Hydraulic dynamic bearing and spindle motor and rotary assembly provided

A double sleeve type dynamic bearing comprises a fixed shaft having at least one end fixedly mountable to an apparatus in which the bearing is utilized, a rotary sleeve arranged coaxially with the fixed shaft so that a first fine gap is formed therebetween, a fixed sleeve arranged coaxially with the rotary sleeve so that a second fine gap is formed therebetween, and a lubrication oil filled in the fine gaps. The first fine gap and the second fine gap each have an open end exposed to air outside the bearing and an opposite end that is not exposed to the air, the opposite ends being in communication with each other. A holding member holds the fixed shaft and the fixed sleeve and is disposed adjacent to a lower end surface of the rotary sleeve to form a third fine gap between the holding member and the lower end surface of the rotary sleeve. The third fine gap is formed with a thrust dynamic pressure producing groove, and opposite ends of the first and second fine gaps meet each other through the third fine gap. A peripheral surface of at least one of the fixed shaft, the rotary sleeve and the fixed sleeve forming at least one of the first and second fine gaps has a dynamic pressure producing groove formed therein.

BACKGROUND OF THE INVENTION

This invention relates to a fluid dynamic pressure bearing adapted as a bearing for rotary apparatuses such as hard disc drive (HDD) units, a spindle motor used as a drive source for such rotary apparatuses and a fluid dynamic pressure bearing adapted for such spindle motors, and more particularly to a both-end fixed-shaft type fluid dynamic pressure bearing having a shaft to be fixed at its respective ends on a chassis, etc. of an apparatus utilized through screwing or the like.

Air dynamic pressure bearings are broadly used in rotary apparatuses such as an HDD, drive optical disc and light polarizing units because of their excellent merits such as their light weight, clean and smooth rotation, durability to heat and cold, long service life and noncontamination to a recording media such as a disc by virtue of not using lubrication oil. In recent years, however, there has been a significant increase in information to be processed. Particularly, are large capacity HDD apparatuses required to rotationally drive as many as five or more disc. This requirement can no longer be met by an air dynamic pressure bearing. In order to cope with this, fluid dynamic pressure bearings have been adopted in HDD apparatuses to support greater load weight than that supported by the air dynamic pressure bearings.

The conventional fluid dynamic pressure bearings, particularly fluid dynamic pressure bearings of the sleeve rotation, type include two kinds of devices depending on the ways used to fix the shaft onto an apparatus in which it is utilized. One type is a one-end fixed-shaft type fluid dynamic pressure bearing as shown in FIG. 13 , and the other is a both-end fixed-shaft type fluid dynamic pressure bearing as shown in FIG. 14 . First, the fluid dynamic pressure bearing of FIG. 13 is structured by a fixed shaft 1 at its lower end fixed on a chassis 16 or the like through a screw 15 , and a rotary sleeve 2 having an upper end completely covered by a lid member 20 and a lower end having an opening 11 forming a capillary seal. Next, the fluid dynamic pressure bearing of FIG. 14 is structured by a fixed shaft 1 fixed at its opposite ends on a chassis 16 or the like of an apparatus utilized through screws 14 and 15 , and a rotary sleeve 2 having openings 11 a and 11 b respectively forming upper and lower capillary seals.

In FIG. 13 and FIG. 14 , 8 , 8 a , and 8 b are radial dynamic pressure producing grooves while 9 a and 9 b are thrust dynamic pressure producing grooves. 5 , 5 a , 5 b , 17 a , 17 b and 17 c are fine gaps formed between the fixed shaft 1 and the rotary sleeve 2 . These fine gaps are filled therein with lubrication oil 18 . The fine gaps have a width of usually 2 to 15 m, although depending on the size of the fluid dynamic pressure bearing. 13 a is an upper screw hole of the fixed shaft, while 13 , 13 b is a lower screw hole.

In the shaft-one-end fixed type fluid dynamic pressure bearing of FIG. 13 , the lubrication oil 18 filled within the fine gaps 5 , 17 a , 17 b and 17 c contacts with the air at tapered opening 11 . However, the lubrication oil 18 filled in the gaps is prevented from leaking outside the fine gaps by a capillary seal and surface tension due to the opening 11 . In particular, the fine gaps 17 a , 17 b and 17 c form a closed end.

