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DESIGN CONSIDERATIONS FOR THRUST BE ARING APPLIC ATIONS

by
James H. Ball
Applications Engineer

Orion Corporation

Grafton, Wisconsin

performance. Of primary interest is an example of a high speed

James H. Ball received his B.S.M.E. , application where controlled lubrication increased machine effi­
M.S.M.E. , a n d Ph.D. (Vibration Analysis ciency by minimizing parasitic losses. Three dimensional maps of
and Mechanical Design)from University of velocity, fluid mixing, temperature, and viscous heating in the
Wisconsin-Madison (1965, 1971, 1975). pocket result in two dimensional temperature fields at the pad
While at the University of Wisconsin, he entrance. Accurate determination of the pad entrance temperature
designed and analyzed a resonant vibratory distribution allows for calculation of the pad maximum tempera­
machine for marine work. This included tures or bearing capacity. The examples and results presented will
modelling viscous and coulomb damping. be of use to rotating equipment manufacturers and machinery users
Dr. Ball has been working in the bearing in applying thrust bearings to their machines.
industry for 21 years. This includes five
years as manager of product development INTRODUCTION
at Waukesha Bearing Corporation and five years as an analyst at The thrust bearing babbitt maximum temperature is used to
the A llis Chalmers Advanced Technology Center working on fluid determine capacity for many designs. This is determined by the
film bearings. Rolling element bearing work occurred at near fluid film temperature. The fluid film temperature increases
Torrington Bearing Company and at the Bucyrus-Erie Company. by the action of the runner and pad shearing the oil. Thus film tem­
Currently, Dr. Ball is responsible for thrust bearing design and perature is determined by adding an incremental temperature to the
applications at the Orion Corporation, a company that manufac­ pad entrance temperature, which is a function of the energy
tures tilting pad thrust and journal bearings, sleeve bearings, seal, generated in the fluid film between the pad entrance gap and the
and housed bearings systems for the rotating equipment industry loc al maximum film temperature. Thus the film temperature at the
for the last 40 years. pad entrance is a maj or factor in establishing bearing capacity.
Dr. Ball, a member of ASME, STLE, and the Vibration Institute, Historically the pad entrance temperatures have been a calculated
has authored 12 technical papers concerning bearings, seals, single value.
pumps, tribology, machinery dynamics, and vibrations. He has five Ettles [ 1 ] describes the pad entrance temperature in terms of a K
patents concerning fluidfilm bearings, seals, pumps, and vibration factor in Equation ( 1 ) .
control.

(1)
ABSTRACT
The hydrodynamic thrust bearing load capacity-speed envelope
Since th e runner temperature is not known, i t is usually taken as
is divided into three major regions by speed. Each region has its
the mean of the pad streaml ine temperature. The limits for K are
own set of limiting physical phenomenon that must be included in
between .45 and .75 from field or laboratory tests on many thrust
the analysis and presents a design challenge. For instance, the high
bearing types . Vohr [2] uses an energy balance to determine the pad
speed region has high lubricant shear rates possibly including tur­
e�trance temperature. The shear energy generated is equal to the
bulence that change pad thermal crowning and local babbitt
energy lost as side leakage, energy lost to the groove, pad, and
temperature and pressure, thus limiting the bearing performance in
runner. The bearing program iterates on the pad film properties
terms of load and speed. Various design tools are used to set the
until the entrance temperature satisfies.
bearing geometry such that the phenomenon limits are not
exceeded, and an appropriate design safety factor is present.
(2)
Several applications are presented that illustrate how the various
design modifications such as pivot type, pivot radial and circum­
Heshmat and Pincus [3] describe a mixing function, A, which is
ferential offset, pad spacing, and pad thickness may be varied to
the deviation in percent between an ideal energy balance and
design a thrust bearing for a specific region. Three dimensional
experimented data. The function accounts for groove heat genera­
finite element models for the fluid film generate film thickness,
tion in and far from the runner boundary layer, heat fluxes at the
pressure and temperature distributions, which when coupled with a
walls, three dimensional effects such as radial and axial recircula­
pad finite element model for thermal conduction, convection,
tion, and loss of some pad exit flow. Thus,
thermal deflection, and elastic deflection show thrust bearing pad
performance. Turbulence is included along with hot oil carryover
from the previous pad. Methods are shown that reduce vibratory (3)
motion by using the squeeze film damping attributes of double
thrust bearings.
The benefits of offset pivot, chrome copper pads, and controlled The mixing constant, A, is fit to experimental data which results
lubrication in a load-speed region are reviewed so that the user has a function in terms of mean velocity and supply oil temperature. If
an appreciation of how each feature is affecting his thrust bearing's recirculation is considered the equation becomes:
223
224 PROCEEDINGS OF THE TWENTY-FIFfH TURBOMACHINERY SYMPOSIUM

