U.S. patent application number 12/367466 was filed with the patent office on 2009-06-04 for fluid dynamic bearing system.
This patent application is currently assigned to Seagate Technology LLC. Invention is credited to Troy Michael Herndon, Jeffry Arnold LeBlanc, Robert Alan Nottingham, Mohamed Mizanur Rahman.
Application Number | 20090142009 12/367466 |
Document ID | / |
Family ID | 33162408 |
Filed Date | 2009-06-04 |
United States Patent
Application |
20090142009 |
Kind Code |
A1 |
Herndon; Troy Michael ; et
al. |
June 4, 2009 |
FLUID DYNAMIC BEARING SYSTEM
Abstract
Fluid dynamic bearing systems produced by specified methods are
provided. In one example, a method for designing a fluid dynamic
bearing system includes determining a first stability ratio for a
first journal bearing configuration. The method further includes
determining a second stability ratio for a second journal bearing
configuration. In a further example system, the first configuration
has two sub-journal bearings and the second configuration has three
sub-journal bearings. The method may comprise comparing the two
stability ratios to determine whether adding a sub-journal
increases the stability ratio of the bearing system.
Inventors: |
Herndon; Troy Michael; (San
Jose, CA) ; LeBlanc; Jeffry Arnold; (Aptos, CA)
; Nottingham; Robert Alan; (Santa Cruz, CA) ;
Rahman; Mohamed Mizanur; (San Jose, CA) |
Correspondence
Address: |
SEAGATE TECHNOLOGY LLC;C/O NOVAK DRUCE & QUIGG LLP
1000 LOUISIANA, SUITE 5350
HOUSTON
TX
77002
US
|
Assignee: |
Seagate Technology LLC
Scotts Valley
CA
|
Family ID: |
33162408 |
Appl. No.: |
12/367466 |
Filed: |
February 6, 2009 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
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10792177 |
Mar 2, 2004 |
7493695 |
|
|
12367466 |
|
|
|
|
60464001 |
Apr 18, 2003 |
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Current U.S.
Class: |
384/100 |
Current CPC
Class: |
Y10T 29/497 20150115;
G11B 19/2009 20130101; Y10T 29/49764 20150115; F16C 2370/12
20130101; F16C 17/026 20130101; F16C 33/107 20130101 |
Class at
Publication: |
384/100 |
International
Class: |
F16C 32/06 20060101
F16C032/06 |
Claims
1. A fluid dynamic bearing system, comprising: a gap region between
an inner member and an outer member; an optimal journal bearing
configuration, including at least three sub-journal bearings,
disposed along the gap region, wherein the optimal configuration is
determined by a method, including the steps of: determining a first
stability ratio for a first journal bearing configuration;
determining a second stability ratio for a second journal bearing
configuration; comparing the two stability ratios; and determining
that the second stability ratio is greater than the first stability
ratio.
2. The system of claim 1, wherein the first configuration comprises
two sub-journal bearings and the second configuration comprises
three sub-journal bearings.
3. The system of claim 2, wherein each sub-journal bearing of the
first configuration has a length equal to substantially one-half of
a total journal length and each sub-journal bearing of the second
journal configuration has a length equal to substantially one-third
of the total journal length.
4. The system of claim 1, wherein the method further comprises the
steps of: determining a third stability ratio of a third journal
bearing configuration; and comparing the third stability ratio to
the second stability ratio.
5. The system of claim 4, wherein the first configuration comprises
two sub-journal bearings, the second configuration comprises three
sub-journal bearings, and the third configuration comprises four
sub-journal bearings.
6. The system of claim 5, wherein each sub-journal bearing of the
first configuration has a length equal to substantially one-half of
a total journal length, each sub-journal bearing of the second
journal configuration has a length equal to substantially one-third
of the total journal length, and each sub-journal bearing of the
third journal configuration has a length equal to substantially
one-fourth of the total journal length.
7. The system of claim 1, wherein the first configuration comprises
N number of sub-journals and the second configuration comprises
(N+1) number of sub-journals.
