Spindle motor and disk drive utilizing the spindle motor

Low-profile spindle motor whose entire shaft length is utilized to configure, along an encompassing sleeve, a radial dynamic-pressure bearing section. One end of the shaft is unitary with the rotor, and a cover member closes the other end. Between the sleeve upper-end face and the rotor undersurface a thrust bearing section is configured. Micro-gaps are formed continuing between the sleeve upper-end face and the rotor undersurface; the sleeve inner-circumferential surface and the shaft outer-circumferential surface; and the cover member inner face and the shaft end face, where an axial support section is established. Oil continuously fills the micro-gaps, configuring a full-fill hydrodynamic bearing structure. Hydrodynamic pressure-generating grooves in the radial bearing section are configured either so that no axial flow, or so that a unidirectional flow that recirculates from one to the other axial end of the radial bearing section through a communicating pathway is induced in the oil.

BACKGROUND OF INVENTION

1. Technical Field

The present invention relates to spindle motors and disk-drive devices utilizing the spindle motors; in particular to low-profile spindle motors furnished with hydrodynamic bearings, and to disk-drive devices utilizing the spindle motors.

2. Description of Related Art

In hard-disk drives that drive hard disks and like recording disks, spindle motors utilizing hydrodynamic bearings that, in order to support the shaft and sleeve as either one rotates relative to the other, employ the fluid pressure of a lubricating fluid such as oil interposed between the two are known.

With regard to spindle motors utilizing hydrodynamic bearings of this sort, the applicant in the present application has proposed, in Japanese Laid-Open Pat. App. No. 2000-113582, a spindle motor as illustrated in FIG.1. Between the bottom face of a rotor100and the top-end face of a sleeve102in the spindle motor depicted inFIG. 1, a thrust bearing section104is configured. Likewise, between the outer circumferential surface of a shaft106furnished integrally with the rotor100, and the inner circumferential surface of the sleeve102, radial bearing sections108,108are configured. The thrust bearing section104generates lifting force on the rotor100, and the radial bearing sections108,108function to center-balance in the radial direction, and prevent wobble in, the rotor100.

The spindle motor depicted inFIG. 1makes the thrust plate that would be a component of the thrust bearing in conventional hydrodynamic bearings unnecessary. The consequent advantage is a simplified structure that reduces the cost of the motor and at the same time enables it to be slimmed, without appreciably compromising the bearing rigidity. Nevertheless, with the advent of the application of disk drives in miniature devices such as portable information terminals, demands are on the rise to make the spindle motors used in the disk drives even slimmer. In addition, calls for lowering the cost of spindle motors still further have gone hand in hand with reducing the cost of disk drives.

Running counter to this is the fact that in its sleeve102the spindle motor depicted inFIG. 1is provided with a communicating passage110made up of a through-hole110aand channels110b,110c. The communicating passage110brings outside air into the bearing areas—that is, it enables air to circulate into and out of the bearing areas—and thus would expose the end portions of the radial bearing sections108,108to the air. Due to the pumping action of dynamic-pressure-generating grooves formed in each bearing section, areas in which the internal pressure of the oil retained among the bearing sections becomes negative, i.e., at pressure less than atmospheric pressure, arise. Upon a decrease in the internal pressure of the oil to a negative pressure level, air that is entrained in the oil during the process of charging the bearing sections with oil, or that is present due to being swept in by the dynamic-pressure-generating grooves, appears in the form of bubbles. The volume of the bubbles expands with increasing temperature or decreasing external environmental pressure. The volume expansion of the bubbles brings leaking oil toward the exterior of the bearing sections and impairs the spindle motor's durability and reliability. Furthermore, the dynamic-pressure-generating grooves that are formed in the bearing sections come into contact with the bubbles, which causes vibrations and worsens non-repeatable run-out. The rotational precision of the spindle motor therefore worsens. Accordingly, the spindle motor configuration includes the communicating passage110in order to exhaust bubbles to the exterior of the bearing sections.

To bore the communicating passage110for discharging bubbles in this way a drilling tool is used. The drill bit can only be so small, however, to be strong enough for machining, which limits how small the through-hole110aand the channels110b,110cthat constitute the communicating passage110can be made. Consequently, the axial dimension of the shaft106and the sleeve102must necessarily be at least a given size for boring the communicating passage110and be extensive enough to maintain bearing rigidity in the radial bearing sections108,108. These requirements stand in the way of making the spindle motor slimmer.

What is more, the fact that the through-hole110aas well as the channels110b,110cthat constitute the communicating passage110are formed in the sleeve102complicates that part of the structure and at the same time increases the number of manufacturing processes. An increased-cost spindle motor is the result.