(Qs + Qr) Ts + Q2T2 limiting capacity. The various regions require alternate design
T (4) solutions to optimize bearing capacity. For some applications,
l = (Qs + Qr + Q2)(1 + X)
especially process fluid bearings, a film thickness to surface
Recirculation was observed in bearings with inlet to outlet film parameter ratio is a better parameter than film thickness to set
ratios greater than 2.1. Thirty seven percent recirculation occurred capacity limits. When tilt pad thrust bearings run unloaded, system
when the ratio was seventeen to one. Dmochowski [4], et al., used vibration can occur. Users can make several adjustments to reduce
a mixing equation to characterize conventional and pocket tilting the vibration. When bearing designers size a bearing, the analyses
pad journal bearings. The Heshmat and Pincus mixing function, A, are varied and, in some cases, answers can differ by more than 100
was set to zero so that the pad entrance temperature is: percent. Both users and machinery builders should be aware of the
differences so that a base level comparison can be made for
capacity at speed with a bearing manufacturer's different features
(5) or different bearing manufacturers.
Various bearing features can effect load capacity. Knowledge of
the features and their relationship with capacity would also help in
for conventional bearings. For the pocket bearings the pad entrance bearing selection. The pad ratio, pad spacing, pad crowning both
temperature is: elasticity and thermally, pad height, button diameter, and button
thickness effect capacity. The babbitt pressure and temperature is
related to bearing failure by creep. By making slight changes to the
(6) pivot position bearing temperature and film thickness deficiencies
can be corrected.