8. The system of claim 7, wherein the method further comprises the
steps of: determining a third stability ratio of a third journal
bearing configuration, the third configuration comprising (N+2)
number of sub-journals; and comparing the third stability ratio to
the second stability ratio.
9. A fluid dynamic bearing system, comprising: a first gap region
between an inner member and an outer member; and at least three
sub-journal bearings disposed along the first gap region, the at
least three sub-journal bearings associated with a first stability
ratio, the first stability ratio greater than a second stability
ratio associated with having only two sub-journal bearings disposed
along the first gap region of the fluid dynamic bearing system.
10. The system of claim 9, further comprising a hub coupled to the
outer member and configured to rotate relative to the inner
member.
11. The system of claim 9, further comprising a hub coupled to the
inner member and configured to rotate relative to the outer
member.
12. A fluid dynamic bearing system, comprising: a gap region
between an inner member and an outer member; a journal bearing
disposed along the gap region, wherein a configuration of the
journal bearing is determined by steps of: determining a first
stability ratio for a first journal bearing configuration having at
least two sub-journal bearing; determining a second stability ratio
for a second journal bearing configuration having at least three
sub-journal bearings, wherein each of the at least three
sub-journal bearings provides radial stiffness; and implementing
the second journal bearing configuration to improve the second
stability ratio relative to the first stability ratio.
13. The system of claim 12, wherein each sub-journal bearing of the
first configuration has a length equal to substantially one-half of
a total journal length and each sub-journal bearing of the second
journal configuration has a length equal to substantially one-third
of the total journal length.
14. The system of claim 12, further including a third journal
bearing configuration.
15. The system of claim 14, further including the step of
determining a third stability ratio of the third journal
bearing.
16. The system of claim 14, wherein the first journal bearing
configuration comprises two sub-journal bearings, the second
journal bearing configuration comprises three sub-journal bearings,
and the third journal bearing configuration comprises four
sub-journal bearings.
17. The system of claim 16, wherein each sub-journal bearing of the
first journal bearing configuration has a length equal to
substantially one-half of a total journal length, each sub-journal
bearing of the second journal bearing configuration has a length
equal to substantially one-third of the total journal length, and
each sub-journal bearing of the third journal bearing configuration
has a length equal to substantially one-fourth of the total journal
length.
18. The system of claim 12, wherein the first journal bearing
configuration comprises 2+N number of sub-journal bearings and the
second journal bearing configuration comprises 3+N number of sub
journal bearings.
19. The system of claim 18, further including a third journal
bearing configuration comprising 4+N number of sub journal
bearings.
20. The system of claim 1, wherein the second stability ratio is
greater than the first stability ratio.
Description
CROSS-REFERENCE TO RELATED APPLICATIONS
[0001] This divisional application claims priority from U.S. patent
application Ser. No. 10/792,177, filed on Mar. 2, 2004, and which
is herein fully incorporated by reference as if fully set forth and
for all purposes. U.S. patent application Ser. No. 10/792,177
claims priority from U.S. Provisional Pat. App. No. 60/464,001,
filed on Apr. 18, 2003, and which is herein fully incorporated by
reference as if fully set forth and for all purposes.
BACKGROUND
[0002] 1. Field
[0003] The present invention relates to fluid dynamic bearing
motors and, more specifically, to a multi-journal fluid dynamic
bearing motor assembly.
[0004] 2. Description of Related Art
[0005] FIG. 1 provides a perspective view of a disc drive assembly
150. In this arrangement, a plurality of discs 110' are stacked
vertically within the assembly 150, permitting additional data to
be stored, read and written. The drive spindle 151 receives the
central openings 105 of the respective discs 110. Separate
suspension arms 156 and corresponding magnetic head assemblies 158
reside above each of the discs 110. The assembly 150 includes a
cover 130 and an intermediate seal 132 for providing an air-tight
system. The seal 132 and cover 130 are shown exploded away from the
disc stack 110' for clarity.
