Tilting pad thrust bearing with optimized tilt axis location

A tilting pad thrust bearing in which the sector-shaped pads are supported on individual disks. A runner runs on top of the pads. The disks have a spherical element projecting from their bottoms. The elements make tangential contact with a stationary support ring and the pads and disks tilt about an axis on which the point of tangency lies. The tilt axis of each pad is located at 70% to 80% of the length of an arc extending from the leading edge to the trailing edge of the pad midway between the radially outermost and innermost edge of the pad.

BACKGROUND OF THE INVENTION 
Tilting pad thrust bearings contain a plurality of bearing pads which are 
each supported on a disk for tilting in a retainer ring on a hardened 
spherical pivot button which extends from the bottom of the disk. The pads 
are flooded with circulating lubricant. A rotating thrust collar or runner 
bears on the top surfaces of the pads and rides on an oil film on the top 
surface of the pads. The bearing pads are usually made of steel that is 
faced with a low friction material such as babbitt. 
As the runner rotates, it shears the oil film that separates the runner 
from the bearing pad faces and generates heat at the surface of the pads. 
This results in a temperature differential across each pad and through 
each pad. The top surface is the hottest so it assumes a slightly convex 
shape. The convexity or downward bending of the leading and trailing edges 
of the pads distorts the oil film on the bearing surface and influences 
its load-bearing capacity. The load on the pad causes a further downward 
bending of the pad about its pivot point and this also affects load 
bearing capacity, but the bending due to temperature differences is 
normally several times greater than that due to pressure. 
The term "leading edge" as used herein is the edge of a bearing pad that 
would be traversed first by any line on the runner moving in the direction 
of rotation or translation over the series of circularly arranged or 
linearly arranged bearing pads. The "trailing edge" then is the edge of 
the pad over which said point or line is second to pass in the direction 
of rotation or translation of the runner. 
Because of hydrodynamic forces generated in the liquid lubricant, it has 
been found desirable to mount the pads for tilting on a point of contact 
between a spherical surface and a planar surface. The tilting action 
results in the maximum lubricant film thickness (h.sub.1) developing, of 
course, at the leading edge of the pads and a minimum film thickness 
(h.sub.2) developing at the trailing edge of the pads. In the design of 
tilting pad bearings, it is an objective, as in the case of the present 
invention, to reduce the film pressure, maximize film thickness and 
minimize the temperature of the lubricant film. The downward deflection or 
convexity developed by the top working surface of the bearing pads 
distorts the lubricating oil film between the bearing surface of the pad 
and the runner and causes, in general, a significant decrease in 
load-bearing capacity. In actual practice, the deflection for a 
babbitt-faced steel bearing pad is a few thousandths of an inch, generally 
varying with bearing size. This is a small amount but it affects the load 
bearing capacity of the bearing significantly. 
Most tilting pad bearing assemblies adapted for use with a rotating runner 
have a plurality of sector or pie-shaped pads arranged in a circle. The 
theory of hydrodynamic film lubricated tilting pad thrust bearings is well 
known to those involved in designing and using large thrust bearings such 
as for hydraulic turbines and the like. The theory that has been generally 
accepted as valid indicates that maximum load capacity results when the 
pad pivot location is offset circumferentially in the direction of runner 
rotation to an optimum position approximately 0.6 of the length of the pad 
from its leading edge when the upper bearing surface of the pad is flat 
and the runner rotates in a single direction. In applications where the 
runner is rotationally reversible, the pivot point must necessarily be on 
a line that is centered or midway between the leading and trailing edges 
of the bearing pads. Theoretical analysis of the hydrodynamic properties 
of bearing pads that are supported centrally for bidirectional rotation 
and flat bearing surfaces indicates that an oil film would be developed 
that has no load bearing capacity at all. Theory and reality do not agree 
in this case. In actual practice the flat bearings develop some distortion 
due to heat and load which results in a load bearing capacity by the 
lubricant being developed. For unidirectional rotatable runners those 
skilled in the art have accepted establishing the pivot point for the 
tilting pads at 60% of the distance between their leading and trailing 
edges. In accordance with the present invention, however, much to the 
surprise of those who are involved in the design of heavy tilting pad 
bearings, it has been demonstrated recently by theoretical analysis and 
practical tests that when distortion is taken into account, the pivot 
point should be downstream from the leading edge by substantially more 
than 60% of the width of the bearing pad in accordance with the invention 
disclosed herein. This has been found to produce a thicker than heretofore 
obtained lubricant film near the trailing edge which is tantamount of 
saying that the bearing will have a higher load capacity. 