The filled lubrication oil 18 hardly leaks out through the opening 11 due to a fine gap structure having such a closed end, i.e. a fine gap structure with one-side closure. In the both-end fixed-shaft type fluid dynamic pressure bearing of FIG. 19 , on the other hand, the lubrication oil 18 filled within the fine gaps 5 a , 5 b , 17 a , 17 b and 17 c contacts with the air at a tapered upper opening 11 a and lower opening 11 b . However, the filled lubrication oil 18 is prevented from leaking out of the fine gaps by the capillary seal and surface tension due to the openings.

Of the above related-art apparatus, the one-end fixed-shaft type fluid dynamic pressure bearing of FIG. 13 has a closed end in the fine gaps. Accordingly, the apparatus, in case tilted, hardly causes the lubrication oil to leak thus being excellent in sealability. However, there is a disadvantage in that the shaft 1 is fixed at only one point of its lower end and undergoes precession motion during rotation at high speed, resulting in instability in rotation. Conversely, the both-end fixed-shaft type dynamic pressure bearing of FIG. 14 fixes the shaft 1 at its both ends and hence does not undergo precession motion, offering stable rotation. However, there is a problem in that the fine gaps are opened to the air at upper and lower sides thus resulting in insufficient sealability. Even if a surface tension is formed by forming an air reservoir in a fine gap between the upper and lower radial dynamic pressure producing grooves 8 a and 8 b, the surface tension abruptly decreases when the fluid dynamic pressure bearing is tilted and positioned in a horizontal direction. Furthermore, when, in this state, temperature change or external impact is applied, the lubrication oil filled within the fine gap readily leaks to the outside.

SUMMARY OF THE INVENTION

It is an object of the present invention to maintain, in a both-end fixed-shaft type fluid dynamic bearing, a high sealability not only during rotation at high speed but also even upon being tilted in a standstill state.

It is another object of the invention to provide a spindle motor which can stably rotate at high speed.

It is still another object of the invention to provide a rotary apparatus which can stably rotationally drive at high speed a rotary member such as a hard disc.

In brief, the present invention is a double sleeve type fluid dynamic pressure bearing, comprising: a fixed shaft having respective ends to be fixed to an apparatus utilized; a rotary sleeve arranged to provide a first fine gap between an inner peripheral surface thereof and an outer peripheral surface of the fixed shaft; a fixed sleeve arranged to provide a second fine gap between an inner peripheral surface thereof and an outer peripheral surface of the rotary sleeve; and wherein the first fine gap and the second fine gap have one ends made as open ends contacting the air while the first fine gap and the second fine gap have the other ends made as closed ends in direct communication with each other, the fine gaps being filled with lubrication oil, and the first fine gap being formed with a dynamic pressure producing groove.

In the double sleeve type dynamic pressure bearing, the second fine gap is greater in width than the first fine gap within a range of capable of producing a dynamic pressure, thereby removing instability during high speed rotation due to a difference in flowing speed of lubrication oil.

In the double sleeve type fluid dynamic pressure bearing, seal means different from a related art capillary seal is provided at the opening of the second fine gap. The seal means is a resinous collar fitted at an outer end of the fixed sleeve. Alternatively, the seal means may be a curved type annular seal groove formed in the opening of the second fine gap by a first curved wall surface curved radially outward and a second curved wall surface similarly curved radially outward. Furthermore, the seal means may be a multi-staged slant type annular seal groove formed in the opening of the second fine gap by a first plurality slant wall surface having a plurality of annular slant surfaces slanted by stages radially outward and a second plurality slant wall surface having a plurality of annular slant surfaces similarly slanted by stages radially outward.