[0006] In operation, the discs 110 are rotated at high speeds about
the spindle 151. As the discs 110 rotate, an air bearing slider on
the head 158 causes each magnetic head 158 to be suspended relative
to the rotating disc 110. The flying height of the magnetic head
assembly 158 above the disc 110 is a function of the speed of
rotation of the disc 110, the aerodynamic lift properties of the
slider along the magnetic head assembly 158 and, in some
arrangements, a biasing spring tension in the suspension arm
156.
[0007] A servo spindle 152 pivots about pivot axis 140. As the
servo spindle 152 pivots, the magnetic head assembly 158 mounted at
the tip of its suspension arm 156 swings through arc 142. This
pivoting motion allows the magnetic head 158 to change track
positions on the disc 110. The ability of the magnetic head 158 to
move along the surface of the disc 110 allows it to read data
residing in tracks along the magnetic layer of the disc. Each
read/write head 158 generates or senses electromagnetic fields or
magnetic encodings in the tracks of the magnetic disc as areas of
magnetic flux. The presence or absence of flux reversals in the
electromagnetic fields represents the data stored on the disc.
[0008] Fluid dynamic bearings tend to generate less vibration and
non-repetitive run-out in the rotating parts of motors than ball
bearings and other types of bearings. For this reason, fluid
dynamic bearing motors are oftentimes used in precision-oriented
electronic devices to achieve better performance. For example,
using a fluid dynamic bearing motor in a magnetic disc drive, such
as magnetic disc drive 150 described above in conjunction with FIG.
1, results in more precise alignment between the tracks of the
discs and the read/write heads. More precise alignment, in turn,
allows discs to be designed with greater track densities, thereby
allowing smaller discs and/or increasing the storage capacity of
the discs.
[0009] As persons skilled in the art are aware, an ongoing
challenge in fluid dynamic journal bearings is balancing the
tradeoff between motor performance and power consumption. For
example, increasing the stiffness of the fluid dynamic journal
bearings results in less vibration in a motor's rotating parts and,
therefore, increased motor precision and performance. However, an
increase in the stiffness of the bearings is usually accompanied by
an increase in the power consumption of the motor. Therefore, there
exists a need for a technique to increase the stability of a fluid
dynamic bearing without increasing the amount of power consumed by
the fluid dynamic bearing.
SUMMARY
[0010] One embodiment of a method for designing and manufacturing a
fluid dynamic bearing system includes determining a first stability
ratio for a first journal bearing configuration. The method further
includes determining a second stability ratio for a second journal
bearing and installing the second journal bearing into the fluid
dynamic bearing system. The two stability ratios may then be
compared. Preferably, the first configuration has two sub-journal
bearings and the second configuration has three sub-journal
bearings. A third stability ratio for a third configuration may
then be determined if the second stability ratio is greater than
the first stability ratio. The third configuration may have four
sub-journals.
[0011] The disclosed method is especially useful for designing
fluid dynamic bearing systems. One advantage of the disclosed
method is that a journal arrangement designed according to the
disclosed method has substantially greater stability than a journal
arrangement not designed according to the disclosed method.
Further, neither the radial stiffness nor the power consumption of
the journal arrangement designed according to the disclosed method
decreases appreciably.
BRIEF DESCRIPTION OF THE DRAWINGS
[0012] So that the manner in which the above recited features of
the present invention can be understood in detail, a more
particular description of the invention, briefly summarized above,
may be had by reference to the appended drawings. It is to be
noted, however, that the appended drawings illustrate only typical
embodiments of this invention and are therefore not to be
considered limiting of its scope.
[0013] FIG. 1 illustrates a perspective view of an exemplary disc
drive assembly as might employ the improved spindle motor
arrangement of the present invention,
[0014] FIG. 2 illustrates a cross-sectional view of an improved
spindle motor, according to one embodiment of the invention, in
which three distinct journal bearings are provided,
[0015] FIG. 3 illustrates the effects of the radial stiffness and
the cross-coupled stiffness of a typical fluid dynamic journal
bearing on the motion of a shaft, according to one embodiment of
the invention,
[0016] FIG. 4 illustrates the results of a simulation comparing the
vibration responses of motors employing two, three, and four
journal bearings, according to one embodiment of the invention,
and
[0017] FIG. 5 is a flow chart of method steps for using the
stability ratio in the design of a fluid dynamic bearing, according
to one embodiment of the present invention.