Further still, a ring element112that constitutes a retainer for the rotor100is fitted onto the end portion of the shaft106on the side opposite the rotor100. In short, because the thrust bearing section104; the radial bearing sections108,108; the through-hole110aas well as the channels110b,110cthat constitute the communicating passage110; and the ring element112are arranged in the axial direction stacked along the same axis, they create an impediment to making the spindle motor slimmer.

SUMMARY OF INVENTION

An object of the present invention is to simplify and slim down the structure of a spindle motor while maintaining its rotational stability.

Another object is in a spindle motor to maintain the internal pressure of the oil retained within the bearing gaps at or above atmospheric pressure, to enable preventing the occurrence of air bubbles within the oil.

Yet another object is balancing the internal pressure of the oil retained within the bearing gaps of a spindle motor.

A different object of the present invention is to enable preventing particulate matter from being produced due to contact between the rotor and stator components in a spindle motor.

Moreover, the present invention provides a low-profile, low-cost disk drive that can spin recording disks stably; and another object of the present invention accordingly is to enable preventing the occurrence of read/write errors that originate in oil leaking out from, or in particulate matter being produced by, the spindle motor in a disk drive device.

One example of a spindle motor under the present invention is configured with a radial dynamic-pressure bearing section, in between the inner circumferential surface of the sleeve and the outer circumferential surface of the shaft, that induces hydrodynamic pressure in oil during rotation of the rotor. On at least either one of the upper-end face of the sleeve, or the bottom face of the rotor, the motor is also furnished with dynamic-pressure-generating grooves, configuring a thrust bearing section, that impart radially inward-heading pressure to the oil during rotation of the rotor. In addition, at its tip end the shaft is configured with an axial support section in which pressure that essentially balances with the oil pressure within the thrust bearing section is utilized.

Likewise, in another example of a spindle motor under the present invention, the shaft is formed unitarily with the rotor, wherein a round tubular casing member whose outer peripheral surface functions as a radial bearing surface is attached to the outer peripheral surface of the shaft. A communicating pathway is formed in between the outer circumferential surface of the shaft and the inner circumferential surface of the casing member, enabling axial upper and lower ends of a radial bearing section formed in between the outer peripheral surface of the casing member and the inner peripheral surface of the sleeve to communicate.

Moreover, in a different example of a spindle motor under present invention, a thrust bearing is configured in between the upper-end face of the sleeve, and the bottom-face of the hub, and a radial dynamic bearing is configured in between the inner circumferential surface of the sleeve and the outer circumferential surface of the shaft. Along its outer circumferential surface the sleeve is furnished with a radially flaring annular flange portion, while on the inner circumferential surface of a round-cylindrical wall on the rotor, an annular member whose surface at least is harder than the sleeve is fixedly fitted. The flange portion and the annular member engage with each other to form a rotor retainer.

In one example of a disk drive under the present invention, the spindle motor that spins recording disks includes: a radial dynamic-pressure bearing section, in between the inner circumferential surface of the sleeve and the outer circumferential surface of the shaft, that induces hydrodynamic pressure in oil during rotation of the rotor; and also a thrust bearing section provided with dynamic-pressure-generating grooves, on at least either one of the upper-end face of the sleeve or the bottom face of the rotor, that impart radially inward-heading pressure to the oil during rotation of the rotor. In addition, at its tip end the shaft has an axial support section in which pressure that essentially balances with the oil pressure within the thrust bearing section is utilized.

Likewise, in another example of a disk drive under the present invention, the shaft is formed unitarily with the rotor in the disk-drive spindle motor for spinning recording disks, wherein a round tubular casing member whose outer peripheral surface functions as a radial bearing surface is attached to the outer peripheral surface of the shaft. A communicating pathway is formed in between the outer circumferential surface of the shaft and the inner circumferential surface of the casing member, enabling axial upper and lower ends of a radial bearing section formed in between the outer peripheral surface of the casing member and the inner peripheral surface of the sleeve to communicate.

Moreover, in a different example of a disk drive under present invention, the spindle motor that spins recording disks includes: a thrust bearing configured in between the upper-end face of the sleeve, and the bottom-face of the hub; and a radial dynamic bearing configured in between the inner circumferential surface of the sleeve and the outer circumferential surface of the shaft. A radially flaring annular flange portion is furnished on the outer circumferential surface of the sleeve, while on the inner circumferential surface of a rotor round-cylindrical wall, an annular member whose surface is at least harder than the sleeve is fixedly fitted. The flange portion and the annular member engage with each other to form a rotor retainer.