Also, the present invention is, in a spindle motor structured by a rotor including a rotor magnet, a stator including a stator coil and a fluid dynamic pressure bearing for rotatably supporting the rotor with respect to the stator, the spindle motor adopting for the fluid dynamic pressure bearing a double sleeve type fluid dynamic pressure bearing. The invention is furthermore a rotary apparatus having, as drive source to a rotary member, the spindle motor structured by a rotor including a rotor magnet, a stator including a stator coil and a fluid dynamic pressure bearing for rotatably supporting the rotor with respect to the stator.

BEST MODE FOR PRACTICING THE INVENTION

Referring to FIG. 1 , there is shown a sectional view of a double sleeve structure both-end fixed-shaft type dynamic pressure bearing according a first embodiment of the present invention, and a spindle motor having this fluid dynamic pressure bearing. In FIG. 1 , the fluid dynamic pressure bearing includes a fixed shaft 1 fixed at respective ends to an apparatus utilized, a rotary sleeve 2 providing a first fine gap 5 cooperatively with the fixed shaft 1 , a fixed sleeve 3 providing a second fine gap 6 cooperatively with the rotary sleeve 2 , and a holder member 4 providing a third fine gap 7 cooperatively with the rotary sleeve 2 .

The fixed shaft 1 is formed with screw holes 13 a and 13 b at respective ends. The fixed shaft 1 is firmly fixed to a chassis 16 of an apparatus utilized such as an HDD apparatus through screws 14 and 15 screwed to the screw holes 13 a and 13 b . The rotary sleeve 2 is a member formed by a sleeve portion 2 a having inner and outer peripheral surfaces, a cup-like hub 2 b for holding a rotary member such as a disc, and a disc formed extended portion 2 c for firmly fixing the cup-like hub 2 b at an upper end of the sleeve 2 a . The disc-formed extended portion 2 c is a portion in a disc form that is horizontally radially outwardly extended from an upper end of a sleeve portion 2 a of the rotary sleeve 2 , and formed integral with the sleeve 2 a . The cup-like hub 2 b serves also as a rotor member for the spindle motor having a rotor magnet 22 mounted on an inner peripheral surface thereof. The fixed sleeve 3 is a member arranged standing on the base plate of the bearing or spindle motor. In the apparatus of the FIG. 1 embodiment without using a base plate, the fixed sleeve 3 is provided standing adjacent the holder member 4 with its inner peripheral surface fitted liquid-tight to an outer peripheral surface of the disc-like holder member 4 coaxially fixed to the fixed shaft 1 . The fixed sleeve 3 also serves as a stator member for the spindle motor, and has a stator coil 23 mounted on an outer peripheral surface thereof.

A tapered opening 11 is provided at a top end of a first fine gap 5 formed between an outer peripheral surface of the fixed shaft 1 and an inner peripheral surface of the rotary sleeve 2 . Similarly, a tapered opening 12 is also provided at a top end of a second fine gap 6 given between an outer peripheral surface of the rotary sleeve 2 and an inner peripheral surface of the fixed sleeve 3 . A third fine gap 7 is given between an lower end surface of the rotary sleeve 2 and an upper surface of the disc-like holder member 4 , which has one end communicated with a lower end of the first fine gap 5 and the other end communicated with a lower end of the second fine gap 6 . In brief, the third fine gap 7 serves as a closed end with respect to the openings 11 and 12 . Lubrication oil 18 is filled within the first fine gap 5 , second fine gap 6 and third fine gap 7 .

Each of these fine gaps, although exaggeratedly shown in FIG. 1 , is actually a fine gap of a size of approximately 5 to 200 m. Due to this, the lubrication oil 18 has its liquid levels respectively kept at bottom portions of the tapered openings 11 and 12 by a surface tension and capillary phenomenon, being prevented from leaking to the outside in a usual use state. Moreover, the first fine gap 5 and the second fine gap 6 at their lower ends are communicated through the third fine gap 7 , forming a closed end. Accordingly, even in case the fluid dynamic pressure bearing of the invention is tilted, the lubrication oil 18 filled within these fine gaps hardly leaks to the outside.

