Title: Spindle motor and disk drive utilizing the spindle motor
Abstract: 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.
Patent Number: 6,888,278 Issued on 05/03/2005 to Nishimura, et al.
| Inventors:
|
Nishimura; Hideki (Ohmihachiman, JP);
Oku; Yoshito (Yasu-gun, JP);
Yano; Shinya (Ohmihachiman, JP)
|
| Assignee:
|
Nidec Corporation (Kyoto, JP)
|
| Appl. No.:
|
604975 |
| Filed:
|
August 29, 2003 |
Foreign Application Priority Data
| Sep 13, 2001[JP] | 2001-277355 |
| Sep 14, 2001[JP] | 2001-279021 |
| Nov 19, 2001[JP] | 2001-352669 |
| Aug 29, 2002[JP] | 2002-251256 |
| Current U.S. Class: |
310/90; 360/99.08; 369/269; 384/107; 384/119; 384/132 |
| Intern'l Class: |
F16C 032/06; H02K007/08 |
| Field of Search: |
360/88,970.1,990.4,990.8
369/258,258.1,261,269
720/695
310/90
384/107,119,132
|
References Cited [Referenced By]
U.S. Patent Documents
Primary Examiner: Schuberg; Darren
Assistant Examiner: Jones; Judson H.
Attorney, Agent or Firm: Judge; James
Parent Case Text
This application CIP of U.S. Ser. No. 10/063,929 filed May 27, 2002, now U.S.
Pat. No. 6,836,388.
Claims
1. A spindle motor comprising:
a shaft;
a sleeve formed with a through-hole for rotary-play insertion of the shaft;
a rotor having a round top plate in the rotational center of which the shaft
is furnished united, and a circular cylindrical wall depending from the top plate
along its outer rim;
a cover member for closing over one end of the through-hole formed in the sleeve;
a circular cylindrical casing member fitted to said shaft over its outer circumferential
surface;
micro-gaps formed continuing between an upper-end face of said sleeve and a bottom
face of said rotor top plate, an inner circumferential surface of said sleeve and
an outer circumferential surface of said casing member, and an inner face of said
cover member and end faces of said shaft and said casing member;
oil retained continuously without interruption within said micro-gaps throughout
their entirety;
a radial dynamic-pressure bearing section configured intermediarily by at least
one surface of either said sleeve inner-circumferential surface or said casing
member outer-circumferential surface, and by said oil when said rotor spins, and
provided with, as dynamic-pressure-generating striations, herringbone grooves for
inducing into said oil when said rotor spins hydrodynamic pressure;
a thrust bearing section configured on at least one of either said sleeve upper-end
face or said top plate bottom face, and furnished with dynamic-pressure-generating
striations for imparting to said oil radially inward-heading pressure when said
rotor spins;
an axial support section, formed between said cover member inner face and said
shaft end face, having pressure essentially balancing radially inward-heading pressure
generated in said thrust bearing section, wherein said rotor is lifted through
cooperation of said thrust bearing section end said axial support section; and
a communicating pathway formed in between said shaft along its outer circumferential
surface and said casing member along its inner circumferential surface, for communicating
said oil retained in, and enabling it to circulate between, the micro-gap formed
between said sleeve upper-end face and the bottom face of said rotor top plate,
and the micro-gap formed between said cover member inner face and said shaft and
casing member end faces.
2. A spindle motor as set forth in claim 1:
a helical groove being formed on the outer circumferential surface of said shaft
in a single path running from its upper-end portion to its lower-end portion;
wherein fitting said casing member over the outer circumferential surface of
said shaft defines said communicating pathway between said helical groove and the
inner circumferential surface of said casing member.
3. A spindle motor as set forth in claim 1:
an outer circumferential surface of said sleeve and an inner circumferential
surface of said rotor circular cylindrical wall opposing via a radial gap;
and said sleeve outer circumferential periphery being provided with a taper surface
constricting in outer diameter according as it separates from said rotor top plate;
wherein and said oil is retained by a meniscus forming in between said taper
surface and the inner-circumferential surface of said rotor circular cylindrical
wall.
