U.S. patent application number 10/775120 was filed with the patent office on 2004-08-12 for dynamic pressure bearing with improved starting characteristics.
This patent application is currently assigned to SUMITOMO ELECTRIC INDUSTRIES, LTD.. Invention is credited to Komura, Osamu, Murabe, Kaoru, Otsuki, Makoto, Takeuchi, Hisao.
Application Number | 20040156569 10/775120 |
Document ID | / |
Family ID | 27335239 |
Filed Date | 2004-08-12 |
United States Patent
Application |
20040156569 |
Kind Code |
A1 |
Takeuchi, Hisao ; et
al. |
August 12, 2004 |
Dynamic pressure bearing with improved starting characteristics
Abstract
A hydrodynamic bearing assembly with improved activation
features is provided. The opposing surfaces in the radial and
thrust bearings have grooves 2 and 5 with depths shallower
gradually towards the downstream flow of the fluid passing
therethrough, for generating the uniform dynamic pressure
distribution. This allows the dynamic pressure distribution to be
leveled so as to increase the bearing supporting force and prevent
the dew grom being generated. Any one or both of the opposing
surfaces in the thrust bearing has the inclined surface from the
inner portion towards the outer portion so that the gap between the
opposing surfaces is extended to about 2 microns. This causes the
contacting points thereof when halted to be closer to the axis so
that the friction can be reduced and the driving torque can be
reduced when restating the bearing assembly. This also prevent the
contact in the thrust bearing due to the external oscillating
motion. Further, a second thrust plate 11 is secured on the other
end surface of the sleeve 3 opposite to the thrust plate 4 so that
the total weight of the rotational member is supported between the
second thrust plate 11 and the end surface of the shaft 1 when
halted, thereby reducing the friction at the time of
activating.
Inventors: |
Takeuchi, Hisao; (Itami-shi,
JP) ; Komura, Osamu; (Itami-shi, JP) ; Murabe,
Kaoru; (Itami-shi, JP) ; Otsuki, Makoto;
(Itami-shi, JP) |
Correspondence
Address: |
McDermott, Will & Emery
600 13th Street, N.W.
Washington
DC
20005-3096
US
|
Assignee: |
SUMITOMO ELECTRIC INDUSTRIES,
LTD.
Osaka
JP
|
Family ID: |
27335239 |
Appl. No.: |
10/775120 |
Filed: |
February 11, 2004 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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10775120 |
Feb 11, 2004 |
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09856093 |
Aug 16, 2001 |
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6702464 |
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09856093 |
Aug 16, 2001 |
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PCT/JP00/06297 |
Sep 14, 2000 |
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Current U.S.
Class: |
384/107 ;
G9B/19.028 |
Current CPC
Class: |
G02B 26/121 20130101;
F16C 33/107 20130101; G11B 19/2009 20130101; F16C 2370/12 20130101;
F16C 17/04 20130101; H02K 7/086 20130101; F16C 17/045 20130101 |
Class at
Publication: |
384/107 |
International
Class: |
F16C 032/06 |
Foreign Application Data
Date |
Code |
Application Number |
Sep 17, 1999 |
JP |
11-263614 |
Nov 11, 1999 |
JP |
11-321086 |
Nov 22, 1999 |
JP |
11-331211 |
Claims
1. A hydrodynamic bearing assembly, comprising: a rotational and
stationary members arranged with predetermined gaps to each other,
the gaps including a radial gap defined therebetween in a radial
direction perpendicular to an axis or a thrust gap defined
therebetween in a thrust direction parallel to the axis, the gap
containing a fluid for generating a radial or thrust dynamic
pressure due to the relative rotation between the rotational and
stationary members so that said rotational member rotate relative
to said stationary member without any contact; wherein said
rotational and stationary members have opposing surfaces to each
other, and any one of the opposing surfaces has grooves formed
thereon, the grooves having the depth modified in accordance with
its position so that the dynamic pressure generated across the gap
is substantially even.
2. The hydrodynamic bearing assembly according to claim 1: wherein
the grooves on one of opposing surfaces of said rotational and
stationary members for generating thrust dynamic pressure to keep
both members away from each other in the thrust direction are
formed such that the grooves are shallower towards the downstream
flow of the fluid passing therethrough.
3. The hydrodynamic bearing assembly according to claim 1: wherein
the grooves on one of opposing surfaces of said rotational and
stationary members for generating radial dynamic pressure to keep
both members away from each other in the radial direction are
formed such that the grooves are shallower towards the downstream
flow of the fluid passing therethrough.
4. The hydrodynamic bearing assembly according to any one of claims
1 to 3: wherein the grooves have the depth modified gradually and
smoothly towards the downstream flow of the fluid passing
therethrough.
5. The hydrodynamic bearing assembly according to any one of claims
1 to 3: wherein the grooves have the depth modified step-by-step
towards the downstream flow of the fluid passing therethrough.
6. A hydrodynamic bearing assembly, comprising: a disk-shaped
thrust plate extending in a radial direction perpendicular to a
bearing axis; a circular thrust opposing surface extending in the
radial direction and opposing to said thrust plate; and a thrust
bearing for generating the thrust dynamic pressure in a thrust
direction parallel to the bearing axis due to a relative rotation
between said thrust plate and said thrust opposing surface; wherein
at least one, or both of opposing surfaces of said thrust plate and
said thrust opposing surface are inclined such that a distance
between both opposing surfaces thereof becomes greater from an
inside portion towards an outer portion of the thrust bearing.
7. The hydrodynamic bearing assembly according to claim 6: wherein
the opposing surface is inclined linearly in the cross section
parallel to the bearing axis so that it is in a frustum form.
8. The hydrodynamic bearing assembly according to claim 6: wherein
the opposing surface is curved as being arc-shaped in the cross
section parallel to the bearing axis so that it is in a spherical
form.
9. The hydrodynamic bearing assembly according to claim 6: wherein
the opposing surface is drawn, in the cross section parallel to the
bearing axis, as a partial circle having the center provided on a
line beneath the innermost portion of the thrust bearing and
parallel to the bearing axis.
10. The hydrodynamic bearing assembly according to any one of
claims 6 to 9: wherein a gradient d defined by a thrust distance
difference between the opposing surfaces in the thrust bearing is
approximately 2 microns or less.
11. The hydrodynamic bearing assembly according to claim 6, further
comprising: a radial bearing defined by a column shaft having an
outer surface parallel to the axis and a cylindrical hollow sleeve
rotatably arranged around the outer surface, said radial bearing
for generating the radial dynamic pressure in the radial direction
perpendicular to the axis due to the relative rotation between the
shaft and the sleeve; wherein a radial distance s on one side of
the thrust bearing defined between the innermost and outermost
portions of the thrust bearing, a gradient d defined by a thrust
distance difference between the opposing surfaces in the thrust
bearing, a radial length L defined by a length along the axis of
said radial bearing, and a total radial gap F of a pair of side
radial gaps along a diameter in said radial bearing satisfy the
following condition; F/L<d/s.
12. A hydrodynamic bearing assembly, comprising: a radial bearing
including a column shaft having an outer surface parallel to an
axis, and a hollow cylindrical sleeve having an inner surface
rotatably arranged around the outer surface of said shaft, said
radial bearing for generating a radial dynamic pressure due to a
relative rotation between said sleeve and said shaft; a thrust
bearings including a thrust plate formed or secured on one end
surface of said shaft along the axis, and a thrust opposing surface
formed or secured on one end surface of said sleeve along the axis,
said thrust bearing for generating a thrust dynamic pressure due to
the relative rotation between said thrust plate and said thrust
opposing surface; and a second thrust plate covering said hollow
cylindrical sleeve at the other end surface along the axis; wherein
a first gap a defined parallel to the axis between the other end
surface of the shaft and said second thrust plate, and a second gap
b defined parallel to the axis between said thrust plate and thrust
opposing surface satisfy the following condition; a<b.
13. The hydrodynamic bearing assembly according to claim 12:
wherein the first and second gaps a, b further satisfy the
following condition; b-a.ltoreq.2 microns.
14. The hydrodynamic bearing assembly according to claim 12:
wherein the first and second gaps a, b further satisfy the
following condition; b-a.ltoreq.0.5 microns.
15. The hydrodynamic bearing assembly according to any one of
claims 12 to 14: wherein a perpendicularity between said shaft and
said thrust plate formed or secured thereon is approximately 0.7
microns/20 millimeters.
16. The hydrodynamic bearing assembly according to any one of
claims 12 to 15: wherein grooves for generating a thrust dynamic
pressure due to the relative rotation between said second thrust
plate and the other end surface of said shaft are formed on any one
of said second thrust plate and the other end surface of said
shaft.
