Note: Descriptions are shown in the official language in which they were submitted.
BEARING ASSEMBLY
Backqround of the Invention
The present invention relates to ~ bearing
assembly which is capable of attaining high load
capabilities while maintaining low power losses for both
flooded and starved conditions.
Hydrodynamic bearinys in the form of pillow
blocks have been used for at least a hundred years and
still remain in wide use as lubricated bearing systems for
general applications. Initially, such bearings were
utilized for relatively low-speed operations at moderate
radial and thrust loads. More recently, however, the range
of use of such bearings has been extended due to the
introduction of new lubricants, improvements in oil
delivery systems, overall improvements to bearing
efficiency, and the like.
Hydrodynamic bearings depend on positive pressure
generation of an oil film for successful operation. In
considering proper bearing operation, many factors come
into play such as film thickness, lubricant temperature,
operational speeds, load, and the like.
Generally speaking, bearing operations may vary
considerably under different operating conditions such as
the parameters noted above, and a bearing designed for one
particular set of operating conditions may not perform
effectively for a different set of conditions. Heretofore,
such differences have led manufacturers to generally design
separate bearing assemblies for specific operational
parameters. Such, of course, increases not only the
expense of an individual bearing assembly, but also
requires the manufacturer to maintain significant
inventories to meet the needs for each separate
application.
It has been stated that a bearing should be
capable of operation at maximum load capabilities with
minimum power losses under both flooded and starved
lubricant conditions. The bearing system according to the
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present invention approaches the aforementioned optimized
bearing operation. Particularly, bearings according to the
present invention have the ability to carry both radial and
thrust loads at reduced power losses under both fluid film
starved and flooded conditions. Furthermore, the bearing
system of t~e present invention may be presented in a self-
contained unit that requires very little maintenance, and
in the same form can utilize an external source of
lubricant and/or an internal circulating lubricant system.
Summary of the Invention
It is thus an object of the present invention to
provide an improved bearing assembly.
Another object of the present invention is to
provide an improved bearing assembly that is capable of
carrying radial and thrust loads while operating at high
load capabilities and low power losses for both flooded and
starved conditions.
Yet another object of the present invention is to
provide an improved bearing assembly which is self-
contained, requires little maintenance and may operate in
the same form under both external or internal lubricant
supply arrangements.
Still further, another object of the present
invention is to provide an improved bearing assembly that
is modular in concept such that the assembly may serve as a
simple combination journal-thrust bearing, or a simple
thrust bearing.
Another object of the present invention is to
provide an improved bearing assembly that performs
effectively under both starved and flooded lubricant
conditions.
Yet another object o~ the present invention is to
provide improved thrust bearings that operate under starved
or flooded lubricant conditions with less power loss than
conventional thrust bearings. In general, the improved
bearing assembly according to the present invention
comprises a housing, said housing defining a shaft
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receiving opening therethrough; and a liner received in
said housing and defining a sha~t receiving bore opening
therethrough and an enlarged annulus at opposite ends of
said bore opening, said liner further defining at least one
5 lubricant delivery groove adjacent said bore opening and
extending axially therealong, said lubricant delivery
groove having restrictor means at opposite ends of same,
said bore opening having a length ~o diameter ratio of from
about 0.87 to about 0.92. A thrust bearing may also be
10 secured to opposite sides of said liner. The thrust
bearings have a shaft receiving bore opening therethrough
concentric with said liner bore opening and include a
plurality of thrust pads located in a thrust surface around
the bore opening. The thrust pads have radially extending
15 lubricant grooves or passageways located therebetween, and
with oil bleeders or throttles at outer ends of same.