The difference being that for the pocket bearing, it is assumed CONSIDERATIONS IN TlLT PAD BEARING
that all the supply flow, Q8, is taken by the pad entrance gap; SELEC TION PROCEDURES
whereas, in the conventional bearing all of the pad exit flow, Q2, is
Applications
taken by the pad entrance gap to 90 percent of the entrance flow,
Q1. Dmochowski explains pad entrance temperature equations for There are numerous considerations to resolve when reviewing a
the two bearings in terms of a stratified temperature of fluid at the thrust bearing application (Table 1). The ratio of minimum film
pad entrance gap. His hypothesis for Equation ( 1 ) and Equation (2) thickness to surface finish should be reviewed for process fluid
was based on noting that experimentally the outlet temperature for bearings. The degree of misalignment present especially for non­
the conventional bearing did not change with variation in flow as it equalized bearings can load less than the full complement of pads.
did for the pocket bearing. Pad height will effect crowning, which influences load
With the increasing use of computational fluid dynamics (CFD) capacity. On large pads, can button recess diameter improve
the pad entrance temperature could be considered a two dimen­ crowning? What is the safety factor with babbitt creep? What
sional array that would include consideration of the hotter design features can be used to augment capacity in the same
boundary layer oil near the runner and between pad mixing. Pad envelope? How do offset pads, copper pads, and controlled inlet
entrance backflow and recirculation effects that occur because of lubrication affect thrust capacity. How can the maximum pad
the stationery pad and the wedge shape of the film could be temperature be reduced by changing the oil feed supply? What are
included. The thermal and elastic pad deflection at the front edge the parasite power losses for a flooded design? Would a thin
would cause the gap to vary with radius; the array could include section bearing be applicable here? What is the safety factor for
this effect. Prepad features such as wipers and directors could be stress at the pad pivot and link pivots? What is the shock capabili­
modelled from a fluids view and their effects on pad entrance tem­ ty of the bearing, time duration, film thickness, and babbitt
perature considered. The CFD tool appears to refine the pressure? For a double thrust bearing, how does the damping
determination of the pad entrance temperatures by modelling fluids reduce unloaded vibration? What is the film and mechanical
phenomenon in greater detail. This approach is used for pad stiffness of the bearing and how do they add to give an equivalent
entrance temperature calculations in a controlled inlet tilt pad bearing stiffness? What thermal crown exists on the pad in each of
thrust bearing design. the three regions of the load speed map? The pivot radial location
A controlled inlet thrust bearing that features a cool oil cavity for is important to minimize film temperature or maximize film
the pad entrance oil along with a hot oil carryover director and thickness. The circumferential location changes film temperature,
wiper is discussed. The oil feed system includes a hydraulically film thickness, and film pressure. The pad length (parallel to
actuated piston between the pad and retainer to transfer the oil to rotation) to width ratio can change film temperature by shortening
the cool oil cavity. The thrust bearing is thin because of the retainer the streamline length under shear in the film. Changing the pad
and equalizing link design. The design features augmented bearing entrance temperature affects the minimum film thickness. Various
capacity and efficiency. pad support designs can reduce pad elastic deflection. Is film
Design and analysis of equalized tilting pad thrust bearings is turbulence occurring? At what level is it, especially for additional
reviewed so that rotating equipment users can operate machines at power loss estimates? If the bearing is running starved, how did
higher reliability. Rotating equipment manufactures with an film pressures increase towards the trailing edge of the pad?
increased awareness of current tilt pad thrust bearing performance The thrust bearing will operate in one of three liquid lubrication
parameters and design features could lower machine bearing costs regions. The region location is shown in Figure 1 for boundary,
or increase machine safety factors. Users with an increased under­ mixed, and hydrodynamic modes. Boundary lubrication is defined
standing of how tilt pad thrust bearing performance parameters by metal to metal contact and wear. The PV value, or unit pressure
were obtained can review bearing reserve capacity or decide on times velocity value, with pressure and velocity limits establish
whether or not a higher capacity bearing is needed for a rerate. If a wear limits for this region. A material pair that runs hard on hard,
capacity increase is needed that could occur in the same envelope such as CVD diamond coated pads and runners, would be suitable
by adding bearing features to the same bearing envelope, that here. Hydrostatic lubrication is an alternative to wear designs at
would save housing and shaft modifications. these speeds.
The bearing load speed map should be used as a guide to Hydrodynamic lubrication is complete separation of the runner
determine which tilt pad bearing performance parameters are and bearing pad by a film of oil. Mixed lubrication is aptly titled
 DESIGN CONSIDERATIONS FOR THRUST BEARING APPLICATIONS 225

Table 1. Application Variables Affecting the Load/Speed Map. 1 Q3

REGION 1 REGION 2 REGION 3

UNIT I
5
FILM PRESSURE
NO. VARIABLES THICKNESS ILOADI TEMPERATURE
Vi
Q._ 4 L
APPLICATION:
-o
0
0
__J 3 L
1 ROUGHNESS, SURFACE FINISH 1

I\
�

i
2 FLATNESS/WAVINESS, SURFACE 1 c
FINISH
::::>
3 RUN IN 1

v
4 MATERIAL PAIR 1
5 LUBRICANT VISCOSITY 1 2 1.5
6 ALIGNMENT 2 1
7 STARVED OIL SUPPLY 1

1 Q2
5 1 Q2 2 3 4 5 1 Q3 2 3 4 5 1 04 2 3 4
DESIGN:
Speed, R P M
8 OIL SUPPLY, QUANTITY 2 1
.
9
10
OIL SUPPLY, TYPE
TURBULENCE
1
Figure 2. Thrust Bearing Load V s Speed Capacity Map.
1
11 NUMBER OF PADS/PAD SPACING 2 1
12 PAD RATIO 2 1
13 PIVOT FATIGUE 1
the load. For the intermediate region that is load limited, the energy
14 ELASTIC CROWNING 1 2
15 THERMAL CROWNING 1 and Reynolds equation need to be used in the solution, since tem­
16 CREEP LIMIT 1 2 perature effects viscosity greatly. The energy equation yields film
temperature so that the local viscosity can be used in the Reynolds
17 COPPER PADS 2 1
18 OFFSET PIVOT 1
equation for the load solution. When speeds are higher, turbulence
needs to be accounted for in the analysis. This is accomplished by
INTENSITY AFFECT UPON DESIGN adding Hirs constants to the Reynolds equation.
1 =HI 2=MEDIUM 3=LOW The low speed region that is governed by the minimum film
thickness has a positive slope. Thus, load is reduced or limited. The
nominal limit is 250 micro in. The limit is based on surface finish
produced by the different manufacturing process. Assume that the
pad will mirror the surface finish of the runner after run in and that
the film thickness should be greater than the irregularities in both
100 surfaces. Process surfaces are characterized by root mean square or
average profilometer roughness, predominant peak roughness, and
peak to trough maximum roughness. Tarasov, as noted by Ocvirk
::l. • BOUN DARY/M I X E D [5], found that processes had different ratios of predominate peaks
c •
0 10-1
MIXE D/ELASTO-HD roughness/RMS roughness ratios. Martin [6], used this value with
:;:; • ELASTO-HD/HD
a safety factor of two for minimum film thickness design criteria.
�
�
c
For example, if the runner was ground, the minimum film
•<£ thickness would be: 1 6 RMS X 4.5 X 2.0 X 2.0 288 micro in.
u
=