4. A spindle motor as set forth in claim 3, wherein:
a stepped portion continuous with said taper surface is provided in said sleeve
by recessing its outer circumferential surface radially inwardly;
an annular member projecting radially inward corresponding to the stepped portion
is fixedly fitted into the inner-circumferential surface of said rotor circular
cylindrical wall, and a rotor retainer is constituted by engagement of the stepped
portion and the annular member;
a micro-gap smaller than the minimum clearance of the radial gap formed between
the taper surface of said sleeve and the inner-circumferential surface of said
rotor circular cylindrical wall, is formed to function as a labyrinth seal between
the annular member along its upper face and said sleeve stepped portion along its
undersurface.
5. A spindle motor as set forth in claim 1, said radial dynamic-pressure bearing
section being configured between said shaft outer-circumferential surface end said
sleeve inner-circumferential surface as an axially separated pair of radial dynamic-pressure
bearing constituents, wherein as the dynamic-pressure-generating striations in
each radial bearing constituent, herringbone-groove forming contiguous pairs of
spiral grooves for inducing into said oil when said rotor spins hydrodynamic pressure
whose pressure gradient becomes axially symmetrical are provided.
6. A spindle motor as set forth in claim 1, said radial dynamic-pressure bearing
section being configured between said shaft outer-circumferential surface and said
sleeve inner-circumferential surface as an axially separated pair of radial dynamic-pressure
bearing constituents, wherein as the dynamic-pressure-generating striations in
at least one of either said pair of radial dynamic-pressure bearing constituents,
asymmetrically configured herringbone grooves for inducing into said oil when said
rotor spins hydrodynamic pressure acting unidirectionally in the axial direction
are provided.
7. A spindle motor as set forth in claim 1, wherein said rotor is urged in a
direction toward said cover member by axially acting magnetic force.
8. A disk-drive device including a housing, a spindle motor fixed inside said
housing for spinning recording disks, and an information access means for writing
information into and reading information out from needed locations on the recording
disks, wherein said spindle motor comprises:
a shaft;
a sleeve formed with a through-hole for rotary-play insertion of the shaft;
a rotor having a round top plate in the rotational center of which the shaft
is furnished united, and a circular cylindrical wall depending from the top plate
along its outer rim;
a cover member for closing over one end of the through-hole formed in the sleeve;
a circular cylindrical casing member fitted to said shaft over its outer circumferential
surface;
micro-gaps formed continuing between an upper-end face of said sleeve and a bottom
face of said rotor top plate, an inner circumferential surface of said sleeve and
an outer circumferential surface of said casing member, and an inner face of said
cover member and end faces of said shaft and said casing member;
oil retained continuously without interruption within said micro-gaps throughout
their entirety;
a radial dynamic-pressure bearing section configured intermediarily by at least
one surface of either said sleeve inner-circumferential surface or said casing
member outer-circumferential surface, and by said oil when said rotor spins, and
provided with, as dynamic-pressure-generating striations, herringbone grooves for
inducing into said oil when said rotor spins hydrodynamic pressure;
a thrust bearing section configured on at least one of either said sleeve upper-end
face or said top plate bottom face, and furnished with dynamic-pressure-generating
striations for imparting to said oil radially inward-heading pressure when said
rotor spins;
an axial support section, formed between said cover member inner face and said
shaft end face, having pressure essentially balancing radially inward-heading pressure
generated in said thrust bearing section, wherein said rotor is lifted through
cooperation of said thrust bearing section and said axial support section;
and a communicating pathway formed in between said shaft along its outer circumferential
surface and said casing member along its inner circumferential surface, for communicating
said oil retained in, and enabling it to circulate between, the micro-gap formed
between said sleeve upper-end face and the bottom face of said rotor top plate,
and the micro-gap formed between said cover member inner face and said shaft and
casing member end faces.
9. A disk-drive device as set forth in claim 8:
a helical groove being formed on the outer circumferential surface of said shaft
in a single path running from its upper-end portion to its lower-end portion;
wherein fitting said casing member over the outer circumferential surface of
said shaft defines said communicating pathway between said helical groove and the
inner circumferential surface of said casing member.
10. A disk-drive device as set forth in claim 8:
an outer circumferential surface of said sleeve and an inner circumferential
surface of said rotor circular cylindrical wall opposing via a radial gap;
and said sleeve outer circumferential periphery being provided with a taper surface
constricting in outer diameter according as it separates from said rotor top plate;
wherein and said oil is retained by a meniscus forming in between said taper
surface and the inner-circumferential surface of said rotor circular cylindrical
wall.