A paper published by the United States Navy Department reports on tests 
made on two different types of thrust bearings one of which was a tilting 
pad bearing and the other was another type of thrust bearing. The 
performance characteristics were compared. The data show that performance 
of the tilting pad bearing improved as the point on which the pads pivot 
increased from 50% to 60% to 70% of the distance from the leading to the 
trailing edge of the pads. An optimized pivot point was not determined nor 
suggested. The traditional 60% pivot point was not positively challenged. 
The study was primarily for comparing two types of bearings. The 
publication is--Nathan T. Sides and Thomas L. Daugherty, "Performance 
Characteristics of Oil Lubricated Swing-Pad Thrust Bearings with Different 
Radii of Curvature"--Report No. DTNSRDC-80/122 (David Taylor Naval Ship 
Research and Development Center), published December, 1980. Government 
Accession No. AD-A093173. 
SUMMARY OF THE INVENTION 
An object of the present invention is to provide for reducing lubricant 
film temperatures, reducing lubricant pressures and increasing the 
lubricant film thickness in tilting pad bearings. How this general object 
is achieved and how other more specific objects are achieved will be 
evident in a more detailed description of a preferred embodiment of the 
invention which will be set forth in detail hereinafter. Briefly, the main 
feature of the invention is to prescribe the preferred range of 
percentages and the optimum percentage of the distance between the leading 
and trailing edges of the pad for the radially extending line on which the 
pivot should be located to obtain reduced film temperatures, reduced 
lubricant pressure and increased film thickness at the trailing edge of 
the pad. More particularly, the invention resides in locating the pad 
pivots in the range of 70% to 80% of the distance between the leading and 
trailing edges of the pads. The optimum point of pivot location is 
revealed to be at about 75% of the distance between the leading and 
trailing edges of the pad. 
In addition to providing evidence corroborating the efficacy of having the 
pivot point between 70% and 80% of pad width, a device is disclosed for 
making the tests that confirm the validity of the pivot location 
prescribed herein. 
A more detailed discussion of an illustrative development of the invention 
will now be set forth in reference to the drawings.

DESCRIPTION OF A PREFERRED EMBODIMENT 
A typical tilting pad bearing assembly is depicted in FIG. 1. It comprises 
a circular casing marked 10. There are a plurality of tiltable bearing 
pads 11 arranged in a circle in casing 10. Casing 10 is ordinarily mounted 
in a machine base such as would be the case in a hydraulic turbine. The 
load carrying shaft is marked 12 and it has a disk-like runner or slider 
13 fastened to it. A section of the runner is cut away to expose the top 
surfaces of the bearing pads 11. These top surfaces are, typically and for 
the purposes of the present invention, coated with a layer of babbitt 14 
which can be seen in FIG. 3. In FIG. 1, the assumption can be made that 
the runner 13 is rotating in the direction of the arrow 15. In such case, 
the edge 16 of the first exposed bearing pad 11 is treated as the leading 
edge and the other edge will then be the trailing edge which is marked 17. 
Attention is now invited to the diagrams in FIGS. 2 and 3. FIG. 2 is a plan 
or top view of a tilting pad 11 that is used in a rotating thrust bearing. 