4=
Q;
The value varies with pad width or bearing size when based on pre­
10-2
0 dominant peek ratios where at a pad width of 0.5 in for a 4.0 in
u
bearing, the lim i t value is 1 80 micro in. This increases to a recom­
mended minimum film thickness of 550 micro in for a 25 in
bearing with a 6.25 in pad width. Using a higher viscosity lubricant
10-3 L-L-�ww���������������-7. in this region can raise the film to an acceptable value, since the
0 2 4 6 8 10 1 2 14 16 1 8 20

Film Parameter, "A

film is not running at high temperatures. Changing from a full pad
complement to a partial pad complement can reduce the pad
Figure 1. Lubrication Regime Map. entrance temperature, thus raising film thickness. For instance, a
complement of six vs twelve pads reduces the pad inlet tempera­
ture 1 7 percent.
and occurs between the two regions. Notice that the map is a plot The intermediate speed region is limited by bearing unit
of friction coefficient vs the film ratio, which is a ratio of liquid pressure or load. The load affects the limits of the components. For
film to pad and runner surface finish. The map is the same if the instance, as load increases, elastic crowning increases, which
abscissa is a duty parameter, J.L NIP, or the Sommerfeld number. decreases bearing capacity. As load increases, pivot radii need to be
When thrust bearings are applied in the hydrodynamic region, as increased or pivot fretting will occur. As load increases, the babbitt
speed increases the bearings are limited by different phenomenon creep limit is approached due to the material local pressure-tem­
(Figure 2). At low speeds adjacent to the mixed film region, film perature relationship. Misalignment can cause fewer pads to share
thickness limits the load. A central region exists where the bearing the load in nonequalized bearings, which can overload the working
is load limited by the bearing features. At high speeds, capacity is pads. A starved bearing causes a percentage of the pad to support
film and bearing material thermal limited. For petroleum lubri­ the load rather than the whole pad, thus changing the babbitt
cants, the oxidation limit is about 3oo·F and babbitt has a limit of pressure temperature profile with regard to the creep limit.
265"F. The high speed region loads need to be reduced due to local
The film exhibits primarily viscous forces in the low speed babbitt temperature where wiping starts at about 265.F. Thermal
region. For different wedge shapes, Reynolds equation can be crowning is dominant when compared with elastic crowning in this
solved for the pressure distribution the film generates to support region which reduces capacity. Increasing the pad spacing can
 226 PROCEEDINGS OF THE TWENTY-FIFI'H TURBOMACHINERY SYMPOSIUM

lower babbitt temperatures by dropping pad entrance temperature, • Isoviscous hydrodynamic film solution
which occurs because the hot oil carryover boundary layer oil
• Thermohydrodynamic film solution
has more time to mix with the feed oil. Turbulence is occurring
in this region that can retard babbitt temperatures for several • Thermohydrodynamic film solution with turbulence
thousand rpm before continuing to rise as a function of speed. • Thermohydrodynamic film solution with turbulence and pad
Power dissipation also increases with speed. At a Reynolds number thermal effects
of 550, turbulent power loss equals the laminar shear loss; at
1 1 00, the turbulence loss is double the shear loss. • Thermohydrodynamic film solution with turbulence and pad
High speed machinery with double thrust bearings running in an elastic effects
unloaded condition can have axial vibration due to a forced • Thermohydrodynamic film solution with pad thermal and elastic
response. The vibration occurs at the shaftlbearing natural effects
frequency and can have a double amplitude less than, or equal to,
Thermohydrodynamic film solution with pad and runner thermal
the bearing float or clearance. By reducing the clearance, the thrust
•