DETAILED DESCRIPTION

First Embodiment

(1) Spindle Motor Configuration

Reference is made toFIG. 2, which illustrates a spindle motor in a first embodiment of the present invention. Set forth inFIG. 2, the spindle motor is furnished with: a rotor6made up of a rotor hub2—composed of an approximately disk-shaped top wall portion2a, and a round-cylindrical peripheral wall portion2bdepending downward from the outer rim of the top wall portion2a—and of a shaft4one end portion4aof which is perimetrically inserted into the central portion of the top wall portion2aof the rotor hub2; a hollow, round cylindrical sleeve8rotatively supporting the shaft4; a sealing cap10opposing the end face of the shaft4along its free end, and closing over the lower portion of the sleeve8; and a bracket12formed integrally with a round cylindrical portion12afor anchoring the sleeve8.

The bracket12has a round cupped portion centered on the round cylindrical portion12a; and a stator14having a plurality of teeth that project radially inward is arranged on the inner circumferential surface12bof a peripheral wall that the outer circumferential edge of the cupped portion defines. Likewise, a rotor magnet16that opposes the stator14via a radially inward clearance therefrom is fixedly fitted to the outer circumferential surface of the peripheral wall portion2bof the rotor hub2.

A flange-shaped disk-mounting portion2cfor carrying recording disks on which information is recorded (inFIG. 6, represented as recording disks53) is formed on an outer circumferential portion of the peripheral wall portion2bof the rotor hub2. A threaded hole4bis formed in the upper-end portion of the shaft4(its end at the top wall portion2aof the rotor hub2). The recording disks are loaded onto the disk-mounting portion2c, and after being retained by a clamp (not illustrated), the recording disks are fixedly secured to the rotor hub2by fastening a screw (not illustrated) into the threaded hole4b.

An unbroken series of micro-gaps is formed in between the upper-end face of the sleeve8and the undersurface of the top wall portion2aof the rotor hub2, and—continuing from the top wall portion2aof the rotor hub2—in between outer circumferential surface of the shaft4and the inner circumferential surface of the sleeve8, and continuous therewith, in between the end face of the shaft4and the inner face of the sealing cap10. Oil continuously fills the micro-gaps without interruption, configuring a so-called full-fill hydrodynamic bearing structure. In this respect, the configuration of the bearings and their supporting function will be described in detail later.

The upper-end portion of the sleeve8outer circumferential surface is made into an annular flange portion8athat flares radially outward and that is contoured into an incline such that the outer circumferential surface contracts parting away from the upper-end face of the sleeve8. The flange portion8aradially opposes, without being in contact with, the inner circumferential surface of the peripheral wall portion2bof the rotor hub2.

Because as noted above the outer circumferential surface of the flange portion8ais contoured into an incline, the gap defined in between the inner circumferential surface of the peripheral wall portion2band the outer circumferential surface of the flange portion8aforms a taper whose radial clearance gradually increases heading axially downward (in the direction toward the distal rim of the peripheral wall portion2b). In particular, the inner circumferential surface of the peripheral wall portion2band the outer circumferential surface of the flange portion8acooperate to configure a taper-seal area18. With regard to the oil retained in the micro-gap series formed (as noted above) in between the upper-end face of the sleeve8and the undersurface of the top wall portion2aof the rotor hub2, and—continuing from the top wall portion2aof the rotor hub2—in between outer circumferential surface of the shaft4and the inner circumferential surface of the sleeve8, and continuous therewith, in between the end face of the shaft4and the inner face of the sealing cap10: the oil-air boundary is in the taper-seal area18alone, and forms a meniscus where the oil surface tension and the outside air pressure balance.

The taper-seal area18functions as an oil reservoir, and in accordance with the amount of oil retained within the taper-seal area18, the location where the boundary forms is movable to suit. Accordingly, attendant on reduction in the amount of oil retained, oil held within the taper-seal area18is supplied to the bearing sections; and meanwhile, expanded oil due to thermal swelling is accommodated within the taper-seal area18.

In this way, the taper-shaped clearance is formed in between the outer circumferential surface of the flange portion8aof the sleeve8, and the inner circumferential surface of the peripheral wall portion2bof the rotor hub2, to configure the taper-seal area18employing surface tension. This configuration makes the taper-seal area18diametrically larger than the bearing sections, and meanwhile lets the axial dimension of the taper-seal area18be relatively large. Consequently, the volumetric capacity within the taper-seal area18is enlarged, making it sufficiently complementary for thermal expansion of the greater amount of oil retained in hydrodynamic bearings having the full-fill structure.

An annular retaining ring20is fixedly attached by means of an adhesive to the peripheral wall portion2bat its end distally beyond the taper-seal area18. The retaining ring20fits into place at the lower-end portion of the outer circumferential surface of the sleeve8without coming into contact against the lower part of the flange portion8a, whereby a structure that keeps the rotor6from coming out from the sleeve8is configured. By thus configuring the rotor6retaining structure along the outer circumferential surface of the sleeve8, a pair of radial bearings, which will later be described in detail, and the retaining structure are not arranged lying in a row along the same axis. This accordingly enables the entire length of the shaft4to be put to effective use as a bearing, and makes it possible to scale down the motor into a lower profile while maintaining bearing rigidity.