FIG. 3 shows how the pad 11 is supported on a stationary base ring 18 
which would be fixed in casing 10 and the latter would be fastened in a 
machine. Pad 11 is supported, in this example, on the annular rim 19 of a 
tiltable disk 20. The disk is usually composed of hard steel. A spherical 
button 21 is formed integrally with and extends from the bottom of disk 
20. As is evident, the periphery of the sphere is tangent to the top 
surface of base ring 18. The diameter of tiltable disk 20 is equal to "d". 
Bearing pad 11 is sector shaped. Its average or central width is labeled 
"L". The width L is an arc extending from leading edge 16 to trailing edge 
17, which passes across a radially extending line on which the pivot point 
falls. The pivot point in this example is halfway between the radially 
outermost and innermost edges 22 and 23 of the pads. The thickness of the 
tilting pad 11 is designated by the letter "t". The load on one of the 
typical pads 11 is imposed through runner 13 as demonstrated in FIG. 1. 
Translation or rotation of the runner in FIG. 3 is indicated by the arrow 
marked 15. In use, the parts shown are immersed in lubricating oil, not 
shown, that may be forcibly circulated. When the runner 13 is in motion, 
all the pads 11 in the circular array of pads tilt and a load carrying 
film of oil develops, due to well known hydrodynamic principles, between 
the upper surface of babbitt layer 14 and the lower planar surface of 
runner 13. The minimum thickness of the film at or near the trailing edge 
17 of the pad is designated "h.sub.2 ". Because the pad will develop a 
slightly convex upper surface, called crowning, due to loading and 
heating, the minimum oil film thickness may occur inward of the trailing 
edge of the pad. An objective of the invention is to attain a film 
thickness h.sub.2 that is as thick as possible. The load carrying capacity 
of the bearing is governed, to a large extent, by the thickness, h.sub.2 
of the lubricant film. The thickness of the film at the leading edge 16 of 
pad 11 is designated "h.sub.1 ". The distance from the leading edge 16 of 
the pad 11 to the pivot point or point of tangency of the spherical button 
21 is designated by the letter "P". 
Up to the time the inventive concepts of the present invention were 
confirmed, conventional wisdom among designers of heavily loaded pad 
thrust bearings was to have the pivot point or radially directed line on 
which pivoting occurs located so that the distance P was equal to about 
60% of the pad length L. A surprise factor in the present invention is 
that the generally accepted practice of having P at 50% to 60% of L does 
not result in optimum bearing performance. As will be demonstrated 
subsequently, in accordance with the invention, desired performance 
characteristics of the bearing are improved at 70% to 80% and optimum when 
P is in the range at about 75% of L. Having a pivot location as prescribed 
by the invention results in an increased film thickness h.sub.2. This 
minimizes heat generated by shearing action on the lubricant which, in 
turn, results in reduced film temperatures and pressures. In accordance 
with the invention, the factors that maximize the load carrying capacity 
of tilting pad bearings when the pivot location P is 70% to 80% of L are 
synergistic or act in a positive feedback mode with each other. Reducing 
the pressure in the film in the area of contact when the pivot location is 
within the prescribed range results in increased film thickness and 
reduced film temperatures and vice versa. Reduced film pressures result 
from more effective use of the pad surface because of less crowning when 
pad temperatures are reduced. 
In a paper presented by Rightmire, D. K., et al, "An Experimental 
Investigation of a Tilting-pad, Compliant-Surface, Thrust Bearing," 
Journal of Lubrication Technology, Trans. ASME, presented Oct. 21-23, 
1975, the authors reported experiments with tilting pads whose support 
surfaces were coated with a resilient rubber-like compound instead of 
babbitt. They noted that moving the pivot point of the pads over a range 
of 55% to 85% of pad length peak oil film pressure increases. Calculations 
of the applicant herein predicted decreasing oil film pressure in going 
from 50% of L to about 65% to 70% and then an increase. This correlates 
with the temperature data obtained by applicant. They do not specifically 
relate pad pivot location to bearing performance. Because of the subtlety 
of the phenomena involved in high load thrust bearings, no deductions can 
safely be made that their data would be applicable to babbitt coated 
bearing pads. The geometry of the bearing pads and test arrangement of 
Rightmire et al is also markedly different from what is described by the 
inventor of this application. 