11. A disk-drive device as set forth in claim 10, wherein:
a stepped portion continuous with said taper surface is provided in said sleeve
by recessing its outer circumferential surface radially inwardly;
an annular member projecting radially inward corresponding to the stepped portion
is fixedly fitted into the inner-circumferential surface of said rotor circular
cylindrical wall, and a rotor retainer is constituted by engagement of the stepped
portion and the annular member;
a micro-gap smaller than the minimum clearance of the radial gap formed between
the taper surface of said sleeve and the inner-circumferential surface of said
rotor circular cylindrical wall, is formed to function as a labyrinth seal between
the annular member along its upper face and said sleeve stepped portion along its
undersurface.
12. A disk-drive device as set forth in claim 8, wherein:
said radial dynamic-pressure bearing section is configured between the outer-circumferential
surface of said casing member and said sleeve inner-circumferential surface as
an axially separated pair of radial dynamic-pressure bearing constituents; and
as the dynamic-pressure-generating striations in each radial bearing constituent,
herringbone-grooves forming contiguous pairs of spiral grooves for inducing into
said oil when said rotor spins hydrodynamic pressure whose pressure gradient becomes
axially symmetrical are provided.
13. A disk-drive device as set forth in claim 8, wherein:
said radial dynamic-pressure bearing section is configured between the outer-circumferential
surface of said casing member and said sleeve inner-circumferential surface as
an axially separated pair of radial dynamic-pressure bearing constituents; and
as the dynamic-pressure-generating striations in at least one of either said
pair of radial dynamic-pressure bearing constituents, asymmetrically configured
herringbone grooves for inducing into said oil when said rotor spins hydrodynamic
pressure acting unidirectionally in the axial direction are provided.
14. A disk-drive device as set forth in claim 8, wherein said rotor is urged
in a direction toward said cover member by axially acting magnetic force.
Description
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 rotor
100 and the top-end face of a sleeve
102
in the spindle motor depicted in FIG. 1, a thrust bearing section
104 is
configured. Likewise, between the outer circumferential surface of a shaft
106
furnished integrally with the rotor
100, and the inner circumferential surface
of the sleeve
102, radial bearing sections
108,
108 are configured.
The thrust bearing section
104 generates lifting force on the rotor
100,
and the radial bearing sections
108,
108 function to center-balance
in the radial direction, and prevent wobble in, the rotor
100.
The spindle motor depicted in FIG. 1 makes 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 sleeve
102 the spindle
motor depicted in FIG. 1 is provided with a communicating passage
110 made
up of a through-hole
110a and channels
110b,
110c.
The communicating passage
110 brings 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 sections
108,
108
to 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
passage
110 in order to exhaust bubbles to the exterior of the bearing sections.
To bore the communicating passage
110 for 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-hole
110a and
the channels
110b,
110c that constitute the communicating
passage
110 can be made. Consequently, the axial dimension of the shaft
106 and the sleeve
102 must necessarily be at least a given size
for boring the communicating passage
110 and be extensive enough to maintain
bearing rigidity in the radial bearing sections
108,
108. These requirements
stand in the way of making the spindle motor slimmer.
What is more, the fact that the through-hole
110a as well as the
channels
110b,
110c that constitute the communicating
passage
110 are formed in the sleeve
102 complicates 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 element
112 that constitutes a retainer for
the rotor
100 is fitted onto the end portion of the shaft
106 on
the side opposite the rotor
100. In short, because the thrust bearing section
104; the radial bearing sections
108,
108; the through-hole
110a as well as the channels
110b,
110c that
constitute the communicating passage
110; and the ring element
112
are 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.
BRIEF DESCRIPTION OF DRAWINGS
FIG. 1 is a sectional view that illustrates the configurational outline of a
conventional spindle motor;
FIG. 2 is a sectional view that illustrates the configurational outline of a
spindle motor in a first embodiment of the present invention;
FIG. 3 is a fragmentary enlarged sectional view that schematically illustrates
the configuration of herringbone grooves formed in a radial bearing section of
the spindle motor in the first embodiment of the present invention;
FIG. 4 is a plan view that schematically illustrates the configuration of spiral
grooves formed in a thrust bearing section of the spindle motor in the first embodiment
of the present invention;
FIG. 5 is a conceptual pressure-distribution diagram schematically illustrating
pressure distribution in the spindle motor hydrodynamic bearing oil;
FIG. 6 is a sectional view that illustrates the configurational outline of a
spindle motor in a second embodiment of the present invention;
FIG. 7 is an elevational, fragmentary view showing a portion of the shaft enlarged
from the spindle motor depicted in FIG. 6;
FIGS. 8A through 8D are fragmentary enlarged sectional views that schematically
illustrate modified examples of the configuration of herringbone grooves formed
in a radial bearing section of the spindle motor in a second embodiment of the
present invention; and
FIG. 9 is a sectional view schematically illustrating the internal configuration
of a disk drive.