Radial dynamic pressure producing grooves 8 a and 8 b are provided vertically separated in an inner peripheral surface of the sleeve portion 2 a of the rotary sleeve 2 forming the first fine gap 5 . A thrust dynamic pressure producing groove 9 is provided in a lower end surface of the sleeve portion 2 a of the rotary sleeve 2 forming the third fine gap 7 . The radial dynamic pressure producing grooves 8 a and 8 b are herringbone grooves but may be of other forms. The thrust dynamic pressure producing groove 9 are herringbone grooves in an annular arrangement but may be of other forms.

The spindle motor thus constructed, when supplied by an energizing current to its stator coil 23 , is rotated due to electromagnetic action caused by the current and magnetic field of the rotor magnet 22 . Thereupon, a radial dynamic pressure is caused in the first fine gap 5 through the radial dynamic pressure producing grooves 8 a and 8 b , and a thrust dynamic pressure is caused in the third fine gap 7 through the thrust dynamic pressure producing groove 9 . Thus, the spindle motor smoothly maintains rotation at high speed while supporting a rotary member, such as a hard disc, by these dynamic pressures.

The double sleeve structure both-end fixed-shaft type fluid dynamic pressure bearing of the invention explained above was structured by adopting a circular columnar member for the shaft 1 , a cylindrical member for the rotary sleeve 2 , and a cylindrical member for the fixed sleeve 3 . Now, among various modifications, four embodiments will be described with reference to FIG. 2 to FIG. 5 exaggeratedly showing the fine gaps.

Referring to FIG. 2 , a double sleeve structure bothend fixed-shaft type fluid dynamic pressure bearing is shown as a modification for the fluid dynamic pressure bearing of FIG. 1 . That is, the double sleeve structure both-end fixed-shaft type fluid dynamic pressure bearing of FIG. 2 is structured by adopting a circular columnar member for a fixed shaft 1 , a conical frustum member for a rotary sleeve 2 , and a conical-inner-perpheral-surfaced member for a fixed sleeve 3 to cooperate with an conical outer peripheral surface to provide therebetween a second fine gap 6 . A first fine gap 5 is given between an outer peripheral surface of the fixed shaft 1 and an inner peripheral surface of the rotary sleeve 2 . A third fine gap 7 is given between a bottom surface of the rotary sleeve 2 and a top surface of the holder member 4 , to communicate between respective lower ends of the first fine gap 5 and the second fine gap 6 serving as a closed end. The first fine gap 5 and the second fine gap 6 respectively have, at their upper ends, tapered openings 11 and 12 serving as capillary seals for the lubrication oil filled within the fine gaps.

Referring to FIG. 3 , a double sleeve structure both-end fixed-shaft type dynamic pressure bearing is shown as a second modification for the fluid dynamic pressure bearing of FIG. 1 . That is, the double sleeve structure both-end fixed-shaft type fluid dynamic pressure bearing of FIG. 3 is structured by adopting a circular columnar member for a fixed shaft 1 , a member with a semi-spherical portion for a rotary sleeve 2 , and a member with an inner peripheral surface providing a second fine gap 6 cooperatively with the semispherical surface of the rotary sleeve 2 for a fixed sleeve 3 . The second gap 6 includes a semi-circular gap portion 6 a and a horizontal gap 6 b . A first fine gap 5 is given between an outer peripheral surface of the fixed shaft 1 and an inner peripheral surface of the rotary sleeve 2 . The first fine gap 5 and the second fine gap 6 at their lower ends are directly communicated with each other thereby forming a closed end. The first fine gap 5 and the second fine gap 6 respectively have, at their upper ends, tapered openings 11 and 12 serving as capillary seals for the lubrication oil filled within the fine gaps.

In the meanwhile, it will be understood from FIG. 1 , FIG. 2 and FIG. 3 that a variety of double sleeve structure both-end fixed-shaft type fluid dynamic pressure bearing can be realized depending on a member shape of the rotary sleeve 2 even with the fixed shaft 1 using a circular columnar member. Also, it will be apparent that the rotary sleeve 2 is not limited to the cylindrical member of FIG. 1 , the conical frustum member of FIG. 2 or the semi-spherical-portion-having member of FIG. 3 but may adopt various shapes of members including deformed versions of cylindrical, conical frustum and disc members.

