Note: Descriptions are shown in the official language in which they were submitted.
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133766~
sackground of the Invention
ThQ pre~ent invention relates to hydrodynamic
bearings. In such bearin~s, a rotating ob~ect such as a
shaft is supported by a stationary bearing pad via a
pressurized fluid such ac oil, air or water.
Hydrodynamic bearings take advantage of the fact that
when the rotating object moves, it does not slide along
the top o f the f luid. Instead the f luid in contact with
the rotating object adheres .tightly to the rotating
object, and motion is accompanied by slip or shear
between the fluid particles through the entire height of
the fluid film. Thus, if the rotating object and the
contacting layer of fluid move at a velocity which is
known, the velocity at intermediate heights of the fluid
thicknes~ decreases at a known rate until the fluid in
contact with the stationary bearing pad adheres to the
bearing pad and is motionless. When, by virtue of the
load resulting from its support of the rotating object,
the bearing pad is deflected at a small angle to the
1337663
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rotating member, the fluid will be drawn into the
wedge-shaped opening, and sufficient pressure will be
generated in the fluid film to support the load. This
fact is utilized in thrust bearings for hydraulic
turbines and propeller shafts of ships as well as in the
conventional hydrodynamic journal bearing.
Both thrust bearings and radial or journal bearings
normally are characterized by shaft supporting pads
spaced about an axis. The axis about which the pads are
spaced generally corresponds to the longitudinal axis of
the shaft to be supported for both thrust and journal
bearings. This axis may be termed the major axis.
In an ideal hydrodynamic bearing, the hydrodynamic
wedge extends across the entire bearing pad face, the
fluid film is just thick enough to support the load, the
major axis of the bearing and the axis of the shaft are
aligned, leakage of fluid from the ends of the bearing
pad surface which are adjacent the leading and trailing
edges is minimized, the fluid film is developed as soon
as the shaft begins to rotate, and, in the case of
thrust bearings, the bearing pads are equally loaded.
While an ideal hydrodynamic bearing has yet to be
achieved, a bearing which substantially achieves each of
these objectives is said to be designed so as to
optimize hydrodynamic wedge formation.
The present invention relates to hydrodynamic
bearings that are also sometimes known as movable pad
bearings and methods of making the same. Generally
these bearings are mounted in such a way that they can
move to permit the formation of a wedge-shaped film of
lubricant between the relatively moving parts. Since
excess fluid causes undesirable friction and power
losses, the fluid thic~ness is preferably just enough to
support the maximum load. This is true when the
formation of the wedge is optlmized. Essentially the
pad displaces with a pivoting or a swing-type motion
about a center located in front of the pad surface, and
1337G63
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~earing friction tends to open the wedge. When the
formation of the wedge is optimized, the wedge extends
across the entire pad face. Moreover, the wedge is
formed at the lowest speed possible, ideally as soon as
the shaft begins to rotate.
In known radial pad type bearings, it has
heretofore been believed necessary to provide an
accurately determined clearance between the bearing and
the rotating object supported so as to allow the
appropriate deflection of the bearing pads to form the
hydrodynamic wedge. The requirement of close tolerances
is particularly troublesome in the manufacture of gas
lubricated bearings. Another problem with gas
lubricated bearings is the breakdown of the fluid film
at high speeds. These problems have limited the use of
gas lubricated hydrodynamic bearings.
U.S. Patent No. 3,107,955 to Trumpler discloses one
example of a bearing having beam mounted bearing pads
that displaces with a pivoting or swing-type motion
about a center located in front of the pad surface.
This bearing like many prior art bearings is based only
on a two dimensional model of pad deflection.
Consequently, optimum wedge formation is not achieved.
In the Hall patent, U.S. No. 2,137,487, there is
shown a hydrodynamic moveable pad bearing that develops
its hydrodynamic wedge by sliding of its pad along
spherical surfaces. In many cases the pad sticks and
the corresponding wedge cannot be developed. In the
Greene Patent, U.S. No. 3,930,691, the rocking is
provided by elastomers that are subject to contAm;nAtion
and deterioration.
U.S. Patent 4,099,799 to Etsion discloses a non-
unitary cantilever mounted resilient pad gas bearing.
The disclosed bearing employs a pad mounted on a
rectangular cantilever beam to produce a lubricating
~edge between the pad face and the rotating shaft. Both
1337663
thrust ~earings and radial or journal bearings are
disclosed.
There is shown in the Ide patent, U.S. No.
4,496,251 a pad which deflects with web-like ligaments
so that a wedge shaped film of lubricant is formed
between the relatively moving parts.
U.S. Patent No. 4,515,486 discloses hydrodynamic
thrust and journal bearings comprising a number of
bearing pads, each having a face member and a support
member that are separated and bonded together by an
elastomeric material.
U.S. Patent No. 4,526,482 discloses hydrodynamic
bearings which are primarily intended for process
lubricated applications, i.e., the bearing is designed
to work in a fluid. The hydrodynamic bearings are
formed with a central section of the load carrying
surface that is more compliant than the remainder of the
bearings such that they will deflect under load and form
a pressure pocket of fluid to carry high loads.
It has also been noted in Ide U.S. Patent No.
4,676,668, that bearing pads may be spaced from the
support member by at least one leg which provides
flexibility in three directions. To provide flexibility
in the plane of motion, the legs are angled inward to
form a conical shape with the apex of the cone or point
of intersection in front of the pad surface. Each leg
has a section modulus that is relatively small in the
direction of desired motion to permit compensation for
misalignments. These teachings are applicable to both
journal and thrust bearings. While the disclosure of
this patent represents a significant advance in the art,
it has some shortcomings. One such shortcoming is the
rigidity of the support structure and bearing pad which
inhibits deformation of the pad surface. Further, the
bearing construction is not unitary.
The last two patents are of particular interest
1337663
because they demonstrate that despite the inherent and
significant differences between thrust and journal
bearings, there is some conceptual similarity between
hydrodynamic journal bearings and hydrodynamic thrust
bearings.
This application relates in part to hydrodynamic
thrust bearings. When the hydrodynamic wedge in such
bearings is optimized, the load on each of the
circumferentially spaced bearings is substantially
equal.
Presently, the most widely used hydrodynamic thrust
bearing is the so-called Kingsbury shoe-type bearing.
The shoe-type Kingsbury bearing is characterized by a
complex structure which includes pivoted shoes, a thrust
collar which rotates with the shaft and applies load to
the shoes, a base ring for supporting the shoes, a
housing or mounting which contains and supports the
internal bearing elements, a lubricating system and a
cooling system. As a result of this complex structure,
Kingsbury shoe-type bearings are typically
extraordinarily expensive.
An alternative to the complex Kingsbury shoe-type
bearing is the unitary pedestal bearings shown in
Figures 19-20. This bearing has been employed in, among
other things, deep well pumps. This relatively simple
structure is typically formed by sand casting or some
other crude manufacturing technique because heretofore,
the specific dimensions have not been deemed important.
As shown in Figures l9 and 20, the bearing is
structurally characterized by a flat base 36PA having a
thick inner circumferential projection 38PA, a plurality
of rigid pedestals 34PA extending transversely from the
base and a thrust pad 32PA centered on each rigid
pedestal.
Figure 20(A) illustrates schematically, the
deflection of the bearing of Figures 19-20 in response
1337663
-- 6
to movement of the opposing thrust runner in the direction of
arrow L. In Figure 20(A) the deflected position (greatly
exaggerated) is illustrated in solid lines and the non-deflected
position is illustrated in phantom. The curve PD in Figure 20(A)
illustrates the pressure distribution across the face of the pad.
Under load, the thrust pads deflect around the rigid pedestals in
an umbrella-like fashion as shown in Figure 20(A). By virtue of
this umbrella-like deflection, only a partial hydrodynamic wedge
is formed. Consequently, there is an uneven distribution of
pressure across the face of the pad as illustrated in
Figure 20(A). Thus, the bearing has proportionately less
hydrodynamic advantage compared to a bearing in which a
hydrodynamic wedge is formed across the entire thrust pad face.
Moreover, the rigidity of the pedestals and flat inflexible base
prevent the deflections necessary to optimize wedge formation.
The foregoing may explain why bearings of the type shown in
Figures 19-20, while far less expensive than Kingsbury bearings,
have proved less efficient and capable and consequently less
successful than the shoe-type bearings.
The present inventor has also discovered that the
center pivot nature of both the bearing shown in Figures 19-20 and
the Kingsbury shoe-type bearing contributes to bearing
inefficiency. It should also be noted that, because of their
rigid center pivots, neither the Kingsbury shoe-type bearings nor
the bearing shown in Figures 19-20 can deflect with six degrees of
freedom to optimize wedge formation. Thus, while, in some
instances, the prior art bearings are capable of movement with six
degrees of freedom, because the bearings are not modeled based
upon or designed for six degrees of freedom, the resulting
performance capabilities of these bearings are limited.
Prior art hydrodynamic bearings often suffer from
fluid leakage which causes breakdown of the fluid film. In radial
bearings, the leakage primarily occurs at the
_7_ 1 33 7G63
axial ends of ~he bearing pad surface. In thrust
bearings, the leakage primarily occurs at the outer
circumferential periphery of the pad surface as a result
of centrifugal forces action on the fluid. When wedge
formation is optimized, fluid leakage is minimized.
Summary of the Invention
The present invention discloses a pad type bearing
and methods of making the same. The pad type bearing,
which is preferably unitary, is preferably formed from a
single piece of heavy walled tubing or a cylindrical
journal that has been machined or formed with small
grooves and slits, bores or cuts through or on the
bearing wall to define a flexible journal or thrust pad
and a support structure capable of supporting the pad
for movement in the six degrees of freedom (i.e.,
translation or movement in the +x, -x, +y, -y, +z and -z
directions) and rotation about the X, Y, and Z axes so
as to optimize formation of the hydrodynamic wedge.
The bearings of the present invention are designed
in three dimensions to provide deflection with six
degrees of freedom so as to ensure optimum wedge
formation at all times. Specifically, it has been
discovered that a hydrodynamic bearing operates most
effectively when the hydrodynamic wedge has several
characteristics. In particular, the wedge should extend
across the entire pad surface; the wedge should have an
appropriate thickness at all times; the wedge should be
shaped so as to minimize fluid leakage; the wedge should
accommodate misalignment such that the major axis of the
bearing is colinear or substantially parallel to the
axis of the shaft; and the wedge should be formed at the
lowest speed possible to prevent damage to the wedge
forming surface which generally occurs as a result of
shaft to pad surface contact at low speeds. Moreover,
with thrust bearings, the loading among the spaced
bearing pads should be equal.
1337663
With regard to thickness of the fluid film, it
should be understood that the optimum thickness varies
with loading. Under high or heavy loading, a relatively
thick fluid film is desirable to adequately support the
load. However, thicker films increase friction and
power loss. Thus, the bearings are preferably designed
to provide the minimum thickness necessary to support
the shaft at maximum load.
The support structure is preferably unitary and
comprises support stubs, beams, and/or membranes
connected to a housing which is sometimes defined by the
radially outermost portion of the bearing in the case of
a journal bearing or, in the case of thrust bearings, a
housing into which the bearing is mounted.
The inventor has discovered that in many specific
applications such as in high speed applications, it is
necessary to P~;ne and evaluate the dynamic
flexibility of the entire system consisting of the shaft
or rotor, the hydrodynamic lubricating film and the
bearing. In computer analysis of this system using a
finite element model, it has been determined that it is
necec-cary to treat the entire bearing as a completely
flexible member that changes shape under operating
loads. By adding more or less flexibility via machining
of the basic structure, bearing characteristics may be
achieved that provide stable low friction operation over
wide operating ranges. A number of variables have been
found to substantially affect the bearing's performance
characteristics. Among the most important variables are
the shape, size, location and material characteristics
(e.g. modulus of elasticity etc.) of the pad and support
members defined by the bores, slits or cuts and grooves
formed in the bearing. The shape of the support members
has been found to be particularly important. Also by
providing a fluid backing to the flexible members, a
high degree of damping may be achieved that further adds
to system stability. In some instances, this damping
has replaced secondary squeeze film dampening that is
133766~
g
present when the oil film is present between the casing
of the bearing and the housing.
