Note: Descriptions are shown in the official language in which they were submitted.
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This is a division of our co-pending Canadian Patent
Application No. 574,570 filed 12 August 1988.
This invention relates to precision hydrodynamic bear-
ings.
One important limitation to increasing track density of
computer disk drives is spindle bearing performance. A disk
drive whose spindle bearing has low runout can accommodate
higher track densities which results in more data storage
capacity per disk.
The kinematics of the spin axis of a spindle bearing
determine the precision of the bearing. As the journal spins
relative to the sleeve, the spin axis may trace out a path or
orbit. The motion of this axis typically has components that
are synchronous with the spin and repetitive in nature. These
motions are termed repetitive runout. Other components of
spin axis motion may be asynchronous and nonrepetitive with
respect to spin. These components are termed nonrepetitive run-
out. As a general rule, spindle bearing precision is increased
as repetitive and nonrepetitive runouts are decreased.
Ball bearing spindle systems make up the majority of
prior art disk drives. The kinematics of the rolling elements
in ball bearings result in relatively large nonrepetitive runout.
This results from the fact that the lubricant film thicknesses
in ball bearings are very thin providing little attentuation of
geometric defects in the bearing. In addition, ball bearings
produce forces on the disk drive structure to which it is
attached which are of relatively high frequency and large
amplitude.
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Hydrodynamic spindle bearing designs are also known.
The Hewlett-Packard Model No. 9154A, 3.5 inch micro-Winchester
disk drive incorporates a hybrid hydrodynamic-ball bearing
spindle. The performance of this bearing is degraded by the
incorporation of the ball bearings. The Phillips video 2000
videocassette recorder utilizes a hydrodynamic bearing which
employs grease as the lubricant limiting operation to low
speeds. Other known hydrodynamic spindle bearings for disk
drives employ a ferromagnetic fluid as the lubricant for the
bearing. This fluid is retained or sealed in the bearing by
magnetic fields set up in pole pieces at each end of the bear-
ing. Unless the magnetic fields and clearances are very
precisely matched at each end of the bearing, one seal will be
stronger than the other and when the bearing heats up, the
lubricant can be spilled. See United States Patent No.
4,526,484.
Summary of the Invention
The present invention provides method for introducing
lubricant into a hydrodynamic bearing having clearance spaces
filled with a lubricant comprising: placing the bearing in a
vacuum chamber above a liquid lubricant; evacuating the chamber
to a pressure below atmospheric pressure; submerging the bearing
into the lubricant; and raising the pressure in the chamber to
atmospheric pressure.
The invention provides a method for introducing lubricant
into the hydrodynamic bearing to avoid incorporating air.
Raising the pressure in the chamber to atmospheric pressure
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forces the lubricant into the clearance spaces in the bearing.
After the bearing is filled, it can be exposed to ultrasonic
energy to expel any residual air. The vacuum chamber can also
be repeatedly cycled between a high and low pressure to expel
residual air.
In general, the hydrodynamic bearing disclosed herein
includes a rotatable shaft/thrust plate combination disposed
within a sleeve forming a first clearance space between the
shaft and the sleeve and a second clearance space between the
thrust plate and the sleeve. The external faces of the thrust
plate are exposed to air and the clearance spaces are filled
with a liquid lubricant. The sleeve includes pressure equaliza-
tion ports connecting the first clearance space and the second
clearance space.
The bearing includes surface tension dynamic seals
between axially extending surfaces of the thrust plate and
sleeve. These axially extending surfaces of the thrust plate
and sleeve diverge toward the ends of the bearing to form the
dynamic seal. The divergence may be a straight taper having an
angle of approximately 2.
The pressure equalization ports include axially extend-
ing passageways in communication with radially extending
passageways to connect the first and second clearance spaces.
The radially extending passageways may be located near the center
of the bearing. The bearing may also include relief patterns in
opposed sleeve/thrust plate faces to generate inwardly directed
radial forces.
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In one embodiment the bearing includes a cylindrical
sleeve including a portion having a smaller inside diameter. A
shaft including a portion having a diameter adapted to form a
first clearance space with respect to the smaller diameter
portion of the sleeve fits within the sleeve. A pair of thrust
plates are disposed on the shaft to form second clearance spaces
with respect to radially extending faces of the smaller diameter
portion of the sleeve, the external faces of the thrust plate
being exposed to the air. The clearance spaces are filled with
a liquid lubricant. The smaller diameter portion of the sleeve
includes plural axially extending passageways in liquid
communication with radially extending passageways interconnecting
the first and second clearance spaces. Surface tension seals are
provided between the thrust plates and sleeve.
