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73-LubS-16
1).8 Soolety shall nol lie (IJ.Sponslble for slstefl\C!nts or opInIons In lloPllfs or 1'1 dls�'on at nteln'!ut!l of th" Soolety 0, ot Its DIvisions or Sections, or prInted In ·its publlca110ns,
Performance of Thrust Bearings at High
R. S. GREGORY Manager, Research and Development, Kingsbury, Inc., Philadelphia, Po.
Operating Speeds
Mem. ASME
As part of a continuing research program, a standard 101/2 in. dia thrust bearing, of the tilting-pad, self-equalizing type, was tested at shaft speeds up to 11,000 rpm and bearing loads ranging up to 400 psi. The bearing and lube oil system were instru mented to measure bearing performance under laminar and turbulent operating condi tions. The effects of varying the oil feed rate on bearing temperature and power loss are discussed in this paper. Some observations on the laminar to turbulent transition regiol� are included.
Significance
THE significant contribution of this paper is the publi
bearing power loss and bearing temperature is desc�ibed in de
cation of experimentally measured values of bearing power loss
tail.
and pad temperatures under variable load", speed, and oil flow.
contribution to underst.anding the phenomenon of turbulence in
This paper is intended solely to present new test data as a
The operating conditions range from laminar to turbulent, and
bearings.
information of this nature for this popular bearing-type has been
until the entire series of tests are complete.
Analysis and theoretical predictions . are postponed
Most, if not all experimental work in the field of bearing tur
heretofore unavailable to designers and analysts.
bulence has been performed on single element thrust bearings. It was felt that testing of a double thrust bearing would be more
Introduction
truly representative of actual machine applications.
This paper presents the initial results of a current research program investigating the performance of a standard
101;'
in.
Kingsbury double thrust bearing operating at shaft speeds in excess of ] 5,000 surface ft per min.
A petroleum-based, light
turbine oil with a viscosity of 150 SUS at lubricant for all test work.
100 deg F was used
as a
During the course of this experi
mental study, shaft speeds ranged from
4000
to
11,000
rpm, and
bearing loading was varied from "no load" to 400 psi, based upon
a bearing area of
55.1
sq in.
The tests were pelformed in a new
research and development facility recently constructed to in
The double
thrust bearing consists of a loaded, or "active," thrust bearing designed to absorb the thrust load imposed by the parent machine. On the other side of the shaft collar is the slack-side, or "inactive," thrust bearing which serves to carry any transient loads that possibly might develop in the other direction.
The two bearings
(loaded and slack) that comprise the double thrust bearing under going test are identical in design and size.
One of the single
element thrust bearings utilized to assemble the double thrust bearing is shown in Fig.
1.
It is a conventional design with stan
dard dimensions and centrally pivoted pads.
vestigate all aspects of high speed bearing performance. The results reported in "this paper are considered incomplete since additional testing is currently under way at still higher shaft speeds. future time.
These additional data will be published at some However, it was felt that information on the critical
laminar to turbulent transition region would be of particular value
if presented now.
Several interesting examples of bearing power
loss in the transitional region are discussed in this text.
More
over, the considerable effect of a variable oil supply rate on both
Contributed by the Lubrication Division for presentation at the Lubrication Symposium, E vanston . Ill., June 4-6, 1973, of THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS. l\-Iunuscript re ceived at ASME Headquarters, March 14, 1973. Paper No. 73LubS-16.
Copies will be available until March, 1974.
Fig. 1
101/2 in. standard, single element thrust bearing
1
Discussion on this paper will be accepted at ASME Headquarters until July 9, 1973
"
During the course of normal machine operation, the loaded
bearing plus the rotating seal rings and stationary oil control ring.
bearing absorbs the imposed thrust load and operates with a rela
Each thrust bearing is a standard 6-pad Kingsbury equalizing
tively thin film thickness, on t.he order of 0.001 in. or less.
Under
type, with a babbitt O.D. of 10 1/2 in. and an I.D. of 5 1/. in.
Pad
this condition, the slack side bearing operates with only the in
arc length is approximately 5 1 deg and the total bearing area is
ternally generated hydrodynamic load due to collar rotation, and
55.1 sq in.
experiences a large film thickness, equivalent to the hot end play
Load is applied to the test bearings by means of an external
of the bearing installation less the film thickness of the loaded
hydraulic system which transmits a force directly to the sliding
bearing.
For the test bearing, this slack side film thickness is
housing.