17. The hydrodynamic bearing assembly according to any one of
claims 12 to 15: wherein any one of said second thrust plate and
the other end surface of said shaft has a spherical, cone, or
frustum boss.
18. The hydrodynamic bearing assembly according to any one of
claims 1 to 17: wherein one or more of opposing portions among said
shaft, said sleeve, said thrust plate, and said thrust opposing
surface, and said second thrust plate are made of ceramics
material.
19. The hydrodynamic bearing assembly according to claim 18:
wherein the ceramics material is selected from a group consisting
of alumina, zirconia, silicon carbide, silicon nitride, and
sialon.
20. A spindle motor incorporating the hydrodynamic bearing assembly
according to any one of claims 1 to 19.
21. A memory device or a bar code scan reader incorporating the
spindle motor according to claim 43 or 44.
22. A process for preventing a dew generated in a hydrodynamic
bearing assembly, in which a rotational and stationary members
having opposing surfaces are arranged with predetermined gaps
therebetween in a radial direction perpendicular to an axis and in
a thrust direction parallel to the axis, and a fluid intervening in
the gaps generates dynamic pressures due to the relative rotation
between the rotational and stationary members so that the
rotational member can be rotated without any contact to the
stationary member, said process comprising the steps of: modifying
a depth of the grooves formed on either rotational and stationary
members in accordance with a position of the grooves to cause a
dynamic pressure distribution to be kept substantially even, to
reduce the peak pressure required to obtain a floating force,
thereby to prevent the dew from being generated because of the
compression in the opposing surfaces.
Description
TECHNICAL FIELD
[0001] This invention relates to a hydrodynamic bearing assembly
incorporated with a spindle motor used for driving a memory device
such as a hard disk drive (referred to as a "HDD", hereinafter), or
a bar code scan reader, and in particular, relates to the
hydrodynamic bearing assembly, which has an improved activation
feature.
BACKGROUND ART
[0002] The hydrodynamic bearing assembly for use in a spindle motor
of the memory device such as the HDD and a drive unit for driving a
polygonal mirror in the bar code reader has been required to attain
a high rate, stable rotation under a high load, to have a high
bearing rigidity that prevents the rotational member from making
contacts with the stationary member even under the existence of
external vibrations, thereby to have a reduced starting torque, and
to have an improved activation feature with a reduced wear caused
by the frictional rotations.
[0003] FIG. 8 illustrates one example of a conventional spindle
motor. In the drawing, a column shaft 1 and a disk-shaped thrust
plate 4 are secured on a base 10. The thrust plate 4 is attached
perpendicularly to the shaft. 1. The shaft includes an outer
surface parallel to the axis of the shaft 1. A cylindrical hollow
sleeve 3 is rotatably arranged around the outer surface of the
shaft 1 with a predetermined gap so that a hydrodynamic bearing is
defined between the shaft 1 and sleeve 3. Thus, the radial bearing
is defined between the outer surface of the shaft 1 and the inner
surface of the sleeve 3 for generating a radial dynamic pressure in
the radial direction perpendicular to the axis. Also, a thrust
bearing is defined between a bottom surface of the sleeve 3 (which
is referred to as a thrust-opposing surface of the sleeve 3,
hereinafter) and the thrust plate 4 for generating a thrust dynamic
pressure in the thrust direction parallel to the axis. Grooves 5
for generating the thrust dynamic pressure are formed on a surface
of the thrust plate 4 opposing to the thrust-opposing surface of
the sleeve 3. A rotor 17 is attached with the sleeve 3 such that it
can rotate together with the shaft 1 around the sleeve 3.
Information media (in case of the HDD) or a polygonal mirror (in
case of the bar code scan reader) is mounted on the outer surface
of the rotor 17. A rotor magnet 18 is attached on the inner surface
of the rotor 17 and opposes to a stator coil 19 mounted on a base
10.
[0004] FIG. 9 illustrates a detail of the grooves 5 formed on the
thrust plate 4 for generating the thrust dynamic pressure. A
plurality of spiral grooves is formed on the thrust plate 4, and
each groove is designed to be inclined with a predetermined angle
relative to the circle and generally has a width of several microns
(approximately 1 to 5 microns). Although FIG. 9 shows the grooves 5
formed on the thrust plate 4, the grooves 5 may be formed on the
thrust-opposing surface 13.
[0005] In the rotation of the spindle motor so constructed, the
stator coil (not shown) energized by the electric flow generates
the attraction/repulsion force. This provides a rotation driving
force with the rotor 17 having the rotor magnet 18 to rotate both
of the rotor 17 and the sleeve 3 secured thereto around the shaft
1. A relative movement between the shaft 1 and the sleeve 3 due to
the rotation generates the radial dynamic pressure through a fluid
such as air intervened therebetween. Also, the relative movement
between the thrust plate 4 and thrust-opposing surface 13 of the
sleeve 3 in cooperation of the grooves 5 generates the thrust
dynamic pressure. The radial and thrust dynamic pressures keep the
rotational member such as the sleeve 3 and the rotor 17 away from
the stationary member such as shaft 1 and the thrust plate 4 during
the rotation.
[0006] FIG. 10 is an enlarged perspective view of the hydrodynamic
bearing assembly in isolation used for the spindle motor of FIG. 8.
In the drawing, the thrust plate 4 is secured on one end of the
shaft 1 perpendicular to the axis. The sleeve 3 indicated by a
phantom line is rotatably arranged around the outer surface of the
shaft 1. When the spindle motor is energized to activate, the
rotational member such as sleeve 3 starts to rotate in contact with
the thrust plate 4 due to its own weight. The spiral grooves 5 in
cooperation with the rotation of the sleeve 3 indicated by the
arrow 6 conducts the fluid such as air into the thrust bearing
between the sleeve 3 and the thrust plate 3 and forces the fluid
towards the center of the thrust plate 4 along the direction
indicated by the arrow 7. A land portion 9 is defined between the
spiral grooves 5 and the outer surface of the shaft 1, in which the
forced fluid are compressed between the land portions 9 and inner
end portions of the grooves 5 so as to generate the dynamic
pressure for supporting the sleeve 3. Thus, according to the
conventional hydrodynamic bearing assembly, the thrust dynamic
pressure to be generated has peaks localized adjacent to the land
portion 9.
[0007] FIG. 10 shows another example of the hydrodynamic bearing
assembly, having the shaft 1 with another grooves 2 offset to the
axis, which are formed on the outer surface and opposes to the
inner surfaces of the sleeve 3. The grooves 2 are not, essential to
generate the radial dynamic pressure. However, the rotational
member such as the sleeve 3 in the drawing rotates around the
bearing axis of the stationary member such as the shaft 1, and also
it may whirl (revolve) around another axis offset to the bearing
axis, which is referred to as a half-whirl phenomenon. The
half-whirl results in whirling of the functional components such as
the information media and the polygonal mirror mounted on the rotor
17, thereby to cause malfunctions in utilizing the components. The
grooves 2 formed on the outer surface of the shaft 1 advantageously
avoid the half-whirl. The grooves may have various configurations
for avoiding the half-whirl, including the offset grooves as shown
in the drawing, grooves parallel to the axis, and the
herringbone-shaped grooves. However, when the grooves 2 are offset
to the axis, advantageously, the rotation of inner surface of the
sleeve 3 in the direction indicated by the arrow 8 forces the fluid
from the top end to the bottom end due to its viscosity, thereby
further increasing the dynamic pressure in the thrust bearing.
Also, the grooves 2 may be formed on the inner surface of the
sleeve 3, rather than on the outer surface of the shaft 1.
[0008] The dynamic pressure distribution in the radial bearing is
generated similar to that of the thrust bearing. Thus, the rotation
of the sleeve 3 in the direction opposing to the grooves as
indicated by the arrow 8, in cooperation with its viscosity, forces
the fluid in the grooves from the upper end (right side of the
drawing) to the lower end (left side). To this end, it is assumed
that the dynamic pressure distribution is uneven, increasing the
dynamic pressure adjacent the bottom ends of the groove 2.
[0009] FIG. 11 illustrates the hydrodynamic bearing assembly having
the sleeve 3, which is whirled and inclined relative to the shaft 1
and the thrust plate 4 due to the external factors applied to the
bearing assembly. The sleeve 3 is inclined counterclockwise
relative to the shaft 1, the shaft 1 moves closer to the sleeve 3
at the upper right portion A and the lower left portion B in the
drawing. Also, the thrust plate 4 moves closer to the
thrust-opposing surface at the leftmost portion C.