DETAILED DESCRIPTION
[0018] FIG. 2 presents a partial cross-sectional view of an
improved spindle motor arrangement 200 in one embodiment, in which
three journal bearings 275, 280, and 285 are provided. The motor
200 first comprises a hub 210. The hub 210 includes an outer radial
shoulder 205 for receiving a disc (not shown in FIG. 2). Disposed
within the hub 210 is a sleeve 250. During operation, the sleeve
250 and hub 210 rotate together. In this arrangement, the sleeve
250 resides and rotates on a non-rotational thrust washer 265. A
shaft 255, which is coupled to a base 230, is provided along the
inner diameter of the sleeve 250 to provide lateral support to the
sleeve 250 while it is rotated. Joined to a bottom surface of the
sleeve 250 is a shield 260. The shield 260 encloses an outer
portion of the thrust washer 265. Between an inner side of the
shield 260 and the enclosed portion of the thrust washer, a gap 295
is formed. The motor 200 also includes a stator 245, which is
mounted on a base 230. The stator 245 typically defines an electric
coil that, when energized, creates a magnetic field. The energized
coil cooperates with magnets 215, which are mounted from an inner
surface of the hub 210 on a back iron 240, to cause the hub 210 to
rotate relative to the shaft.
[0019] As persons skilled in the art will recognize, the motor 200
includes a hydrodynamic bearing system. More specifically, the
thrust washer 265 is disposed proximally to a bottom surface of the
sleeve 250, and fluid is injected in gaps maintained between the
sleeve 250 and surrounding parts, e.g., the shaft 255 and the
thrust washer 265, through a fill hole 270 disposed through the
shield 260. The fluid defines a thin fluid film that supports
relative movement of the parts. The interface between the bottom of
the sleeve 250 and the top of the thrust washer 265 thus defines a
thrust bearing 290. Liquid lubricant is provided along the thrust
bearing gap 291 to provide a fluid bearing surface. Either a top
face of the thrust washer 265 or a bottom surface of the sleeve 250
may include a grooved pattern (not shown) for receiving and holding
liquid lubricant when the motor 200 is at rest. When the motor 200
is at rest, the sleeve 250 presses directly on the thrust washer
265. Fluid is then extruded around the outer diameter of the shaft
155 and into a shaft-sleeve gap 292 and/or a shield-thrust washer
gap 295.
[0020] When the motor 200 is energized and the sleeve 250 and
adjoining hub 210 are rotated, lubricating fluid is drawn into the
thrust bearing region 290 to support relative rotation between the
bottom end of the sleeve 250 and the facing surface of the thrust
washer 265. To limit the axial displacement of the sleeve 250 and
adjoining hub 210 during operation, an axial bias force is
typically introduced. In this embodiment, a bias ring 235 is
utilized to prevent the thrust bearing gap 290 from becoming too
large and reducing the stiffness of the thrust bearing. In such a
configuration, an axially downward magnetic force results that
pulls magnet 215 (and therefore hub 210) towards base 230. The
magnitude of this force is a function of, among other things, the
size of a gap between the magnet 215 and the bias ring 235.
Alternatively, base 230 may comprise a magnetic metal such as a
Series 400 steel or a low carbon steel. In other alternative
embodiments, the biasing force may be created in any other feasible
way such as, for example, by applying a spring force or a
downward-acting pressure force on hub 210. When the rotating sleeve
250 comes to rest, the sleeve end will rest on the thrust washer
265. Although the volume of fluid is very small, it will tend to be
forced back out into the gap(s) 295 and/or 292. Therefore, space is
preferably allowed in this gap(s) 295 and/or 292 for this fluid. To
inhibit the loss of liquid lubricant from the gaps 290 and 292,
optional capillary seals are provided along gaps 295 and 294,
respectively. When the motor 200 is idle, the capillary seals 295,
294 aid in maintaining fluid within the bearing system.