In another particularly preferred embodiment the bearing
incorporates both external and internal surface tension seals at
each end of the bearing. In this embodiment, there is an air
space between the two ends of the bearing. This embodiment
results in a reduced evaporation rate from the seals, improved
moment stiffness, and faster thermal transient response.
The shaft and sleeve may include mating tapered portions
at each end of the bearing defining lubricant filled clearance
spaces for supporting radial and axial loads. Each clearance
space is sealed by an internal and an external surface tension
dynamic seal and pressure equalization ports are provided to
connect the internal and external seals. In this embodiment,
the shaft is a continuous unit without a separate thrust plate
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portion. No O-ring seals are required.
The hydrodynamic bearing disclosed herein achieves
lower levels of runout than ball bearings as a result of a
thick film of lubricant which separates the sliding metal
surfaces. This film provides a high degree of viscous damping
which significantly attenuates nonrepetitive runout to levels
which are less than state of the art rolling element bearings.
In addition, the bearing generates forces on the structure
attached to it which are low frequency and low amplitude
relative to ball bearings. This reduction in the forcing
function bandwidth and amplitude minimizes other vibrations in
the disk drive and further improves tracking performance. The
pressure equalization ports reduce pressure differentials which
are caused by pumping actions inside the bearing. Because of
the pressure balancing, the bearing does not tend to pump
lubricant in a preferential manner through one seal or the
other. Thus, only the external pressure differential across
the bearing influences the position of the dynamic seal inter-
faces. The surface tension seals of the present invention do
not leak nor do they generate solid debris.
Brief Description of the Drawings
Figure 1 is a cross-sectional view of the bearing of
the invention;
Figure 2 is an elevational view of the sleeve portion of
the bearing;
Figure 3 is an expanded view of a portion of Figure l;
Figure 4 is an expanded view of a portion of Figure l;
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1338270
Figure 5 is an expanded view of a portion of Figure 4;
Figure 6 is a schematic illustration of the method of
filling the bearing with lubricant;
Figure 7 is a cross-sectional view of a particularly
preferred embodiment of the present invention; and
Figure 8 is a cross-sectional view of an embodiment of
the invention utilizing a tapered shaft.
Description of the Preferred Embodiment
A hydrodynamic bearing 10 shown in Figure 1 includes a
sleeve 12 including a portion of smaller inside diameter 14. A
journal or shaft 16 fits within the sleeve 12 forming a first
clearance space 18. The journal 16 may include a recess 20.
Thrust plates 22 and 24 rest on the journal 16 and are sealed
by means of O-ring seals 26. The thrust plates 22 and 24 form
second clearance spaces 28 with respect to radially extending
surfaces of the smaller inside diameter portion 14 of the
sleeve 12. The portion 14 of the sleeve 12 also includes
axially extending passageways 30 and radially extending passage-
ways 32. As shown in Figure 2 the passageways 30 and 32 are
arranged around the circumference of the sleeve 12. Four sets
of passageways 30 and 32 are shown in Figure 2 but more or fewer
may be employed. Figure 2 also shows spiral relief patterns 34.
These relief patterns cooperate with patterns on the journal to
generate radially directed inward hydrodynamic pressure.
Relative rotation between the journal 16 and the sleeve
12 is provided for by the clearance spaces 18 and 28. Suitable
dimensions for the clearance spaces 18 and 28 are 0.0002 to
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0.001 inches and 0.0005 to 0.002 inches, respectively. These
clearance spaces are filled with a lubricant such as oil which
reduces wear between the journal and sleeve and provides a
medium through which a hydrodynamic pressure field may be
generated. Relative rotation or radial motion between the
journal 16 and sleeve 12 is required to set up the hydrodynamic
pressure field. The hydrodynamic bearing 10 supports loads by
metal-to-metal contact when there is no relative motion. During
normal operation, the spinning of the journal 16 sets up a
steady pressure field around the clearance spaces which pushes
the journal and sleeve apart and thus prevents metal-to-metal
contact. The hydrodynamically pressurized film provides the
stiffness needed to support the radial load of the disk, motor
and associated hardware. Note that the hydrodynamic film
stiffness is a measure of the resistance of the clearance space
to change size under the influence of a load.