Here, arranging the rotor6retaining structure external to the bearings, as is the case with the spindle motor illustrated inFIG. 2, in order to slim the profile of the motor means that the retainer is disposed within the air (referred to hereinafter as “the dry area”).

In a hard disk drive, for example, in order to shorten seek time the heads and the recording surface of the recording disks are separated by a clearance of as little as 1 □m or less. Therefore, even micro-particles can get caught in the clearance between a head and a recording surface, becoming the causative source of a so-called head crash. For spindle motors employed under such environments, this sort of particle spatter is a serious problem in terms of quality.

If the retainer were to be configured inside the bearings, metal abrasion dust that would be produced during rotation by contact occurring in the retainer section due to exteriorly acting vibrations and shock would be captured by the oil retained in the bearing sections. The dust therefore could not be scattered away to the spindle motor exterior. In contrast, configuring the retainer section in the dry area means that particulate matter produced in the retainer section readily gets scattered away to the exterior of the spindle motor.

The production of particulate matter during contact becomes even more pronounced in those particular situations in which the rotary-side components and the stationary-side components that compose the retaining section are made from the same type of metal.

Under these circumstances, making the retaining ring20harder at least on its surface than the sleeve8makes scaling down the motor profile while gaining desired rotational precision a reality. At the same time, the at least superficially harder retaining ring20enables preventing as much as possible the production of particulate matter due to contact between the retaining ring20and the sleeve8that together constitute the retainer. Accordingly, even if exterior vibrations and shock have an impact on the spindle motor when the rotor6spins, and contact between the retaining ring20and the sleeve8occurs, the generation of particulate matter will be prevented.

In this instance, forming the retaining ring20from a ceramic material makes surer prevention of the production of particulate matter possible, without increasing the manufacturing process steps.

Likewise, generation of particulate matter due to contact between the sleeve8and the retaining ring20can be prevented by forming the retaining ring20from, e.g., a stainless-steel material and carrying out a surface-hardening process on the surface thereof. Nickel plating, DLC (diamond-like carbon) coating, or nitriding treatments are preferable as surface treatments in this case.

As far as forming the retaining section is concerned, in either of the foregoing cases, the sleeve8and retaining ring20can be made from raw materials that differ—formed using a stainless-steel material or a copper raw material.

The upper face of the retaining ring20opposes the undersurface of the flange portion8aacross an axial gap that is continuous with the taper-seal area18and whose clearance is smaller than the minimum clearance of the radial gap in the taper-seal area18.

By establishing the clearance of the axial micro-gap defined between the upper face of the retaining ring20and the undersurface of the flange portion8ato be as small as possible, it functions as a labyrinth seal when the spindle motor is spinning. The difference between the air current speed in the axial micro-gap and the air current speed in the radial clearance defined in the taper-seal area18is thus enlarged, and the resistance to outflow of oil vapor occurring due to gasification is made greater. This keeps the vapor pressure in the vicinity of the oil boundary surface high, so as further to prevent vapor dispersal of the oil.

Setting up a labyrinth seal in this way in association with the taper-seal area18not only checks outflow of oil as a fluid, but makes it possible to deter outflow to the motor exterior of oil mist produced by the oil gasifying due to elevation in the exterior ambient temperature of the motor. This consequently works to prevent decline in the retained amount of oil and maintain stabilized bearing performance over the long term, making the bearings highly durable and reliable.

(2) Bearing Configuration

Herringbone grooves22aas illustrated inFIG. 3are formed on the inner circumferential surface of the sleeve8by its upper-end face so as to induce hydrodynamic pressure in the oil when the rotor6spins. Each of the herringbone grooves22ais configured by a pair of linked spiral grooves22a1and22a2inclining into each other from mutually opposing directions with respect to the rotary direction. An upper radial hydrodynamic bearing section22is constituted between the inner circumferential surface of the sleeve8where the grooves22aare formed and the outer circumferential surface of the shaft4.

Likewise, herringbone grooves24aare formed on the inner circumferential surface of the sleeve8by the free-end portion of the shaft4so as to induce hydrodynamic pressure in the oil when the rotor6spins. Each of the herringbone grooves24ais configured by a pair of spiral grooves24a1and24a2inclining into each other from mutually opposing directions with respect to the rotary direction. A lower radial hydrodynamic bearing section24is constituted between the inner circumferential surface of the sleeve8where the grooves24aare formed and the outer circumferential surface of the shaft4.