Referring to FIG. 4 , a double sleeve structure both-end fixed-shaft type dynamic pressure bearing is shown as a third modification for the fluid dynamic pressure bearing of FIG. 1 . That is, the double sleeve structure both-end fixed-shaft type fluid dynamic pressure bearing of FIG. 4 is structured by adopting a circular columnar member having upper and lower large diameter portions for a fixed shaft 1 , a member having an inner peripheral surface to provide a first fine gap 5 cooperatively with an outer peripheral surface of the fixed shaft 1 for a rotary sleeve 2 , and a cylindrical member for a fixed sleeve 3 . The first fine gap 5 includes a plurality of vertical gap portions 5 a and a plurality of horizontal gap portions 5 b . A second fine gap 6 is given between an outer peripheral surface of the rotary sleeve 2 and an inner peripheral surface of the fixed sleeve 3 . The first fine gap 5 and the second fine gap 6 at their lower ends are communicated with each other to provide a third fine gap 7 serving as a closed end between a bottom surface of the rotary sleeve 2 and a top surface of the holder member 4 . The first fine gap 5 and the second fine gap 6 respectively have, at their upper ends, tapered openings 11 and 12 serving as capillary seals for the lubrication oil filled within the fine gaps.

Referring to FIG. 5 , a double sleeve structure both-end fixed-shaft type dynamic pressure bearing is shown as a fourth modification for the fluid dynamic pressure bearing of FIG. 1 . That is, the double sleeve structure both-end fixed-shaft type fluid dynamic pressure bearing of FIG. 5 is structured by adopting a circular columnar member having at its intermediate portion a small diameter portion for a fixed shaft 1 , a member having an inner peripheral surface to provide a first fine gap 5 cooperatively with an outer peripheral surface of the fixed shaft 1 for a rotary sleeve 2 , and a cylindrical member for a fixed sleeve 3 . The first fine gap 5 includes a plurality of vertical gap portions 5 a and a plurality of horizontal gap portions 5 b . A second fine gap 6 is given between an outer peripheral surface of the rotary sleeve 2 and an inner peripheral surface of the fixed sleeve 3 . The first fine gap 5 and the second fine gap 6 at their lower ends are communicated with each other to provide a third fine gap 7 serving as a closed end between a bottom surface of the rotary sleeve 2 and a top surface of a holder member 4 . The first fine gap 5 and the second fine gap 6 respectively have, at their upper ends, tapered openings 11 and 12 serving as capillary seals for the lubrication oil filled within the fine gaps.

Referring to FIG. 6 , there is shown a sectional view of a spindle motor having a second embodiment of a fluid dynamic pressure bearing of the present invention.

The fluid dynamic pressure bearing of FIG. 6 is the same in basic structure as the fluid dynamic pressure bearing of FIG. 1 , but different in structure of the second fine gap 6 . That is, the fluid dynamic pressure bearing of FIG. 6 has a second fine gap 6 that is wider in gap width than the second fine gap 6 of the fluid dynamic pressure bearing of FIG. 1 .

Because the second fine gap 6 is located radially outward of the first fine gap 5 , the speed of flowing lubrication oil is not equal between these fine gaps. The speed is greater in the second fine gap 6 . It was found that the difference in speed impedes stability during high speed rotation in an actual apparatus and that a high flow speed of the lubrication oil 18 through the second gap 6 increases friction and hence loss.

Accordingly, the second fine gap 6 was made wider than the first fine gap 5 as shown in an exaggerated manner in FIG. 6 . That is, the first fine gap 5 has a width of approximately 5 to 20 m while the second fine gap 6 has a width of approximately 50 to 500 m. This removed the instability at high speed rotation due to a difference in speed of lubrication oil. The width of the second fine gap is selected on an empirical basis as best suited depending on the structure and size of the fluid dynamic pressure bearing, lubrication oil property, and so on. Such a size includes a height of the first fine gap 5 and a height of the second fine gap 6 . Incidentally, a width of the third fine gap 7 is selected as approximately same as the first fine gap 5 .