More specifically, two or more lubricant delivery
grooves are preferably located in the liner bore opening,
along opposite sides .o~ same, and are preferably provided
20 with openings therein for communication with an external
source of lubricant which may be supplied therethrough for
lubrication of the bearing. Likewise, utilizing an
internal lubricant delivery system, one or more oil rings
or the like may be received about the liner to rotate with
25 rotation of a shaft received therethrough and with a
portion of the ring passing through a lubricant reservoir
where lubricant is removed therefrom and is transported to
the rotating shaft. In a preferred arrangement, the oil
ring is present with supplemental lubricant feed from an
30 external source, if required. Likewise, with both an oil
ring delivery system and an external oil delivery system,
the oil ring system serves as an emergency back-up system
in the event of failure of the external system.
A bearing assembly according to the present
35 invention may include a journal bearing and thrust
bearings, thrust bearings only or thrust bearings in
combination with roller bearings or the like. Thrust
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bearings utilizable with the assembly o~ the present
invention are secured to opposite ends of the bearing liner
and define a sha~t bore opening therethrough, concentric
with the boxe opening ~hrough the liner. A plurality of
thrust pads are provided around the bore openiny of the
thrust bearing and extend radially outwardly therefrom.
Each of the thrust pads is provided with a suitable
configuration for receipt and subsequent delivery of
lubricant to a thrust runner employed in conjunction
therewith. In one embodiment, the thrust pads may be
scalloped at opposite ends defining a compound taper with a
lubricant groove located along a lower surface of adjoining
scalloped areas and with a lubricant bleeder being located
between said groove and an outer periphery of the thrust
surface of the bearing. Likewise, in another form the
thrust pads include a flat land area and a tapered area
with radial oil grooves located between thrust pads.
Insofar as the overall bearing assembly of the
present invention is concerned, it has been determined
surprisingly through experimentation that when certain
liner bore and thrust bearing characteristics are
maintained within certain ranges, the bearing assembly is
suitable for very wide-spread use approaching the optimized
bearing assembly as discussed hereinbefore. In other
words, a bearing assembly according to the present
invention, and in the correct configuration is capable of
maintaining high thrust and radial loads with minimal power
loss under both flooded and starved lubricant conditions,
and thus capable of achieving higher rotational speeds.
Important characteristics for the journal bearing
liner include axial length of the bore, the length of the
axial oil groove, location of oil feeding holes inside the
axial oil groove, width and shape of the axial oil groove,
undercutting of the upper liner bore in a slot area for
receipt of oil rings, and dams located on the split surface
of the liner at the entrance to the axial oil groove.
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The thrust face of a liner so e~uipped, as noted
above, is also important insofar as enhancement of thrust
bearing characteristics under the overall load conditions.
Features deemed to be of importance for the thrust bearings
are the number of pads per thrust surface, shape of the
pads, slope angle of the pads, location of the pads on the
thrust surface, a separating shroud between the pads and
the exterior of the bearing, and a circumferential groove
delivering oil from the journal bearing to the thrust
surface pads.
Additionally, modular design of the bearing
assembly according to the present invention accommodates
both thrust units and expansion bearings while as noted
below, optimized geometry of the bore and thrust surfaces
permit operation with minimum power losses for both flooded
and starved lubricant conditions for ring type and/or
external lubricant circulation systems. Still further
symmetrical configuration of the thrust pads for one
embodiment of the thrust bearing permits bi-directional
bearing operation while a large number of pads per thrust
surface leaves a substantial flat area on the thrust
surface, contributing to an increase in total load
capacity.
Tapering of the pads on one embodiment of the
thrust surface in both a circumferential and radial
direction and with no peripheral openings for oil in the
thrust surface therearound, limits exit of oil therefrom
only through the radial bleeders, whereby there is greater
effective use of the amount of lubricant present. A
circumferential qroove geometry on the thrust surface at
the inner radius of the thrust bearing permits oil flow
from the axial oil groove in the liner into each feeding
groove of the thrust surface. Further, the particular
design of the axial oil grooves in the bore of the journal
bearing and the location of the oil inlets provides a most
effective mixing of oil whereby an extended film length is
realized.
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d ~r~ 33
Dams may also be provided, located on the split
surface of the liner, at tha entrance to the axial oil
groove which restrict the flow o~ oil and direct oil into
the axial oil groove. Moreover, circumferential
5 undercutting of the upper liner portion bore in a slot area
allows penetration of the oil into the clearance gap and
protects from oil losses.