The inventor has also discovered that with respect
to gas or air lubricated deflection pad bearings, there
are instances where loads or speeds exceed the
capability of a gas film. In these cases, it is
necessary to introduce a liquid type lubricant into the
converging wedge without providing a liquid reservoir or
bath. The present invention provides a bearing which
solves this problem by providing liquid lubricant when
necessary.
Specific applications of the bearings of the
present invention include electric motors, fans,
turbochargers, internal combustion engines, outboard
motors, and compressors/ expanders. Test speeds have
exceeded 300,000 r.p.m. It is noted that the cuts,
grooves and openings in addition to allowing the bearing
pad to move to form a converging wedge for hydrodynamic
lubrication, allow the pad itself to deflect and change
shape -by for example flattening. This improves
operating performance by, among other things, changing
the eccentricity of the bearing.
The bearings may be formed of metals, powdered
metals, plastics, ceramics or composites. When
manufactured in small quantities, the bearings are
typically machined by facing, turning, and milling the
blanks to form larger grooves or openings; smaller
grooves are formed by water-jet cutting, electrical
discharge or laser machining methods and allow total
design flexibility to tune the bearing to provide
desired characteristics. Tuning will essentially change
the stiffness that in turn eliminates vibration.
Manufacture of larger quantities of a single type
bearing is preferably accomplished through injection
molding, extrusion, powdered metal die casting,
investment casting or some similar manufacturing
technique. In accordance with one aspect of the present
-
1337663
--10--
invention, intermediate quantities of bearings are
manufactured according to a novel method combining
machining and investment casting techniques. The
present invention also contemplates easily moldable
bearings which include no hidden openings such that they
can be molded in a simple two-piece mold. In general,
the bearings of the present invention can be
manufactured at a fraction of the cost of competitive
bearings.
Unlike prior pad type bearings which have a support
structure that is essentially oriented in the direction
of load, the present invention provides an orientation
that allows for comparable deflections within a smaller
envelope (i.e., the difference between the radially
inner journal surface and the radially outer journal
surface in journal bearings) especially in journal
bearings; allows for movement of the bearing pad in any
direction (i.e., six degrees of freedom) to form a
converging wedge shape; allows for the pad itself to
change shape (e.g., flatten) to improve performance;
allows for development of a membrane damping system for
improved stability; and allows the bearings to
compensate for misalignment of the supported part or
shaft and to equalize loading among the bearing pads in
a thrust bearing. All of these characteristics
contribute to formation of an optimum hydrodynamic
wedge.
While there are numerous arrangements of bores,
grooves, cuts, or slits there are primarily two modes of
deflections, namely one or more ligaments or membranes
which deflect in the general direction of load in a
bending mode and secondly by torsional deflection in a
beam or membrane in a direction extending away from the
pad along the longitudinal axis of the shaft in journal
bearings. The degree of deflection in the bending mode
is, in part, a function of the stiffness of the support
structure in the radial direction. The pad itself may
be made to deflect under a load to form a different
1337663
shape by providing internal cuts beneath the pad or by
undercutting the edges of the pad. In either case the
cuts are specifically made to result in a predetermined
shape under load. By surrounding or backing certain
ligaments or membranes with lubricating fluid, a damping
element may be added to t~e design.
Similar cuts are used for journal bearings and
thrust bearings. The primary determinant is the
deflections desired for optimum performance. However,
since journal and thrust bearings perform significantly
differently functions there are inherent differences in
desired performance requiring different desired
deflections. Consequently, despite the general
conceptual similarity between the journal bearings and
thrust bearings of the present invention there are also
significant conceptual differences and plainly evident
structural dissimilarities.
The bearing of the present invention includes a pad
that may change shape and move in any direction (i.e.,
is supported for movement with six degrees of freedom).
The bearing also may have a built-in damping system and
is preferably of unitary or single piece construction
for high volume economical manufacture. The journal
bearings of the present invention also fits in a
relatively small envelope (i.e., spacing between the
housing outer diameter and the pad inner diameter).
In accordance with the present invention, the need
for close tolerances between the bearing pad and the
shaft portion to be supported can be obviated by
dimensioning the bearing so as to eliminate the spacing
between the bearing pad and the shaft portion to be
supported while at the same time dimensioning the
support structure such that the radial (in the case of a
journal bearing) or axial (in the case of a thrust
bearing) stiffness of the bearing is less that the
corresponding fluid-film stiffness of the supporting
fluid. Either the entire pad or only a portion thereof
-12- 133766~
can be pre-biased into contact with the shaft. For
instance, with extremely flexible bearings it may be
desirable to pre-torque the entire bearing pad into
contact with the shaft. On the other hand, in some
instances it is advantageous to pre-torque only the
trailing edge of the bearing pad into contact with the
shaft so as define a hydrodynamic wedge. Thus, the
bearings of the present invention can be designed to
have an interference fit when installed on the shaft.
In one embodiment, as the bearing is forced onto the
shaft, the pad support structure deflects slightly to
form a converging wedge shape while in the installed,
static position with contact between the bearing pad and
the shaft at the trailing edge. In such an instance
where the bearing is designed to provide a statically
loaded wedge, an appropriate spacing between the pad and
the shaft will be established instantaneously upon
rotation of the shaft by virtue of the stiffness of the
fluid-film. This is because the fluid film enters the
wedge and builds up fluid pressure causing separation of
the shaft and pad. Specifically, the relatively stiff
fluid causes the relatively flexible beam support
structure to deflect until the stiffness of the support
structure is equal to the fluid film stiffness. The
instantaneous formation of the fluid film protects the
surface of the bearing pad from damage which occurs at
low start-up speeds when there is direct contact between
the shaft.
Interference fit bearings of the aforementioned
type allow a much larger variation in machining
tolerances. For example, a relatively large (e.g. .003
inch) variation in the interference fit can be designed
to have an insignificant impact on the wedge. This is
particularly critical for gas lubricated bearings where
alternate bearing forms require extraordinarily precise
machining for proper operation. The present invention
relaxes machining requirements.
Similarly the thrust bearings of the present
1337663
- 13 -
invention can be designed to provide a statically loaded wedge.
Specifically, the thrust bearings of the present invention can be
designed such that the bearing pads are biased so that the inner
circumferential edge of the bearing pad extends away from the
shaft and so that the trailing edge extends toward the shaft.
With this arrangement, in the static loaded condition, the bearing
pad slopes toward the shaft in the radial direction (when moving
outwardly from the axis). Further, the bearing pad slopes toward
the shaft from the leading edge to the trailing edge. In this way
a statically loaded wedge approximating the optimum wedge is
formed and appropriate spacing between the pads and shafts is
established instantaneously upon rotation of the shaft.
In the bearings of the present invention, the pad
movement may be directed toward the shaft to hold shaft location
and to give the pad the ability to adjust for misalignment of the
shaft and unequal loading among pads. Of course, the present
invention may apply to any radial, thrust or combined radial and
thrust form of bearings and may be one or two directional in
nature, depending on the configuration of the bearing. More
specifically, if the bearing support structure is symmetrical
about the bearing's pad circumferential centerline, the bearing
will be bidirectional, i.e., capable of supporting a shaft for
rotation in two directions in an identical fashion. However, if
the bearing support structure is non-symmetrical about the
bearing's pad circumferential centerline, the bearing will deflect
differently when supporting a shaft for rotation in a first
direction as compared to rotation in the opposite direction. For
both journal or radial bearings and thrust bearings, the major
axis is the central axis of the cylindrical blank from which the
bearing is formed.
In accordance with another important aspect of the
bearings of the present invention, the bearing pads can be
supported for deflection so as to retain the hydrodynamic fluid,
thus obviating the problem of fluid
~1_
1337663
-14-
leakage. With respect to radial or journal bearings,
the support structure is designed such that, under load,
the bearing pad deflects to form a fluid retaining
pocket. Generally, such a support is achieved when the
primary support portion is connected to the bearing pad
proximate the axial edges of the bearing pad and the
center of the bearing pad is not directly supported,
i.e., is free to deflect radially outward. With respect
to thrust bearings, the pad is supported so as to tilt
toward the bearing's inner diameter under load so as to
prevent centrifugal leakage. Generally, this is
achieved when the pad support surface at which the
primary support structure supports the bearing pad is
located closer to the bearing outer diameter than to the
bearing inner diameter. When the primary support
structure includes two or more radially spaced beams,
the overall support structure must be designed to cause
deflection of the bearing pad at the inner end.
Further, when the bearing pad is supported by a
plurality of radially spaced beams and the region
between the beams is not directly supported, the pad
will tend to deflect so as to form a concave fluid
retaining channel.
In accordance with the present invention, a number
of methods of manufacturing the bearings of the present
invention are also contemplated. The selection of a
particular method of manufacturing depends largely on
the volume of the particular bearing to be manufactured
and the materials used. In low volume applications, or
when it is desired to produce prototypes for testing
and/or production of molds or the like, the bearings are
preferably manufactured from metallic cylindrical blanks
such as heavy wall tubing or other journals which are
machined to provided radial and/or facing bores or
grooves and formed with radial cuts or slits through
either numerically controlled electrical discharge
manufacturing techniques, numerically controlled laser
cutting techniques, or numerically controlled water-jet
cutting. In intermediate volumes, the bearings of the
13~766~
present invention are preferably manufactured using an
investment casting method in accordance with the present
invention. In high volume applications, the bearings of
the present invention can be manufactured using a wide
variety of materials such as plastics, ceramics,
powdered and non-powdered metals, and composites. In
high volume applications, a number of manufacturing
methods including injection molding, casting, powdered
metal, die casting, and extrusion can be economically
employed. The bearings of the present invention can be
formed in a shape which is easily moldable.
.
In short, the present invention relates to radial,
thrust and compound radia. and thrust hydrodynamic
bearings which perform significantly better than known
bearings and can be manufactured at a fraction of the
cost of competitive bearings.
Brief Description of the Drawings
The details of the invention will be described in
connection with the accompanying drawing, in which:
Figure 1 is a sectional view of a journal bearing
illustrating a sector thereof embodying one form of the
invention;
Figure 2 is a schematic view of a single pad made
in accordance with the example illustrated in Figure l;
Figure 3 is an edge view of the pad of Figure 2
illustrating the pad orientation with the support
structure in the loaded state:
Figure 4 is a sectional view of a sector of a
second example of a journal bearing made in accordance
with the present invention;
Figure 5 is a view partly in section of a single
pad of Figure 4;
1~37663
-16-
Figure 5A is a perspective view of a section a
modified form of the bearing of Figure 4;
Figure 5B is a perspective vie.w of a modified form
of the bearing shown in Figure 4;
Figure 6 is an end view of the bearing of Figure 4;
Figure 7 is a diagramatic view of the torsional
deflection of a beam, greatly enlarged;
Figure 8 is a sectional view of a journal bearing
illustrating an example of a bearing incorporating the
features of the present invention which includes two
beams;
Figure 9 is an edge view of the pad of Figure 1
illustrating local deflection of the pad surface without
support structure deflection, greatly exaggerated;
Figure 10 is an edge view of the pad of Figure 8
illustrating the pad orientation with the support
structure in the loaded state.
Figure lOA is an edge view of the pad of Figure 8
illustrating local deflection of the pad surface greatly
exaggerated.
Figures llA and llB are cross sectional views of a
cylindrical journal or blank prior to machining;
Figures 12A and 12B are cross sectional views of a
machined journal or blank;
Figures 13A and 13B are cross-sectional views of a
further machined journal or blank;
Figures 14A and 14B are cross sectional view of a
modified machined journal or blank; and
1337663
-17-
Figures 14C and 14D are cross sectional views of a
bearing constructed from the modified machined journal
or blank Figures 14A and 14B.
Figure 15 is top view of a thrust bearing having
beam mounted bearing pads.
Figure 16 is a side cross section of the thrust
bearing of Figure 15.
Figure 17 is a bottom view of the thrust bearing of
Figure 15.
Figure 18 is a perspective view of a portion of the
thrust bearing of Figure 15.
Figure 19 is a top view of a prior art thrust
bearing.
Figure 20 is a cross-section of the prior art
thrust bearing of Figure 19.