DETAILED DESCRIPTION
First Embodiment
(1) Spindle Motor Configuration
Reference is made to FIG. 2, which illustrates a spindle motor in a first
embodiment of the present invention. Set forth in FIG. 2, the spindle motor is
furnished with: a rotor
6 made up of a rotor hub
2—composed
of an approximately disk-shaped top wall portion
2a, and a round-cylindrical
peripheral wall portion
2b depending downward from the outer rim
of the top wall portion
2a—and of a shaft
4 one end
portion
4a of which is perimetrically inserted into the central portion
of the top wall portion
2a of the rotor hub
2; a hollow, round
cylindrical sleeve
8 rotatively supporting the shaft
4; a sealing
cap
10 opposing the end face of the shaft
4 along its free end, and
closing over the lower portion of the sleeve
8; and a bracket
12
formed integrally with a round cylindrical portion
12a for anchoring
the sleeve
8.
The bracket
12 has a round cupped portion centered on the round cylindrical
portion
12a; and a stator
14 having a plurality of teeth that
project radially inward is arranged on the inner circumferential surface
12b
of a peripheral wall that the outer circumferential edge of the cupped portion
defines. Likewise, a rotor magnet
16 that opposes the stator
14 via
a radially inward clearance therefrom is fixedly fitted to the outer circumferential
surface of the peripheral wall portion
2b of the rotor hub
2.
A flange-shaped disk-mounting portion
2c for carrying recording
disks
on which information is recorded (in FIG. 6, represented as recording disks
53)
is formed on an outer circumferential portion of the peripheral wall portion
2b
of the rotor hub
2. A threaded hole
4b is formed in the
upper-end portion of the shaft
4 (its end at the top wall portion
2a
of the rotor hub
2). The recording disks are loaded onto the disk-mounting
portion
2c, and after being retained by a clamp (not illustrated),
the recording disks are fixedly secured to the rotor hub
2 by fastening
a screw (not illustrated) into the threaded hole
4b.
An unbroken series of micro-gaps is formed in between the upper-end face of the
sleeve
8 and the undersurface of the top wall portion
2a of
the rotor hub
2, and—continuing from the top wall portion
2a
of the rotor hub
2—in between outer circumferential surface of
the shaft
4 and the inner circumferential surface of the sleeve
8,
and continuous therewith, in between the end face of the shaft
4 and the
inner face of the sealing cap
10. 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 sleeve
8 outer circumferential surface is
made into an annular flange portion
8a that 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 sleeve
8. The flange
portion
8a radially opposes, without being in contact with, the inner
circumferential surface of the peripheral wall portion
2b of the
rotor hub
2.
Because as noted above the outer circumferential surface of the flange portion
8a is contoured into an incline, the gap defined in between the inner
circumferential surface of the peripheral wall portion
2b and the
outer circumferential surface of the flange portion
8a forms a taper
whose radial clearance gradually increases heading axially downward (in the direction
toward the distal rim of the peripheral wall portion
2b). In particular,
the inner circumferential surface of the peripheral wall portion
2b and
the outer circumferential surface of the flange portion
8a cooperate
to configure a taper-seal area
18. With regard to the oil retained in the
micro-gap series formed (as noted above) in between the upper-end face of the sleeve
8 and the undersurface of the top wall portion
2a of the rotor
hub
2, and—continuing from the top wall portion
2a of
the rotor hub
2—in between outer circumferential surface of the shaft
4 and the inner circumferential surface of the sleeve
8, and continuous
therewith, in between the end face of the shaft
4 and the inner face of
the sealing cap
10: the oil-air boundary is in the taper-seal area
18
alone, and forms a meniscus where the oil surface tension and the outside air pressure balance.
The taper-seal area
18 functions as an oil reservoir, and in accordance
with the amount of oil retained within the taper-seal area
18, 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 area
18 is
supplied to the bearing sections; and meanwhile, expanded oil due to thermal swelling
is accommodated within the taper-seal area
18.