The data supporting the unexpected results reported herein were obtained 
with the testing device depicted in the FIG. 4 exploded view. The device 
is composed of two semi-circular channel shaped sections having inside and 
outside rims 26 and 27 that define a circular channel whose flat bottom or 
base surface is marked 28. Two split rings 29 and 30 are adapted for being 
set in the channel. They have a plurality of holes such as the one marked 
31 for receiving tilting disks 20 of various diameters. When the test 
device in FIG. 4 is in use the two channel sections are locked together to 
form a circular channel by means of socket-headed screws that fit into 
holes such as the one marked 32. During any given test, only one identical 
pair of tilting disks 20 are used. One disk in the pair is diametrically 
opposite of the other. In FIG. 4 the disk that is marked 20 and is shown 
inverted exposes the supporting spherical button 21. Its counterpart at 
the opposite side of the channel ring is shown in upright position as it 
would be during a test. Two tilting pad sectors 11 are used to make a test 
run. One is in place in the base ring in FIG. 4 and the other, which would 
be diametrically opposite, has been removed. Typically, the disks 20 fit 
in mating holes 31 in the sectors 30. The sectors 30 have notches 33 on 
the rims. These notches lie adjacent outside rim 27 of the base. There is 
a slot in the side rim 27 in which a straight key 34 can be inserted to 
engage with any of the notches 33. This permits locking the sector 30 in 
one position during a test. When the key 33 is removed, the sectors can be 
rotated in the channel to position the pair of disks 20 being used 
relative to the bearing pads 11. In other words, setting the angular 
position of notched sectors 30 rotationally causes the pivot point or 
contact point of the spherical button to be located at any selected 
distance P relative to the length L of the pad 11. In FIG. 4 there are 
three bearing pads shown adjacent the testing device to suggest that pads 
11', 11" and 11'" of different thicknesses were used to corroborate that 
optimum pivot location prescribed by the present invention was independent 
of a practical range of bearing pad 11 thicknesses "t". The runner for 
imposing a load on the upper surfaces of test pads 11 has been omitted 
from FIG. 4. 
Some of the conditions that were established for the tests are: to keep 
support disks 20 always fully within the trailing and leading edge of the 
pads; not to exceed runner speeds of 4000 rpm to avoid non-laminar film 
conditions; limit maximum load to 600 psi to avoid failing any bearing and 
preserving the ability to get valid data; limiting maximum pad temperature 
to 121.degree. C. (250.degree. F.); the lubricant would be turbine oil 
type ISO VG32; oil inlet temperature would be held between 48.6.degree. C. 
(119.5.degree. F.) to 49.2.degree. C. (120.5.degree. F.); and oil would be 
circulated at about 57 liters per minute. The criteria for bearing 
performance would be inferred from heating effects. Accordingly, all of 
the test tilting pads 11 had several thermal sensitive detectors such as 
thermocouples, not visible mounted in contact with the babbitt facing at 
various locations. The electric leads from these sensors are shown as 
coming out of opposite ends of the pads in FIG. 4 and are marked 35 and 
36. 
The pads 11 are held against rotation in the channel shaped base ring by 
means of pins 37 and 38 that extend radially outwardly and inwardly and 
nest in notches in rim 27. The notches are deep enough so that the pins 37 
do not rest on their bottoms in which case the pads tilt exclusively on 
the circular support disks 20 and their pivoting spherical buttons 21. 
The charts presented as FIGS. 5-12 present some of the test data that 
corroborates the concepts of the present inventor that the pivot location 
P should be in the range of 0.70 to 0.80 of L. In these charts, there are 
different independent variables. 