[0010] The parallel lines indicated in the drawing schematically
illustrates the dynamic pressure in the radial and thrust bearings
when the shaft 1 is inclined relative to the sleeve 3. As the shaft
1 moves closer to the sleeve 3 adjacent to the portion A, the wedge
effect due to the convolution of the fluid therebetween generates
the higher dynamic pressure. The same effect is observed adjacent
to the portion B. Therefore, the counter forces due to the dynamic
pressure adjacent to the portions A and B are generated against the
contacting force between the shaft 1 and sleeve 3, and the contact
between the shaft 1 and the sleeve 3 is avoided unless the external
oscillation force overcomes the counter forces.
[0011] Meanwhile, the fluid is guided from the circumference of
thrust plate 4 towards the axis so that the dynamic pressure
between the thrust plate 4 and the thrust-opposing surface 13 is
lower adjacent to the portion C and greater towards the bearing
axis. Thus, the dynamic pressure around the portion C is low even
if the thrust plate 4 and the thrust-opposing surface 13 moves
closer. This may cause the thrust plate 4 and the thrust-opposing
surface 13 to contact with each other around the portion C, when
the external force such as the oscillating motion is applied. Once
they contact with each other, the friction force therebetween
results the unstable rotation of the rotational member. Further,
the rebound followed by the contact causes the undesired impact,
which could bring the malfunction of the magnetic head used for the
HDD, or could result an extensive damage to the spindle motor.
[0012] As can be seen from the above description, the conventional
hydrodynamic bearing assembly has following several disadvantages.
Firstly, the high dynamic pressure distribution has peaks localized
in certain portions within the gap defined by the hydrodynamic
bearings, and if the compressed fluid is gas such as air, then the
air locally compressed in the portions may generate a dew in the
portions, because the water vapor in the air is compressed. The dew
may cause no adverse effect while the bearing assembly keeps
rotating so that the continuous flow of the fluid blows off the
dew. However, when the electric power is interrupted and the
rotation is halted, the sleeve 3 stops and contacts with the thrust
plate 4 while the dew is remained therebetween. To this end, the
dew between the sleeve 3 and the thrust plate 4 causes them to
closely fit with each other, resulting some activating
disadvantages when the bearing assembly is restated.
[0013] Secondary, when the external forces is applied to the
bearing assembly to tilt the shaft relative to the sleeve, the
thrust bearing generates the dynamic pressure (particularly in the
circumference thereof) insufficiently to bear the external forces.
Thus, the sleeve 3 and the thrust plate 4 are likely to contact
with each other due to such external forces. This comes from the
fact that the dynamic pressure distribution has the peak in the
portions adjacent to the bearing axis.
[0014] Thirdly, when the spindle motor is halted, the weight of the
rotational member causes the thrust opposing surface 13 as the
bottom surface of the sleeve 3 to contact with the thrust plate 4.
Both of the thrust opposing surface 13 and the sleeve 3 have even
surfaces allowing a full contact therebetween. When the spindle
motor is restated, the greater activating torque is required enough
to overcome the friction force of the full contact. Thus, the
capability of the motor should be increased, resulting more energy
consumption. Further, the rotational member rotates relative to and
in contact with stationary member until the dynamic pressure
generated therebetween is enough for floating the rotational member
away from the stationary member. The rotation in contact causes
both members to wear and generate the abraded particles, which
brings an adverse effect to the precise bearing assembly. Even
worse, the seizure of the bearings may be caused due to the greater
friction force. Thus, the rotation in contact reduces the endurance
and the reliability of the hydrodynamic bearing assembly or the
spindle motor incorporating thereof.
[0015] The conventional techniques have proposed various approaches
in order to address those disadvantages. For example, Japanese
Patent Laid-Open Publication Nos. 11-18357 and 11-55918 disclose,
in particular, the technique for preventing the contact in the
thrust bearing due to the tilt of the shaft. According to the prior
art techniques, the coil is arranged eccentrically with the rotor
magnet so that the shaft is biased against the sleeve in a
predetermined direction, allowing the rotation in a stable manner.
However, the art requires the coil to be positioned concentrically
with the bearing assembly for biasing the shaft to the sleeve in
parallel. This arrangement is often impossible because of the
design restriction.
[0016] Japanese Utility Model Laid-Open Publication No. 55-36456
discloses the stationary permanent magnet attached to the housing
opposing to the rotor magnet for tilting the rotor towards the
predetermined direction for the rotation. However, this technique
decreases the lifetime of the bearing assembly because the distal
edge of the shaft contacts with the sleeve.
[0017] Also, Japanese Patent Laid-Open Publication No. 60-234120
discloses the technique for reducing the torque when activated, in
which at least one of the thrust opposing surface and the thrust
plate has a convex configuration in the thrust bearing as
illustrated in FIG. 12. In the drawing, a plurality of grooves 32
for generating the dynamic pressure are formed on the thrust plate
31, and the shaft 33 rotates around the axis in a direction
indicated by the arrow w. A thrust member 34 is secured on the
bottom end, which opposes to the thrust plate 31. The thrust member
34 and the thrust plate 31 together define the thrust bearing. The
thrust member is designed such that it has a spherical surface 35
with a predetermined radius R opposing to the thrust plate 31 and
the spherical surface 35 protrudes by the protruding thickness
N.
[0018] The rotation keeps the thrust plate 31 and thrust member 34
of the thrust bearing away from each other, and the halt of the
rotation causes them in a small region. However, since the
contacting area is a pinpoint, the rotational member can be
activated and floated without any excessive torque nor galling.
[0019] However, the central portion of the spherical surface 35 may
contact with the thrust plate 31 and the rotational member may not
float away therefrom depending upon the radius R and the protrusion
N of the spherical surface 35. Also, the thrust member has to be
processed such that it has the spherical surface 35, in which it is
difficult to form such a small convex surface.
[0020] Also, Japanese Patent Laid-Open Publication No. 9-328381
discloses the technique for reducing the frictional coefficient
between the thrust contacting surfaces by forming one of the thrust
contacting surfaces of an amorphous hard carbon film and the other
of ceramics material having the void occupying rate of 6% or less
and the maximum diameter of 10 microns or less.
[0021] However, while the reduced torque corresponding to the
reduced frictional coefficient due to the fixed lubricant film is
observed, no further improvement can be expected by this technique,
because the thrust contacting surfaces are kept in full contact
therebetween, the circumference thereof, in particular, has the
longer radius, which can be harmful to contribute the reduction of
the activation torque.
SUMMARY OF THE INVENTION
[0022] This invention is to provide a hydrodynamic bearing
assembly, which eliminates the disadvantages of the conventional
technique, realizes an even dynamic pressure distribution without
extreme pressure peaks localized in the grooves for generating
dynamic pressure to avoid the malfunctions because of the dew in
the hydrodynamic bearings, and endures against the external
oscillation applied to the thrust bearing. Also, this invention
includes a process for preventing the dew from being generated in
the hydrodynamic bearings by causing the dynamic pressure
distribution in the grooves to be kept substantially even.
[0023] Further, this invention is to provide a hydrodynamic bearing
assembly, in which the full contact between the thrust plate and
the thrust opposing surface is avoided so that the friction and the
energy consumption are reduced in comparison with the prior art
technique, a reduced activating torque ensures the rotational
member to float away from the stationary member, and the high
rigidity against the external oscillation is achieved.
[0024] Even further, this invention is to provide a spindle motor
rotating in a stable manner with an improved activation feature and
the tilt rigidity as described, and also to provide a memory device
and a bar code scanning device with the improved endurance and the
reliability.
[0025] In particular, one aspect of the present invention is to
provide the hydrodynamic bearing assembly comprises: a rotational
and stationary members arranged with predetermined gaps to each
other, the gaps including a radial gap defined therebetween in a
radial direction perpendicular to an axis or a thrust gap defined
therebetween in a thrust direction parallel to the axis, the gap
containing a fluid for generating a radial or thrust dynamic
pressure due to the relative rotation between the rotational and
stationary members so that the rotational member rotate relative to
the stationary member without any contact; wherein said rotational
and stationary members have opposing surfaces to each other, and
any one of the opposing surfaces has grooves formed thereon, the
grooves having the depth modified in accordance with its position
so that the dynamic pressure generated across the gap is
substantially even. The depth of the grooves is modified to level
the dynamic pressure so that the counter force against the tilting
force in the thrust bearing is improved, and the dew generated
under the peak pressure in the thrust bearing can be reduced. In
order to level the dynamic pressure in the grooves, preferably, the
grooves have the depth modified gradually and smoothly towards the
downstream flow of the fluid passing therethrough.