In this way, the taper-shaped clearance is formed in between the outer circumferential
surface of the flange portion
8a of the sleeve
8, and the
inner circumferential surface of the peripheral wall portion
2b of
the rotor hub
2, to configure the taper-seal area
18 employing surface
tension. This configuration makes the taper-seal area
18 diametrically larger
than the bearing sections, and meanwhile lets the axial dimension of the taper-seal
area
18 be relatively large. Consequently, the volumetric capacity within
the taper-seal area
18 is 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 ring
20 is fixedly attached by means of an adhesive
to the peripheral wall portion
2b at its end distally beyond the
taper-seal area
18. The retaining ring
20 fits into place at the
lower-end portion of the outer circumferential surface of the sleeve
8 without
coming into contact against the lower part of the flange portion
8a,
whereby a structure that keeps the rotor
6 from coming out from the sleeve
8 is configured. By thus configuring the rotor
6 retaining structure
along the outer circumferential surface of the sleeve
8, 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 shaft
4 to 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 rotor
6 retaining structure external to the bearings,
as is the case with the spindle motor illustrated in FIG. 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 ring
20 harder at least
on its surface than the sleeve
8 makes scaling down the motor profile while
gaining desired rotational precision a reality. At the same time, the at least
superficially harder retaining ring
20 enables preventing as much as possible
the production of particulate matter due to contact between the retaining ring
20 and the sleeve
8 that together constitute the retainer. Accordingly,
even if exterior vibrations and shock have an impact on the spindle motor when
the rotor
6 spins, and contact between the retaining ring
20 and
the sleeve
8 occurs, the generation of particulate matter will be prevented.
In this instance, forming the retaining ring
20 from 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 sleeve
8 and the retaining ring
20 can be prevented by forming the retaining
ring
20 from, 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 sleeve
8 and retaining ring
20 can be made from raw materials
that differ—formed using a stainless-steel material or a copper raw material.
The upper face of the retaining ring
20 opposes the undersurface of the
flange portion
8a across an axial gap that is continuous with the
taper-seal area
18 and whose clearance is smaller than the minimum clearance
of the radial gap in the taper-seal area
18.
By establishing the clearance of the axial micro-gap defined between the upper
face of the retaining ring
20 and the undersurface of the flange portion
8a to 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 area
18 is 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
area
18 not 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 grooves
22a as illustrated in FIG. 3 are formed
on the inner circumferential surface of the sleeve
8 by its upper-end face
so as to induce hydrodynamic pressure in the oil when the rotor
6 spins.
Each of the herringbone grooves
22a is configured by a pair of linked
spiral grooves
22a1 and
22a2 inclining
into each other from mutually opposing directions with respect to the rotary direction.
An upper radial hydrodynamic bearing section
22 is constituted between the
inner circumferential surface of the sleeve
8 where the grooves
22a
are formed and the outer circumferential surface of the shaft
4.
Likewise, herringbone grooves
24a are formed on the inner
circumferential surface of the sleeve
8 by the free-end portion of the shaft
4 so as to induce hydrodynamic pressure in the oil when the rotor
6
spins. Each of the herringbone grooves
24a is configured by a pair
of spiral grooves
24a1 and
24a2 inclining
into each other from mutually opposing directions with respect to the rotary direction.
A lower radial hydrodynamic bearing section
24 is constituted between the
inner circumferential surface of the sleeve
8 where the grooves
24a
are formed and the outer circumferential surface of the shaft
4.
Here, the herringbone grooves
22a,
24a that are
formed in the upper and lower radial hydrodynamic bearing sections
22,
24
are established so that the spiral grooves
22a1 and
22a2,
and
24a1 and
24a2 generate 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 grooves
22a,
24a
are established so as to be axially symmetrical with respect to where the spiral
grooves
22a1 and
22a2, and
24a1
and
24a2 join. Accordingly, in the upper and lower radial
hydrodynamic bearing sections
22,
24 maximum pressure appears in
the axially central portion (where the spiral grooves join) of each bearing section,
meaning that the pumping action by the spiral grooves
22a1
and
22a2, and
24a1 and
24a2
is non-uniform with respect to either direction axially, whereby no axial flow
is generated in the oil.
In addition, as illustrated in FIG. 4 pump-in spiral grooves
26a are
formed on the upper-end face of the sleeve
8 so as to induce radially inward-heading
pressure (toward the shaft
4) in the oil when the rotor
6 spins,
and a thrust bearing section
26 is constituted between the upper-end face
of the sleeve
8 and the undersurface of the rotor hub
2 top wall
portion
2a.