Here, the herringbone grooves22a,24athat are formed in the upper and lower radial hydrodynamic bearing sections22,24are established so that the spiral grooves22a1and22a2, and24a1and24a2generate essentially equal pumping force—so that the groove fundamentals, which are axial dimension, inclination angle with respect to the rotary direction, or groove width and depth, will be the same. That is, the herringbone grooves22a,24aare established so as to be axially symmetrical with respect to where the spiral grooves22a1and22a2, and24a1and24a2join. Accordingly, in the upper and lower radial hydrodynamic bearing sections22,24maximum pressure appears in the axially central portion (where the spiral grooves join) of each bearing section, meaning that the pumping action by the spiral grooves22a1and22a2, and24a1and24a2is non-uniform with respect to either direction axially, whereby no axial flow is generated in the oil.

In addition, as illustrated inFIG. 4pump-in spiral grooves26aare formed on the upper-end face of the sleeve8so as to induce radially inward-heading pressure (toward the shaft4) in the oil when the rotor6spins, and a thrust bearing section26is constituted between the upper-end face of the sleeve8and the undersurface of the rotor hub2top wall portion2a.

Accordingly, structuring the spindle motor to be a full-fill type bearing configuration while maintaining desired bearing rigidity and—in not requiring a thrust plate to configure the thrust hydrodynamic bearing—retaining a simplified, reduced-cost enabling structure makes it possible further to slim the motor profile and lower its cost.

Likewise, an axial support section28that, as will later be described in detail, employs oil internal pressure heightened by the spiral grooves26aof the thrust bearing section26, is configured in between the free-end end face of the shaft4and the inner face of the sealing cap10as a hydrostatic bearing section.

(3) Manner in Which Rotor is Supported

How the rotor6is supported by the bearings configured as described in the foregoing will be detailed with reference to FIG.5. Here,FIG. 5is a pressure-distribution chart schematically representing relative relationships in pressure distribution, developing from bearing to bearing, of the oil retained in the micro-gap formed in between the upper-end face of the sleeve8and the undersurface of the top wall portion2aof the rotor hub2, and—continuing from the top wall portion2aof the rotor hub2—in between outer circumferential surface of the shaft4and the inner circumferential surface of the sleeve8, and continuous therewith, in between the end face of the shaft4and the inner face of the sealing cap10. Because the pressure distribution in the spindle motor is axially symmetrical, however, the pressure distribution with respect to the rotational center axis, indicated by the dotted-dashed line inFIG. 5, for the region that would be on the opposite side of a vertical section through the spindle motor is omitted. Further, the numbers shown inFIG. 5are the same numbers that mark each of the bearing sections in FIG.2.

Accompanying rotation of the rotor6, the pumping force from the herringbone grooves22a,24ain the upper and lower radial hydrodynamic bearings22,24is heightened, producing hydrodynamic fluid pressure. As indicated by the distribution graph inFIG. 5, the pressure through the herringbone grooves22a,24ain the upper and lower radial hydrodynamic bearings22,24rises abruptly at their either ends, becoming maximal in the places where the spiral grooves22a1and22a2, and24a1and24a2join. Utilizing the hydrodynamic pressure generated in the upper and lower radial hydrodynamic bearings22,24, the shaft4is supported axially along its upper/lower ends, and actions that center the shaft4and restore it from deviations are borne.

Accompanying rotation of the rotor6, radially inward-heading pressure is induced in the oil in the thrust bearing section26by the pump-in spiral grooves26a. The flow of the oil is accelerated by the radially inward-heading pressure, raising the oil internal pressure and generating hydrodynamic pressure acting in a lifting direction on the rotor6. As indicated inFIG. 5, the hydrodynamic pressure induced in the thrust bearing section26does not rise abruptly as is the case with the upper and lower radial hydrodynamic bearings22,24; rather, at maximum it is at a level exceeding atmospheric pressure to a certain degree.

Owing to the pressure generated in the thrust bearing section26, pressure-wise the oil retained—continuing from the top wall portion2aof the rotor hub2—in between outer circumferential surface of the shaft4and the inner circumferential surface of the sleeve8, and continuous therewith, in between the end face of the shaft4and the inner face of the sealing cap10is essentially sealed. Likewise, the fact that the herringbone grooves22a,24aformed in the upper and lower radial hydrodynamic bearings22,24have an axially symmetrical form, and that the dynamic pressure generated is balanced in the axial direction means that, as described above, axially directed flow is not induced in the oil. Thus, the internal pressure of the oil retained in between the outer circumferential surface of the shaft4and the inner circumferential surface of the sleeve8, and continuous therewith, in between the end face of the shaft4and the inner face of the sealing cap10balances with the internal pressure of the oil retained in the thrust bearing section26. Accordingly, as indicated inFIG. 5, in either of these areas the internal pressure of the oil will be on par with that of the oil retained in the thrust bearing26. Negative pressure, wherein the internal pressure would go below atmospheric pressure, will not be generated in the oil retained within these micro-gaps.