[0026] Another aspect of the present invention is to provide the
hydrodynamic bearing assembly comprises: a disk-shaped thrust plate
extending in a radial direction perpendicular to a bearing axis; a
circular thrust opposing surface extending in the radial direction
and opposing to the thrust plate; and a thrust bearing for
generating the thrust dynamic pressure in a thrust direction
parallel to the bearing axis due to a relative rotation between the
thrust plate and the thrust opposing surface; wherein at least one,
or both of opposing surfaces of the thrust plate and the thrust
opposing surface are inclined such that a distance between both
opposing surfaces thereof becomes greater from an inside portion
towards an outer portion of the thrust bearing. The gradient is
preferably formed such that it hardly has an influence to the
thrust dynamic pressure but avoids the contact due to the tilt. In
particular, preferably, either one or both of opposing surfaces in
the thrust bearing have the frustum or spherical configuration, and
the distance difference of the gradients are within approximately 2
microns or less.
[0027] Further another aspect of the present invention is to
provide the hydrodynamic bearing assembly comprises: a radial
bearing including a column shaft having an outer surface parallel
to an axis, and a hollow cylindrical sleeve having an inner surface
rotatably arranged around the outer surface of the shaft, the
radial bearing for generating a radial dynamic pressure due to a
relative rotation between the sleeve and the shaft; a thrust
bearings including a thrust plate formed or secured on one end
surface of the shaft along the axis, and a thrust opposing surface
formed or secured on one end surface of the sleeve along the axis,
the thrust bearing for generating a thrust dynamic pressure due to
the relative rotation between the thrust plate and the thrust
opposing surface; and a second thrust plate covering the hollow
cylindrical sleeve at the other end surface along the axis; wherein
a first gap a defined parallel to the axis between the other end
surface of the shaft and the second thrust plate, and a second gap
b defined parallel to the axis between the thrust plate and thrust
opposing surface satisfy the following condition;
a<b.
[0028] The second thrust plate is provided so that the total weight
of the rotational member is supported between the second thrust
plate and the shaft when the bearing assembly is halted. This
reduces the arm length of the friction when restarting (activating)
the bearing assembly, thus eventually improving the activation
feature of the hydrodynamic bearing assembly. Preferably any one of
the second thrust plate and the other end surface of the shaft has
a spherical, cone, or frustum boss, because the contacting points
between the second thrust plate and the shaft can be closer to the
axis.
[0029] Further another aspect of the present invention is to
provide the hydrodynamic bearing assembly, in which one or more of
opposing portions among the shaft, the sleeve, the thrust plate,
and the thrust opposing surface, and the second thrust plate are
made of ceramics material. The ceramics material is selected from a
group consisting of alumina, zirconia, silicon carbide, silicon
nitride, and sialon. Usage of the ceramics material having high
anti-abrasion reduces the friction during the rotation with contact
and provide the hydrodynamic bearing assembly having the high
rigidity and the high accuracy.
[0030] Further another aspect of the present invention relates to
provide a spindle motor incorporating the hydrodynamic bearing
assembly having the improved activating feature and the high
rigidity as described above, as well as to provide a memory device
or a bar code scan reader incorporating the spindle motor. This
invention is to provide such products having the high stability,
the high reliability, and the high endurance.
BRIEF DESCRIPTION OF DRAWINGS
[0031] FIG. 1 is a schematic view illustrating the relation between
the configuration of the groove and the dynamic pressure
distribution.
[0032] FIG. 2 is a computer analysis chart of the dynamic pressure
distribution generated by the conventional thrust groove.
[0033] FIG. 3 is a computer analysis chart of the dynamic pressure
distribution generated by the thrust groove according to the
present invention.
[0034] FIG. 4 is a cross section of one embodiment of the
hydrodynamic bearing assembly according to the present
invention.
[0035] FIG. 5 is a cross section of an alternative of the
hydrodynamic bearing assembly in FIG. 4.
[0036] FIG. 6 is a cross section of another embodiment of the
hydrodynamic bearing assembly according to the present
invention.
[0037] FIG. 7 is a cross section of an alternative of the
hydrodynamic bearing assembly in FIG. 6.
[0038] FIG. 8 is a cross section of the spindle motor incorporating
the conventional hydrodynamic bearing assembly.
[0039] FIG. 9 is a perspective view of the prior art thrust plate
having thrust grooves for generating dynamic pressure.
[0040] FIG. 10 is a enlarged perspective view illustrating only the
hydrodynamic bearing assembly in FIG. 8.
[0041] FIG. 11 is a schematic view of the dynamic pressure
distribution in the hydrodynamic bearing shown in FIG. 8.
[0042] FIG. 12 is a partially fragmentary side view of the
embodiment of the prior art hydrodynamic bearing assembly having
the activating feature.
PREFERRED EMBODIMENTS OF THE INVENTION
First Embodiment
[0043] The first embodiment of the hydrodynamic bearing assembly
according to the present invention will be described hereinafter
with reference to the drawings. As described above, a plurality of
spiral grooves conducts the fluid such as air from the
circumference so that the thrusting dynamic pressure is generated
in the thrust bearing. The relative rotation in the thrust bearing
pumps the conducted fluid towards the inner portion (adjacent to
the outer surface of the shaft) so that the dynamic pressure
adjacent to the shaft is increased, thereby to keep the rotational
member floating. The computation of the dynamic pressure
distribution indicates the high value adjacent to center of the
axis, or adjacent to the land portion 9 and the edges of the
grooves 5 when the land portion 9 is provided as shown in FIG. 10.
Also, the computation of the dynamic pressure distribution
indicates the gradually decreasing value towards the circumference.
Thus, the conventional hydrodynamic bearing assembly generates the
peaks of the dynamic pressure distribution localized in some
portions in the thrust bearing.
[0044] FIG. 1(a) shows a schematic cross section of the groove 5
formed on the conventional thrust plate 4. As described above, the
groove 5 has a spiral configuration when viewing from the top,
meanwhile, FIG. 9 shows a cross section taken in radial direction,
on which the groove 5 is projected. In the drawing, X and Y
represent the radial positions of the outer surface of the shaft
(or the inside portion of the thrust plate 4) and the outer surface
of the thrust plate 4, respectively. A land portion 9 is extended
from the outer surface of the shaft by a predetermined distance. In
FIG. 1(a), the groove 5 having a depth h is formed on the thrust
plate 4, followed by the land portion 9. The lines a-a and b-b
represent the top surface of the thrust plate 4 and the bottom
surface of the groove 5, respectively. The arrow 7' represents the
direction of the fluid flow in the groove 5, and the arrow 7 shown
in FIG. 10 is illustrated on the radial cross section as the arrow
7'. The groove 5 shown in FIG. 1(a) has a radially outer end open
to the circumstance. The conventional spiral groove 5 has an even
depth of h as illustrated in FIG. 1(a).
[0045] The dotted line J in FIG. 1(a) represents the dynamic
pressure distribution generated by the conventional grooves 5
during the rotation, which is also illustrated on the radial cross
section. The transverse S-axis in FIG. 1(b), corresponding to FIG.
1(a), indicates the distance from the inside portion of the thrust
plate 4 towards the circumference, and also the vertical P-axis
shows the dynamic pressure. Both of the inner portion and the outer
portion of the thrust bearing are presumably open to the
circumstance so that the dynamic pressure in each portion has the
same dynamic pressure. The dynamic pressures at the inner and outer
portions equal to one of the circumstance, i.e., zero. When the
radial bearing connects with the thrust bearing, the dynamic
pressure distribution has a tendency similar to that indicated by
the dotted line J in FIG. 1(b) except that the dynamic pressure at
the inner portion is increased. The dynamic pressure distribution
of the prior art thrust bearing has the peak pressure Pa at the
position adjacent to the groove edge and near the bearing center,
and gradually decreases towards the circumference, as shown by the
dotted line J in FIG. 1(b). The computations that the present
inventors have conducted revealed the peak pressure Pa is
approximately 1.59 atmospheric pressure (which is referred to
simply as "atm"), in which the diameter of the thrust bearing is 20
millimeters, the thrust inner diameter is 14 millimeters, the
groove width h is 2 microns, the bearing clearance during the
rotation is 1.5 microns, and the rotation rate is 16,000 rpm. It is
difficult to directly measure the dynamic pressure based upon the
experimental data, however, the computation of the dynamic pressure
is consistent with the floating height (clearance) and the load
capacity (rotational mass), thus the computation is presumably
acceptable.
[0046] According to the present invention, the groove depth is
modified according to the position between the inner portion and
the outer portion along the transverse S-axis so that the uniform
dynamic pressure distribution is generated in the radial direction.
For example, the bottom surface b'-b' inclined relative to the top
surface of the thrust plate 4 so that the groove depth is gradually
decreased from h1 at a position of the outer portion to h2 at the
inner end of the groove 5 interfacing with (in the vicinity of) the
land portion 9. Like reference numerals are used as in FIG. 1(a).