Figure 20(a) is a schematic representation of a
segment of the prior art thrust bearing of Figures 19
and 20 showing the pressure distribution across the
surface of a bearing pad.
Figure 21 is a top view of a thrust bearing
according to the present invention having two legged
support.
Figure 22 is a side cross-section of the thrust
bearing of Figure 21.
Figure 23 is a bottom view of the bearing of Figure
21.
Figure 23(A) is a bottom view of a modified version
of the bearing of Figure 21.
1337663
-18-
Figure Z4 is a perspective view of a segment of the
bearing of Figure 21.
Figure 25 is a cross-section of another bearing
according to the present invention.
Figure 26 is a cross-section of another bearing
according to the present invention.
Figure 27 is a side cross-section of another
bearing construction according to the present invention.
Figure 28 is a top cross-section of the bearing
construction of Figure 27.
Figure 29 is a side cross-section of another
bearing construction according to the present invention.
Figure 29A is a cross-section of another thrust
bearing construction according to the present invention.
Figure 29B is another cross-section of the bearing
of Figure 29A.
Figure 30 is a top cross-section of the bearing
construction of Figure 29.
Figure 3OA is a top view of the bearing of Figure
29A.
Figure 3OB is a bottom view of the bearing of
Figure 29A.
Figure 31 is a side view of another journal bearing
construction in accordance with the present invention.
Figure 3lA is a radial cross-section of a portion
of the bearing illustrated in Figure 31.
Figure 32 is a side view of another journal bearing
133766~
--19--
construction in accordance with the present lnvention.
Figure 32A is a radial cross-section of the bearing
of Figure 32.
Figure 32B is a perspective view of the bearing of
Figure 32.
Figure 33 is a side view of another journal bearing
construction in accordance with the present invention.
Figure 33A is a detail of a portion of the outer
periphery of the bearing of Figure 33.
Figure 33B is a cross-section of the bearing of
Figure 33.
Figure 33C is another cross section of the bearing
of Figure 33.
Figure 34 is a side view of another journal bearing
according to the present invention.
Figure 34A is a detail of a portion of the outer
periphery of the bearing of Figure 34.
Figure 34B is a cross-section of the bearing of
Figure 34.
Figure 34C is another cross-section of the bearing
of Figure 34.
Figure 34D is another cross-section of the bearing
of Figure 34.
Figure 35 is a side view of a combined radial and
thrust bearing according to the present invention.
Figure 35A is a cross-section of the bearing of
Figure 35.
1~37663
-20-
Figure 35B is another cross-section of the bearing
of Figure 35.
Figure 36 is a side view of another combined radial
and thrust bearing according to the present invention.
Figure 37 is a diagramatic cross-section of the
bearing of Figure 36 illustrating the forces acting on
the bearing pad.
Figure 38A is a top view of an easily moldable
thrust bearing according to the present invention.
Figure 38B is a bottom view of the bearing of
Figure 38A.
Figure 38C is an exploded cross-section along the
lines indicated in Figure 38A.
Figure 38D is a bottom view illustrating
modifications of the bearing illustrated in Figures
38A-C.
Figure 39A is a top view of another easily moldable
thrust bearing according to the present invention.
Figure 39b is a bottom view of the bearing of
Figure 39A.
Figure 39C is partial cross-section showing the
support structure for the bearing pads in the bearing of
Figures 39A and 39B.
Figure 40 is a side view of a self-lubricating
bearing according to the present invention.
Figure 40A is a cross-section of the bearing of
Figure 40.
Figure 41 is a side view of a self-lubricating
1337663
- 21 -
combined radial and thrust bearing according to the
present invention.
Figure 41A is a cross-section of the bearing of
Figure 41.
Detailed Description of the Illustrative Embodiments
In describing the bearings of the present
invention in an understandable way, it is helpful to
describe the bearing structures as being formed from a
cylindrical blank by providing grooves, slits, bores and
other openings in the cylindrical blank. As noted below,
this is sometimes a useful technique for manufacturing a
prototype bearing. However, the reference to the
cylindrical blank is primarily intended to assist
understanding of the present invention. It should be
noted that although many of the bearings of the present
invention could be manufactured from a cylindrical blank,
it is not necessary that any of them be so manufactured.
Indeed the bearings can be manufactured in numerous ways,
some of which are discussed hereinafter.
Referring first to Figure 1, the structure
therein illustrated is a sector of a journal bearing
assembly having grooves and slits formed therein so as to
define a housing 10 and a plurality of circumferentially
arranged bearing pads 12 each of which is supported by a
support structure which includes the housing, a beam 14
and a stub section 16. The bearing defined by the grooves
and slits is non-symmetrical about the pad circumferential
centerline (Figs. 3, 13a). Accordingly, the bearing
illustrated is a radial unidirectional bearing, i.e., it
is adapted for radially supporting a shaft for rotation in
only one direction. In the illustrated embodiment, the
bearing supports the shaft 5 only for rotation in the
counterclockwise direction illustrated by the arrow. On
the other hand, if the bearing were symmetrical about the
centerline of the pad, it would be capable of supporting
the shaft 5 for either clockwise or counterclockwise
rotation, i.e., the bearing would be bidirectional.
-~"
1337663
- 22 -
Each bearing pad 12 includes a leading edge 15
and a trailing edge 17. The leading edge is defined as
the edge first approached by a point on the circumference
of the shaft as it continues to rotate. Similarly, the
trailing edge is defined as the edge approached
circumferentially later by the same point on the shaft as
it continues to rotate. When the shaft 5 is rotating in
the proper direction, it moves, on a fluid film, from the
leading edge across the bearing pad and off the trailing
edge. Optimum performance is obtained when the stub-
section 16 supports the bearing pad 12 and hence any load,
at a point 16a (Figure 3) between the circumferential
center line 13a of the pad 12 and the trailing edge 17
preferably, closer to the center line 13a. The beam 14
should also pivot about a point 14a which is located
angularly between the leading edge and the trailing edge
so that as a result of deflection of the beam 14, the
trailing edge 17 deflects inwardly. Of course, the degree
of deflection depends on, among other things, the shape of
the beam and the length of the cuts or slits formed in the
bearing.
Although specific reference is made to either
journal bearings or thrust bearings to facilitate an
understanding of this invention, some of the same
principles of bearing design apply regardless of the
specific form of bearing being designed. For example,
both types of bearings operate on the principle of
formation of a hydrodynamic wedge. Further, the major
axis of both journal bearings and thrust bearings is the
central axis of the cylindrical blank from which the
bearing is formed. The circumferential pad centerline is
the radially extending line passing through the geometric
center of the pad and the major axis of the bearing.
Accordingly, if either a thrust bearing or a journal
bearing is sy~metrical about this centerline axis, i.e.,
the major axis, the bearing will be
-23- 1337 663
bidirectional.
There are significant differences between thrust
bearings and journal or radial bearings. The most
prominent difference is, of course, the portion of the
shaft supported and consequently the orientation and/or
attitude of the bearing pad supports. For instance,
while journal bearings support circumferential portions
of shafts, thrust bearings support shoulder or axial end
portions of shafts. Other differences follow from this
fundamental difference. For example, in a radial or
journal bearing the pads in the direction of the load
take or support the load; whereas, in a thrust bearing,
all pads normally share load. Moreover, a journal
bearing generally has a built-in wedge due to
differences in the shaft and bearing diameters;
conversely, there is no such built-in wedge in thrust
bearings. Additionally, while a journal or radial
bearing controls rotational stability as well as load; a
thrust bearing typically only carries load. It should
also be understood that the design of journal bearings,
particularly hydrodynamic journal bearings, is
significantly more complicated than the design of thrust
bearings. In part, this is because of the constraints
imposed by the need to limit the radial envelope of the
journal bearings. In order to accommodate these
differences the configuration of the thrust bearings is
naturally somewhat different than that of journal
bearings. Nevertheless, as is evident from this
disclosure, many of the principles discussed herein are
applicable to either thrust or journal bearings.
Referring now to Figures 2 and 3, it will be seen
that the pad 12 is provided with an arcuate face 13
which corresponds essentially to the radius or arc of
the outer diameter of the shaft which the pad will be
supporting (via the fluid film) and each pad is defined
by axially extending and radially extending edges. The
axially extending edges comprise the leading and
trailing edges. The beam is shown both in a static
13376C3
-24-
position (solid lines) and in a deflected position
(phantom lines) in Figure 3. The basic construction of
the support structure as illustrated in Figure 1, is
created by the use of small slits or cuts through the
wall. Typically these slits or radial cuts are between
0.002 to 0.12S" wide. The degree of deflection can be
varied by varying, among other things, the length of the
cuts. Longer cuts provide a longer moment arm which
yields greater deflection. Shorter cuts yield beams
having less flexibility and higher load carrying
ability. In selecting a length of cut or slit, care
must be taken to avoid resonance.
By locating the end of beam 14 as shown, the
deflection downward about the connection point 16a will
result in inward movement of the trailing edge 17 of the
pad 12 outward movement of the leading edge 15 and a
slight flattening of the pad 12 as seen in the dotted
lines of Figure 9,. As a result of this deflection, the
gap between the pad face 13 and the outer surface of the
shaft 5, through which fluid flows, becomes wedge shaped
to yield the well-~nown hydrodynamic support effect.
Ideally the ratio of the spacing between the trailing
edge and the shaft versus the spacing between the
leading edge and shaft is between 1:2 to 1:5. In other
words, the spacing between the leading edge and shaft
should be between 2 to S times greater than the spacing
between the trailing edge and the shaft. In order to
attain this ideal spacing or wedge ratio for any
specific application, appropriate deflection variables
including number, size, location, shape and material
characteristics of the unitary element must be selected.
A computer aided finite element analysis has proven to
be the most efficacious means of optimizing these
variables. Computer aided analysis is particularly
useful in a bearing such as the type described above
which permits movement in all six directions (six
degrees of freedom).
Referring to Figures 4 and S, there is shown a
1337663
-2S-
second illustrative example of a bearing incorporating
features of the present invention in which the bearing
is formed with slits or cuts and grooves to define a
bearing housing 30 with a ~earing pad 32 that is
supported from the housing by a support structure which
includes a beam having a pair of beam portions 34a, 34b
which extend substantially in a single line away from
the pad. Moreover, the pad may be undercut so that it
is supported by the beams only on a pad support surface
34ps. Referring to Figure 5, it will be seen that the
beams 34, 34a have a convenient stub beam end as is 36,
36a which acts as a cantilever support for the beam.
As is evident from Figure 4, the perspective view
of Figure 5 shows only a portion of the pad 32. The
complete pad is illustrated in Figures 5A and 5B which
show possible modifications of the bearings illustrated
in Figure 4. As is clear from the drawings, the pad
support surface 34ps is located closer to the trailing
edge 37 than the leading edge 35. With this
construction, twisting of the beam, as illustrated in
Figure 7, will take place intermediate the beam and
create the torsional deflection illustrated. Again the
primary flexibility is developed by small cuts or slits
through the bearing housing wall. These cuts provide
the bearing pad with six degrees of freedom (i.e., the
pad can translate in the +x,-x, +y,-y, +z and -z
directions as well rotate about the x, y and z axes) and
are designed to optimize hydrodynamic wedge formation.
If the cuts or slits were terminated before breaking
through to form beam portions 34a and 34b, the pad 32
would be supported by a continuous cylindrical membrane
34m as shown in Figure 5A. The membrane acts as a fluid
damper upon which the pad 32 is supported. The
termination of the cuts would occur at Point A and Point
B of Figure 4. The flexibility of the membrane combined
with the fluid lubricant, provides a means to vary the
damping action and to isolate the pad from the housing.
The damping takes the form of a dashpot that exhibits
high damping characteristics. As with the bearing
I33766~
- 26 -
illustrated in Figures 1-3, the bearing illustrated in
Figures 4-7 is non-symmetrical about its pad centerline
and is therefore a unidirectional bearing. Accordingly,
the bearing has a leading edge 35 which deflects outward
and a trailing edge 37 which deflects inward to form a
wedge. Again, the wedge ration (ratio of spacing between
the trailing edge and the shaft to the spacing between the
leading edge and the shaft) should be between 1:2 to 1:5.