In FIG. 5 and in the other charts of FIGS. 6-12 too, the maximum measured 
temperature of the tilting pads during a particular test run is plotted 
against pad pivot location. Maximum temperature, wherever detected in the 
pad, is indicative of bearing performance. In FIG. 5, curves 1, 2 and 3 
are for different bearing pad thicknesses. Curve 1 is based on a pad 11 
thickness of 0.50 inch (12.7 mm). Curve 2 is based on a pad of 0.75 inch 
(19.05 mm) and curve 3 is based on a pad 1.00 inch (25.4 mm) thick. The 
load on the pads was 600 psi, the button disk diameter was 1.50 inches 
(38.1 mm) and rotational speed of the runner was 4000 rpm. Note that 
maximum measured temperature spot in the bearing pads drops rapidly for 
all pad thicknesses and stays low in a pivot location range of 0.70 to 
0.80 of L. Minimum pad temperature occurs for all pad thicknesses where 
the ratio P/L is about 0.75. This is the preferred location for the pivot 
with any combination of test conditions. The range between the limits of 
0.70 to 0.80 of L is designated as preferred because in all cases maximum 
measured temperature of the pad at 0.80 L had increased again to the same 
temperature from which the pad dropped after 0.70 L. 
The FIG. 6 chart has all of the conditions of the FIG. 5 test held constant 
except that the button disk 20 in FIG. 6 has an outside diameter of 0.75 
inch (19.05 mm) in FIG. 6 as opposed to 1.50 inch (38.1 mm) in FIG. 5. 
Here again, the temperature in the zone of maximum measured temperature of 
the bearing pad begins to fall off sharply where the pivot location ratio 
of P/L is about 0.70 and as in the previous test, maximum temperature did 
not rise to where it was at about 0.70 L until a pivot location of 0.8 L 
is reached. In FIG. 6, curve 1 is based on use of a pad 11 having a 
thickness of 0.50 inch (12.7 mm). Pad 2 had a thickness of 0.75 inch (25.4 
mm) and curve 3 relates to a pad thickness of 1.0 inch (25.4 mm). 
In FIG. 7, pads having the same thickness, 0.75 inch (19.05 mm), as in the 
FIG. 6 test are used. The button disk diameter is also the same for the 
two tests and so is the rpm. In the FIG. 7 test different load pressures 
were applied. The curve marked 1 is based on a pressure of 200 psi. Curves 
2 and 3 are based on pressures or loads of 400 psi and 600 psi, 
respectively. Note that in all cases, performance, as represented by low 
maximum measured temperature in the pads, improves substantially where the 
pivot location P/L is in the range of 0.70 to 0.80 of L and the lowest 
maximum temperature occurred at about where P/L is 0.75. 
In FIG. 8, the load on the pads was held at 300 psi, pad thickness "t" was 
0.75 inch and the diameter of the support disk 20 was 0.75 inch. Here the 
pressure was held constant, the pad thickness and button diameter were the 
same as in the FIG. 7 charts. Rotational speed in rpm is the variable in 
the FIG. 8 test. The curve marked 1 is based on rotating the runner at 
1000 rpm. Curves 2, 3 and 4 are for increasing speeds of 2000, 3000 and 
4000 rpm. Note again that maximum measured temperature on the pad was 
lowest when P/L is 0.75. Again, the maximum temperature at P/L at 0.8 is 
about the same as it was at 0.7 and the lowest temperature occurred when 
P/L was 0.75. 
In FIG. 9, the tilting disk 20 is 0.75 inch (19.05 mm) thick and the load 
force is 300 psi as was true in the FIG. 8 test. In the FIG. 9 test 
rotational speed is 4000 rpm as it was in curve 4 of the FIG. 8 test. In 
FIG. 9, pad thickness is the independent variable. In the curve marked 1, 
pad thickness is 0.50 inch (12.7 mm). In curves 2 and 3, pad thickness is 
0.75 inch (19.05 mm) and 1.00 inch (25.4 mm). As was demonstrated in the 
preceding graphs, maximum measured temperature on the tilting pads 11 was 
lowest in the range of 0.70 to 0.80 for P/L. Under the test conditions of 
the FIG. 9 chart, the benefit of having the P/L in the range of 0.70 to 
0.80 is achieved regardless of pad thickness. As in other cases, the 
maximum measured pad temperature at P/L of 0.70 and 0.80 is the same. 