The conventional thrust bearing conducts and then compresses the
fluid adjacent the interface portion between the land portion 9 and
the groove 5, thereby to generate the dynamic pressure distribution
with the peak in certain portions. Meanwhile, the thrust bearing
according to the embodiment includes the grooves gradually
shallower from h1 to h2 as the position moves towards inner portion
so that the dynamic pressure distribution is increased, in
particular, adjacent to the outer end of the groove.
[0047] The hydrodynamic bearing assembly according to the present
invention has the groove depth designed as described above so that
the dynamic pressure is substantially leveled. When the groove 5
has the depth h1 of 2 microns at the outer end of the groove 5 and
the depth h2 of 0.4 microns at the inner end of the groove 5
adjoining to the land portion 9, and the other dimensions and the
rotation rate are identical to those indicated above, the peak
pressure Pc is about 1.31 according to the computations conducted
by the present inventors. The uniform dynamic pressure is
substantially generated in the radial direction. Thus, although the
peak pressure Pc is less than the peak pressure Pc of the
conventional thrust bearing, the dynamic pressure of the embodiment
is evenly distributed and the total bearing force thereof is the
same as that of the conventional thrust bearing. Rather, the
hydrodynamic bearing assembly of the embodiment has an advantage
that it realizes a good counter force bearing an external vibration
because the dynamic pressure at the portions away from the center
is relatively high.
[0048] FIGS. 2 and 3 are schematic views of the computations that
the present inventors have conducted for the dynamic pressure
distribution in the thrust bearing. In particular, FIG. 2(a)
illustrates the computer analysis chart of the dynamic pressure
distribution generated by the grooves of the prior art thrust
bearing, and FIG. 2(b) is a graph of FIG. 2(a) illustrating some
area of the dynamic pressure distribution. Both drawings illustrate
the thrust bearing viewed from the above. As shown in FIG. 2(b),
the highest pressure areas including the peak pressure point of
1.59 atms appear adjacent to the interface portions between the
grooves 5 and the land portion 9, and another narrow pressure areas
of approximately 1.4 atms appear around the highest pressure areas.
Further, outside the narrow pressure areas, the dynamic pressure is
gradually decreased. Thus, the areas of the high dynamic pressure
are localized so that the configuration of the groove can hardly
traced from the drawings.
[0049] Although similar to FIGS. 2(a) and 2(b), FIGS. 3(a) and 3(b)
shows the present invention. FIG. 3(a) is the computer analysis
chart of the dynamic pressure distribution generated by the grooves
of the thrust bearing according to the present invention, and FIG.
3(b) is a graph of FIG. 3(a) illustrating some areas of the dynamic
pressure distribution. As can be seen also from FIG. 3(b),
according to the present invention, while the peak pressure is 1.31
atms that is relatively low, the highest pressure areas are evenly
distributed so that the configuration of the groove can readily be
recognized. Also, the highest pressure areas are further surrounded
by another pressure areas of 1.2 atms. Those figures show that
while the conventional thrust bearing supports the weight of the
bearing assembly mainly and/or exclusively adjacent to the
interface portions between the land portion 9 or the shaft and the
grooves, the thrust bearing of the present invention supports the
weight across the whole thrust surface.
[0050] The conventional hydrodynamic bearing assembly is activated,
for example, with use of the fluid such as air containing humidity
of 60% at the room temperature, the dew is very likely caused due
to the maximum pressure of 1.59 atms. However, the dew is hardly
caused under the maximum pressure of 1.31 atms according to the
present invention. Thus, this reduces the risk that the
hydrodynamic bearing assembly cannot be restarted due to the dew
remaining in the thrust bearing.
[0051] The hydrodynamic bearing assembly of the embodiment is
pump-in type, in which the fluid for generating the dynamic
pressure is conducted from the circumference towards the inside
portion of the thrust bearing. The spiral grooves 5 have
configurations designed such that the rotation of the sleeve 3
directs the fluid from the circumference to the inside portion.
Alternatively, a pump-out hydrodynamic bearing assembly may be
configured by reversing the spiral direction of the grooves 5 or
the rotational direction of the sleeve 3. The pump-out hydrodynamic
bearing assembly, in general, conducts the fluid from the inside
portion adjacent to the shaft towards the circumference of the
thrust plate to generate the dynamic pressure increasing radially
outwardly. Therefore, if the present invention is applied to the
pump-out hydrodynamic bearing assembly, the grooves 5 is designed
to have the gradient of the depth formed in a reversed manner
relative to one indicated in FIG. 2(c), on the radially outer
surface when the land portion 9 is provided.
[0052] Another dotted line M in FIG. 1(c) illustrates the
alternative of the embodiment, in which the groove 5 is designed to
have steps discretely decreasing the depth towards the inside
portion, rather than the gradient gradually decreasing. Such
stepped groove 5 may readily be formed with use of the shotblasting
or the laser machining in combination with masks. The resultant
stepped grooves 5 have the advantage similar to that with slope
described above. However, for the smooth flow of the fluid, the
more preferably, the more steps are formed and the less depth by
one step is provided.
[0053] Although the embodiment has been described above, in which
the grooves 5 for generating the dynamic pressure are formed on the
thrust plate 4, the grooves may be formed on the thrust opposing
surface 13 opposing to the thrust plate 4. Further, the groove
depths are illustrated in the drawings that they are designed to
have linear slope or steps with regular intervals and heights,
however, the purpose of the present invention is to generate the
substantially uniform dynamic pressure distribution in the thrust
bearing. Therefore, the groove depth may be formed such that they
follow, for example, on a smoothed curve or on steps with different
intervals and heights for leveling the dynamic pressure
distribution.
[0054] The description has been made for the embodiments with
respect to the thrust bearing, the present invention can also be
applied to grooves 2 in the radial bearing. While the groove of the
pump-in thrust bearing is designed as being shallower from the
circumference towards the shaft, the radial bearing has the groove
2 designed such that it is gradually shallower from the top end to
the bottom end. Also, the groove 2 may be formed on the inner
surface of the sleeve 3 rather than the outer surface of the shaft
1.
EXAMPLE 1
[0055] As illustrated in FIG. 9, twelve grooves, each of which has
the maximum depth of 2 microns and the width of 15 degrees, are
formed on an area for generating the dynamic pressure in the thrust
bearing, which has an outer diameter of 20 millimeters and an inner
diameter of 14 millimeters. The land portion without any grooves
formed thereon is provided in a region 0.75 millimeters away from
the shaft. The spiral groove 5 is designed such that angle defined
by the line between the inner end thereof and the center of the
axis and the line between the outer end thereof and the center of
the axis is 45 degrees.
[0056] (Sample A)
[0057] Sample A has the groove formed by the laser process, of
which depth is linearly decreasing towards the axis. The groove
includes the depth hi (shown in FIG. 1(c)) of 2 microns at the
outer end for conducting the fluid and the depth h2 (shown in FIG.
1(c)) of 0.4 microns at the inner end interfacing the land portion
9.
[0058] (Sample B)
[0059] Sample B has the groove formed by the blasting with use of a
plurality of masks, of which depth is stepped as illustrated by the
dotted line M in FIG. 1(c). The groove includes three steps with
the regular intervals and the heights. As shown by the dotted line
M in FIG. 1(c), the outer step has the depth h1 of 2 microns at the
outer end for conducting the fluid and the inner step has the depth
h2 of 0.4 microns at the inner end interfacing the land portion
9.
[0060] (Sample C)
[0061] Sample C is a comparative sample and has the groove formed
by the blast process with use of one mask, of which depth is 2
microns across the groove, i.e., the step height of the step at the
inner end interfacing the land portion 9 is 2 microns.
[0062] The hydrodynamic bearing assemblies incorporating those
Samples supported a weight of 200 grams in total, and rotated on
the plain plate at the rotation rate of 16,000 rpm for 10 minutes.
The hydrodynamic bearing assemblies were evaluated as to whether
the dew is generated therein by checking whether the hydrodynamic
bearing assemblies were able to be rotated again soon after they
were halted. Also, the floating heights during the stable rotation
were measured. The room temperature was kept at 25 degrees Celsius,
and the humidity were changed from 50% to 100%. Those tests were
repeated 5 times at each humidity.
[0063] The tests revealed the highest humidities capable of
rotating the hydrodynamic bearing assemblies in a reliable manner
were 90%, 85%, and 75% for Sample A, B, and C, respectively. Each
of the samples had the floating heights of approximately 1.5
microns during the reliable rotation, which are consistent with the
computations.