Stable high speed rotation was thus realized in the spindle motor adopting the fluid dynamic pressure bearing that is stable at high speed rotation and low in loss. Accordingly, a large capacity HDD apparatus having this spindle motor as a drive source for a rotation member can rotate a hard disc at high speed smoothly and stably.

Referring to FIG. 7 , there is shown a sectional view of a spindle motor having a third embodiment of a fluid dynamic pressure bearing of the present invention.

The fluid dynamic pressure bearing of FIG. 7 is the same in basic structure as the fluid dynamic pressure bearing of FIG. 1 , but different in fluid seal structure at an opening of the second fine gap 6 . That is, the fluid dynamic pressure bearing of FIG. 1 had, at the opening 12 of the second fine gap 6 , the fluid seal structure made as a capillary seal structure utilizing a slant type annular seal groove having a groove width continuously increasing from its minimum groove portion toward maximum groove width portion, i.e. seal groove formed by the slanted inner peripheral surface of the fixed sleeve 3 extended outward in a continuous fashion and the vertical outer peripheral surface of the rotary sleeve 2 . On the contrary, the fluid dynamic pressure bearing of FIG. 7 has, at an opening 12 A of a second fine gap 6 , a fluid seal structure made in a fluid seal structure utilizing a resin collar 24 . This resin collar 24 is a ring-formed Teflon resin formed member that is nearly same in inner and outer diameters as the fixed sleeve 3 .

The resin collar 24 , as shown in the magnified view of FIG. 8 , is formed at its axially lower side with a fitting annular leg 24 A smaller in outer diameter than its upper side. on the other hand, the fixed sleeve 3 which provides the second fine gap 6 cooperatively with the rotary sleeve 2 is formed at its end with a fitting annular step 3 A having an inner diameter equal to or slightly smaller than the outer diameter of the fitting annular leg 24 A and a length equal to that of the fitting annular leg 24 A. The resin collar 24 at its fitting annular leg 24 A is press-fitted into the fitting annular step 3 A of the fixing sleeve 3 , thereby closely fitting and fixing the resin collar 24 in the end of the fixing groove 3 .

The provision of the resin collar 24 improves wettability of lubrication oil 18 in an area close to the opening 12 A of the second fine gap 6 . Consequently, lubrication oil 18 reaches to a position close to the opening 12 A of the second fine gap 6 . The lubrication oil 18 contacting an inner peripheral surface of the resin collar 24 has an arcuate recess 18 B in a surface smaller in recessing than an arcuate recess 18 A in a surface of a lubrication oil 18 reaching close to the opening 12 A of the second fine gap 6 but not contacting with the inner peripheral surface of the resin collar 24 . In other words, an angle 1 defined between a tangent line of the recess 18 B at a point where the lubrication oil 18 contacts the wall surface of the annular seal groove and the wall surface of the annular seal groove is greater than an angle 0 defined between a tangent line of the recess 18 A at a point where the lubrication oil 18 contacts the vertical wall surface of the fine gap 6 and the vertical wall surface of the fine gap 6 . This is because, where contacting the inner peripheral surface of the resin collar 24 higher in wettability than metal, an increased surface tension acts on the lubrication oil 18 than in a case other than the above. Due to the action of such an increased surface tension, the sealability was improved for the lubrication oil 18 at the opening 12 A of the second fine gap 6 .

Referring to FIG. 9 , there is shown a magnified view of a modification to the resin collar 24 . A resin collar 24 is the same in basic structure as that of FIG. 8 but different in its inner peripheral surface. That is, the resin collar 24 of FIG. 8 had the inner peripheral surface as a cylindrical surface having its the same inner diameter as the inner peripheral surface of the fixed sleeve 3 . On the contrary, the resin collar 24 of FIG. 9 has an inner peripheral surface formed by a lower vertical inner peripheral portion 24 B and a upper stanted inner peripheral surface 24 C. The vertical inner peripheral portion 24 B has a same inner diameter as that of an inner peripheral surface of the fixed sleeve 3 . The slanted inner peripheral portion 24 C is a tapered surface having an inner diameter increasing continuously from a boundary to the vertical inner peripheral surface portion 24 B toward an opening end that is opened to the air.

