but different rotation directions by simply inverting the bearing.

The fluid dynamic bearing device of the present invention comprises the dynamic bearing, a shaft member inserted into the inner periphery 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 periphery of the dynamic bearing and the outer periphery of the shaft member.

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

A 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 .

Even if the impeller size is increased, it can still fully support the load on the bearing, improving cooling performance.

The present invention improves load capacity, allowing the bearing to fully function 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 position of the rotor that constitutes it, and there is no risk of compromising the versatility of the bearing.

1 is a cross-sectional view of a hydrodynamic bearing according to the present invention. FIG. 2 is an enlarged simplified view of a main part of FIG. 1; 1 is a cross-sectional view of a dynamic pressure bearing device using a 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. 1 shows the embodiment 1, where (a) is a groove specification diagram and (b) is a pressure distribution diagram. 1 shows a comparative product 1, where (a) is a groove specification diagram and (b) is a pressure distribution diagram. 1 shows comparative product 2, where (a) is a groove specification diagram and (b) is a pressure distribution diagram. 1 shows a conventional product 1, where (a) is a groove specification diagram and (b) is a pressure distribution diagram. FIG. 10 is a cross-sectional view of a hydrodynamic bearing used in Example 2. FIG. 10 is a cross-sectional view of a hydrodynamic bearing used in conventional product 2.

Embodiments 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 essential 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.

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 facing 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, generating an airflow in the axial or outer diameter direction.

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

The shaft member 2 is formed into a cylindrical shape from a metal material such as stainless steel. The shaft member 2 has a cylindrical outer surface 2a and a spherical protrusion 2b at its lower end.

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 bottom portion 7b are injection molded integrally from resin. The casing 5 and 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 located above the inner diameter portion is provided at the outer diameter end of the upper end surface 7b1 of the bottom portion 7b, 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 support 10 is disposed in the center of the upper end surface 7b1 of the bottom portion 7b. Furthermore, 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.

The bearing sleeve 8 is cylindrical and is fixed to the inner circumferential surface 7a1 of the side portion 7a of the housing 7 by appropriate means such as gap welding, press fitting, or press fitting (press fitting with adhesive). 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.

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 multiple 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, while 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 an upside-down shape.

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 outer 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 axially inner dynamic pressure groove 11b is continuous with the cylindrical surface 13 provided axially between the two dynamic pressure generating portions 20 (20A, 20B).

Sloped hills 11c and 11d are provided between the dynamic pressure grooves 11a and 11b, respectively. Furthermore, in each dynamic pressure generating section 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 sloped hills 11c and 11d and the annular hills 11e and 11e are indicated by cross-hatching. The annular hills 11e and 11e and the sloped hills 11c and 11d rise from the bottom surfaces of the dynamic pressure grooves 11a and 11b toward the inner diameter. The inner diameter surfaces of the annular hills 11e and 11e and the sloped hills 11c and 11d are provided on the same cylindrical surface. The annular hills 11e and 11e and all of the sloped hills 11c and 11d are provided continuously.

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) of the annular hill portion 11e, 11e, C, so that A = B > C. Furthermore, 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 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.

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.

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 copper-based powder (copper powder or copper alloy powder) and iron-based powder (iron powder or iron alloy powder). The raw material powder may also contain 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 low-melting-point metal powder such as tin powder, carbon powder such as graphite powder, or a molding lubricant. This 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 surface of the sintered body is subjected to a sealing process such as rotary sizing. The bearing sleeve 8 is completed by impregnating the internal pores of this sintered body with lubricating oil.

The bearing sleeve 8 has a density ratio of 80 to 95%. The bearing sleeve 8 has interconnected pores that connect the interior and surface; specifically, interconnected pores with an oil content of 4% or more. In other words, the molding conditions for the bearing sleeve 8 (e.g., the compression ratio in the compacting and sizing processes) are set so that interconnected pores with an oil content of 4% or more are formed. Sizing of the bearing sleeve 8 results in the surface open area ratio of the inner peripheral surface 8a (radial bearing surface) being equal to or less than the porosity (= 100% - density ratio) of the bearing sleeve 8; specifically, 10% or less, preferably 8% or less, and more preferably 5% or less. By keeping the open area ratio at 10% or less, dynamic pressure loss from the inner diameter surface 8a can be effectively prevented.