Second Embodiment
[0064] Next, the second embodiment of the hydrodynamic bearing
assembly according to the present invention will be described
hereinafter with reference to the drawings. FIG. 4 illustrates the
hydrodynamic bearing assembly of the embodiment. The sleeve 3 is
rotatably arranged around the cylindrical shaft 1 with a
predetermined gap. Also the shaft 1 has the bottom end secured on
the thrust plate 4a perpendicularly to the axis. The thrust plate
4a includes the grooves for generating the dynamic pressure, formed
on the surface thereof opposing to the bottom surface of the sleeve
3. The bottom surface of the sleeve 3 opposing to the thrust
grooves 5 is referred to as the thrust opposing surface 13.
[0065] During the rotation of the hydrodynamic bearing assembly so
constructed, the rotational driving force between the stator coil
and the rotor magnet (not shown) of the spindle motor rotates the
rotor and the sleeve 3 secured with the rotor, so that the radial
dynamic pressure is generated between the shaft 1 and the sleeve 3
due to the relative rotation therebetween. Also, the relative
rotation between the thrust opposing surface 13 of the sleeve 3 and
the thrust plate 4a generates the thrust dynamic pressure. The
radial and the thrust dynamic pressures rotate the rotational
member such as the sleeve 3 around the shaft 1 without any
contact.
[0066] According to the embodiment, the surface of the thrust plate
4a having the thrust grooves is inclined relative to the thrust
opposing surface 13 such that the surface of the thrust plate 4a
deviates from the thrust opposing surface 13 towards the radial
circumference. The inclined surface keeps the thrust plate 4a away
from the sleeve 3 because the circumference H of the thrust plate
has the greater gap from the thrust opposing surface 13, even if
the external force is applied. To this end, the sleeve 3 is
inclined relative to the shaft 1 and the thrust plate 4a as
illustrated by the solid line in FIG. 4.
[0067] Needless to say, if the inclining angle of the thrust plate
4a relative to the sleeve 3 is greater, then preferably, the risk
of the contact between the sleeve 3 and the thrust plate 4a is
smaller. However, since the thrust plate 4a generates the thrust
dynamic pressure in cooperation with the thrust opposing surface
13, when the inclining angle is so great, the sufficient thrust
dynamic pressure can hardly expected. Therefore, preferably, the
inclining angle is great enough to avoid the contact and small
enough to generate the thrust dynamic pressure sufficiently.
[0068] In the vertical cross section of FIG. 4, the thrust plate 4a
is linearly formed such that it has the inclining angle of the
gradient d as the height along the axis for the thrust width s from
the innermost portion of the thrust bearing (outer surface of the
shaft 2 arranged on the thrust plate 4) to the outermost portion
thereof (the circumference of the thrust plate 4). The thrust plate
4a having the thrust grooves 5, as whole, has in a frustum shape.
As illustrated by the dotted lines in FIG. 4, the shaft 1 operates
in a normal manner, and radial gaps between the shaft 1 and the
sleeve 3 and the total gap thereof are referred to as f1, f2, and F
(f1+f2), respectively, also the radial bearing has the length L
along the axis. Then, the maximum gradient of the sleeve relative
to the shaft in the radial bearing is;
F/L
[0069] Therefore, in order to prevent the contact of the
circumference H of the thrust plate 4a, the maximum gradient of the
thrust plate 4a relative to the sleeve in the thrust bearing is
less than that in the radial bearing. That is, the contact in the
thrust bearing is avoided if the following condition is
satisfied;
F/L.ltoreq.d/s
[0070] The experiments that the present inventors have conducted
revealed that if the length L of the thrust bearing is 15
millimeters, the total radial gap F is 4 microns, the diameter D of
the thrust plate 4 is 20 millimeters, and the gradient d is 2
microns, any contact in the thrust bearing can be avoided
satisfactorily. Preferably, the gradient d of approximately 2
microns or less ensures the thrust dynamic pressure
sufficiently.
[0071] Another advantage of the hydrodynamic bearing assembly
according to the embodiment is capability of reducing the torque
required at the activation of the hydrodynamic bearing assembly. As
can be seen also from FIG. 4, when the hydrodynamic bearing
assembly according to the embodiment is halted, the total weight of
the rotational member secured with the sleeve 3 is applied to a
portion of the thrust plate 4 where the thrust opposing surface 13
of the sleeve 3 and the thrust plate 4 contact with each other. On
the contrary, in the conventional hydrodynamic bearing assembly,
the full contact across the thrust opposing surface 13 of the
sleeve 3 and the thrust plate 4 is caused when halted. Thus, the
embodiment limits the contact surface thereof to a ring-shaped area
of the thrust opposing surface 13 closer to the bearing axis, which
is located only inside end portion.
[0072] Therefore, when restarting the hydrodynamic bearing assembly
according to the present invention, since the weight is applied
mainly to the portion adjacent to the rotation center, the arm
length from the center weighted portion is reduced so that the
torque can be minimized. This causes various advantages, including
for example, reducing the capacity and the dimension of the motor,
reducing the electric power consumption, and minimizing the
rotation with contact by quickly increasing the rotation rate so as
to reduce the abrasion between the contacting members, eventually
extending the effective lifetime of the bearing assembly.
[0073] FIG. 5 is a partially enlarged view of the alternative
embodiment of the hydrodynamic bearing assembly, illustrating one
side of the thrust bearing. Similar reference numerals denote the
components similar to those in FIG. 4. In the vertical cross
section, the thrust plate 4 is designed to have the configuration
of a partial circle (arc-shaped) rather than the line, thus as a
whole, the thrust plate 4 has an inclined surface with the
spherical configuration.
[0074] The parallel lines indicated in the thrust plate 4b in FIG.
5 schematically illustrate the dynamic pressure generated in the
thrust bearing. As described above, the thrust bearing conduct the
fluid from the circumference of the thrust plate 4b, thus, the
dynamic pressure adjacent to the circumference is low. Rather, the
dynamic pressure is increased towards the outer side of the shaft,
i.e., the inside portion of the thrust bearing. Most of the thrust
dynamic pressure is generated adjacent to the shaft 1.
[0075] According to the alternative embodiment, the arc-shaped
inclined surface of the thrust plate 4 causes the gradient thereof
less adjacent to the bearing axis and more towards the
circumference. This avoids the reduction of the thrust dynamic
pressure adjacent the inside portion because of the close gap to
the thrust opposing surface 13, and also keeps the outside portion
of the thrust plate 4, which does not much contributes the dynamic
pressure, away from the thrust opposing surface 13. Thus, the
thrust plate 4 has the inclined surface with the arc-shaped
configuration so that the thrust dynamic pressure is ensured and
the contact therebetween is avoided.
[0076] The experiment that the present inventors have conducted
revealed that, preferably, the gradient d of the thrust plate 4b is
approximately 2 microns or less. The radius of curvature may be
calculated from the gradient of 2 microns and the distance s
between the innermost portion of the thrust bearing (the outer
surface of the shaft 1) and outermost portion of the thrust bearing
(the circumference of the thrust plate 4).
[0077] As described above, the cross section and the whole
configuration of the inclined surface of the thrust plate 4 are
referred to as being "arc-shaped" and "spherical", respectively.
However, as can be seen from FIG. 5, a partial circle drawn between
the innermost and outermost portions of the thrust bearing has the
center offset from the bearing axis, thus, the whole configuration
of the inclined surfaces on both sides is not completely spherical.
While the true spherical configuration must have the center on the
bearing axis, the shape of the both inclined surfaces of the thrust
plate 4b is referred to be "spherical", for the convenience. In
case where the innermost portion of the thrust bearing is close to
the bearing axis, the inclined surfaces may be traced as the true
partial circle with the center on the axis of the shaft 1. The
circle has the radius calculated based upon the gradient d and the
outer diameter of the thrust plate 4b. It should be noted that the
thrust plate may have any configurations other than a true circle,
for example, ellipse and parabola, if the surface of the thrust
plate has the gradient d less adjacent to the bearing axis (the
innermost portion of the thrust bearing) and more towards the
circumference. The other configuration of the inclined surface
leads the advantages similar to those of the second embodiment.
[0078] Also, the inclined surface is formed on the thrust plate 4
according to the embodiment including the above-mentioned
variations, however, it may be formed on the thrust opposing
surface 13 of the sleeve 3, which opposes to the thrust plate 4b,
so as to achieve the same advantages. Further, the inclined surface
may be formed on a thrust opposing plate as another component used
for the different type of the bearing assembly. Even further, both
of the thrust plate 4b and thrust opposing surface 13, which are
opposing to each other, may be provided with the inclined surface.