In FIG. 9 , the resin collar 24 at its fitting annular leg 24 A is press-fitted into the fitting annular step 3 A of the fixing sleeve 3 , thereby closely fitting and fixing the resin collar 24 to an end of the fixed sleeve 3 .

The provision of the fluid seal of FIG. 9 using the resin collar 24 to the opening 12 of the second fine gap 6 further improved the sealability for the lubrication oil 18 in the second fine gap 6 . That is, the sealability for the lubrication oil 18 in the second fine gap 6 was further improved by the seal due to an increased surface tension given by a good wettability of the resin collar 24 and the capillary seal utilizing a slant type annular seal groove formed by the slanted inner peripheral surface portion 24 C and the corresponding vertical outer peripheral surface of the rotary sleeve 2 .

Incidentally, the material of the resin collar 24 may be a polyimide based resin or fluorine based resin, instead of the Teflon based resin. Polyimide based resin exhibits equivalent wettability to that of fluorine based resin, and is excellent in formability.

Referring to FIG. 10 , a sectional view of a spindle motor is shown having a fourth embodiment of a fluid dynamic pressure bearing of the invention.

The fluid dynamic pressure bearing of FIG. 10 is the same in basic structure as the fluid dynamic pressure bearing of FIG. 1 . The different point lies in a structure of an annular seal groove formed in an opening in a second fine gap 6 to have a groove width continuously extending radially outward. That is, in the fluid dynamic pressure bearing of FIG. 1 , the fixed sleeve 3 at its end inner peripheral surface was made as the annular slant inner peripheral surface and further the corresponding outer peripheral surface of the rotary sleeve 2 to the end inner peripheral surface of the fixed sleeve 3 was made in the vertical outer peripheral surface, whereby the slant annular seal groove was formed having a groove width continuously extending radially outward toward the opening 12 of the second fine gap 6 . On the contrary, the fluid dynamic pressure bearing of FIG. 10 adopts a curved-type annular seal groove 12 B different from the above.

That is, the fluid dynamic pressure bearing of FIG. 10 has a fluid seal structure formed by a curved-type annular seal groove 12 B in an opening of a second fine gap 6 and having a groove width continuously extending radially outward, and structured by a first curved wall surface 25 and second curved wall surface 26 both curved radially outward. As shown in FIG. 11 as a partly magnified view of FIG. 10 , the first curved wall surface 25 is formed at an end of the fixed sleeve 3 while the second curved wall surface 26 is formed in an opposed outer peripheral surface of the rotary sleeve 2 to the first curved wall surface 25 , i.e. in a wall surface connecting a vertical outer peripheral surface of the rotary sleeve 2 and an underside of a horizontal disc-formed extending portion 2 c . The first curved wall surface 25 formed in the fixed sleeve 3 and the second curved wall surface 26 formed in the rotary sleeve 2 have respective radii of curvature determined such that the groove width continuously increases from a minimum groove width portion toward a maximum groove width portion of the curved type annular seal groove 12 B.

The provision of the curved-type annular seal groove 12 B at the opening of the second fine gap 6 improved the lubrication oil sealability at that portion. That is, the lubrication oil 18 reaching the curved-type annular seal groove 12 B has an arcuate recess 18 C in a surface in smaller in recessing than an arcuate recess 18 A in a surface of a lubrication oil 18 reaching close to the opening 12 of the second fine gap 6 but not reaching the curved-type annular seal groove 12 B. In other words, an angle 2 defined between a tangent line of the recess 18 C at a point where the lubrication oil 18 contacts the curved wall surface of the curved-type annular seal groove 12 B and the curved wall surface of the curved-type annular seal groove 12 B is greater than an angle 0 a defined between a tangent line of the recess 18 A at a point where the lubrication oil 18 contacts the vertical wall surface of the fine gap 6 and the vertical wall surface of the fine gap 6 . This is because of increase in surface tension at this portion by curving the pair of wall surfaces of the curved-type annular seal groove 12 B with a determined radii of curvature so that the groove width continuously increases from a minimum groove width portion to a maximum groove width portion. Due to the action of such an increased surface tension, the sealability was improved for the lubrication oil 18 at the opening 12 of the second fine gap 6 .