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 section
28 that, as will later be described
in detail, employs oil internal pressure heightened by the spiral grooves
26a
of the thrust bearing section
26, is configured in between the free-end
end face of the shaft
4 and the inner face of the sealing cap
10
as a hydrostatic bearing section.
(3) Manner in Which Rotor is Supported
How the rotor
6 is supported by the bearings configured as described in
the foregoing will be detailed with reference to FIG.
5. Here, FIG. 5 is
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 sleeve
8 and
the undersurface of the top wall portion
2a of the rotor hub
2,
and—continuing from the top wall portion
2a of the rotor hub
2—in between outer circumferential surface of the shaft
4
and the inner circumferential surface of the sleeve
8, and continuous therewith,
in between the end face of the shaft
4 and the inner face of the sealing
cap
10. 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 in FIG. 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 in FIG. 5 are the same numbers that mark each
of the bearing sections in FIG.
2.
Accompanying rotation of the rotor
6, the pumping force from
the herringbone grooves
22a,
24a in the upper and lower
radial hydrodynamic bearings
22,
24 is heightened, producing hydrodynamic
fluid pressure. As indicated by the distribution graph in FIG. 5, the pressure
through the herringbone grooves
22a,
24a in the upper
and lower radial hydrodynamic bearings
22,
24 rises abruptly at their
either ends, becoming maximal in the places where the spiral grooves
22a1
and
22a2, and
24a1 and
24a2
join. Utilizing the hydrodynamic pressure generated in the upper and lower radial
hydrodynamic bearings
22,
24, the shaft
4 is supported axially
along its upper/lower ends, and actions that center the shaft
4 and restore
it from deviations are borne.
Accompanying rotation of the rotor
6, radially inward-heading
pressure is induced in the oil in the thrust bearing section
26 by the pump-in
spiral grooves
26a. 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 rotor
6. As indicated in FIG.
5, the hydrodynamic pressure induced in the thrust bearing section
26 does
not rise abruptly as is the case with the upper and lower radial hydrodynamic bearings
22,
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 section
26, pressure-wise
the oil retained—continuing from the top wall portion
2a of
the rotor hub
2—in between outer circumferential surface of the shaft
4 and the inner circumferential surface of the sleeve
8, and continuous
therewith, in between the end face of the shaft
4 and the inner face of
the sealing cap
10 is essentially sealed. Likewise, the fact that the herringbone
grooves
22a,
24a formed in the upper and lower radial
hydrodynamic bearings
22,
24 have 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 shaft
4 and the inner circumferential surface of the sleeve
8,
and continuous therewith, in between the end face of the shaft
4 and the
inner face of the sealing cap
10 balances with the internal pressure of
the oil retained in the thrust bearing section
26. Accordingly, as indicated
in FIG. 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 bearing
26. 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 bearing
26 it is
at
a level exceeding atmospheric pressure to a certain degree, but this pressure alone
is unlikely to lift the rotor
6 sufficiently. Nevertheless, the internal
pressure of the oil retained in the axial support section
28 formed between
the free-end end face of the shaft
4 and the inner face of the sealing cap
10 as described above will be pressure equal to the oil internal pressure
heightened by the hydrodynamic pressure induced in the thrust bearing section
26.
That is, although a hydrodynamic bearing is not configured between the inner face
of the sealing cap
10 and the end face of the shaft
4, the axial
support section
28—which functions as a so-called hydrostatic pressure
bearing, and which in cooperation with the thrust bearing
26 allows the
rotor
6 to be lifted—is configured.
Accordingly, the thrust bearing section
26 and the axial support
(hydrostatic bearing) section
28 cooperate to enable the rotor
6
to be sufficiently lifted.
Here, as illustrated in FIG. 2 an annular thrust yoke
30 made of a ferromagnetic
material is disposed in a position on the bracket
12 opposing the rotor
magnet
16. This generates axially directed magnetic attraction between the
rotor magnet
16 and the thrust yoke
30 that balances the lifting
pressure on the rotor
6 generated in the thrust bearing section
26
and the axial support section
28, stabilizes the thrust-directed support
of the rotor
6, and controls occurrence of over-lift that would buoy the
rotor
6 more than necessary. A thus magnetically urging force can also be
made to act on the rotor
6 by, for example, displacing the magnetic centers
of the stator
14 and the rotor magnet
16 in the axial direction.