and elastic effects

bearing pads add damping to the single degree of freedom
shaft/bearing vibratory model. Some reduction in the clearance • CFD coupled thermohydrodynamic film solutions with pad and
adds sufficient damping, usually to reduce the vibration to accept­ runner thermal and elastic effects
able levels. A graph is presented in Figure 3 of thrust bearing
The various analyses can be one, two, or three dimensional. In
damping ratio vs clearance used to adjust the axial float to reduce
today's bearing market, most of the effects should be included in
vibration.
the analyses that is part of the tilt pad thrust bearing application
work.
. 105
A comparison was made of bearing film thickness for various
5 analyses types shown in Table 2 for a 10-112 in eight pad bearing
4 running at 1 0,000 rpm and loaded to 635 psi (35,000 pounds). The
3 \.
c program used to generate the performance parameters uses finite
z:::..
u element methods to solve a two dimensional Reynolds equation
Q)
"'
I
.0
104
1"'- with turbulence and groove mixing. The energy equation is solved
...J in three dimensions and is also finite element which allows for
,;, 5 cross film viscosity variation. The pad uses a three dimensional
c
4
·o.. ....... thermal conduction/convection finite element model and an elastic
E 3
0
...... finite element model. Heat flux from the film to the pad uses tem­
0
...... perature gradients from the energy equation solution. Varying
103
' geometry can occur in the pad. The runner model is axisymmetric
for both the thermal and elastic case. Turbulence is included in
5 Reynolds equation by Reichardt's Formula for the eddy viscosity.
0.000 0.002 0.004 0.006 0.008 0.0 10 0.01 2 0.0 14 0.01 6
The energy equation has modified local conductivity and viscosity
Film Thickness, Inch
for turbulence.
Notice that the energy equation is used to find maximum film or
Figure 3. Thrust Bearing Axial Damping.
babbitt temperature which is an indication of bearing capacity.
Crowning can effect capacity; therefore, pad deflection both
Bearing Analysis Procedures thermal and elastic should be added. This reduces the minimum
film thickness by 35 percent when compared to the isoviscous
There is an abundance of analysis procedures for tilt pad thrust case, Table 2, Case 1 vs 6. All these effects should probably be
bearings. Historically closed form equations with a constant included in a tilt pad thrust bearing application review. By allowing
viscosity were used a� with the one dimensional infinite slider the viscosity to vary with temperature, the minimum film thickness
equations for load capacity, friction,· and center of pressure. With decreases by 47 percent and the horsepower by 25 percent for a
computers came the ability to include two or three dimensional rigid pad when compared to the isoviscous case, Table 2, Case 1 vs
modelling, modelling additional phenomena with interactions, and 3. The film thickness increases 23 percent from the
reasonable time frames to solve a specific application. All analyses Reynolds/Energy solution when pad effects are added, Table 2,
solve the film Reynolds equation (A- 1 ) for the pad pressure distri­ Case 3 vs 6. The elastic pad causes film thickness to increase.
bution given the wedge geometry. The wedge geometry for the pad Brockett (7] found little change in the film thickness when consid­
is represented by a circumferential and radial tilt. This tilt is ering runner mechanical deformation for a typical runner
iterated until a pad pressure distribution is found _that will.support thickness, The addition of turbulence effects increased the
the load. Since the viscosity temperature gradient is high and maximum temperature and horsepower loss values. Thus, when
nonlinear, a more accurate solution to the hydrodynamic problem comparing, bearing manufacturer's data, the question of analysis
involves solving the energy equation with the Reynolds equation. type should be asked.
This increases the coinputational'requirement that is easily handled Typical pad plots are shown in Figures 4, 5, 6, 7, and 8. The
with today's personal computers and engineering work stations. plane view is shown in Figure 4 of the film pressure distribution.
Now other phenomena can be modelled to give results that The film thickness. with isoclines is shown in Figure 5. The three
correlate with experimental data. Pad elastic deflections due to the dimensional film .temperature distribution is shown in Figure 6.
load and thermal differential through the pad are important in The pad temperature distribution and . elastic deformation are
terms of crowning effects on load capacity. This can illso be. the shown in Figures 7 and 8.
case for the runner. Turbulence can be approximated with the Hirs
coefficient which is incorporated into the Reynolds equation. Design
The various analyses are summarized below:
Pad design determines the tilting pad bearing capacity. The pad
• Isoviscous approximation by closed form solutions for film, i.e., crowning ratio and babbitt strength are criteria that are considered
short or long bearing theory along . with minimum film thickness. The maximum bearing
 DESIGN CONSIDERATIONS FOR THRUST BEARING APPLICATIONS 227