Moreover, the location of the center of action of the load
which is primarily determined by the location of pad
support portion 34ps of the beam 34 with respect to the
pad should, again, be between the circumferential center
of the pad face and the trailing edge, preferably closer
to the circumferential center of the pad face.
As shown in Figure 5B, the beam may be defined
more simply than shown in Figure 5 by simply extending the
cuts or slits downward from points A and B.
Referring to Figure 8, there is shown a third
illustrative example of a bearing incorporating features
of the present invention. In this example, internal slits
or cuts are provided to create a beam on beam support
structure. Specifically, the bearing is formed with
grooves and slits or cuts to define a pad 40 which is
supported from a housing by beams 42 and 44. The pad is
connected to the beams at support stubs 4Oa and 4Ob. Beam
attachment to the housing is at support stubs 46 and 48.
Again the bearing consists of the thin cuts or slits shown
cut through the bearing wall. The cut or slit 60 below
the pad surface introduces additional flexibility such
that under load the pad changes shape to form an airfoil
for the introduction of lubricant. Thus, as a result of
the beam on beam two point support, the pad acts as a
spring like membrane.
Figure lOA shows the deflected shape of the pad
40 under load. As shown in the drawings (exaggerated) the
pad can be formed and supported so as to deflect to an
airfoil shape under load. The airfoil dramatically
1337663
improves performance. As is evident from the drawings,
the pad is capable of displacement in the x, y, and z
directions as well as rotation about the x, y, and z axes,
that is, the pad has six degrees of freedom. Again, the
structure allows optimal hydrodynamic wedge formation.
Referring to Figure 9, there is shown the local
inherent deflection of the face pad 50 where the pad
flattens under load. These deflections are combined with
the support structure deflection shown in Figures 3 and 10
but are of a lower magnitude. The net result is the shape
shown in Figures 3 and 10 but with a face curvature that
has been minutely flattened.
Figures 31 and 31A illustrate another example
of a journal bearing in accordance with the present
invention. The bearing construction illustrated in
Figures 31 and 31A differs from the previously described
journal bearing constructions in that the bearing is
bidirectional, i.e., the bearing is capable of supporting
a shaft for either clockwise or counterclockwise rotation
as viewed in Figure 31. The bearing is bidirectional
because the pads are symmetrical about their centerline,
which are defined as the radial extending line passing
through the bearing, major axis (606), and the geometric
center of the pad. Like the previously described journal
bearings, the bearing of Figures 31 and 31A is formed with
a plurality of thin radial and circumferential slits to
define a plurality of circumferentially spaced bearing
pads 632.
The support structure for each of the bearing
pads 632 is somewhat similar to the support structure for
the journal bearing illustrated in Figure 8. In
particular, each bearing pad 632 is supported by a beam
support structure at two pad support surfaces 632ps. The
beam network connected to the bearing pads at each pad
support surface 632ps is identical yielding the
symmetrical construction of the bearing which makes the
bearing bi-directional. For purposes of simplifying this
i
1337663
- 28 -
description, only the network of beams which supports the
bearing at one pad support surface will be described since
the other pad support surface is supported in an identical
fashion. Thus, as shown in Figure 31, a first, generally
radially extending, beam 640 is connected to the bearing
pad 632 at the pad support surface 632ps. A second,
generally circumferential, beam 642 is connected to the
radially outermost end of beam 640. A third, generally
radial, beam 644 extends radially inward from the beam
642. A fourth, generally circumferential, beam 646
extends from the radially innermost portion of the beam
644. A fifth, generally radial beam 648 extends radially
outwardly from a beam 644 to the housing portion of the
support structure. In summary, each bearing pad 632 and
the bearing illustrated in Figure 31 is supported by ten
beams and the bearing housing. Further, as discussed
below, by forming radially extending circumferentially
spaced grooves or continuously extending circumferential
grooves in the housing portion of the support structure,
the housing portion of the support structure can be
designed to act as a plurality of beams or membranes. It
should also be noted that, like the bearing in Figure 8,
the cut or slit formed below the pad's surface introduces
additional flexibility such that under load the pad
changes shape to form an air foil for the introduction of
lubricant. Thus, as a result of the beam on beam two
point support, the pad acts like a spring-like membrane.
Figure 31A is a radial cross-section of
Figure 31 showing the third beam 644, the bearing pad 632
and the housing.
Figures 32, 32A and 32B illustrate another
journal bearing construction in accordance with the
present invention. This bearing construction differs from
the previously described bearing constructions in that the
bearing pads and support structure are defined by
relatively large grooves and openings formed in a
cylindrical blank. Normally, this type of construction
1337663
- 29 -
would be formed by milling the blank rather than
electrical discharge machining or some other similar
technique for forming small grooves as with the previously
described embodiments. An advantage of the bearing
construction illustrated in Figure 32 is that in
applications requiring extremely small bearings, it is
easier to form precisely the proportionately larger cuts
and openings required to form a bearing of the type
illustrated in Figures 32, 32A and 32B as compared to the
proportionately smaller cuts and openings required by the
construction of, for example, Figures 1, 4 and 8.
Moreover, the large grooves or openings are generally
easier to mold or extrude. Bearings formed by larger cuts
also find use in applications requiring extremely large
bearings with stiff bearing pad support structures.
The bearing pads shown in Figure 32 are
symmetrical about their pad centerline, 706A. Hence, the
bearing is bidirectional. Moreover, as best shown in the
perspective view of Figure 32B, the bearing has a
continuous cross-section with no hidden openings. Hence,
it is easily extrudable and easily moldable. Naturally,
the support structure can be altered by providing
discontinuities in the cross-section, e.g., by providing
radially extending circumferential grooves or non-
symmetrically disposed radially extending openings to
alter the support structure and thereby alter the
performance characteristics. The bearing's major axis is
706.
As shown in Figure 32, the bearing includes a
plurality of circumferentially spaced bearing pads 732.
Each bearing pad 732 is supported by a support structure
which includes a pair of generally radial beams 740
connected to the bearing pad 732 at a pad support surface.
A second, generally circumferentially extending, beam 742
supports each of the beams 740. Beams 742 are connected
to the housing or support stubs 744 in a cantilever type
fashion. In this bearing, the beams 740 can be regarded
1337663
- 30 -
as a primary support structure; the beams 742 can be
regarded as a secondary support structure; and the beams
744 can be regarded as a tertiary support structure.
The second beams 742 shown in Figure 32 are
defined by forming a plurality of axially extending
circumferential grooves 750 in the housing of the support
structure. In order to maintain the symmetry of the
bidirectional bearing, these grooves are circumferentially
spaced about pad centerlines 706A in a manner identical to
the circumferential spacing of the bearing pads 732.
Naturally, similar circumferentially spaced radial grooves
could be provided in any of the previous bearing
constructions. For instance, as noted above, such grooves
could be formed in the periphery of the bearing
construction illustrated in Figures 31 and 31A to provide
a further beam-like support.
Figure 32A is a radial cross-section of a
portion of the bearing illustrated in Figure 32. In this
cross-section, the bearing pad 732 and first beam 740 are
visible.
Figure 32B is a perspective view of the bearing
of Figure 32. It should be noted that although the
peripheral, circumferential and cylindrical portions of
the bearing are depicted in a somewhat segmented fashion
to emphasize the curvature, these curved surfaces are in
fact continuously curved.
Figure 33 illustrates a journal bearing
construction according to the present invention. Like the
bearing of Figure 32, the bearing of Figure 33 is formed
by proportionately large grooves and bores. In
particular, a plurality of equally spaced radially
extending circumferential grooves define a plurality of
circumferentially spaced bearing pads 832. The bearing
pads 832 are further defined by a pair of axially
extending circumferential grooves which extend
symmetrically from the planar faces of the cylindrical
blank and are best seen in Figures 33B and 33C in which
1337663
the grooves are indicated by the reference numerals 834
and 835. The bearing support structure is defined by the
aforementioned structural features and by a plurality of
circumferentially spaced symmetrically disposed shallow
bores 838 and a plurality of circumferentially spaced
symmetrically disposed deep bores 837. Because of the
presence of the "hidden" bores 837, 838, the bearing
construction of Figure 33 is not extrudable and not
moldable in a simple two-piece mold, i.e., easily
moldable.
As best shown in Figure 33A, the deep bores 837
intersect the axial grooves 836 so as to define support
structures for each bearing pad. The support structure is
further defined by a circumferential groove 839 extending
from the outer periphery of the cylindrical blank.
With reference to Figures 33-33C, it will be
understood that the provision of the structural members as
discussed above provide a support structure for the
bearing pad 832 which includes a beam 840 directly
supporting the pad, i.e., a primary support structure.
Two continuous beams 882, i.e., a tertiary support
structure and a secondary support structure comprising a
plurality of beams defined in part by bores 837 and 838
connecting the beam 840 to the continuous beams 882.
Because the support structure of the bearing
illustrated in Figures 33-33C is non-symmetrical about the
pad centerline 806A extending from the major axis 806, it
is unidirectional. Further, like the bearing of Figure
32, this bearing is particularly well suited to
applications requiring extremely small bearings since the
proportionately larger grooves and bores which define this
bearing and its support structure are more easily
manufactured.
Figures 34 and 34A-34D illustrate another
journal bearing construction in accordance with the
present invention. The bearing construction of Figure 34
is similar to that of Figure 33 insofar as the bearing
1337663
- 32 -
pads and their support structures are defined by
proportionately large grooves and bores as shown in the
drawings. The support structure for the bearing pads 932
is like the support structure for the bearing pads 832.
In particular, while the support structure for each of the
bearing pads 932 is identical, the support structure is
not symmetrical with respect to each bearing pad. Hence,
the bearing illustrated in Figure 34 is unidirectional.
Moreover, because the support structure includes "hidden"
openings, the bearing is neither extrudable or moldable in
a simple two-piece mold.
As shown in the drawings, the bearing support
structure includes a primary support structure comprising
a pair of beam-like members 940 which are connected to the
bearing pads 932 and defined in part by symmetrically
disposed openings 942. A shallow circumferential groove
formed on the outer periphery of the bearing defines a
tertiary support structure comprising a pair of continuous
beam-like elements 982. A secondary support structure
comprising a beam and membrane network 960 for connecting
the beams 940 to the continuous beams 982 is defined by
the provision of a plurality of large symmetrically
disposed bores 944, the provision of smaller symmetrically
disposed bores 946, and the provision of small non-
symmetrically disposed bores 948. By virtue of the
provision of the non-symmetrically disposed bores 948, the
support structure is more flexible, and thus biased, in
the direction of those bores.
Figures 15-18 illustrate a unitary hydrodyna~ic
thrust bearing in accordance with the present invention.
As noted earlier, thrust bearings in accordance with the
present invention incorporate some of the same features as
journal bearings in accordance with the invention. For
instance, like journal bearings, the thrust bearings of
the present invention have a major axis defined as the
central axis of the blank from which the bearing is
formed. Also the bearing pads have a circumferential
133766~
- 33 -
centerline extending radially from the major axis through
the geometric center of the pad. When the thrust bearing
is symmetrical about its circumferential centerline, it is
bidirectional; when the bearing is non-symmetrical about
its circumferential centerline, it is unidirectional.
However, by nature of their different function, the thrust
bearings have a slightly different configuration. For
example, the thrust bearing shown in Figures 15-18
includes a plurality of bearing pads 132 of substantially
identical configuration. Figure 18 shows the
circumferential dividing line CDL and radial dividing line
RDL of the bearing pad 132. The bearing pad surfaces of
the bearing pads 132 lie in a plane which is essentially
transverse to the axis of the shaft to be supported and
the bearing's major axis. Of course, when the pad
surfaces are deflected under load, or if it is desired
that the bearing be skewed slightly so as to contact the
shaft in the installed or static state, the surface of the
bearing pads may be somewhat nonplanar and somewhat skewed
with respect to the major axis or the axis of the shaft to
be supported.