Second Embodiment
(4) Spindle Motor Configuration
Next, using FIGS. 6 through 8 the 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 in FIG. 6, the spindle motor includes: a rotor
6′ made
up of a rotor hub
2′—composed of an approximately disk-shaped
top wall portion
2′
a, and a round-cylindrical peripheral wall
portion
2b depending downward from the outer rim of the top wall
portion
2a—and of a shaft
4′ formed integrally
with the central part of the top wall portion
2′
a of the rotor
hub
2′; and a round-cylindrical casing member
5 that is fitted
to the outer circumferential surface of the shaft
4′.
(5) Configuration and Function of Communicating Pathway
Reference is made now to FIG. 7, which is an elevational view representing
the shaft
4′ enlarged. As illustrated in FIG. 7, a single helical
groove
4a′ (represented in part by dashed lines) is furnished
on the outer circumferential surface of the shaft
4′, running in
the axial direction from its upper to its lower end.
The helical groove
4a′ 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
groove
4′
a into the outer circumferential surface of the shaft
4′, the process can be carried out in a single chucking.
With the casing member
5 fitted onto the outer circumferential surface
of the shaft
4′, in between it and the inner circumferential surface
of the casing member
5′ a helix-shaped communicating pathway
7
is defined by the helical groove
4′
a. The communicating pathway
7 runs along the inner circumferential surface of the casing member
5′
from the upper to the lower end portion in the axial direction, i.e., the pathway
7 is continuous with the micro-gaps formed in thrust bearing section
26
and the axial support section
28. Within the communicating pathway
7,
oil is retained continuously with the oil held in the thrust bearing section
26
and in the axial support section
28. Likewise, the internal pressure of
the oil retained within the communicating pathway
7 balances 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 sleeve
8 and the outer circumferential
surface of the casing member
5 becoming combined in the worst case scenario
within tolerances, or due to the impact of stresses that occur in fastening the
screw into the threaded hole
4′
b provided in the shaft for
retaining recording disks on the disk disk-mounting portion
2′
c
of the rotor hub
2′, the clearance of the micro-gap formed in
between the inner circumferential surface of the sleeve
8 and the outer
circumferential surface of the casing member
5 will be non-uniform between
the upper-end and lower-end sides in the axial direction. Should the micro-gap
formed between the sleeve
8 inner circumferential surface and the casing
member
5 outer 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 sleeve
8 inner circumferential surface and the casing member
5 outer circumferential surface—i.e., a pressure disparity between
the thrust bearing section
26 and the axial support section
28—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 section
28. 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 section
28 more than is necessary and producing over-lift
on the rotor
6.
Countering this, the communicating pathway
7 that is continuous
with micro-gaps formed in the thrust bearing section
26 and the axial support
section
28, and that retains oil continuously with the oil retained in these
thrust-bearing and bearing sections
26 and
28, 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 sleeve
8 inner circumferential surface
and the casing member
5 outer circumferential surface, because a flow of
oil passing through the communicating pathway
7 from 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 pathway
7 as in the foregoing in a spindle
motor of the second embodiment means that in the herringbone grooves
22a
and
24a in the radial bearing sections
22 and
24
formed in between the inner circumferential surface of the sleeve
8 and
the outer circumferential surface of the casing member
5, configurations
such as indicated in FIGS. 8A through 8D for the spiral grooves
22a1
and
22a2, and
24a1 and
24a2
that form the herringbone grooves
22a and
24a, 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
(6-1) Modification Example 1
In the modification example diagrammed in FIG. 8A herringbone grooves
22a′
formed in an upper radial bearing section
22′ have an asymmetrical
configuration in the axial direction, while the herringbone grooves
24a
formed in the lower radial bearing section
24 have 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 grooves
22a′ formed
in the upper radial bearing section
22′, spiral grooves
22a′
1
located toward the upper end of the sleeve
8 (thrust bearing section
26)
are established so as to be longer in axial dimension than spiral grooves
22a′
2
located toward the lower radial bearing section
24. Consequently, the place
in which the pairs of spiral grooves
22a′
1 and
22a′
2
join is lower than the center of the upper radial bearing section
22′—i.e.,
is located biased toward the lower radial bearing section
24. Therefore,
the pumping action by the spiral grooves
22a′
1 on the
oil when the rotor
6 spins surpasses the pumping action by the spiral grooves
22a′
2, which in