Referring to FIG. 12 , a multi-staged slant type annular seal groove 12 C is shown in a partially magnifying view which is a modification for the curved-type annular seal groove 12 B of FIG. 11 . The multi-staged slant type annular seal groove 12 C has a groove width increasing by stages from its minimum groove width portion toward maximum groove width portion, and formed by a first curved wall surface 25 having a plurality of annular slant surface 25 a , 25 b slanted by stages adially outward and a second curved surface 26 having a plurality of annular slant surface 26 a , 26 b also slanted by stages adially outward. The first curved wall surface 25 is formed in an end of a fixed sleeve 3 , while the second curved wall surface 26 is formed in an opposed outer peripheral surface of a rotary sleeve 2 to the first curved surface, i.e. in a wall surface of the rotary sleeve 2 connecting a vertical outer peripheral surface and an underside of a horizontal disc-formed extending portion 2 c.

The provision of the multi-staged slant type annular seal groove 12 C at the opening of the second fine groove 6 improved the lubrication oil sealability at this portion. That is, the lubrication oil 18 reaching the multi-staged slant type annular seal groove 12 C has a in arcuate recess 18 D in surface in smaller in recessing than in arcuate recess 18 A an surface of a lubrication oil 18 reaching close to the opening 12 of the second fine gap 6 but not reaching the multi-staged slant type annular seal groove 12 C. In other words, an angle 3 defined between a tangent line of the recess 18 D at a point where lubrication oil 18 contacts the annular slant wall surface of the multi-staged slant type annular seal groove 12 C and the annular slant wall surface is greater than an angle 0 defined between a tangent line of the recess 18 A at a point where the lubrication oil 18 contacts the vertical wall surface of the fine gap 6 and the vertical wall surface. This is because of increase in surface tension at this portion due to the multi-staged slant type annular seal groove 12 C having a groove width increasing by stages from a minimum groove width portion to a maximum groove width portion structured by a first plurality slant wall surface having a plurality of annular slant surfaces slanted by stages radially outward and a second plurality slant wall surface having a plurality of annular slant surfaces similarly slanted by stages radially outward. Due to the action of such an increased surface tension, the sealability was improved for the lubrication oil 18 at the opening of the second fine gap 6 .

The multi-staged slant type seal groove 12 C functions as above and hence constitutes so-called a multi-staged capillary seal. Accordingly, the angle 3 defined between a tangent line of the recess 18 D at a point where lubrication oil 18 contacts the annular slant wall surface of the multi-staged slant type annular seal groove 12 C and the annular slant wall surface is increased at a boundary to an adjacent capillary seal stage. Thus, the surface tension is increased at this boundary.

In case that the curved-type annular seal groove 12 B of FIG. 11 or multi-staged slant type annular seal groove 12 C of FIG. 12 is provided in the opening of the second fine gap 6 , it is possible to enhance the mechanical strength for a base portion 2 d of the horizontal disc-formed extended portion 2 c integrally coupling a cup-like hub portion 2 b to a sleeve portion 2 a of the rotary sleeve 2 . For example, in the fluid dynamic pressure bearing, the rotary sleeve 2 at its upper end inner peripheral surface 2 e is made as an annular slant surface, in order to provide a capillary seal in the opening 11 of the first fine gap 5 , as is clear by reference to FIG. 8 . Due to this, the base portion 2 d is reduced in thickness T 0 . Contrary to this, the base portion 2 d is enhanced in mechanical strength because greater than the thickness T 0 is a thickness T 1 of the base portion 2 d where a curved type annular seal groove 12 B is provided in the opening of the second fine gap 6 and also a thickness T 2 of the base portion 2 d where a multi-staged slant type annular seal groove 12 C is provided in the opening of the second fine gap 6 .

Although the present invention was explained hereinabove by way of various embodiments, the scope of the invention should never be limited to these embodiment but be defined by the inventions set forth in the claims attached herewith and those of their equivalencies.