Brief Description of the Drawings
The construction designed to carry out the
10 invention will be hereinafter.described, together with
other features thereof.
The invention will be more readily understood
from a reading of the following specification and by
reference to the accompanying drawings forming a part
15 thereof, wherein an example of the invention is shown and
wherein.
Figure 1 is a partial cut-away view of a bearing
assembly according to the present invention illustrating a
partial vertical cross-section of an interior configuration
20 of the bearing assembly.
Figure 2 is a side elevational view of a bearing
assembly illustrated in Figure 1.
Figure 3 is a vertical cross-sectional view of
one embodiment of a journal bearing liner according to
25 teachings of the present invention.
Figure 4 is an end elevational view of an end of
the liner as shown in Figure 3 illustrating one embodiment
of a thrust bearing secured to same.
Figure 4A is a planar view of a portion of the
30 ~ thrust bearing as illustrated in Figure 4.
Figure 5 is an enlarged cross-sectional view of
an axial oil delivery groove present in the bore of a
journal bearing liner according to the present invention.
Figure 6 is an enlarged view of a portion of an
35 enlarged annulus provided at opposite ends of the shaft
receiviny boxe of a bearing liner according to the present
invention.
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3~
Figure 7 is an enlarged view of a portion of a
thrust pad according to the thrust bearing embodiment
illustrated in Figure 4.
Figure 8 is an enlarged view of a further
embodiment of a thrust pad as illustrated in Figure 4.
Figure 9 is a verkical cross-sectional viaw of a
further embodiment of a bearing liner according to the
present invention.
Figure lo is an elevational view of a further
embodiment of a thrust bearing secured to opposite ends of
the liner as illustrated in Figure 9.
Figure 11 is a further detailed planar view of a
portion of the thrust bearing embodiment as illustrated in
Figure 10.
Figure 12 is cross-sectional view of a portion of
the thrust bearing surface illustrated in Figure 11 taken
along a line XII-XII, and with a portion of a thrust runner
in cross-section adjacent thereto.
Descrlption of the Preferred Embodiment
Making reference to the figures, preferred
embodiments of the present invention will now be described
in detail.
Figures 1 and 2 illustrate an overall bearing
assembly, arranged for use as a free bearing, i.e. one
not intended to handle thrust loads. A pedestal housing
generally 10 is provided that serves as a general enclosure
for the elements of the overall assembly and defines an
opening 11 therethrough for supporting receipt of a
rotatable shafts as shown in phantom.
An upper housing portion 8 and a lower housing
portion 9 are interconnectable to provide housing lO with
stiffness in a radial direction from a bore opening 18
located through housing 10 for receipt of shaft S. Housing
10 defines an oil reservoir 12 in lower housing portion 9
to preferably permit an oil ring utilized therewith to be
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submerged in oil at a level of about 15 percent of the
diameter of the oil ring.
A liner, generally 30 is received within housing
10, being properly located and maintained by a plurality
of liner mounts 14, 15 that are de~ined by an inside of
housing portions 8 and 9, respectively, above and below
liner 30. When the bearing assembly is properly received
and secured about a sha~t S, liner 30 is thus restricted
from movement. Also as illustrated in Figure 1, a spac~ 17
is defined adjacent liner 30 on opposite sides of same
(only one shown) for receipt of a thrust runner 80
(indicated in phantom) when the bearing assembly is to be
employed as a fixed bearing, and as described in more
detail hereinafter. A sealing ring 16 is also provided on
opposite sides of housing 10 within opening 11 and defines
shaft receiving bore opening 18 therethrough. Sealing ring
16, in conjunction with appropriate seals thus basically
encloses housing 10 and thus the bearing assembly by
inclusion of proper seals at relevant locations. Moreover,
as generally illustrated but not described, housing 10 is
also provided with conventional means for mounting same,
ports for the addition of lubricant, means to facilitate
transfer of the assembly and the like.