Table 2. Thrust Bearing Performance by Analysis Type.

1-
() 1-
w ()
1- u. w
2 () 1- u. u.
0 w () w u.
2 u. w ...1 w
i=
< 0 u. u. < ()
::J i= w u. :a:
w i=
a < ...1 0::
�
w ::J w () w
U) a
w
()
2 �
0::
i= :r::
t- w
c
...1 > :s
::J
w
:r:: � 0::
w
0::
w
0 C)
2 0:: lXI t- w 2 2
2
� w
2
0:: c
�
c
::J
2
::J
0:: w
::J
t- � 0:: 0::

1 10
20

2 20 30

3 20 30 X Figure 5. Typical Pad Film Thickness Distribution.

L. E.
4 20 30 X X

5 20 30 X X
PAD

6 20 30 X X X

7 20 30 X X X X

8 20 30 X X X X

9 20 30 X X X X X

THRUST RUNNER
rm

T. E.

Figure 6. Typical Pad Film Temperature Distribution.

Figure 4. Typical Pad Film Pressure Distribution. Figure 7. Pad Temperature Distribution.
 228 PROCEEDINGS OF THE TWENTY-FIFrH TURBOMACIDNERY SYMPOSIUM

(Figure 1 0) . The low film tangential shear forces create babbitt

flow due to sub surface shear at high yield stresses. A bearing load
line can be superimposed on this graph to determine at what load
and speed the babbitt will creep and a safety factor can then be
applied. This assumes that the same pad surface node is always the
node of maximum temperature and pressure. This is not the case,
in fact the maximum pressure and temperature are occurring at
different nodes. Also, as the pad load or speed increases, the
thermal and elastic deflections cause the positions to change. Thus,
each node with its respective temperature and pressure should be
compared to the pressure/temperature creep limit for each load and
speed condition.

8000.0

7000.0 f'-.

i7i 6000.0
�
:�
(L
ui
Figure 8. Pad Elastic Deformation. V1 5000.0
�
capacity is related to the pad crowning ratio and at this load the
li1
�

:.0
4000.0 G ""-...
babbitt has a pressure/temperature creep limit that is considered. .0
0
The pad thickness and pivot type and geometry influence CD 3000.0 b.-
crowning. Pad spacing affects pad entrance temperature, thus "-.,
r--...
maximum capacity. The maximum temperature and minimum film 2000.0
thickness positions on the pad are affected by the circumferential �
t'-.....
and radial pivot position. The pad ratio which is pad width, length 1 000i0oo.o
transverse to the flow, over mean circumferential length, shear 1 50.0 200.0 250.0 300.0 350.0 400.0
Temperature, ° F
streamline length, which is tangent to the flow direction, affects
maximum babbitt temperature. Horsepower loss reflected in the
Figure 10. Grade 2 Babbitt Yield V s Limiting Temperature.
bulk temperature rise and required oil supply and temperature are
system variables that the machinery builder and user require to
specify the bearing lubrication system. Pad thickness affects load capacity. At maximum capacity the
Thrust bearing capacity depends on pad distortion and minimum ratio of crown to minimum film thickness is 0.6. Assuming a film
film thickness. Distortion is the maximum height of the pad surface thickness of 0.001 in, the pad elastic crown should be 0.0006 in.
curvature from a line connecting the opposite pad edges. Raimondi With the criteria and deflection equations, the pad height may be
found that the load variable was maximum at a crown to minimum determined. This results in pad heights approximately one-fourth
film thickness ratio of 0.6. For center pivot thrust pads that depend of the equivalent pad diameter.
on elastic curvature to support the load at start up, crown ra!ios The pad ratio affects the bearing parameters. The pad ratio,
below 0.5 show a steep reduction in load capacity. For crown ratios which is the mean circumferential pad length over pad radial
above 0.6 the reduction in capacity occurs but at a slower rate. The width, influences the temperature rise across the pad. Using a one
load capacity is shown in Figure 9 as a function of crowning ratio. dimensional energy equation and Reynolds equation in the
direction of motion for an adiabatic film will give the temperature
0.07 rise of a laminar film. For turbulent films, the temperature rise is