Problems with leakage of oil out to the bearing exterior, with vibrations, or with worsening of non-repeatable run-out, which arise due to air bubbles residing within the oil, are accordingly prevented from occurring. Thus, a communicating passage for communicating the bearing interior with the external air is thereby rendered unnecessary.

As noted above, the pressure generated in the thrust bearing26it is at a level exceeding atmospheric pressure to a certain degree, but this pressure alone is unlikely to lift the rotor6sufficiently. Nevertheless, the internal pressure of the oil retained in the axial support section28formed between the free-end end face of the shaft4and the inner face of the sealing cap10as described above will be pressure equal to the oil internal pressure heightened by the hydrodynamic pressure induced in the thrust bearing section26. That is, although a hydrodynamic bearing is not configured between the inner face of the sealing cap10and the end face of the shaft4, the axial support section28—which functions as a so-called hydrostatic pressure bearing, and which in cooperation with the thrust bearing26allows the rotor6to be lifted—is configured.

Accordingly, the thrust bearing section26and the axial support (hydrostatic bearing) section28cooperate to enable the rotor6to be sufficiently lifted.

Here, as illustrated inFIG. 2an annular thrust yoke30made of a ferromagnetic material is disposed in a position on the bracket12opposing the rotor magnet16. This generates axially directed magnetic attraction between the rotor magnet16and the thrust yoke30that balances the lifting pressure on the rotor6generated in the thrust bearing section26and the axial support section28, stabilizes the thrust-directed support of the rotor6, and controls occurrence of over-lift that would buoy the rotor6more than necessary. A thus magnetically urging force can also be made to act on the rotor6by, for example, displacing the magnetic centers of the stator14and the rotor magnet16in the axial direction.

Second Embodiment

(4) Spindle Motor Configuration

Next, usingFIGS. 6 through 8the configuration of a spindle motor in a second embodiment of the present invention will be described. Here, components in the second embodiment that are identical with the first embodiment are marked with the same reference numerals, and explanation thereof is omitted. Likewise, the bearing configuration is essentially identical with the first embodiment, as is the way in which the bearings support the rotor, and the configuration is therefore marked with the same reference numerals.

Set forth inFIG. 6, the spindle motor includes: a rotor6′ made up of a rotor hub2′—composed of an approximately disk-shaped top wall portion2′a, and a round-cylindrical peripheral wall portion2bdepending downward from the outer rim of the top wall portion2a—and of a shaft4′ formed integrally with the central part of the top wall portion2′aof the rotor hub2′; and a round-cylindrical casing member5that is fitted to the outer circumferential surface of the shaft4′.

(5) Configuration and Function of Communicating Pathway

Reference is made now toFIG. 7, which is an elevational view representing the shaft4′ enlarged. As illustrated inFIG. 7, a single helical groove4a′ (represented in part by dashed lines) is furnished on the outer circumferential surface of the shaft4′, running in the axial direction from its upper to its lower end.

The helical groove4a′ is formed by a machining process to have a sectional contour that is approximately rectangular or triangular, or else semicircular. Here, when carrying out the process of machining the helical groove4′ainto the outer circumferential surface of the shaft4′, the process can be carried out in a single chucking.

With the casing member5fitted onto the outer circumferential surface of the shaft4′, in between it and the inner circumferential surface of the casing member5′ a helix-shaped communicating pathway7is defined by the helical groove4′a. The communicating pathway7runs along the inner circumferential surface of the casing member5′ from the upper to the lower end portion in the axial direction, i.e., the pathway7is continuous with the micro-gaps formed in thrust bearing section26and the axial support section28. Within the communicating pathway7, oil is retained continuously with the oil held in the thrust bearing section26and in the axial support section28. Likewise, the internal pressure of the oil retained within the communicating pathway7balances with the internal pressure of the oil retained in the bearing sections.

It can sometimes happen that on account of manufacturing discrepancies in the inner circumferential surface of the sleeve8and the outer circumferential surface of the casing member5becoming combined in the worst case scenario within tolerances, or due to the impact of stresses that occur in fastening the screw into the threaded hole4′bprovided in the shaft for retaining recording disks on the disk disk-mounting portion2′cof the rotor hub2′, the clearance of the micro-gap formed in between the inner circumferential surface of the sleeve8and the outer circumferential surface of the casing member5will be non-uniform between the upper-end and lower-end sides in the axial direction. Should the micro-gap formed between the sleeve8inner circumferential surface and the casing member5outer circumferential surface be non-uniform, an abnormal flow will be induced in the oil. As a consequence, a disparity in internal pressure of the oil in the upper end and in the lower end axially of the micro-gap formed between the sleeve8inner circumferential surface and the casing member5outer circumferential surface—i.e., a pressure disparity between the thrust bearing section26and the axial support section28—will arise. If this oil internal pressure difference is left as is, oil will happen to flow from the lower to the upper end axially, giving rise to negative pressure in the axial support section28. Likewise, oil will happen to flow from the upper to the lower end axially, raising the internal pressure of the oil in the axial support section28more than is necessary and producing over-lift on the rotor6.