Axial grooves 8d1 are formed on the outer peripheral surface of the bearing sleeve 8. The number of axial grooves 8d1 is arbitrary; for example, three grooves 8d1 may be formed at equal intervals in the circumferential direction.

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 the upper end face 8b of the bearing sleeve 8. The inner circumferential surface 9a of the seal member 9 faces radially opposite the outer circumferential surface 2a of the shaft member 2, forming a seal space S between them. When the shaft member 2 rotates, the seal space S prevents 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.

The hydrodynamic bearing device 1 is assembled as follows. First, the thrust bearing 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 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 on 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 (including the internal cavity of the bearing sleeve 8), which is sealed by the seal member 9, is filled with lubricating oil, and the oil level is maintained within the seal space S.

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. 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 the radial direction without contact. Furthermore, the protrusion 2b at the lower end of the shaft member 2 comes into contact with and slides against the thrust receiver 10, forming a thrust bearing portion T that supports the shaft member 2 in the thrust direction.

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 (hydrodynamic pressure). Furthermore, the recess 11e1 of the annular hill portion 11e generates pressure (hydrodynamic pressure) downstream of the recess 11e in the direction of rotation. This increases the pressure (hydrodynamic pressure) generated on the annular hill portion 11e, improving load capacity. In other words, improved load capacity allows the bearing to function fully even when miniaturized. This means that a bearing with improved load capacity can be provided 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 compromising the versatility of the bearing.

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.

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

The motor of the present invention can adequately support the load on the bearings even if the size of the impeller 4 is increased, thereby improving cooling performance.

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.

A fluid dynamic bearing device incorporating the oil-impregnated sintered bearing 8 of 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 into other information devices, polygon scanner motors in laser beam printers, color wheels in projectors, and cooling fan motors.

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

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. Also, while the shape of the recess 11e1 is rectangular in the embodiment, it may also be square, a polygonal shape with pentagons or more sides, etc.

The bearing sleeve 8 is a cylindrical body with a porous structure, and is formed, for example, from sintered metal as in the embodiment, but it can also be formed from a porous body made from a non-metallic material such as resin or ceramic. In addition to porous bodies, it can also be formed from a structure that has no internal pores or has pores large enough to prevent the inflow and outflow of lubricating oil.

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, centered on fluid analysis functionality using the finite volume method. It also includes fluid-related physics functions, structural analysis functionality using the finite element method, and particle analysis functionality using the DEM method. STAR-CCM+ handles everything from CAD geometry creation to evaluation of calculation results in a single package, facilitating automation of work processes. The workflow required for simulation setup, calculation execution, and result evaluation are all integrated into a single GUI, complete with all the necessary automation functions. STAR-CCM+'s extensive physical models and functions enable the comprehensive analysis of a variety 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.

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.

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

In this case, as shown in Figure 5 of 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 positioned so as not to be offset from the confluence point U. This was used as Example 1. Calculations were then carried out for two comparison products 1 and 2, which differed in the position and size of the recess 11e1, as well as a conventional product that did not have a recess. Therefore, for Comparison 1, the circumferential length of the ridge portion was half that of Example 1, and for Comparison 2, the axial length of the ridge portion was approximately 0.1 mm shorter than for Example 1, and the ridge portion was offset approximately 0.22 mm in the direction of rotation. The rotation speed of shaft member 2 was set to 5,500 rpm.

Figure 6 shows the implementation product 1, where (a) is the groove specification diagram and (b) is the pressure distribution diagram; Figure 7 shows the comparison product 1, where (a) is the groove specification diagram and (b) is the pressure distribution diagram; Figure 8 shows the comparison product 2, where (a) is the groove specification diagram and (b) is the pressure distribution diagram; and Figure 11 shows the conventional product 1, where (a) is the groove specification diagram and (b) is the pressure distribution diagram.

As can be seen from each pressure distribution diagram, Example 1, Comparative Example 1, and Comparative Example 2, which are provided with recess 11e1, exhibited an increase in pressure downstream of the groove in the direction of rotation. 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 was the highest, 1.3 times that of the Conventional Example.

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.

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

In this case, three items, No. 1 to No. 3, are produced as the implementation product 2, and No. 1 to No. 3 are produced as the conventional product 2.
Three examples were produced. Examples 1 to 3 of Example 2 were as shown in Figure 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 misaligned with the confluence 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, Examples 1 to 3 of Example 2 were identical in design.