Axial loads along the journal 16 spin axis are supported
by the hydrodynamic pressure field in the clearance spaces 28
between the thrust plate faces and the sleeve portion 14. The
amount of separation between the thrust plate faces and sleeve
is controlled by the hydrodynamic film stiffness and the applied
axial load (usually the weight of the entire rotating assembly).
Pressure building geometries such as the relief pattern 34 shown
in Figure 2 are employed to generate film stiffness of sufficient
magnitude.
The sealing of the lubricant within the hydrodynamic
bearing 10 will now be described in conjunction with Figures 1,
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4 and 5. There are two types of seals in the bearing 10,
namely, static and dynamic seals. Static seals 26 which are
preferably O-rlng seals prevent lubricant leakage between the
thrust plates 22 and 24 and the journal 16. They are called
static seals in that there is no relative rotation or sliding
between the thrust plates 22 and 24 and the journal 16. Dynamic
sealing is required in the clearance space 36 between the thrust
plates and the sleeve. These seals must not leak or generate
solid debris. Sealing is provided by surface tension-capillary
seals in which a lubricant-air interface 38 provides the
surfaces forces.
As shown in Figure 5, two components, the liquid-gas
(lubricant-air) interface 38 and the solid surfaces of the
thrust plates and sleeve make up each seal. Surface tension
forces directed axially away from each end of the bearing
indicated by the arrows 40 balance the forces due to pressure
differentials which may be applied across each interface as
indicated by the arrows 42 and a force due to gravity. The
magnitude of the axial surface tension forces depends on the
wetted perimeter of the liquid-gas interface 38, the surface
tension (a property of the liquid lubricant), the taper angle
and the contact angle. The forces due to pressure differentials
are dependent on the pressure differentials and the lubricant-
air interface area. Since the solid boundaries of the seal are
tapered, the wetted perimeter and area of the interface vary
with the axial position of the interface. As a result, the
axial position of the interface varies with pressure differences
75917-llD
1~3~270
applied to the bearing until the surface tension forces and
pressure forces balance. Stability of the interface is
sensitive to the angle of taper. A taper angle of approximately
2 has been experimentally determined to be optimum for insuring
interface stability.
During bearing 10 operation, it is necessary that the
pressures be nearly the same at the lubricant side of each
lubricant-air interface 38. This pressure balance is provided
by the pressure equalization ports 30 and 32 which connect the
clearance spaces 18 and 28. Without the equalization ports,
pumping actions inside the bearing may set up pressure
differentials. For example, the thrust plates 22 and 24 produce
an inwardly directed radial pumping action. The equalization
ports tend to equalize the pressures. Furthermore, the passages
should maintain a constant radial position in the neighborhood
of the thrust plates. This requirement prevents large pressure
gradients from developing in the passages due to the centrifugal
pumping effects caused by the thrust plates. The bearing 10 is
thus pressure balanced and does not tend to pump the lubricant
in a preferential manner through one seal or the other. Only
the external pressure differential across the bearing, therefore,
influences the position of the interfaces. The equalization
ports coupled with the surface tension dynamic seals result in
a hydrodynamic bearing of higher precision with respect to runout
relative to conventional bearing designs.
Lubricant must be introduced into the bearing in such way
that a minimal amount of air is trapped in the bearing. This is
1~ 3 8 2 7 0 75917-llD
necessary because trapped air in the bearing expands as the
bearing heats up and tends to push the lubricant out of the
bearing. A method for filling the bearing with lubricant so as
to minimize the amount of trapped air will be described in
conjunction with Figure 6. First of all, the bearing 10 is
placed within a vacuum chamber 50 above the level of a liquid
lubricant 52. The vacuum chamber 50 is then evacuated to a
suitable pressure below atmospheric such as 5 ~ of mercury. The
bearing 10 is then submerged within the lubricant 52, after which
the pressure in the chamber 50 is allowed to rise to atmospheric
pressure. As the pressure rises, lubricant is forced into the
bearing through the clearance spaces between the thrust plates
and sleeve. Residual air bubbles in the bearing may be removed
by applying ultrasonic energy to the chamber 50 within an ultra-
sonic tank 54. If necessary, additional residual air may be
removed by repeatedly cycling of the pressure in the chamber 50
between a high and a low pressure.