Countering this, the communicating pathway7that is continuous with micro-gaps formed in the thrust bearing section26and the axial support section28, and that retains oil continuously with the oil retained in these thrust-bearing and bearing sections26and28, is provided. Therefore, even if the above-noted axial flow is induced in the oil, and a disparity arises in the internal pressure of the oil in the upper end and in the lower end axially of the micro-gap formed between the sleeve8inner circumferential surface and the casing member5outer circumferential surface, because a flow of oil passing through the communicating pathway7from the internal-pressure high end to the low end will occur, the internal pressure of the oil retained in each of the bearing areas will balance, preventing incidents of negative pressure and over-lift.

The presence of a communicating pathway7as in the foregoing in a spindle motor of the second embodiment means that in the herringbone grooves22aand24ain the radial bearing sections22and24formed in between the inner circumferential surface of the sleeve8and the outer circumferential surface of the casing member5, configurations such as indicated inFIGS. 8A through 8Dfor the spiral grooves22a1and22a2, and24a1and24a2that form the herringbone grooves22aand24a, other than being symmetrical with respect to where they join—as is the case in the first embodiment—are possible.

(6) Modified Examples of Second Embodiment

In the modification example diagrammed inFIG. 8Aherringbone grooves22a′ formed in an upper radial bearing section22′ have an asymmetrical configuration in the axial direction, while the herringbone grooves24aformed in the lower radial bearing section24have a symmetrical configuration with respect to where they join, likewise as is the case in the first embodiment.

To be more specific: In the herringbone grooves22a′ formed in the upper radial bearing section22′, spiral grooves22a′1located toward the upper end of the sleeve8(thrust bearing section26) are established so as to be longer in axial dimension than spiral grooves22a′2located toward the lower radial bearing section24. Consequently, the place in which the pairs of spiral grooves22a′1and22a′2join is lower than the center of the upper radial bearing section22′—i.e., is located biased toward the lower radial bearing section24. Therefore, the pumping action by the spiral grooves22a′1on the oil when the rotor6spins surpasses the pumping action by the spiral grooves22a′2, which in terms of the upper radial bearing section22′ induces in the oil a flow heading toward the lower end of the sleeve8(toward the lower radial bearing section24).

Rendering the herringbone grooves22a′ in the upper radial bearing section22′ in an axially unbalanced configuration in this way keeps the pressure in the region between the upper radial bearing section22′ and the lower radial bearing section24at positive pressure greater than atmospheric, preventing the occurrence of negative pressure. Then on account of the pressuring force generated in the herringbone grooves22a′ the oil always flows toward the lower end of the sleeve8; and oil that has flowed toward the lower end of the sleeve8recirculates through the communicating pathway7from along the lower end to along the upper end of the sleeve8, and is pushed in toward the lower end of the sleeve8all over again by the upper radial bearing section22′, wherein a constant oil circulation path is formed.

Thus the herringbone grooves22a′ causing the oil to flow at all times in a predetermined direction within the bearing gaps provides for stability in the balance of pressure in every region in the oil retained within the bearing gaps, which ensures that occurrences of negative pressure and of over-lift on the rotor6are prevented. What is more, even in instances where manufacturing discrepancies or deformation stress during assembly have occurred, with oil circulation in a constant direction being secured the acceptable range—beyond which is unacceptability whose fault lies in processes and assembly—is markedly enlarged, therefore bettering yields.

Then as diagrammed inFIG. 8B, it is also possible to render in an axially asymmetrical configuration not only the upper radial bearing section22′, but also the lower radial bearing section24′, in a makeup where among the spiral grooves24a′1and24a′2constituting the herringbone grooves24a′ formed therein, establishing the spiral grooves24a′1, located toward the upper radial bearing section22′, so as to be longer in axial dimension than the spiral grooves24a′2, located toward the lower end of the sleeve8, positions the place in which they join biased toward the lower end of the sleeve8.

Thus configuring not only the upper radial bearing section22′ but also the lower radial bearing section24′ so as induce in the oil a flow heading toward the lower end of the sleeve8makes the pressure in the hydrostatic bearing section28higher, and strengthens the lifting force on the rotor6. Because this accordingly makes it so that higher-burden loads can be supported, utilization in situations where a plurality of disks is rotationally driven is made possible. Furthermore, a more active circulation is urged upon the oil, which is effective in preventing occurrences of negative pressure and of over-lift on the rotor6.