Furthermore, as shown in Figure 11, No. 1 to No. 3 of conventional product 2 are products that do not have the recesses in the product shown in Figure 10. In this case, No. 1 to No. 3 of conventional product 2 are identical in design.

The hydrodynamic bearing shown in Figure 11 has a pair of hydrodynamic pressure generating portions 120 (120A, 120B) spaced apart in the axial direction on the inner peripheral surface 108a of the bearing sleeve 108, which serves as the radial bearing surface. Each hydrodynamic pressure generating portion 120A, 120B has multiple hydrodynamic pressure grooves 111a, 111b arranged in a herringbone pattern. The hydrodynamic pressure grooves 111a, 111a on the axially outer side of the hydrodynamic pressure generating portions 120A, 120B form a first hydrodynamic pressure groove group 111A, 111A, and the hydrodynamic pressure grooves 111b, 111b on the axially inner side of the hydrodynamic pressure generating portion 120 form a second hydrodynamic pressure groove group 111B, 111B. The pair of hydrodynamic pressure generating portions 120A, 120B have an upside-down shape.

Furthermore, in each dynamic pressure generating portion 120A, 120B, the dynamic pressure grooves 111a, 111a on the axially outer side and the dynamic pressure grooves 111b, 111b on the axially inner side have different inclination directions. Sloping hill portions 111c, 111d are provided between the dynamic pressure grooves 111a, 111b, respectively. Sloping hill portions 11c, 11d and annular hill portions 11e, 11e are indicated by cross-hatching.

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.

In the conventional products (No. 1 to No. 3), the shaft member and bearing come into contact from 1000 rpm, and
It can be seen that the number of contacts increases dramatically when the rotational speed drops below 200 rpm. In contrast, with Example 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 bearing came into contact with each other 500 times or more in 60 seconds.

The presence or absence of contact between the shaft (shaft member 2) and the bearing (bearing sleeve 8) was measured when the load (load) applied to the shaft (shaft member 2) was gradually increased. In this case, the test pieces 2 (No. 1 to No. 3)
The rotation speed was 1000 rpm.
The presence or absence of contact between the shaft (shaft member 2) and the bearing (bearing sleeve 8) was measured when the load (load) applied to the shaft member was gradually increased, and 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 before this load, and therefore the number of contacts at the subsequent load will also exceed 500.

As can be seen from Table 4, the load at which frequent contact between the shaft member and 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 for implemented product 2 was higher than that of conventional product 2.

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 groups 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 (5)

A hydrodynamic bearing having an inner diameter surface of a bearing facing an outer diameter surface of a shaft member, the inner diameter surface of the bearing being provided with a hydrodynamic pressure generating portion,
a hydrodynamic bearing characterized in that the hydrodynamic pressure generating portion has a plurality of hydrodynamic pressure grooves arranged in a herringbone pattern, the hydrodynamic pressure grooves are provided in first and second hydrodynamic pressure groove groups, and a sloping hill portion is formed between each of the hydrodynamic pressure grooves of the first and second hydrodynamic pressure groove groups, and an annular hill portion is provided between the first and second hydrodynamic pressure groove groups, connecting with each of the sloping hill portions and extending in the circumferential direction, and a recess for generating hydrodynamic pressure is formed in a part of the annular hill portion. 2. A hydrodynamic bearing according to claim 1, characterized in that the circumferential width of the recess is smaller than the circumferential width of the confluence portion , the number of hydrodynamic grooves in the first and second hydrodynamic groove groups is the same, the confluence portion of the inclined hill portion and the annular hill portion in the first hydrodynamic groove group coincides with the confluence portion of the inclined hill portion and the annular hill portion in the second hydrodynamic groove group, and the recess is provided at all confluence portions. The hydrodynamic bearing according to claim 1, characterized in that a pair of upside-down hydrodynamic pressure generating portions spaced apart in the axial direction are formed on the inner diameter surface of the bearing. A fluid dynamic bearing device comprising the hydrodynamic bearing according to any one of claims 1 to 3, a shaft member inserted into the inner periphery of the hydrodynamic bearing, and a radial bearing portion that supports the relative rotation of the shaft member by the hydrodynamic action of the lubricating fluid in the radial bearing gap formed between the inner periphery of the hydrodynamic bearing and the outer periphery of the shaft member. 5. A motor comprising: the fluid dynamic bearing device according to claim 4; 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 .