[0021] In addition to thrust bearing 290, FIG. 2 shows that there
are three sets of grooves disposed on an outer surface of the shaft
255 along a shaft-sleeve interface 292. The grooved portions of the
shaft along with the corresponding outer sections of the sleeve 250
comprise top 275, intermediate 280, and bottom 285 journal
bearings, respectively. Upon rotation of the sleeve 250, the
grooved patterns of each of the journal bearings 275, 280, 285
create a high pressure region at the center of each pattern to
retain fluid and provide stiffness to the motor 200. Although, only
one intermediate journal 285 is depicted in FIG. 2, any number of
intermediate journals 285 may be provided, depending upon design
considerations of a particular motor.
[0022] The substantially chevron-shaped groove patterns of the
journals 275, 280, and 285 shown in FIG. 2 are preferred but other
patterns, such as spiral patterns, would suffice. Preferably, the
groove patterns are disposed on the shaft 255, however, they may
also be disposed on an inner surface of the sleeve 250. It can seen
that the groove patterns are each identical to one another and
spaced equidistantly along the shaft 255. This is only a preferred
embodiment, however. In alternative embodiments, the journal
bearings 275, 280, and 285 may each comprise a different number and
configuration of grooves and may be unequally spaced along the gap
292. Further, the groove patterns may overlap one another as long
as they are substantially separate, i.e., as long as the apex(es)
of each groove pattern is/are longitudinally spaced apart.
[0023] It is understood that the motor seen in FIG. 2 is exemplary
only. The present invention may be employed in various motor
configurations. For example, the motor may be configured so that
the shaft is coupled to the hub and rotates around a stationary
sleeve.
[0024] FIG. 3 illustrates the effects of the radial stiffness 320
and the cross-couple stiffness 325 of a typical fluid dynamic
journal bearing 330 on the motion of a shaft 305, according to one
embodiment of the invention. As shown, the shaft 305 is subject to
an excitation force that causes the shaft 305 to move in the
horizontal direction towards a sleeve 310. As a result of this
horizontal motion 315, the fluid dynamic journal bearing 330 exerts
a force on the shaft 305 in a direction parallel and opposite to
the horizontal motion 315. The radial stiffness 320 (k..sub.xx) of
the fluid dynamic journal bearing 330 causes this parallel and
opposite reaction force. This type of stiffness is a desirable
quality in fluid dynamic journal bearing 330 because the radial
stiffness 320 tends to reduce the amplitude of the horizontal
motion 315. In addition to the radial stiffness 320, the fluid
dynamic journal bearing 330 is configured to have a cross-coupled
stiffness 325 (k.sub.xy), which acts orthogonally to the radial
stiffness 320. As persons skilled in the art will understand, the
cross-coupled stiffness 325 causes the fluid dynamic journal
bearing 330 to exert a force on the shaft 305 in a direction
orthogonal to the horizontal motion 315. This orthogonal force, in
turn, introduces an orthogonal component 316 to the motion of the
shaft 305, thereby decreasing the stability and performance of
fluid dynamic journal bearing 330. For this reason, cross-coupled
stiffness 325 is not a desirable quality in fluid dynamic journal
bearing 330.
[0025] The stability of a bearing may be gauged by the radial
stiffness divided by the cross coupled stiffness (stability ratio).
A high ratio indicates a relatively greater radial stiffness,
meaning that the radial stiffness governs the behavior of the shaft
more than the cross-coupled stiffness. Thus, with a greater ratio,
the propensity of a fluid dynamic journal bearing, through the
radial stiffness, to limit the shaft motion outweighs the
propensity of the journal, through the cross-coupled stiffness, to
add unwanted motion to the shaft.
[0026] As persons skilled in the art are aware, shorter journal
bearings tend to have greater stability ratios than longer ones.