With modification example 3, diagrammed inFIG. 8C, in the herringbone grooves22a′ formed in the upper radial bearing section22′ the spiral grooves22a′1located toward the thrust bearing section26are established so as to be longer in axial dimension than the spiral grooves22a′2located toward the lower radial bearing section24, likewise as is the case in modification examples 1 and 2, so that a flow toward the lower radial bearing section is generated in the oil. What is different, however, is that in herringbone grooves24a″ formed in lower radial bearing section24″, spiral grooves24a″2located toward the lower end of the sleeve8are formed so as to be slightly longer in axial dimension than spiral grooves24a″1located toward the upper radial bearing section22′.

This configuration consequently prompts an oil flow heading from along the lower radial bearing section24′ toward the upper radial bearing section22′, preventing incidence of negative pressure in the regions between the upper radial bearing section22′ and the lower radial bearing section24″. It should be understood that as the dimensional difference between the spiral grooves24a″1and the spiral grooves24a″2making up the herringbone grooves24a″ of the lower radial bearing section24″ is less than the dimensional difference in the herringbone grooves22a′ of the upper radial bearing section22′, oil flow generated in the upper radial bearing section22′ and heading toward the lower radial bearing section24″ is therefore not hindered by the oil flow that the lower radial bearing section24″ generates heading toward the upper radial bearing section22′.

An additionally possible modification, as diagrammed inFIG. 8D, is to render the herringbone grooves in the upper radial bearing section herringbone grooves22alikewise as in the first embodiment, in a configuration symmetrical with respect to the joints, and to render the herringbone grooves in the lower radial bearing section herringbone grooves24a′ likewise as with Modification Example 2 illustrated inFIG. 8B, in an asymmetrical form biased toward the lower end of the sleeve8. In this case, the dimensional difference between the spiral grooves24a′1and24a′2in the lower radial bearing section24′ is less than is the case with the herringbone grooves on the upper radial bearing section side rendered in an asymmetrical configuration. Consequently, a relatively large bearing span between upper and lower radial bearing sections22and24′ can be secured to make it possible to enhance bearing rigidity, while an oil flow heading toward the lower end of the sleeve8is still generated to expand tolerance in terms of manufacturing discrepancy or deformation stress during assembly.

Here, inducing in the oil a flow heading from along the radial bearing sections toward the lower end of the sleeve8, as indicated inFIGS. 8A through 8D, means that the internal pressure of the oil retained in the hydrostatic bearing section28balances to the sum of flow pressure induced in the thrust bearing section26and the flow pressure of the oil from along the radial bearing sections. This enables more stabilized support with increased bearing load capacity.

(7) Configuration of Disk Drive Device

Reference is made toFIG. 9, in which the internal configuration of a disk drive50is illustrated as an exemplary diagram. A clean space where dust and debris are extremely slight is formed inside a housing51, in the interior of which is disposed a spindle motor52on which platter-shaped recording disks for data recordation are fitted. In addition, a head-shifting mechanism57that reads data from and writes data onto the recording disks53is disposed within the housing51. The head-shifting mechanism57is composed of heads56that read/write data on the recording disks53; arms55that support the heads56; and an actuator54that shifts the heads56and arms55over the needed locations on the recording disks53.

Employing a spindle motor under the first or second embodiments of the present invention as the spindle motor52for the disk drive50as such yields desired rotational precision while making it possible to scale the disk drive50down into a lower profile and reduce its cost. In addition, the reliability and durability of the disk drive50may be improved.

While spindle-motor and disk-drive embodiments in accordance with the present invention have been explained in the foregoing, the present invention is not limited to these embodiments. Various changes and modifications are possible without departing from the scope of the invention.

For example, instead of the pump-in type of spiral grooves26athat were described in the foregoing embodiments, herringbone grooves that in the radial direction are asymmetrical in contour would be possible as the means provided in the thrust bearing section for generating pressure that acts radially inward on the oil. This would establish a situation in which pumping force from the spiral, radially outwardly located grooves would exceed pumping force from the spiral, radially inwardly located grooves. The amount of imbalance in pumping force between the spiral groove areas would therefore be pressure acting radially inward on the oil.

Here, in a situation in which the above-described herringbone grooves are furnished in the thrust bearing section, the lifting force imparted to the rotor will be higher than the lifting force generated in the spiral grooves. The load-supporting force from the thrust bearing section therefore will be improved, but a downside tied in with the lifting force being generated in the bearing sections is a concern that over-lift on the rotor will arise. Consequently, this must be controlled by means of magnetic urging force imparted to the rotor.