This load .applicator consists of a manually operated
Using one critical value of
pump, hydraulic cylinder and force multiplier arm, with an oil
Reynolds number (as discussed below) for the criterion of transi
reservoir, accumulator, pressure relief valve, solenoid valve, and
tion into the turbulent regime, it can easily be shown that the
pressure gage completing the system.
slack side bearing will encounter turbulence at a much lower
can best be understood by referring to Fig. 4.
shaft speed than a loaded bearing, owing to its thicker film thick
journal bearings are of the tilting pad type, 5.0 in. dia X 2 1/. in.
ness.
long, and serve only to support the test shaft.
on the order of 0.017 in. or more.
The mechanism of loading The instrumented A force input is
applied to the aft end of the sliding housing causing the housing to move forward a small amount within the limits of bearing end
Test Apparatus Mechanical Arrangement.
A schematic drawing of the test ma
chine is shown in Fig. 2.
A variable speed gas turbine with a
rated output of 1 100 horsepower is the prime mover. trollable test speed range is 4000 to 14,000 rpm.
The con
The turbine is
connected to the test shaft by means of a flexible coupling.
Two
identical bearing housings contain the bearing components under going tests.
The forward housing adjacent to the turbine is
firmly secured to the foundat.ion while the aft housing is re strained, but free to slide axially.
Each housing is equipped
with separate lube oil supply and drain lines.
Each housing
contains a journal bearing to support the test shaft as well as a 101/2 in. double thrust bearing.
In accordance with normal industry practice for high speed thrust bearings, an oil control ring surrounds the collar to divert the bearing discharge oil away from the collar periphery. construction of the oil control ring is shown in Fig. 3.
The
Teflon
oil seal rings are used to contain the oil in the thrust bearing area and prevent leakage either into the journal bearing cavity or out of the housing.
play.
This housing movement forces thl'Ust bearing "D" against
the test shaft collar, thus applying a load to the shaft.
Thrust
bearing "C" moves along with the housing since it is not re strained and it operates as an unloaded, or"slack side," bearing. The movement of the test shaft is limited by thrust bearing"A," located in the fixed housing.
As a result of the single applied
force, both thrust bearing "A" and "D" experience the same loading, while thrust bearings "B" and"C" remain unloaded. This duplicate arrangement permits a check of repeatability when two identical bearings are installed in the test rig.
Con
versely, the installation of two dissimilar thrust bearings permits a fair comparison to be made under identical operating conditions. In any event, the flexible coupling at the turbine end of the test shaft is relatively soft and prevents load transmittal to the turbine bearings, precluding any possibility of "sharing" the applied test load. Instrumentation,
The lube oil supply line to each of the six
bearings is equipped with a turbine flow meter, throttling valve,
This type of bearing arrangement is typical of
many actual machine designs.
All horsepower loss values re
ported in this paper include those for a double 101/. in. thrust
SLIDING HOUSING
JOURNAL 2
FLEXIBLE COUPLING
FORCE
INPUT:
LOADED BEARING
Fig.
2
Schematic of thrust bearing test apparalus
Fig.
4
Descriptian of method of applying thrust load
ROTATION ....
Fig.
2
3
Cross-sectional view of housing interior
Fig.
5
Thermocouple locations on thrust pad
Transactions 01 the AS M E
thermocouple and pressure
transducer.
All
drain
lines
600
are
similarly equipped with thcrmocouples and throttling valves, making it possible not only to determine the bearing power losses
400
by an energy balance method, but also to independently vary the oil flow conditions to each separate bearing.
300
Small thermocouples are imbedded in the babbitt metal of the
oc
The actual thermo
journal bearing and thrust bearing pads.
� 200
couple junction i� positioned within 1/" of an in. of the babbitt surface.
0 a.. w til a: 0
An alTay of nine thermocouples is installed in one pad
each of bearing "A" and bearing "j)" in order to determine the temperature gradient in both the radial and circumferential directions of an operating thrust bearing pad.
:;
en tOO
Fig. 5 shows
til 0 ...J
these thermocouple locations. Additionally, one pad in loaded bearing "A" is equipped with
� 60 0 a.. ·50
Electronic load cells are installed
Cl
in several of the bearing leveling plates to measure the actual
�
applied thrust load carried by each pad of the loaded bearing and
r
to give an indication of the degree of pad equalization.
40
/
/
/
./
/
�::/ %-
It,oOORPM
10,000RPM
9000RPM
8000 RPM
7000 RPM
'/