FIG. 10 is a curve of maximum pad temperature versus pivot position where 
the runner is rotating at 2000 rpm. During this test, a tilting support 
disk 20 having an outside diameter "d" of 1.5 inch (38.1 mm) was used. The 
radial dimension of the sector shaped tilting pad 11 is identified as "b". 
The independent variable in FIG. 10 is the ratio t/b versus maximum 
temperature. Here again, it will be seen that the maximum pad temperature 
drops substantially after the pivot position as a percent of pad arc, L, 
exceeds 65%. The best range is from 0.70 to 0.80 of L. The temperature at 
0.70 of pad arc is just about equal to the temperatures at 0.80 of pad 
arc. The temperature of the pad at the hottest spot is the lowest when the 
pivot position is about 75% of the pad arc length L. In the FIG. 10 test 
run, curve 1 is based on having the ratio t/b be the variable. The 
dimension "b" is actually fixed for any given test run and it is the 
thickness "t" that is actually varied by substituting different pads for 
consecutive test runs. By way of example, in FIG. 10, curve 1 is for a t/b 
equal to 0.13. Curves 2 and 3 are for t/b equal to 0.20 and 0.27. 
In the FIG. 11 test runs, the ratio of the diameter "d" of the support disk 
20 to the pad radius "b" or d/b is the independent variable. For the test 
bearing pads, the dimension "b" is 95 mm. By way of example, d/b for curve 
1 is 0.2. d/b for curves 2 and 3 are 0.4 and 0.6, respectively. In this 
test run, the maximum pad temperature where the pivot position as a 
percent of pad arc is at about 0.70 is equal to the maximum pad 
temperature where the pivot position of the pad arc is equal to 0.80. As 
in other of the test runs, the lowest maximum pad temperature occurs where 
the pivot position is about 75% of the pad arc. 
With the hottest spot on the pads being at the lowest temperature where the 
pivot is located between 70% and 80% of pad arc length, it follows that 
the film thickness at the trailing edge must necessarily be thicker than 
it would otherwise be at different pivot locations. During a test run, the 
load and speed conditions desired were set and maintained for a minimum of 
ten minutes. A reading was taken when the oil supply temperature was 
within the limits of 48.6.degree. C. to 49.2.degree. C. If not, 
adjustments were made to bring it within limits and data was then 
recorded. The data collected consisted of temperatures from 8 
thermocouples embedded in each of the two diametrically opposite test pads 
plus the oil inlet and drain, plus the oil flow, bearing load and shaft 
speed. The data acquisition program averaged the two temperatures at a 
corresponding specific location on each of the two pads and also recorded 
the difference. The difference was typically less than 3.degree. C. and 
often less than 1.degree. C. This gave confidence in the load equalization 
between pads and in the consistency of the thermocouple installations. The 
thermocouple locations were chosen to cover an area where the highest pad 
temperatures have normally been found, both by theoretical analysis and 
tests. It is evident that the effect of the pivot position was clear and 
consistent. The lowest temperatures were found in the pads during tests 
with the pivot at 75% of pad arc. It necessarily follows that reduced unit 
pressure in the lubricant film and increased film thickness result from 
reducing film temperatures. 
An alternative implementation of the invention will now be described in 
reference to FIGS. 12, 13 and 16. 
FIG. 12 is a bottom plan view of an alternative form of a bearing pad. 
As shown in FIG. 12, the bearing pad 50 is comprised of a sector of a ring. 
The pad has a tilting button insert 51 embedded in it. A section through 
the button 51 is shown in FIG. 13 where it is evident that the button is 
fitted snuggly into a recess 52 and makes good contact with the bearing 
pad body 50. As in the FIG. 3 embodiment, the button has a convex 
projection 53 which is curved so as to make substantially a point contact 
at 54 with stationary base ring 55 of the bearing assembly. The face of 
the pad 50 has a layer 56 of babbitt bonded to it which interfaces with 
the sliding or rotating runner 57 which would be mounted to a shaft such 
as the one marked 12 in FIG. 1. The hydrodynamic phenomena discussed 
relative to the FIG. 3 embodiment applies to the FIGS. 12 and 13 
embodiment as well. 