However, simply reducing the length of the journal bearings in a
conventional two-journal motor to improve stability would have
negative consequences, such as a loss of radial stiffness. In order
to obtain the stability benefits of a shorter bearing while not
sacrificing the radial stiffness of a longer bearing, the overall
journal length of the two-journal motor may be broken into a
greater number of sub-journals. Doing this, effectively takes a
larger journal with a lower stability ratio and divides it into
smaller journals with greater stability ratios. Because these
smaller sub-journals act in parallel, the overall effective journal
length remains approximately constant, so the radial stiffness
remains relatively constant. Further, because the radial stiffness
doesn't change, power consumption remains constant (or, more
importantly, does not increase). Because smaller-journals with
greater stability ratios are used, the effective stability ratio
for the overall journal increases, thereby decreasing or masking
the influence that the cross-coupled stiffness has on shaft
motion.
[0027] FIG. 4 shows the results of a simulation measuring the
response of a fluid dynamic bearing motor with two 400, three 410,
and four 420 journal configurations at different operating
frequencies, according to one embodiment of the invention. More
specifically, the performance of each motor configuration is shown
as the operational vibration (op-vibe) of the motor as a function
of frequency. The journal gap of each motor configuration is the
same. The total journal bearing length of each motor is the same,
therefore, power consumption is approximately the same. In each of
the three simulations, the total bearing length was divided into
two 400, three 410, and four 420 "sub-journals," respectively. As
FIG. 4 shows, for frequencies between about 200 and about 600
Hertz, the four-journal configuration 420 exhibits stability
superior to either the two 400 or three 410 journal configurations.
Further, for frequencies between 400 and about 800 Hertz, the three
journal configuration 410 exhibits superior stability to the two
journal 400 configuration. The foregoing shows that in certain
circumstances, dividing up the total journal length into smaller
sub-journals may increase performance without increasing the power
consumption of the motor.
[0028] FIG. 5 is a flow chart of method steps 500 for using the
stability ratio in the design of a fluid dynamic bearing, according
to one embodiment of the present invention. Although the method
steps 500 are described in the context of the systems illustrated
in FIGS. 1-4, any system configured to perform the method steps in
any order is within the scope of the invention.
[0029] As shown in FIG. 5, the method of using the stability ratio
starts in step 510 where fluid dynamic journal bearing parameters
are determined according to total system requirements, such as, for
example, power consumption, total radial stiffness, and loading.
These requirements are usually supplied by a customer. Typically,
these requirements fix bearing parameters, such as the shaft
diameter, the journal gap between the shaft and the sleeve, and the
total length of the journal bearings.
[0030] At step 520, the motor is initially designed with a
conventional two journal bearing configuration. At step 530, the
stability ratio of the two journal motor is determined. The
stability ratio may be determined theoretically or empirically. At
step 540, an additional journal bearing is added to the initial
two-journal design. Preferably, this is done by taking the total
journal bearing length of the two journal design and providing
three sub-journals, each one third the length of the total journal
bearing length.
[0031] At step 550, the stability ratio of the motor with the three
journal configuration is determined. As previously discussed
herein, decreasing the individual sub-journal bearing lengths may
improve the stability ratio of a given motor. There are situations,
however, in which this will not be the case. For example, referring
back to FIG. 4, it can be observed that for frequencies less than
about 200 Hertz, the stability ratio of the two journal
configuration is optimal and for frequencies greater than about 600
Hertz and less than about 800 Hertz, the stability ratio of the
three journal configuration is optimal. Thus, at step 560, the
stability ratio of the three journal design is compared to that of
the conventional two journal design to see if the stability ratio
has increased by adding another sub-journal. If so, then the method
returns to step 540, and the process is repeated until adding an
additional sub-journal no longer increases the stability ratio of
the motor. At that point, the stability ratio is optimized and the
design is concluded. If the stability ratio of the three journal
design is not greater than that of the two journal design, then the
two journal configuration may be deemed optimal and the design is
concluded.
[0032] The invention has been described above with reference to
specific embodiments. Persons skilled in the art, however, will
understand that various modifications and changes may be made
thereto without departing from the broader spirit and scope of the
invention as set forth in the amended claims.
* * * * *