In this instance, the total gradient of each gradient of each
inclined surface is preferably corresponding to the above-mentioned
gradient d.
EXAMPLE 2
[0079] In the hydrodynamic bearing assembly having the rotational
sleeve as illustrated in FIG. 4, the radial bearing has the
diameter of 15 millimeters and the length of 15 millimeters, and
the thrust plate has the outer diameter of 20 millimeters and inner
diameter of 15 millimeters, also has the spiral grooves formed
thereon with the depth of 5 microns for generating the thrust
dynamic pressure. The spiral grooves may be formed by the
shot-blasting, the laser abrasion, or the plasma-etching
process.
[0080] Several samples of the motors incorporating the hydrodynamic
bearing assemblies were prepared such that the inclined surface of
the thrust plate had the gradients of 0.1, 0.3, 0.5, 1.0, 1.5, and
2.0 microns, respectively. The gradient is defined herein by the
height difference between the heights along the axis at the
innermost and outermost portions of the thrust bearing. The sample
motors were driven with the electric power of 12V to rotate at the
rotation rate of 12,000 rpm. The electric current for activating
the motor was evaluated as the peak current required for the stable
rotation.
[0081] The pitching for oscillating the hydrodynamic bearing
assembly was tested with use of the oscillation evaluation
equipment including a stepping motor, while the motor rotated at
various rotation rates within the angle of 60 degrees. The contact
was observed by the noise from the rotor, which was monitored by
the microphone. The rotation rate was recorded when the noise was
observed. The higher rotation rate recorded when the noise is
observed can be evaluated that the motor endures more
satisfactorily against the pitching.
[0082] The test results revealed that the gradient d of 0.3 microns
or more improves the features in the activation current and the
pitching.
Third Embodiment
[0083] Next, the third embodiment of the hydrodynamic bearing
assembly according to the present invention will be described
hereinafter with reference to the drawings. FIG. 6 illustrates the
hydrodynamic bearing assembly of the embodiment. In the drawing,
the cylindrical hollow sleeve 3 is arranged around the cylindrical
shaft 1, and the thrust plate 4 connected perpendicularly with the
sleeve 3 opposes to the bottom surface of the sleeve 3. Compared
with the conventional hydrodynamic bearing assembly, the
hydrodynamic bearing assembly of the embodiment further includes
the second thrust plate 11. The second thrust plate 11 in a
disk-shaped form is secured, for example by the adhesive, on the
top end along the axis (the other end away from the thrust opposing
surface 13) of the sleeve 3. The grooves 5 for generating the
thrust dynamic pressure are formed on the surface of the thrust
plate 4, which opposes to the thrust opposing surface 13 as the
bottom surface of the sleeve 3.
[0084] During the rotation of the hydrodynamic bearing assembly so
constructed, similar to those of above-mentioned embodiments, the
sleeve 3 rotates relative to the shaft 1 and the thrust plate 4 so
that the radial and thrust dynamic pressures are generated between
the sleeve 3 and the shaft 1, and between the thrust opposing
surface 13 as the bottom surface of the sleeve 3 and the thrust
plate 4, respectively. Those dynamic pressures allow the sleeve 3
to rotate relative to the shaft 1 and the thrust plate 4 without
any contact.
[0085] During the rotation illustrated in the drawing, when gaps of
a and b are defined between the second thrust plate 11 and the top
surface of the shaft 1 opposing thereto and between the thrust
opposing surface 13 of the sleeve 3 and the thrust plate 4 (thrust
bearing), respectively, the hydrodynamic bearing assembly is
designed such that the gaps of a and b satisfy the following
condition;
a<b
[0086] In the hydrodynamic bearing assembly so constructed, once no
electric power is supplied with the spindle motor, the rotation
rate of the rotational member is decreased, thereby reducing the
dynamic pressures. This causes the floating rotational member to
lower due to its weight. This causes, in turn, the second thrust
plate 11 to contact with the top surface of the shaft 1, because of
the condition of a<b. Eventually, the rotational member is
halted, but the sleeve 3 is still kept away from the thrust plate
4. This means that the total weight of the rotational member such
as sleeve 3 is loaded on the second thrust plate 11 and the top
surface of the shaft 1 while the hydrodynamic bearing assembly is
halted.
[0087] The spindle motor is then energized so that the rotational
member of the hydrodynamic bearing assembly activates to rotate
again. At the beginning, the lower rotation rate generates the
thrust dynamic pressure less sufficiently for floating the
rotational member, keeping the second thrust plate 11 and the top
surface of the shaft 1 in contact with each other. Then, the higher
rotation rate exceeding the predetermined rate allows the
rotational member to float away from the stationary member.
According to the conventional hydrodynamic bearing assembly, the
full contact between the thrust plate 4 and the thrust opposing
surface 13 requires the torque moment against the friction
generated far from the center (i.e., the arm length is long). On
the contrary, the full contact between the second thrust plat 11
and the top surface of the shaft 1 requires the torque much less
than that of the prior art because the friction therebetween is
applied adjacent to the center (i.e., the arm length is much
shorter). To this end, the torque moment for activation can
significantly be reduced.
[0088] In addition, when the rotation rate for floating the
rotational member is unchanged, the relative velocity between the
contacting surfaces according to the embodiment can also be reduced
proportionally to the shorter arm length. This also reduces the
friction between the contacting surfaces, thereby to avoid the risk
of the seizure thereon. The friction between the contacting
surfaces according to the embodiment is minimized in comparison
with that of the prior art, which in turn, facilitates to
accelerate the rotation so as to achieve the floating rate at the
earlier stage.
[0089] Meantime, when the above-mentioned condition i.e., a<b is
met, if the width b is significantly greater than the width a, then
the width difference (b-a) between the thrust opposing surface and
the thrust plate 4, while the rotation is halted, is so great that
the thrust dynamic pressure can insufficiently be generated.
Otherwise, even if the thrust dynamic pressure is generated enough
to float the rotational member, unfavorably, the second thrust
plate 11 is very likely to contact with the top surface of the
shaft 1 due to the oscillating motion applied to the bearing
assembly. Once the contact is happened therebetween, the rotation
becomes unstable and the HDD incorporating the bearing assembly may
malfunction.
[0090] In order to prevent the disadvantage, the gap difference of
a and b should be minimized. It is understood, in general, that if
the gap difference between the two members rotating relative to
each other is 2 microns or less, then the floating force in the
thrust bearing is increased. Thus, because of the above-mentioned
oscillation motion, the second thrust plate 11 moves closer to the
top surface of the shaft 1 and before contact thereto, the gap
between the thrust opposing surface and the thrust plate 3 is less
than 2 microns so that the floating force can be increased. To this
end, the contact is likely to be avoided. Therefore, the following
condition is preferably met;
0<b-a.ltoreq.2 microns.
[0091] To ensure the disadvantage to be prevented, the gap
difference between a and b should be minimized, and the gap of b
between the bottom surface of the sleeve 3 and the thrust plate 4
is reduced so that the floating force in the thrust bearing is
increased. Thus, more preferably, the gap difference satisfies the
following condition;
0<b-a.ltoreq.0.5 microns.
[0092] Further, in order to avoid the contact between the second
thrust plate 11 and the top surface of the shaft 1 due to the
external oscillation, another means for generating the thrust
dynamic pressure may be provided therebetween. Thus, another
grooves similar to grooves 5 for generating the thrust dynamic
pressure may be formed on either surface of the second thrust plate
11 or the top surface of the shaft 1. Such another grooves generate
the thrust dynamic pressure (the repulsion force against the
contact thereof) to avoid the contact therebetween. Since this
floating force is stronger as the gap therebetween is narrower, the
force gives a favorable influence in avoiding the contact.
[0093] FIG. 7 illustrates the alternative hydrodynamic bearing
assembly of the embodiment. In the drawing, while the hydrodynamic
bearing assembly is halted, the top surface 12 of the shaft 1a is
formed as being curved. Other structure is similar to that of the
third embodiment. The top surface 12 of the shaft 1a is curved such
that the second thrust plate 11 and the top surface 12 of the shaft
1a, when halted, contact with each other at a contacting point T
substantially in the center position of the bearing assembly. The
aforementioned arm length of the friction force generated when
activating the rotation can be set as substantially zero, thereby
to further reduce the driving torque.
[0094] The curved top surface 12 of the shaft 1a causes the
rotational member to tilt and contact with the thrust plate 4 at
another contacting point G between the thrust opposing surface of
the sleeve 3 and the thrust plate 4 so that the rotational member
is supported in a stable manner as shown in the drawing. This may
also be happened in the first embodiment, for example if the shaft
1a is too thin or if the outer surface of the shaft 1a has the
offset perpendicularity to the end surface of the shaft 1a.