A particularly important consideration in the
design of thrust bearings of the present invention is the
prevention of fluid leakage. To a large extent this
objective is achieved by designing the support structures
such that under load the inner edge of the bearing pads
deflect downward ~as viewed in Figure 16) and the outer
edge deflects upwardly. All of the thrust bearings
described herein are designed in this manner. For
instance, in the bearing shown in Figure 16, the beam 134
is connected to the pad 132 at a pad support surface 134ps
which is closer to the outer edge of the bearing pad than
it is to the inner edge of the bearing pad. Thus, the pad
support surface 134ps is located radially outward of the
radial dividing line RDL shown in Figure 18. Hence, the
bearing is designed such that, under load, the inner edge
of the bearing deflects downward. In operation, the
1337663
- 34 -
downward deflection of the inner edge of the bearing pad
corresponds to deflection away from the shaft supported
and the upward deflection of the outer edge of the bearing
pad corresponds to deflection toward the shaft. The
deflected orientation of the bearing pad significantly
inhibits the loss of fluid which otherwise occurs as a
result of centrifugal forces action on the fluid.
The loss of hydrodynamic fluid can be further
reduced by supporting the bearing pad such that under
load, the bearing pad deforms to form a lubricant
retaining pocket. Generally, such support is achieved
when the bearing pad is supported by a plurality of
radially or circumferentially spaced beams and the region
between the beams is not directly supported such that the
unsupported central region of the pad will tend to deform
outwardly so as to form a fluid retaining channel. Figure
29, which is discussed below, illustrates an example of a
bearing having the requisite radially spaced beams
therein. A greater pocket is obtained when beams are
spaced farther apart. In a similar manner, a channel can
be formed in a journal bearing by providing axially or
circumferentially spaced beam supports and an unsupported
region between the beams.
As best shown in Figures 15 and 16, each
bearing pad has a chamfer or beveled edge 132b around its
entire periphery. The purpose of the chamfer is reduce
entrance and exit lubricant losses.
Each of the bearing pads 132 is supported by
primary support portion, which in the illustrated
embodiment comprise a beam-like support member 134
supporting the pad at a bearing pad support surface 134ps.
Each beam 134 is in turn supported by a secondary support
portion such as a beam supported beam or membrane 136.
The beam or membrane 136 is in turn supported by a
tertiary support member such as pair of beam-like legs
138a, 138b.
1337663
- 35 -
By providing holes or openings 142 in the beam
or membrane portion 136, the continuous membrane 136
becomes a set of beams 136. Naturally, if holes or
openings 142 are not provided in the membrane 136, the
membrane functions as a continuous membrane.
Alternatively, the inner beam-like leg 138a could be
replaced with short stub-like beams or even eliminated to
define a tertiary support such that the secondary support
is supported in a cantilever fashion. Finally, because
the holes and openings are symmetrically disposed with
respect to the major axis, the bearing is symmetrical
about the major axis and is therefore bidirectional.
As shown in Figures 15, 17 and 18, the holes or
openings 142 which divide the continuous membrane into
separate beams are round. The use of round openings
facilitates manufacture of the bearing prototype because
circular openings can easily be drilled into the bearing
material. This is true of all the bearings described
herein. Once such circular openings are provided, it may
also be advantageous to extend the openings past the beam
or membrane member 136 to the lower portion of the bearing
pads 132 so as to define the beam-like members 134. This
is why in Figure 15, the cross-section of the pad support
surface 134ps and consequently the side walls of the beam
134 have an arcuate appearance.
Although the shape of the beam members may be
dictated by manufacturing convenience, the shape also
effects the performance of the individual bearings. Thus,
although the specific shape of the bearings described
herein, including the thrust bearing shown in Figures 15-
18, is primarily attributable to the ease of manufacturing
a prototype, it also has been found to yield excellent
results for a specific application. Any changes in the
shape would, of course, influence the performance
characteristics of the bearing by, for
- 13~7663
-36-
example, altering the bending or twisting
characteristics of the beams which support the bearing
pad. Thus, while other shapes of beams, pads and
membranes are certainly contemplated, both the ease of
manufacturing and the effect of the beam pad or
membrane's shape on bearing performance must be
considered.
Examples of other thrust bearing shapes are shown
in an Figures 21-30 and 38-39. The difference between
these bearings and the bearing construction shown in
Figures 15-18 primarily resides in different
constructions of the primary support portion, the
secondary support portion and the tertiary support
portion.
One such other bearing shape is illustrated in
Figures 21-24. A top view of the bearing as shown in
Figure 21; a cross-section of the bearing as shown in
Figure 22; a bottom view of the bearing as shown in
Figure 23 and a perspective view of the bearing is shown
in Figure 24. The bearing shown in Figures 21-24 is
similar to the bearing of Figures 15-18 with two notable
exceptions. First, the bearing of Figures 21-24
includes an angled or slanted support beam 134A rather
than a vertical support beam as in Figure 15. Second,
the bearing includes additional holes 144 which extends
through the support beam 136 to form a cylindrical
opening through the slanted or angled beam 134 so as to
form elliptical openings in the support beam. The
elliptical openings divide the beam into a pair of
complex ligaments, the shape of which can be appreciated
with reference to the perspective view of Figure 24.
The provision of the openings 144 and consequent
division of the slanted or angled beams 134A into
complex ligaments significantly increases the
flexibility of the support structure of the bearing
shown in Figures 21-24 as compared to the bearings shown
in Figures 15-18. Thus, the pads 132 of the bearing of
Figures 21-24 deflect to form a hydrodynamic wedge in
1337663
--37--
response to a lighter load than do the pads 132 of the
bearing shown in Figures 15-18. It follows that the
bearing shown in Figures 21-24 is more well suited for
supporting light loads and the bearing shown in Figures
15-18 is more well suited for carrying heavier loads.
Further, the provision of angled or slanted support
beams such as beam 134A, with or without openings to
divide the beam into complex ligaments, increases the
flexibility of the pad in the vertical direction since a
vertically applied load creates a moment which tends to
cause the beam to deflect toward the center or inner
diameter of the bearing and thereby eliminate
centrifugal leakage of the lubricating fluid.
Figure 23A shows a bottom view of a bearing of the
type shown in Figures 21-24 in which additional holes
146 are formed in the membrane or support beam 136 to
enhance the flexibility of the beam or membrane 136 even
further. As illustrated in Figure 23A, the holes 146
are formed nonsymetrically with respect to each bearing
segment. The provision of these holes in such a
nonsymetrical fashion results in a bearing in which the
pads tend to deflect more easily in one direction than
in the other direction. In other words, the bearing
pads are biased in one direction by the provision of
non-symetrical openings in the support structure.
Naturally, such nonsymetrically disposed openings can be
provided in any of the bearing constructions of the
present invention in which it is desired to bias the
bearing pads in one direction. It may even be desirable
to provide the non-symetrically disposed openings or
holes such that only selected ones of the bearing pads
are biased.
Figure 25 is a cross-sectional view of another
bearing according to the present invention. In
accordance with this construction, the bearing pad 132
is supported on a pad support stub 134S which is in turn
supported on a horizontally oriented beam portion 134H
which is in turn supported on an inversely angled beam
- 1337663
-38-
portion 134I. In other respects, the construction is
similar to that of the previously described bearings.
By virtue of this construction, the bearing has a great
deal of flexibility in one direction but it is extremely
rigid in the opposite direction.
A similar construction is illustrated in Figure 26.
The difference between the bearing illustrated in Figure
26 and the bearing illustrated in Figure 25 is that the
bearing illustrated in Figure 26 uses a vertical beam
portion 134V rather than an inversely angled beam
portion 134I. The bearings are similar in all other
respects. The absence of a angled beam in the bearing
of Figure 26 tends to give the bearing more rigidity in
the vertical direction.
Figures 27-28 illustrate another embodiment of the
bearing construction of the present invention.
As shown in the drawings, this bearing includes a
plurality of bearing pads 321-326 (shown in phantom in
Figure 28). Each of the bearing pads 321-326 are
supported on a pad support surface 342 of a bearing
support structure. The bearing support structure
includes a primary support portion composed of a pair of
nested frustums supported on a secondary support portion
which includes a split peripheral membrane 360 which is
supported on a tertiary support portion which includes a
pair of peripheral beams 382. The peripheral beams 380
and 382 are similar to those of the previously described
constructions. The membrane 360 differs from the
membrane in previously described constructions since the
membrane 360 is radially split by the groove formed in
the bottom of the bearing support structure which forms
the nested frustums. The inner frustum is inverted with
respect to the outer frustum such that the mean
centerlines of the frustums merge at a point 350 above
the pad support surface 342 and have a cross-section
which appears similar to an inverted V. Since the
centerlines of the frustums intersect at point 350 above
1337663
- 39 -
the pad surface, the primary support structure supports
the bearing pad for pivoting about a point above the pad
surface. This ensures proper deflection.
The beams 346 and 344 which support the bearing
pad can be angled toward one another at the same angle,
angled toward one another at different angles, one beam
angled and one beam not angled, and angled in the same
direction. Of course, variations in the degree of angling
of the beams in the primary support structure impacts the
deflection characteristics of the bearing.
A plurality of holes or openings 420 disposed
symmetrically about the bearing support structure divide
the nested frustum or inverted V structure into a
plurality of support beams 344, 346 and divide the apex of
the nested frustums so as to define the pad support
surfaces 342. Thus, for example, the bearing pad 321 is
supported on a pad support surface 342 by a pair of
complex support beams 344 and 346 which are tapered toward
one another and have a complex geometrical configuration
defined by the cylindrically extending openings passing
through the nested frustum section. As best shown in
Figure 27, the center lines of the beams 344 and 346
intersect at a point 350 above the pad surface to ensure
proper pivoting support. The individual beams 344 and 346
are supported on a peripheral membrane 360 which is split
by the groove which defines the frustums. The membrane is
supported by peripheral beams 380, 382. Naturally, as
discussed above, the peripheral beams 380, 382 and the
peripheral membrane 360 can be circumferentially split to
define individual beam supports.
Numerous modifications to the bearing support
structure are possible. For example, deflection of the
support structure can be modified by changing the angle of
the beams, changing the location of the holes or openings
which define the legs, varying the length of any of the
beams or membranes, and changing the width or
133766~
-40-
thickness of any of the beams or membranes. In order to
illustrate a number of these possibilities, Figures 27
and 28 depict a different support structure for each of
the bearing pads 321-326. It should be understood that
these various support structures are shown in a single
bearing for purposes of illustrating the present
invention. In normal use, each of the bearing pads
321-326 would have a similar, though not necessarily
identical, support structure to assure uniform
performance.
The support structure for bearing pad 322 differs
from that of bearing pad 321 by virtue of the provision
of a hole or opening 422 which extends through the beam
346 so as to divide the beam 346 into a plurality of
beams or sub-beams 346(a) and 346(b). If, like the
opening 422, the diameter and positioning of the opening
is such that the beam is completely separated, the beam
is divided into separate beams. On the other hand, if
the opening only partially separates the beam (e.g.
opening 423) the beam is divided into sub-beams. As
shown in Figure 27, the opening 422 forms an elliptical
opening in the side of the beam 346 such that as viewed
in Figure 27, radially outer beam 344 is visible. By
virtue of this construction, the pad 322 is supported by
three angled ligaments or beams, 344, 346(A) and 346(B).
Bearing pad 223 is supported by four angled beams
or ligaments 344(a), 344(b), 346(a) and 346(b). This
structure is achieved by providing a hole or opening 423
which extends through both beam 344 and beam 346 and
divides the pad support surface 342 into two sections.
,
It should be noted that with respect to all of the
modifications discussed herein, the size of the openings
should be selected based upon the degree to which the
beams 344 and 346 are to be divided into separate beams.
In some instances it may be desirable to completely
separate the beam sections in which case a larger
opening would be used. In other instances, such as that
1337663
-41-
illustrated with respect to the support of bearing pad
323, it is desirable to subdivide the beam at some point
along the sidewall of the beam. It should also be noted
that although the drawings only show the provision of
one opening for bearing pad support structure to divide
the beams 344 and 346. It is possible that two or more
openings similar to that of the openings 422-426 shown
in Figure 28 could be provided so as to divide the beams
344, 346 into three or more beams or sub-beams. As
always, a determination of the type of support to be
employed depends on the desired performance
characteristics. Generally, dividing the beams into
separate beams or sub-beams makes the support structure
more flexible. By making the support structure more
flexible in one direction as with the support structure
for bearing pads 322, 324 and 326 the bearing pads are
biased in a predetermined direction.