Liner 30 as illustrated in Figure 3 includes an
upper liner portion 30A and a lower liner portion 30B
which are interconnected to define the overall liner
structure. Unless the individual liner portions 30A, 30B
`are necessary for description of a particular feature,
reference will be made hereinafter only to liner 30.
Liner 30 defines a shaft receiving bore opening
32 with an enlarged annulus 33 at opposite ends of same
and with a journal bearing surface 34 received therein.
Liner 30 further defines at least one, and preferably two
or more, axial oil grooves 35 that extend along bore
opening 32 with a dam or restrictor means 36 located at
opposite ends of same, defining a restricted orifice 37
thereat. Oil grooves 35 also preferably define a pair of
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3..~
openings 38 adjacent opposite ends of same for supply o~
oil into groove 35 from an external oil clrculation
system.
Upper liner portion 30A defines at least one
groove 40 therein for receipk of an oil ring 50 (see
Figure 1). Two such grooves are shown in Figure 3 with
undercut areas 41 extending along opposite sides of same.
As can be seen in Figure 1, oil ring 50 is received within
groove 40 and rests atop a shaft S extending through bore
opening 32. A lower portion of ring 50 resides within oil
reservoir 12 preferably with about 15 percent of its
diameter submerged in oil. As shaft S rotates, ring 50
rotates therewith, lifting oil from reservoir 12 and
depositing same atop shaft S. Oil ~rom shaft S then
optimally forms a lubricating film for proper lubrication
between shaft S and bearing assembly 10.
It has been determined that oil ring bearing
lubricant systems do not supply the total lubricant
requirements for full film lubrication, whereby with any
oil ring system, a degree of lubricant starvation condition
exists. While improvement may be achieved by particular
design of an oil ring to enhance oil delivery, further
improvement may be realized by use of an axial oil groove
35 as mentioned above. With an oil ring 50 delivering oil
to shaft S, the undercut areas 41 ad~acent grooves 40 aid
in directing oil from ring 50 into axial oil groove 35.
Normally an oil film around a rotating shaft
tends to diminish in thickness from a primary point of
formation outwardly towards opposite ends of same. It is
therefore highly desirable to attempt to maintain a uniform
thickness across the entire width of the film. Further,
when thrust bearings are utilized on opposite ends of a
journal bearing and thus adjacent the conventional
feathered film areas, inadequate lubrication of the thrust
surface can result. Oil groove 35 with its restricted
orifice 37 assists in maintaining a ready supply of oil
across the intended film width as opposed to conventional
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~ ~ ~3 ~ ~ ~ 3 ~ 3
structures whereby an improved film results. Such
construction further improves the supply of oil to a thrust
bearing, particularly where additional oil is supplied to
groove 35 through openings 38.
A general flow of oil from groove 35 and
restricts orifice 37 is indicated by the arrows in
Figure 3.
It has been determined that ~or operation under
both starved and flooded conditions, as will be described
in more detail hereinafter, that the ratio of the length
(L) of bore 32 to diameter (D) of bore 32 as illustrated in
Figure 3 should range from about 0.87 to about 0.92. In
like fashion, to minimize the feathering of the oil film
outwardly from a central location from groove 35, oil inlet
feeder openings 38 are preferably specifically located
along groove 35. Making reference to Figure 3, Zo
indicates the distance between center lines through
openings 38. It has been determined that oil feeder
openings 38 should be located along oil groove 35 to
achieve a ratio of Zo/L of from about 0.76 to about 0.87,
again to afford a bearing capable of operating under both
starved and flooded conditions without excess power loss.
Further, referring to Figure 5, it is seen that lines drawn
from e~tremities of groove 35 to a point of intersection at
a center line axial to the bore of liner 30 form an angle -
0 while similar lines drawn from extremities of orifice 37
form an angle a ~ . For operation under starved and flooded
conditions, angle O is preferably about 30# while angle 0'
is preferably about 20.