Figure 7 is a particularly preferred embodiment of the
invention having several advantages as compared to the embodiment
of Figure 1. A bearing 70 includes a shaft 72 with thrust plates
74 and 76. The shaft 72 with attached thrust plates 74 and 76
rotates within a sleeve 78. The sleeve 78 includes a portion
having increased inside diameter to create an air space 80. The
bearing 70 includes external surface tension seals 82 and
internal surface tension seals 84. The external surface tension
seals 82 and internal surface tension seals 84 are connected by
pressure equalization ports 86. The surface tension seals 82
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and 84 and the pressure equalization ports 86 are filled with a
lubricant. As with the embodiment of Figure 1, the surface
tension seals are created by diverging, axially extending
surfaces.
The embodiment of Figure 7 results in reduced evapora-
tion rate of the lubricant from the seals. When the orientation
of a bearing changes, the position of the surface tension seals
along the spin axis also changes. In the case in which the oil-
air interface moves into the bearing, a film of oil is left on
the region of the metal which was previously covered by the
lubricant of the seal. This film of oil is then exposed to air
and has a large amount of surface area compared to the seal oil-
air interface area. As a result of this increased surface area,
the evaporation of the oil is increased and the life of the
lubricant supply is reduced.
When a bearing is not operating, the position of the
seals is determined by the pressure difference between the two
sealed regions of the bearing which are connected together by
the pressure equalization port or balance tube. The internal
fluid pressure difference is controlled by the elevation
difference between the two regions of the bearing and the
specific weight of the lubricant fluid. The external pressure
differences due to variations in air pressure around the bearing
are usually negligible. Thus seal position and the change in
seal position are controlled primarily by the elevation changes
in the bearing. Splitting the lubricated regions of the bearing
of Figure 7 into two separate and shorter zones reduces the
133827 ~ 75917-llD
range of possible elevation differences and also the resulting
range of seal position changes. This design thus reduces the
wetted area of the bearing and the evaporation rate.
The bearing of Figure 7 also provides higher moment
stiffness. The higher stiffness results from the fact that the
length of the bearing can be made longer relative to the bearing
of Figure 1. Moment stiffness is proportional to the length of
the bearing squared when all of the other bearing characteristics
are held constant. The bearing of Figure 7 can be longer than
the bearing of Figure 1 because the seal areas are split into
separate zones so that the central region of the bearing can be
lengthened without affecting the behavior of the seals.
Another advantage of the embodiment of Figure 7 is
faster thermal transient response of the lubricant. It is
desirable to have the lubricant come up to temperature as fast
as possible during start up. I~hen the lubricant oil is warm, it
has a lower viscosity than when it is cold and thus the torque
requirements are less when the oil is warm. Accordingly, when
the oil can be made to heat up quickly, a shorter period of high
load on the driving motor results which is very desirable for
some applications. The faster thermal response of the bearing
of Figure 7 results from the reduction of oil volume in this
bearing design and the resulting increase in bearing power to
oil volume ratio.
Figure 8 is yet another embodiment of the present
invention. A bearing 100 includes a spindle shaft 102 which has
tapered portions 104 and 106. These tapered portions mate with
12
133827 0 75917-llD
tapered bearing shells 108 and 110 which reside within a spindle
housing or sleeve 112. The spaces between the tapered shaft and
tapered bearing shells are filled with a liquid lubricant. The
lubricant is sealed by external surface tension or capillary
seals 114 and 116, and internal capillary seals 118 and 120. An
equalization port 122 connects the seals 114 and 118, and an
equalization port 124 connects the seals 116 and 120.
Because of the tapered surfaces, both radial and axial
loads are supported by the bearing. The spindle housing and
shaft surfaces are a single contiguous unit without any parting
line. No O-ring seals are required since no secondary leakage
is possible with the tapered arrangement. The tapered portions
of the bearing shaft or the tapered bearing shell surfaces
include herringbone patterns which generate a net liquid flow due
to machining tolerances. This net liquid flow in the bearing is
compensated for by a flow in the opposite direction through the
equalization ports 122 and 124.
The bearing shells 108 and 110 have grooves on their
outer surfaces. These bearing shells are shrink fitted into the
spindle housing 112 and the grooves cooperate with the housing
112 to create the equalization ports.