The support structure for bearing pad 324 is
similar to that for bearing pad 322 except that the
opening 424 extends through the outer support beam 344
rather than the inner support beam 346. Thus, like the
bearing pad 322, the bearing pad 324 is supported by
three angled legs.
The support structure for bearing pad 325 is
similar to that for bearing pad 321 except that an
opening 425 is provided through the outer peripheral
beam 380 and peripheral membrane 360 in a nonsymetrical
position. Thus, the bearing pad 325 is biased in a
predetermined direction, i.e., the direction of greatest
flexibility caused by the provision of the opening 425.
The support structure for the bearing pad 326 is
similar to that of bearing pad 322 except that the
opening 426 which divides the beam 346 is provided in a
nonsymetrical fashion so as to bias a bearing pad 326 in
the direction of greater flexibility, i.e., the
direction of the smaller, more flexible beam.
1337~63
- 42 -
Naturally, any combination of the support
structures illustrated in Figures 27, 28 could be employed
to achieve desired performance characteristics.
Figures 29-30 illustrate another embodiment of
the bearing of the present invention. As shown in the
drawings, this bearing includes a plurality of bearing
pads 521-526 (location shown in phantom in Figure 30).
Each of the bearing pads 521-526 are unitary with, and
supported on, a bearing pad support structure. Generally,
the bearing pad support structure includes at least a
primary support structure including an inner
circumferential support beam 546 and an outer
circumferential support beam 544, a secondary support
portion including an inner peripheral membrane 362 and a
tertiary support portion including an outer peripheral
membrane 364 and an inner peripheral support beam 382 and
an outer peripheral support beam 380. As best shown in
Figure 29, the circumferential support beams 544, 546 are
defined in part by a deep circumferential channel
extending from the bottom of the bearing to the bearing
pad. The support beams are further defined by a plurality
of holes or openings 620 disposed symmetrically about the
bearing pad support structure which separate the beams
544, 546 from adjacent beams. Thus, for example, the
bearing pad 521 is supported on a pair of beams 544 and
546, which beams have generally arcuate side walls. As
mentioned earlier, the beam support structure also
includes membranes 364, 362 and peripheral beams 380, 382.
Numerous modifications to the bearing support
structure are possible. In order to illustrate a number
of these possibilities, Figures 29 and 30 depict a
different support structure for each of the bearing pads
521-526. As with the previously described embodiment of
Figures 27-28, these various support structures are shown
in a single bearing for the purpose of illustrating the
present invention. In normal use, each of bearing pads
521-526 would have a similar, though not
13~7663
-43-
necessarily identical, support structure to assure
uniform performance.
The support structure for bearing pad 522 differs
from that of bearing pad 521 by virtue of the provision
of a hole or opening 622 which extends through the inner
circumferential beam 546 so as to divide the beam 546
into a plurality of beams s46a and 546b. By virtue of
this construction, the pad 522 is supported by three
vertically extending beams or ligaments 544,546a and
546b.
The bearing pad 523 is supported by four vertically
extending beams or ligaments 544a, 544b 546a and 546b.
This structure is achieved by providing a hole or
opening 623 which extends through both beam 544 and beam
546. The thinner beams which result from this
modification would naturally have greater flexibility
than the support structure for bearing pads 522 and 521.
The bearing pad 524 is supported by five,
relatively thin vertically extending beams or ligaments.
This structure is achieved by providing a hole or
opening 624 to divide the inner beam 546 into two beams
and providing two holes 624 to divide the outer beam 544
into three beams.
The support structure for bearing pad 525 is
similar to that for bearing pad 522 except that an
additional opening 635 non-symmetrically divides the
outer beam 544 into two beams. By virtue of the non-
symmetrical division of the outer beam 544, the bearing
pad is biased in the direction of greater flexibility.
The support structure for bearing pad 526 is
similar to that for bearing pad 522 except that the
outer beam 544 is split rather than the inner beam 546.
Further, the opening 626 is somewhat larger than the
opening 622 such that a groove is formed on the outer
periphery of the inner beam 546 so as to make the inner
1 3~ 766~
-44-
beam 546 somewhat more flexible.
Naturally, any combination of the support
structures illustrated in Figures 29, 30 could be
employed to achieve desired performance characteristics.
Figures 29A, 29B, 30A and 30B illustrate in detail
a thrust bearing in which each of the bearing pads 52lA
of the support structure are very similar to that used
to support bearing pad 521 in Figures 29 and 30. The
bearing construction is different, however, insofar as
the beams 544A and 546A are circumferentially narrower
and vertically shorter than their counterparts in the
bearing illustrated in Figures 29 and 30. Naturally,
shorter beams are more rigid than the comparatively
longer beams and narrow beams are less rigid than
comparatively wider beams. Moreover, the beam 544A is
radially narrower than the beam 546A; whereas in the
bearing illustrated in Figures 29 and 30, the beams 544
and 546 have equal widths. The difference in radial
thickness is compensated for since the large opening 620
which defines the circumferential extent of the beams
544A and 546A is arranged such that beam 544A is
significantly wider in the circumferential direction
than is beam 546A. Finally, it should be noted that the
openings 620 are significantly larger than the
corresponding openings 620 in the bearing construction
of Figures 29 and 30. Naturally, the larger openings
increases the flexibility of the support structure
defined thereby.
Figures 35-37 illustrates a combined thrust and
radial hydrodynamic bearing in accordance with the
present invention. The bearing illustrated in Figure 35
is quite similar to the bearing illustrated in Figure 34
and similar numerals are used to designate similar
structure. Similarly, as viewed in the cross-section in
Figure 37, the bearing of Figures 36-37 is somewhat
similar to the radial bearings illustrated in Figures 4
and 14D except that the bearing pad 1032 and the bearing
- 1337663
- 45 -
pad support structure, which includes beams and/or
membranes 1034, 1036 and 1038, are defined by
proportionately larger slits and grooves. However, the
radial-thrust bearings differ from radial-only bearings in
that the bearing pad surface 1032ps is angled with respect
to the major axis 1006. By virtue of its angled pad
surface, the bearings of Figures 35-37 support loads
acting both along the major axis 1006 and radially from
the axis 1006.
In order to be supported by the angled pad
support face 1032ps, the shaft must be fitted with a
runner which is angled at an angle complementary to the
angle of the pad support face. The portion of the axial
load taken by the bearing and the portion of the radial
load taken by the bearing depends on the angle of the pad
surface 1032ps. If the pad is angled at an angle a with
respect to the major axis 1006, the axial load applied to
the bearing can be determined by the following equation:
Applied Axial Load = Total Axial Load (Sin a) .
Similarly, the radial load applied to the bearing can be
determined by the following equation:
Applied Radial Load = Total Radial Load (Cos a).
The support structure for the bearing shown in
Figure 35 is similar to the support structure for the
bearing shown in Figure 34.
The support structure for the bearing
illustrated in Figures 36 and 37 includes a primary
support structure for the spaced bearing pads 1032 having
a beam 1034 which supports the bearing pad 1032, a
tertiary support structure which comprises a pair of
circumferential beams 1038 which may be continuous. The
secondary support structure comprises a membrane 1036 or a
network of beams 1036 for connecting beam 1034 to the
beams 1038. As shown most clearly in Figure 36, the
1337663
-46-
support structure for each of the plurality of bearing
pads 1032 is nonsymmetrical. Accordingly, the bearing
illustrated in Figures 36 and 37 is unidirectional.
rene-ally, any of ~he gel,erai bearing constructions
described in this application can be employed in the
design of combined radial-thrust bearings of the type
illustrated in Figures 36 and 37. Of course in order to
achieve the combined radial and thrust bearing
characteristic, the bearing pad surface must be angled
at an angle between 0 and 90 degrees with respect to the
major axis. Moreover, the need to accommodate both
radial and axial loads necessarily will impact the
design of the bearing pad support structure.
An important aspect of the present invention is the
disclosure of machinable bearing shapes. In other
words, bearing shapes which can be produced by machinina
a piece of heavy walled tubing or similar cylindrical
journal using standardly available machining techniques.
Such bearings are characterized by the fact that they
are formed from a piece of heavy walled tubing or
cylindrical or similar cylindrical journal through the
provision of bores, slits and grooves. The advantage of
such bearings is that it is easy to manufacture
prototypes and to modify these prototypes after testing.
Naturally, when the bearings are to be mass produced,
using, for example, molding or casting techniques,
different manufacturing considerations may dictate
different shapes. It is important to recognize that
changes in shape affect bearing performance.
Another manufacturing consideration is ease of
molding. Naturally, most of the bearing constructions
of the present invention are capable of being molded by
some molding technique. However, only certain shapes
can be injection molded in a simple two-piece mold,
i.e., a mold which does not include cams. Another
advantage of the bearings of the present invention is
that the bearings can be constructed with easily
1~37663
-47-
moldable shapes which are defined as shapes which can be
injection molded using a simple two-piece mold. An
easily moldable shape generally is characterized by the
absence of "hidden" cavities which require cams for
molding. For instance, with respect to radial bearings,
an easily moldable shape includes no radially extending
grooves in the inner and outer diameter and a continuous
axial cross section. The bearing shown in Figures 32,
32A and 32B is an example of an example of an easily
moldable radial or journal bearing.
Similarly, easily moldable thrust bearings are
characterized by the fact that they can be molded with a
single seam line such that, for example, when viewed
only from the top and bottom, all surfaces are visible.
Figures 38A-38C illustrates an easily moldable
thrust bearing. The bearinq includes a plurality of
circumferentially spaced bearing pads 132m and a support
structure supporting each of the bearing pads 132m. The
support structure includes a primary support portion
which includes circumferential beams 134mb and 134ma, a
secondary support portion which includes radially
extending beam 136m and a tertiary support portion which
includes the stub-like pair of beams 138m. It should be
noted that in Figures 38A-38C the dimensions of the
support structure are somewhat distorted to provide
clarity. For instance, as shown in Figure 38C, the
circumferential beams 134ma and 134mb are shown as
extremely thick. Such a beam structure would provide a
very rigid support for the bearing pads 132m and in
practice, such a rigid support would probably not be
necessary or desirable.
Variants of the specific moldable beam structure
illustrated are possible. For instance, either or both
of the spaced circumferential beam segments 134ma or
134mb could be formed as a continuous circumferential
beam element. Additionally, the secondary support
portion could include a plurality of radially extending
1337663
-48-
beams between each bearing pad 132m. Further, the
primary support structure could be modified to include
three or more circumferential beam segments connecting
each pair of adjacent bearing pads and/or
circumferential beam segments of different radial widths
could be used. Further, the stub-like beam portions
138m could be provided along the radially extending
edges of the beams 136 rather than the circumferentially
extending ends. Finally, as with any bearing in
accordance with the present invention, the structure
could also be varied by varying the length or thickness
of any of the elements in the support structure to
modify the deflection characteristics of the support
structure.
In order to illustrate a number of possible support
structure constructions, Figure 38D depicts a different
support structure for each of the bearing pads 321m-
326m. In particular, Figure 38D is a bottom view with
the modifications illustrated herein. It should be
understood that these various support structures are
shown in a single bearing for purposes of illustrating
the present invention. In normal use, each of the
bearing pads 321m-326m would have a similar, though not
necessarily identical, support structure to assure
uniform performance.
The support for bearing pad 32lm differs from that
for the bearing pads 132m in that a oval shaped
projection extends from the back of the bearing pad
surface to provide a rigid support for the outer
circumferential edge of the bearing pad 321m. By virtue
of this construction, the bearing pad 321m would be
extremely rigid at its outer circumferential end.
The support for bearing pad 322m is similar to that
to 321m except that rather than a single large
projection, two smaller projections 122m extend from the
bottom of the bearing proximate the outer
133766~
-49-
circumferential edge of the bearing pad. Like the
projection 120m, these two projections 122m provide
rigidity to the outer circumferential edge of the
bearing pad 322m. However, this construction allows the
bearing to deflect in the unsupported region between the
projections.
The bearing pad 323m is supported by modified
support structure which includes a continuous
circumferential beam 134ma in the primary support
portion. Similarly, the bearing pad 324m includes a
continuous inner circumferential beam 134mb. The
provision of such continuous beams increases the
rigidity of the bearing support structure.