Oil groove 35 as noted above, further feeds a
thrust bearing 60, if same is employed. When a fixed
bearing arrangement is desired, a thrust bearing 60 (see
Figure 4) is secured to opposite sides of liner 30.
Additionally, a thrust runner 80 illustrated in phantom in
Figures 1 and 3 is located adjacent bearings 60 secured to
shaft S within space 17, thus affording a thrust load
capability for the bearing assembly.
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3~
Referring to Figures 4, 4A, 7, and 8 it may be
seen that a thrust bearing generally 60 may be provided
with the bearing assembly o~ the present invention. Thrust
bearing 60 is located on a base 61 defining a shaft
receiving bore opening 62 therethrough. Located about bore
opening is an annular oil groove 63. Located about bore
opening 62 and annular groove 63 are a plurality of thrust
pads 64 which include a flat land portion 65 and composite
tapered or scalloped portions 66 at opposite ends of same.
As may be specifically seen in Figure 4A, a thrust pad 64
includes the area between lines X/ X, thus incorporating
two scalloped areas 66. Scalloped areas 66 of adjacent
thrust pads 64 have a radially extending oil groove 67
therebetween that communicates with an oil bleeder 68 at an
outer end of same. Radial grooves 67 and oil bleeder 68
extend from annular oil groove 63 through bearing surface
60.
In operation, with shaft S rotating, oil is
supplied to axial oil groove 35 of liner 30 via one or more
oil rings 50 or oil rings 50 and an external oil
circulation system (not shown) that is in communication
with axial groove oil inle~ feeder openings 38. Oil then
escapes groove 35, forming an oil film on shaft S for
lubrication of same.
As shaft S rotates, oil is provided to thrust
bearings 60 by passing from orifices 37 into annular groove
63 about bore opening 62 and then into thrust pad scalloped
areas 66 and oil grooves 67. With a thrust runner 80
adjacent thrust bearing 60, lubricant located in radial
grooves 67, feeds scalloped areas 66 and is generally
confined therein for improved bearing lubrication. In
fact, excess oil may escape from thrust pads 64 only
through bleeders 68 after which it returns to reservoir 12
of housing 10. Composite tapered portions 66 of thrust
pads 64 are symmetrical, leading to reduced costs of
production, and also a capability of bidirectional
rotational operation.
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,3.h~,3~.3
A most preferred embodiment of thrust bearing 60
may be defined according to the following characteristics
as best described in Fiyures 3, 7, and 8. Scalloped areas
66 as mentioned above include a compound taper in both
radial and circumferential directions. In Figure 7, the
dimension S represents the distance between an inner radius
of scalloped areas 66 and the central point of location of
a cutting tool used to produce areas 66. l'he distance S
may, in fact, be a positive number (in a direction
outwardly of area 66) or a negative number in an opposite
direction. Preferably, the value of S ranges from about
0.03 to about -0.04 inch. Also illustrated in Figure 7 is
a dimension LK indicative of the length of oil bleeder 68
which preferably ranges from about 0.05 to about 0.10 inch.
Figure 8 illustrates the angle alpha associated with
scalloped areas 66 as well as the depth XK of scalloped
area 66 down to the top of radial grooves 67. Such is
measured along a line Ac coincident with a deepest portion
of scalloped areas 66. Preferably angle alpha should be
about 0.5 degrees, while XK should range from about 0.005
to about 0.010 inch. A further important characteristic of
bearing of bearing 60 is the ratio of the radial length of
the thrust surface (L) to the length of an arcuate line B
extending across the middle of the thrust pad. Preferably
L/B ranges from about 0.65 to about 1.04. The thrust
bearing characteristics noted above are most preferred
characteristics for specific bearings having a number of
thrust pads ranging from about 8 to about 12.