However, in general, since the shaft 1a has the diameter of
approximately 10 to 15 millimeters and the length of approximately
10 to 20 millimeters, and the gap in the diameter direction is
approximately 1 to 5 microns, the degree of the tilt of the sleeve
3 is very small.
[0095] In case where the thrust opposing surface contacts with the
thrust plate 4 even at a point, the friction force applied at the
point is inevitable. However, as can be imagined from the drawing,
since the center of gravity is located adjacent to the rotation
axis, almost all weight of the rotational member is applied on the
contacting point T between the top surface of the shaft 1a and the
second thrust plate 11, and slight weight thereof is applied to the
thrust plate 4 on the contacting point G. To this end, the friction
force on the contacting point G in activating is very limited
because of the slight weight thereon. Particularly, in comparison
with the conventional thrust bearing in which the thrust opposing
surface 13 and the thrust plate 4 are in full contact, the
activating torque according to the present invention is
significantly reduced even where the contacting point G contributes
the friction force to some extent.
[0096] In the drawing, the curved top surface 12 of the shaft 1a is
formed as being spherical, any geometrical configuration such as
the cone and frustum configurations can be adapted, if the top
surface 12 causes the contacting point to keep closer to the
beating axis. Also, any types of configurations may be shaped with
the curved surface on the second thrust plate 4 (rather than the
top surface of the shaft 1a) opposing to the top surface of the
shaft 1a.
Fourth Embodiment
[0097] Next, the fourth embodiment of the hydrodynamic bearing
assembly according to the present invention will be described
hereinafter. As discussed above with reference to FIG. 6 of the
third embodiment, the perpendicularity between the shaft 1 and the
thrust plate 3 secured thereon is critical for narrowing the gap
between the thrust opposing surface of the sleeve 3 and thrust
plate 4, thereby increasing the floating force. The sleeve 3 is
arranged around the shaft 1, and rotates relative to the shaft 1 so
as to generate the dynamic pressure. Thus, the precise
perpendicularity ensures the gap even between the thrust opposing
surface 13 as the bottom surface of the sleeve 3 and the thrust
plate 3 for the sufficient thrust dynamic pressure. Otherwise, the
imprecise perpendicularity causes the insufficient dynamic
pressure.
[0098] The experiment the present inventor have conducted revealed
that for example, when the disk-shaped thrust plate of the diameter
of 20 millimeters is used, if the circumference of the thrust plate
4 has the perpendicularity deviation of 0.7 microns or less
relative to the shaft 1, then dynamic pressure in the thrust
bearing is generated sufficient enough to achieve the satisfactory
tilt rigidity. In general, the trust plate 4 has the disk-shape
having the diameter of approximately 20 millimeters. By keeping the
perpendicularity deviation of the gap to be 0.7 microns or less,
the second thrust plate 11 can be kept away from the top surface of
the shaft 1. To generalize the relations between the diameter and
the perpendicularity deviation, the perpendicularity between the
shaft 1 and the thrust plate 4 secured thereon is referred that it
has approximately 0.7 microns/20 millimeters. This ratio can be
applied to another thrust plate having diameters rather than 20
millimeters, also to another thrust plate, which is not directly
secured to the end of the shaft, but indirectly secured through
another component.
Fifth Embodiment
[0099] Next, the fifth embodiment of the hydrodynamic bearing
assembly according to the present invention will be described
hereinafter. The components of the hydrodynamic bearing assembly
are made of ceramics material having the high anti-abrasion, the
high endurance, the compact dimensions, and the high rigidity. The
applicable ceramics material includes, for example, alumina,
zirconia, silicon carbide, silicon nitride, and sialon.
[0100] For example, an alumina-based ceramics has the Young's
modulus within the range of approximately 300 to 400 giga-Pascals,
which is about double to the steel, and the specific gravity of 3.9
which is about half of the steel. Thus, briefly speaking, the
alumina-based ceramics provides the rigidity double with half mass
in comparison with steel. If the bearing assembly is made of
ceramics material, it can reduce its volume and weight, and can
improve its anti-abrasion and endurance in comparison with
stainless steel. In the hydrodynamic bearing assembly, since the
rotational member starts rotating in contact with the stationary
member when halted, the friction and the seizure may be caused
between the members. By making those members of ceramics material
showing the good anti-abrasion, such problems can readily be
avoided. The hydrodynamic bearing assembly used for the spindle
motor incorporated in the HDD is required to be formed and
assembled with high accuracy. Ceramics materials are less
susceptible to the plastic deformation and the elastic deformation
than metal materials. Thus, usage of the ceramics materials reduces
the deformations in processing so as to provide a precise
hydrodynamic bearing assembly.
[0101] The hydrodynamic bearing assembly shown in FIG. 6 has
components, which may all be replaced with ones made of ceramics
material. However, only surfaces opposing to the another rotating
parts may be replaced with ones made of ceramics material taking
consideration of the cost-performance. For example, the core member
of the shaft 1 may be made of stainless steel, and the cylindrical
hollow member made of ceramics material may be formed on the outer
surface of the shaft 1 by shrink fitting, cooling fitting or
bonding with the adhesive.
Sixth Embodiment
[0102] Next, the sixth embodiment of the hydrodynamic bearing
assembly according to the present invention will be described
hereinafter. The sixth embodiment according to the present
invention relates to a spindle motor, and a memory device and bar
code scan reader incorporating the spindle motor. As above, usage
of the hydrodynamic bearing assemblies according to the present
invention provides the spindle motor minimizing the friction
between the contacting members when activated without requiring the
less torque, thereby realizing the efficient operation thereof.
Also, the abrasion on contacting surfaces can be minimized, the
spindle motor having the high endurance can be realized. And by
incorporating the spindle motor, the memory device and the bar code
scan reader which are efficient and reliable, are obtained.
[0103] As described above for various embodiments of the present
invention, the exemplary drawing is illustrated, in which the
stationary member includes the shaft and the thrust plate, and the
rotational member includes the sleeve. However, this invention is
not limited thereto, and also applied to the hydrodynamic bearing
assembly, in which the stationary member includes the sleeve and
the rotational member includes the shaft and the thrust plate.
[0104] (Advantages of the Invention)
[0105] To summarize the advantages of the present inventions, the
thrust spiral grooves formed on the thrust bearing have the depth
gradually shallower from the circumference towards the bearing
center, and radial spiral grooves formed on the radial bearing have
the depth gradually shallower towards the downstream. This allows
the dynamic pressure generated adjacent to the center in the thrust
bearing and the downstream in the radial bearing to increase
locally without the keen peak, thereby maintaining the high dynamic
pressure substantially in the wide area. In turn, the maximum
pressure required to generate the floating force is reduced so that
the dew generated due to the compression in the hydrodynamic
bearings can be prevented and also the counter force against the
external oscillation force to the thrust bearing can generally be
improved.
[0106] According to the prior art hydrodynamic bearing assembly,
since the thrust plate opposes in parallel to the thrust opposing
surface (or thrust opposing plate), when the external factor of the
oscillation causes them to tilt relative to each other, the thrust
plate may contact with the thrust opposing surface. On the
contrary, according to the present invention, the surface of thrust
plate is inclined from the inside portion towards the radial
circumference so that the gap in the thrust bearing is wider
towards the circumference. This allows the thrust plate to tilt
relative to the thrust opposing surface without any contact, thus,
the hydrodynamic bearing assembly can be obtained, which endures
satisfactorily against the oscillation. Further, while the
conventional hydrodynamic bearing assembly has the thrust plate and
the thrust opposing surface in full contact with each other when
halted, therefore, the activating torque for restarting the bearing
assembly is significant. However, the inclined surface of the
thrust plate according to the present invention allows the
contacting surface of the thrust bearing when halted to became
closer to the inner portion adjacent to the rotation center. This
causes the arm length from the center to the weighting point to be
shorter, thereby to reduce the activating torque.
[0107] In addition, the hydrodynamic bearing assembly according to
the present invention includes the second thrust plate supported on
the top surface of the shaft for bearing the total weight while the
bearing assembly is halted. This avoids the full contact between
the thrust plate and the thrust opposing surface and allows the
total weight of the rotational member to be applied on the top
surface of the shaft which has the shorter arm length from the
center so that the activating torque can be reduced for restarting.
Further, since the contacting point is closer to the axis, the
relative motion rate to the floating rotation rate can be reduced.
To this end, the abrasion between the contacting surfaces can be
reduced and the seizure due to the friction can be avoided.
[0108] Also, the components of the hydrodynamic bearing assembly
are made of ceramics material so that the endurance and the
reliability of the hydrodynamic bearing assembly can be
significantly improved.
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