Bearing pads having the elements of the pads depicted in FIGS. 12 and 13 
have been used before the present invention was made. However, in prior 
usage, the contact point 54 of the button 51 was located at about 60% of L 
where L is the arc length between the leading edge 58 of the bearing pad 
and the trailing edge 59 as illustrated in FIG. 12. In accordance with the 
invention, the bearing pivot point 54 is between 70% and 80% of L and, 
most desirably, at 75% of L. Tests of two variations of the pad in FIG. 12 
were made to confirm that having the pivot point 54 between 70% and 80% of 
L to obtain the lowest operation temperatures with a given load on the 
bearing held true for both variations. In the one variant, the pad 50 was 
composed of steel and in the other variant pad 50 was composed of an alloy 
of copper. In particular, a chromium/copper alloy was used. The FIG. 16 
graph reveals the test results. The two curves are plots of bearing pivot 
location versus bearing pad temperature during test runs with a typical 
load on the bearing such as 600 psi. Curve A demonstrates the relation 
between pivot location and bearing pad temperature where the pad is 
comprised of steel and the insert 51 was comprised of hardened steel. Note 
in curve A, in accordance with the invention, the bearing ran at its 
minimum temperature when the pivot location was at 75% of L and that at 
80% of L pad temperature had risen to about the same temperature 
prevailing in the bearing when it was run at the same load and pivot 
location was at 70% of L, thereby establishing that the permissible range 
of the pivot point would be between 70% and 80% of L to obtain the lowest 
operating temperatures. Incidentally, pivot point 54 in the embodiment of 
the pad illustrated in FIG. 12 is substantially equidistant between the 
radially inward edge 60 of the bearing pad and its radially outward edge 
61. 
Another basically well known type of bearing pad was tested to determine if 
locating the line of pivot of the pad at 70% to 80% of L was valid for 
this type of pad. In FIG. 14, the pad is identified generally by the 
numeral 70 and it has leading and trailing edges 71 and 72, respectively. 
In this design, there is a rib 73 formed integrally with the pad and 
projecting from the bottom of the pad. As is evident in FIG. 15, rib 73 
has a curved face which causes contact between the rib 73 of the pad and 
stationary base ring 74 to occur along a radially extending line 75. 
Contact line 75 is located at 75% of L or, in other words, 75% of the 
distance between the leading and trailing edges of the pads 71, and 72, 
respectively. In the FIGS. 14 and 15 embodiment, the bearing is coated 
with babbitt 76 and the runner 77 is running on the babbitt surface with 
an oil film developed between the babbitt and runner due to the tilt of 
the pad on line 75. 
Tests were made with two variants of the FIG. 14 embodiment. In one series 
of tests, the bearing pad 70 and rib 73 were integral and composed of 
relatively hard steel. In another variant, the bearing pad 70 and rib 73 
were composed of a chromium/copper alloy. The results of the tests are 
summarized in the FIG. 17 graphs. 
In FIG. 17, curve C applies to the all steel bearing pad 70 and curve D 
applies to the all copper alloy bearing pad. FIG. 17 shows that for 
bearing pads having the configuration of the pad in FIG. 14 and made of 
steel or copper alloy, the lowest bearing pad operating temperatures at a 
given load occurred when the tilting axis for line contact 75 was located 
in the range of 70% to 80% of L, the distance between leading and trailing 
edges 71 and 72, respectively. Moreover, FIG. 17 demonstrates again that, 
in accordance with the invention, the optimum location for the pivot axis 
of the pad is at 75% of L as inferred from the fact that that is the pivot 
line location which results in the bearing operating at the lowest 
temperature for a given load such as 600 psi.