The support structure for bearing pad 325 is
modified by the provision of large openings 142m in the
inner beam 134mb and smaller openings 144 in the outer
beam 134ma. The provisions of these openings increases
the flexibility of the beams. Naturally, the larger
openings 142 increase the flexibility of the beams to a
greater extent than the small openings 144. Variants of
this support structure include the use of different
sized openings or a different number of openings to bias
the bearing pad 32Sm in a predetermined direction.
The bearing pad 326m is supported by a modified
structure in which the primary support portion includes
a membrane 134m rather than a pair of beams. In the
illustrated example, one of the membranes is provided
with a opening 146 to bias the bearing pad 326m in a
predetermined direction. Of course, the provision of
the opening 146m is not necessary and if desired, a
number of openings could be provided.
As is evident from these drawings, the moldable
bearings do not include any hidden cavities which would
necessitate the use of a complex mold and/or a mold
including a displaceable cam. In particular, since each
surface of the bearing structure is directly visible in
133~ 663
-50-
either the top view of Figure 38A or the bottom view of
Figure 38B, the bearing can be simply molded using a two
piece mold. Specifically, a first mold piece defines
those surfaces which are directly visible only in the
top view of Figure 38A. The second mold piece defines
those surfaces which are only visible in the bottom view
of Figure 38B. Surfaces having edges visible in both
Figures 38A and 38B can be molded using either or both
molds. In the illustrated bearing, easy moldability is
achieved because the secondary and tertiary support
portions are circumferentially located in the space
between bearing pads. The modifications illustrated in
Figure 38D do not alter the easy moldability of the
bearing.
More complex variants of the moldable thrust
bearing illustrated in Figures 38A-38D are possible. In
particular, any of the previously discussed
modifications of the bearing structure which can be
adapted to easy molding could be employed. For
instance, the primary support beams could be continuous.
Thus, the provision of an easily moldable bearing does
not necessarily require a simple bearing construction.
An example of a more complex bearing structure is
illustrated in Figures 39A-39C.
As illustrated in Figures 39A-C, the bearing
includes a plurality of circumferentially spaced bearing
pads 232m supported by a bearing pad support structure.
The secondary and tertiary portions of the support
structure are similar to corresponding portions of the
bearing support structure of Figure 38. However, the
bearing of Figure 39 differs from the bearing of 38 in
that in the bearing of Figure 39 the primary support
portion includes a plurality of complex beams 234.
Specifically, each bearing pad is supported by a
radially outer continuous complex circumferential beam
234ma. The pads are further supported by the plurality
of spaced circumferential complex beams 234mb. The
complex shapes of the continuous beam 234ma and the beam
1337663
-51-
segments 234mb can be best appreciated with reference to
Figure 39C which shows, somewhat schematically, the
profile of the complex beams 234. In operation, the
beams 234ma and 234mb function as a beam network. Thus,
it can be seen that numerous complex thrust bearing
constructions can be provided while retaining the
ability to mold the bearing with a simple two-piece
mold, i.e., easy moldability. Naturally, each structure
provides unique deflection characteristics which must be
considered in designing the bearing for optimum wedge
formation.
In certain gas or air lubricated deflection pad
bearings, there are cases where loads or speeds exceed
the capability of an air film. In these cases, it is
necessary to introduce a liquid type lubricant into the
converging wedge without providing a liquid reservoir or
bath. Figures 40, 4OA, 41 ~n~ a~
constructions for achieving this purpose. In
particular, these drawings illustrate a novel self
lubricating deflection pad bearing in accordance with
another important aspect of the present invention. The
bearing is essentially a deflection pad bearing of the
type described herein which has been modified to include
lubricating plastic in its various openings.
The plastic employed in the bearing is a
conventional castable porous plastic which is capable of
absorbing lubricating liquid when soaked in such a
liquid. One such plastic is sold under the tradename
POREX. Generally, the porous plastic can be formed from
various plastics by injecting air into the plastic
material to form the pores. In particular, the liquid
is absorbed into the porous plastic in a wi~k like
manner and held in place by the plastic.
The lubricating deflection pad bearing is
constructed by taking a conventional journal, thrust or
combined radial and thrust deflection pad bearing of the
type described above and casting or injecting the
1337663
- 52 -
conventional porous plastic around and into the spaces
between the deflection members. As a consequence of this
construction, during operation, the movement of the shaft
and the compression of the deflection members causes the
lubricating liquid to leave the porous plastic be drawn
into the leading edge of the converging wedge. The
formation of the liquid filled wedge greatly increases the
load and speed capability of the bearing. After the
liquid passes over the pad surface, it is reabsorbed by
the porous plastic after leaving the trailing edge.
An important aspect of the present invention is
the composite structure combining a standard bearing
material with the porous plastic. By virtue of this
composite, it is possible to take advantage of the unique
characteristics of both materials. More specifically,
conventional porous plastics alone make poor deflection
pad bearing materials because the pores in the plastic are
actual voids that are detrimental to the development of
the very thin fluid film. On the other hand, conventional
plastic or metal bearing materials not having the pores
are incapable of absorbing lubricant to any great extent.
However, through the use of both materials in the manner
described, an effective self-lubricating hydrodynamic
bearing can be obtained. Further, there are synergistic
results from the combined use of standard bearing material
and lubricant absorbing porous plastic. For example, the
deflections of the bearing surface assist in forcing the
liquid lubricant into the leading edge. Moreover,
channeling or lubricant ret~;ning deformation of the
bearing surface assists in cont~ining the liquid.
Figures 40 and 41 show two examples of the self
lubricating deflection pad bearing of the present
invention. In particular, these drawings show bearings
similar to bearings described previously which have been
modified to include the liquid absorbing porous plastic
filled into the spaces between the deflection members.
1337663
-53-
To some extent, the bearing acts as a skeletal portion
and the porous plastic portion acts as a lubricant
retaining and releasing sponge.
~ particular, Fig-~res 40 and 40A show a self-
lubricating bearing having an underlying bearing
structure which is essentially identical to the bearing
shown in Figures 32 and 32A. However, the bearing
structure of Figure 40 is modified such that porous
plastic fills the openings between the bearings and the
openings within the support structure which are
continuous with the spaces between the bearing pads 732.
Naturally, the spaces under the bearing pads could be
filled with porous plastic as well. However, unless
there is communication between the porous plastic and
the bearing pad surface, the provision of such porous
plastic areas would be fruitless.
Likewise, Figures 41 and 4lA show a bearing having
a construction virtually identical to the construction
of the combined radial and thrust bearing shown in
Figures 36 and 37. However, porous plastic is again
injected into the interstices or spaces within the
support structure between the end between the pads.
Again, the injection of the porous plastic as
illustrated results in a bearing having a continuous
inner diameter. However, like the bearing of Figure 40,
the material characteristics across the inner diameter
vary significantly.
Specifically, like the bearing of Figure 40, the
inner diameter of the bearing of Figure 41 includes
wedge supporting bearing pad surfaces and
circumferentially spaced lubricant releasing and
absorbing and retaining portions. In operation, the
movement of the shaft and the compression of the
deflection members causes the lubricating liquid to
leave the porous plastic and to be drawn into the
leading edge of the converging wedge. The formation of
the liquid filled wedge greatly increases the load and
1337663
54
speed capability of the bearings.
The manufacturer of the self-lubricating deflection
pad ~ezring in-~olve~ ~hree general steps. First, the
basic bearing or skeletal portion is formed standard
bearing material. Second, the porous plastic is
injected into the desired spaces in the bearing
structure. For purposes of manufacturing convenience,
the plastic is injected to the bearing without
lubricant. Finally, the bearing with the porous plastic
injected into the desired spaces is loaded with liquid
lubricant. To properly load the plastic with liquid
lubricant, it is necessary to wick the lubricant in from
one side. The merging in the liquid results in an
unfilled internal portion. This is caused by not
allowing the pores to vent from one side. In Figure 40
the basic bearing structure is combined radia~ ?.nd
thrust structure similar to that shown in Figure 36.
However, porous plastic fills the interstices within the
support structure. The provision of the porous plastic
yields a composite bearing having a continuous inner
diameter surface. However, the deflection
characteristics cross the surface very greatly.
Specifically, the deflection pads which are formed of
standard bearing materials such as metal or non-porous
plastic is suited for deflection and formation of a
fluid wedge. On the other hand, the porous plastic
portions are suited for compression so as to release
lubricant at the reading edge of the bearing pads and
absorbing lubricant at the trailing edge of the bearing
pads.
As noted with respect to each of the illustrative
examples described above, the bearings of the present
invention can be formed to provide for a wedge ratio of
1:2 to 1:5, have a deformable bearing surface the shape
of which can be modified, allow six degrees of freedom
of the pad, and provide a dashpot type damping action.
The bearings are typically of a unitary construction.
133766~
By virtue of the wedge formed by deflection of the
bearing pad and the ability of the pad to move with six
degrees of freedom, the bearing of the present invention
exhibits exceptional performance characteristics.
Sp~ifically, the ~earln~ dimensions and de~lection
variables including number, size, shape, location and
material characteristics of the elements defined in the
unitary bearing can be tailored for any specific
application to support a wide variety of loads. of
these variables, the shape of the support members is
particularly important. The impact of shape of the
support members on the deflection characteristics of the
support structure can be appreciated when the variable
formula for moment of inertia bh3/12 (English units)
(the main component of sectional modulus for rectangular
section, z = I/c = bh2/6) used in an example. Moreover,
the ability of the pad to move with six degrees of
freedom allows the bearing to compensate for and cnrre~
shaft misalignment. In this regard it is noted that the
bearings of the present invention have a self correcting
characteristic resulting from the tendency of the
bearing to return to its non-deflected state due to the
stiffness of the bearing. Of course, the stiffness of
the bearing is primarily a function of the shape of the
support structure, and to a lesser extent the other
deflection variables including number, size, location,
and material characteristics of the elements defined by
the grooves and cuts or slits formed in the unitary
element. Stiffer bearings have a greater self-
correcting tendency but are less able to adjust for
shaft misalignment.
Tests have shown that bearings incorporating the
features of the present invention exhibit dramatically
improved performance even in comparison to the structure
disclosed in the present inventors prior patent No.
4,496,251. In a recent test the journal bearings of the
present invention were utilized in a radial bearing
with a radial envelope of 0.091"(2.31mm). Inward
deflections of the bearing pad were 0.0003"(.0076mm)
1~37663
-56-
which provides exceptional stability and bearing
performance. A comparable displacement using the
arrangement shown in the present inventor's prior patent
No. 4,496, 251 would have required a radial space of
0.30"~7.6mm).
In conventional hydrodynamic journal bearings, it
is typically necessary to provide a fluid-film clearance
between the bearing pad surface and the shaft portion to
be supported. This requires extremely close
manufacturing tolerances which can present an obstacle
to high volume production.
The bearings of the present invention can be
designed to obviate the need for such close
manufacturing tolerances. Specifically, by providing
appropriate bores, grooves and cuts or slits, it is
possible to define a bearing havina v~ ' ' y any
desired performance characteristic. One such
characteristic is the stiffness or spring characteristic
of the bearing pad in the direction of load, i.e., in
the radial direction (radial stiffness) with respect to
journal bearings and in the axial direction (axial
stiffness) with respect to thrust bearings. It is known
in the bearing art that the fluid film between the shaft
and the bearing may be modeled as a spring since it has
a calculatable radial or axial fluid film stiffness or
spring characteristic. This is true for both
compressible and incompressible fluids but is
particularly useful in regard to gas fluid lubricants.
The fluid film stiffness and the bearing stiffness act
in opposition to one another such that if the fluid film
stiffness or spring characteristic exceeds the bearing
stiffness or spring characteristic, the bearing will
deflect in the direction of the fluid film stiffness
(i.e., radial direction for journal bearings and axial
direction for thrust bearings) until the stiffness of
the fluid and the bearing are in equilibrium. Thus, it
has been found that if a journal bearing is designed
such that radial stiffness of the bearing is less than
l33~6~3
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the radial stiffness of the fluid film, it is not
necessary to provide a precise spacing between the shaft
and the bearing because the radial stiffness of the
fluid film will automatically and instantaneously, upon
rotation of the shaft, cause appropriate radial
deflection of the journal bearing. The virtually
instantaneous wedge formation results in virtually
instantaneous formation of the protective fluid film
thereby preventing damage to wedge forming surface which
typically occurs at low speeds during the formation of
the fluid film.