A further embodiment of a thrust bearing
arrangement is illustrated in Figures 9, lO, 11 and 12. In
Figure 9, a bearing liner 150 is illustrated which is of
like general construction and has the same preferred
characteristics as liner 50 of Figure 3 except for a
plurality of lubricant ports generally 155, received around
opposite sides of same. Lubricant ports 155 include a
first circumferential leg 156 that communicates with a
groove 151 extending around the periphery of liner 150 and
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an axially extending leg 157 that extends in~o
communication with a thrust bearing generally 160.
Lubricant grooves 151 may also communicate with oil
delivery openings 38 of axial oil groove 135.
As illustrated in Fiyure 10, thrust bearing 160
is located on a base 161 and has a plurality of thrust pads
165 located around an inner annular oil groove 164,
extending radially outwardly therefrom. Each thrust pad
165 includes a flat land portion 166 and a tapered land
portion 167. A radially extending oil groove 168 is
located between thrust pads 165 and extends from annular
groove 164 to an oil bleeder 169 which terminates at an
outer annular oil groove 170. Radial oil grooves 168 also
define an oil supply opening 167' in a bottom of same, (See
Figure ll) which communicate with axial legs 157 of the oil
delivery ports 155 whereby oil may be positively supplied
to thrust bearing 160. ~iner 150 and thrust bearings 160
are more precise than the liner-thrust bearing arrangement
of Figures 3 and 4, and are capable of carrying more thrust
load. The liner-thrust bearing arrangement of Figures 9,
10, 11 and '12 are unidirectional bearings intended for
premium performance situations, though reverse rotation is
possible with lesser performance.
Thrust bearing 160 also possesses
characteristics which are preferably controlled within
certain limits for improved bearing operations under both
starved and flooded conditions. Figures 11 and 12
illustrate such characteristics. The ratio of the radial
length L of the thrust bearing 160 to the radius R3
measured from a central point of the bearing to an outer
edge of the thrust surface should range from about 0.253 to
about 0.325 inch. The base angle B of the thrust pads 165
is pre~erably about 27 degrees. B represents the
circumferential distance at about the middle of the thrust
pad determined by the formula R3 - R2 where R3 is
--2
described above and where R2 is the radius of the inner
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3 ~33
circumference of the thrust surface. The ratio of the
radial length L of the thrust pad to the circumferential
length B of the pad preferably ranges from ~bout 0.61 to
about 1 and most preferably is approximately 0~8. The
ratio of the circumferential length of the tapered portion
~1 of pad 165 to the total length B of pad 165 preferably
ranges from about 0.72 to about 0.75. The ratio L/B as
defined with respect to thrust bearing 60 preferably ranges
from about 0.66 to about 0.80 for thrus~ bearing 160.
Also, the ratio .of the distance between the thrust bearing
160 and the thrust runner 180 at the flat land 166 (h1) to
the deepest tapered portion 167 (h2) should be about 2.45.
In arriving at the bearing characteristics noted
above as preferable for bearinys according to the
presention to properly operate under both starved and
flooded conditions, testing was conducted. Two test
bearing assemblies were mounted on a shaft flexibly coupled
to a 100 horsepower vari-drive motor. One of the bearing
assemblies included a thrust face while the other was a
free bearing. The load was applied via a hydrostatic
bearing between the journal bearings. The test bearings
were also constructed for oil ring and external
lubrication. External oil supply was provided via the
holes described herein in the axial oil grooves. Bearing
characteristics as identified herein, were varied to
ascertain ranges that were operable with minimum power
losses under both starved and flooded conditions.
Appropriate pressure taps, thermocouples and eccentric
probes were installed. Lubricant used was SAE 10, SAE 20
and SAE 30 oil. Test results indicated the characteristics
and ranges set forth in Table I to be important for the
bearing liner and for thrust bearing 60. Table II
represents the characteristics for thrust bearing 160.
It will be understood, of course, that while the
form of the invention herein shown and described
constitutes a preferred embodiment of the invention, it is
not intended to illustrate all possible forms of the
- 14 -
invention. It will also be understood that the words used
are words of description rather than of limitation and that
various changes may be made without departing from the
spirit and scope of the invention herein disclosed.
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