The radial stiffness of the bearing is, of course,
primarily a function of the section or flexure modulus
of the support structure which depends on the shape of
the support structure. The radially stiffness of the
pad also depends on the length of the slits or cuts
formed in the bearing. The same is true of th_i~st
bearings except, naturally, the axial stiffness of the
bearing is critical. Accordingly, with the present
invention, it is possible to achieve high performance
without the close manufacturing tolerances typically
required of hydrodynamic bearings.
For example, the bearings of the present invention
may be designed to have an interference fit when
installed on the shaft such that as the bearing is
forced on the shaft the pads deflect, slightly so as to
form a converging wedge shape while in the stationary
installed position. Contact between the bearing pad and
shaft being at the trailing edge. At instantaneous
start up, the fluid film enters the wedge and builds up
fluid pressure causing separation of the shaft and pad.
Thus, in accordance with another important aspect of
this invention, the bearings of the present invention
may be designed and dimensioned such that the trailing
edge of the bearing is in contact with the shaft portion
to be supported when the shaft is at rest.
The thrust bearings of the present invention can
1~37663
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also be designed to provide a statically loaded wedge.
In order to provide a statically loaded wedge, the
support structure for the bearings is designed such that
the bearing pads slope toward the shaft from the
radially inner circumferentiai edge of the bearing pad
to the radially outer circumferential edge of the
bearing pad. Further, the support structure is designed
such that the bearing pad slopes toward the shaft from
the radially extending leading edge to the trailing
edge. In this way, a statically loaded wedge
approximating the optimum wedge is formed. Further, the
pad is sloped toward the shaft at the outer
circumferential edge so as to provide the desired fluid
retaining characteristic. The stiffness of the support
structure can also be designed such that an appropriate
space inbetween the pads and shaft is established
instantaneously upon rotation of the shaft.
Alternatively, the bearing may be designed such
that the entire bearing pad contacts the shaft portion
to be supported when the shaft is at rest. This aspect
of the present invention is particularly useful in high
volume production of the bearings and with bearings
using gas lubricating fluids because it allows a much
larger variation of machining tolerances. In one
example, a .003 inch variation can be designed to have
an insignificant impact on the wedge whereas
conventionally machining of known gas bearings require
.OOOOOx tolerance which can only be attained through the
use of sophisticated and expensive machining techniques
such as micro inch machining via etching.
In small quantities, the bearings disclosed herein
are preferably constructed by electrical discharge
machining or laser cutting methods. The double lines
shown in the drawings are the actual paths of the wire
or beam which is typically 0.002-0.060"(0.50-1.52mm) in
diameter. The lubricant that flows into the electrical
discharge machined paths, acts as a fluid dampener that
reduces any vibration or instability at resonant
1337663
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frequencies. In the situations described above where a
continuous cylindrical membrane is formed, the damping
takes the form of a dashpot that exhibits high damping
characteristics. A significant consideration in the
design is that the support structure length and
direction be oriented to provide the inward deflection
shown in Figure 3. Also minute deflections of the pads
themselves in the direction of load as shown in Figure
9, result in eccentricity changes which further improve
bearing performance. It is noted that in Faires, Design
of Machine Elements the distance between the center of
the bearing and the center of the shaft is called the
eccentricity of the bearing. This terminology is well
known to those skilled in bearing design. With the
novel approach of tuning or modifying the stiffness of
the bearing configuration or structure and particularly
the beam to suit a particular bearing application,
optimum performance is readily obtained. ~ent
computer analysis has demonstrated that virtually any
stiffness or deflection may be accomplished.
As noted above, when manufacturing low volumes or
prototypes of the bearings of the present invention, the
bearings are preferably constructed by electrical
discharge machining or laser cutting methods. Such
small volumes or prototypes are usually constructed of
metal. However, when higher volume production of a
particular bearing is contemplated, other methods of
manufacture such as injection molding, casting, powdered
metal die casting and extrusion are more economical. In
connection with such manufacturing methods, it may be
more economical to employ plastics, ceramics, powdered
metals or composites to form the bearings of the present
invention. It is believed that methods such as
injection molding, casting, powdered metal die casting
with sintering and extrusion are sufficiently well known
that the processes need not be detailed herein. It is
also believed that once a prototype bearing is
constructed, the method of producing a mold or the like
for mass production of the bearing is well known to
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those skilled in the molding and casting art. Moreover,
it is to be understood that only certain types of the
bearings of the present invention are adapted to be made
in high volumes through extrusion. Generally, these are
the bearings that are formed only through the provision
of circumferential grooves and radial and
circumferential cuts or slits which extend axially
throughout the entire bearing. In other words, those
bearings having a constant or otherwise extrudable
cross-section.
In accordance with another aspect of the present
invention, a novel investment casting method has been
found to be particularly useful in the manufacture of
intermediate quantities, e.g., less than 5,000 bearings.
In accordance with this method of manufacture, the first
step of the investment casting procedure is manufacture
of a prototype bearing. As discussed above and detailed
below, the prototype can be manufactured in any number
of ways, but is preferably manufactured by machining a
piece of heavy walled tubing or similar cylindrical
journal. In larger bearings, the cylindrical journal
typically is machined using a lathe for forming face
and circumferential grooves, and a mill for forming
axial and radial bores. In machining smaller
cylindrical journals, techniques such as water-jet
cutting, laser and wire electrical discharge techniques
are generally more suitable. However, in either
application the journals are typically turned and milled
to form the larger grooves.
After the prototype bearing is formed, it may be
desirable to test prototype to confirm that the bearing
functions in the predicted manner. As a result of such
testing, it may be necessary to modify and refine the
prototype to obtain the desired results.
Once a satisfactory prototype is obtained, a rubber
mold of the prototype is formed. Typically, this step
involves encasing the prototype in molten rubber and
13376B3
-61-
allowing the rubber to harden so as to form a rubber
mold of the prototype. The rubber encasing the
prototype is then split and the prototype is removed to
yield an open rubber mold.
Once the rubber mold is obtained, it is used to
form a wax casting. This step typically involves
pouring molten wax into the rubber mold and allowing the
wax to harden to form a wax casting of the bearing.
After the wax casting is obtained, it is used to
form a plaster mold. This step typically involves
encasing the wax casting and plaster, allowing the
plaster to harden around the wax casting so as to form a
plaster mold.
The plaster mold can then be used to form a
bearing. Specifically, molten bearing material, such as
bronze, is poured into the plaster mold so as to melt
and displace the wax casting from the mold. Thus, the
plaster mold is filled with molten bearing material and
the melted wax is removed from the plaster mold.
After the molten bearing material is allowed to
harden, the plaster mold is removed from around the
bearing and a bearing is obtained.
Because this method of manufacture involves the
sacrifice of a wax casting, it is known as investment
casting or sacrificial casting.
Despite the fact that the investment or sacrificial
casting method described above involves sacrifice of a
wax casting and the production of both rubber and
plaster molds, and is quite labor intensive, it has
proven to be cost effective when intermediate
quantities, e.g., less than 5,000 units, of a particular
bearing are required. The cost effectiveness of this
procedure for lower ~uantity bearing requirements is due
to the fact that the molds used in this method are far
1337663
-62-
less expensive to produce than the complex mold required
for injection molding or powdered metal casting.
As noted above, the first step in the investment
casting method, indeed in any m~od, ~I produclng
bearings in accordance with the present invention is the
production of a prototype bearing. In accordance with
another aspect of the present invention, the relatively
complex journal and thrust bearings of the present
invention can be formed using simple manufacturing
techniques. Similar techniques are used for both thrust
and journal bearings.
With the foregoing in mind, it is believed
sufficient to describe the method of making a single
journal bearing through the use of electrical discharge
manufacturing and machining. It is believed that a
description of such manufacture demonstrates the ea~sP
with which the relatively complex bearing shapes of the
present invention can be achieved.
Each bearing is initially in the form of a
cylindrical blank having a cylindrical bore as shown in
Figures llA and llB. The blank is then machined to
provide a radial lubricating fluid groove as shown in
Figures 12A and 12B. For certain applications, it is
desirable to further machine the blank to include facing
grooves which are preferably symmetrically disposed on
the radial faces of the bearings as shown in Figures 13
and 13B. The provision of such facing grooves
ultimately results in a bearing which is easily
torsionally deflected. While the groove shown in
Figures 13A and 13B are cylindrical, it is possible to
provide tapered grooves as shown in Figures 14A and 14B.
As will become evident below, this yields a bearing
which exhibits improved deflection characteristics by
virtue of the angled alignment of the support beams. In
this context, it should be noted that it is preferable
that the support beams as viewed in Figure 14A are
tapered along lines which converge at a point proximate
1337663
- 63 -
the center line of the shaft. This ensures that
flexibility occurs about the shaft center line by
establishing a center of action for the entire system such
that the pads may adjust to shaft misalignment. In
essence, the tapering of the support beams causes the
bearing to act in a manner similar to a spherical bearing
by concentrating the support forces on a single point
about which the shaft may pivot in all directions to
correct any misalignment. The arrows in Figure 14A
illustrate the lines of action of the deflection.
Bearings having cross-sections of the type
shown in Figures 12A and 14A are particularly effective at
retaining the hydrodynamic fluid. This is because the
bearing pad is supported proximate the axial ends of the
bearing pad and the central portion of the bearing pad is
not directly supported. By virtue of this construction,
the bearing pad is supported so as to deform under load to
form a fluid ret~in'ng concave pocket, i.e., the central
portion of the bearing pad deflects radially outward.
This greatly decreases fluid leakage. Naturally, the
degree of pocket formation depends on the relative
dimensions of the bearing pad and support structure. A
larger fluid retaining pocket could be obtained by
providing a thinner bearing pad surface and supporting the
pad surface at the extreme axial ends of the bearing pad.
After the cylindrical blank is properly
machined as shown in Figures 12A and 12B, Figures 13A and
13B, or Figures 14A and 14B, radial and/or circumferential
slits or grooves are formed along the radial face of the
machined blank to define the bearing pads, the beam
supports and the housing. Figures 14C and 14D illustrate
such grooves formed in the machined blank of Figures 14A
and 14B. When manufacturing low volumes of the bearings
or prototypes of the bearings for use in the construction
of a mold, the cuts or slits are preferably formed through
electrical discharge manufacturing or through the use of a
laser. The machining of the cylindrical blanks to achieve
- 64 - 1 3 3 7 6 63
the configurations illustrated in Figures 12A and 12B,
Figures 13A and 13B, Figures 14A and 14B or a similar
shape, can be done through conventional machine tools such
as lathe or the like.
Although the foregoing discussion is
specifically directed to journal bearings, the principles
apply just as well to thrust bearings. For instance, the
thrust bearing shown in Figures 15-18 can be formed by
machining a section of heavy walled tubing to provide
radially inner and outer grooves, facing grooves, axial
bores, radial cuts and chamfers so as to define bearing
pads and support structure.
The performance characteristics of the bearings
of the present invention results from the relative shape,
size, location and material characteristics of the bearing
pads and the beam supports defined by the bores and cuts
or slits formed in the machined blank. These parameters
are largely defined by the dimensions and location of the
radial circumferential bores, cuts or slits formed in the
bearing in conjunction with the shape of the machined
blank in which the bores or slits are formed to yield the
bearing.
As noted above, while the construction of the
bearings of the present invention is most easily
understood by reference to the machining process, larger
quantities are preferably manufactured through the
investment casting method of the present invention, and
even larger scale production of the bearings contemplated
by the present invention could be more economically
performed through injection molding, casting, powdered
metal, die casting, extrusion or the like.
In extruding a large number of bearings from a
pipe-like cylindrical blank, radial lubricating fluid
grooves as shown in Figures 12A and 12B can be provided
1337663
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along the length of the pipe-like cylindrical blank
prior to extrusion. However, if facing grooves were
desired in the bearing, these can be individually
defined after slicing the individual bearings from the
extruded and machined blank. For this reason, extrusion
might not be a preferred method of producing bearings
which require facing grooves to enhance torsional
flexibility.