.
0 06 -- computed with a turbulent fluid friction factor and an equivalent
viscosity that is a function of Reynolds number. When comparing
(
�

ci
CD
::> 0.05 � a six and eight pad 1 0- 1 12 in thrust bearings, the maximum babbitt

.I
:::::: "-.., temperature of the six pad with a mean streamline of 3.5 in is
.£ approximately seven percent higher than the eight pad with a 2.63
C'f
.r: 0.04 '-....._
in mean streamline. The pad ratio affects the load capacity and film
.
>: ......
.,; 1---- thickness because of side leakage. The pad horsepower loss is
0 0 03
0 different since the friction factor changes with pad ratio.
_j
"'
0.02
--� Pad spacing can affect the inlet temperature; thus ultimately
:0
0
·c:
influencing pad maximum temperature or capacity. The relation is
0
> 0.01 shown in Table 3 of different bearing variables to Ettles' hot oil
carryover coefficient. The table was a result of 597 tests with tilt
0.00 pad thrust bearings. By decreasing the number of pads in a thrust
0 2 3 4 5
ring in increments of two, from 1 2 pads to four pads, the carryover
Crown Ratio, 6/hmin coefficient, K1, was reduced from 0.77 to 0.5 1 . Using Equation ( 1 ),
the pad inlet temperature drops by 46'F for the four pad bearing
Figure 9. Load Variable V s Crown Ratio for a Center Pivot Thrust compared to the 1 2 pad. In changing from an eight to six pad, the
Bearing. pad entrance temperature drops 1 2'F. Note that the design change
of the number of pads is as significant as load change for the range
Babbitt creep occurs at elevated local temperatures and local tabulated.
pressures through the hydrodynamic film. Booser [8] correlated Various pivot designs can be used to support the pad: point
wiping or creep with babbitt yield that is temperature dependent contact, ring contact, and line contact. Point contact as a sphere on
 DESIGN CONSIDERATIONS FOR THRUST BEARING APPLICATIONS 229

Table 3. Tilt Pad Hot Oil Carryover Vs Thrust Bearing pad width. The center of pressure of the film is at a greater radius
Parameters. than the pad geometrical center. An outward tilting pad does have
benefits even though a fixed compound taper land thrust bearing
APPLICATION/DESIGN PARAMETER K AK has an inward tilting pad. Varying the radial tilt angle by moving
the pivot, the maximum local film temperature position can be
OVERALL .67
moved and reduced. Also the film thickness can be increased by
4000 REV/MIN .60
6000 REV/MIN .68 .13 moving the pivot radially. By changing the pivot radius, the tem­
8000 REV/MIN .73 peratures and film thickness of the pad trailing edge may be
optimized such that the pad trailing edge values are approximately
NUMBER OF PADS equal between the pivot and the babbitt outside diameter.
4 .51
6 .63 PAD ENTRANCE TEMPERATURE
8 .70 .26
10 .72 DETERMINATION WITH CFD ANALYSIS
12 .77 CFD Analysis and Procedures

FLOW, GAUMIN A computational fluid dynamics analysis of the inlet and near
2.74 .70 geometry was completed to compare the pocket performance under
5.42 .66 .04 laminar and turbulent flow conditions. Tucker and Keogh [9] use
8.25 .66 CFD methods to study shaft thermal bowing in journal bearings
OIL TEMPERATURE RISE ACROSS HOUSING, •F
while forward and backward whirling was occurring. San Andres,
10-20 .51 et a!. [ 1 0], used CFD methods to investigate thermal behavior of
20-30 .59 cryogenic seals.
30-40 .62 Computational fluid dynamics solves the fluid continuity
40-50 .70 .28 Equation (A-2) or (A-6), the momentum equation (A-3) or (A-7),
50-60 .74
and the energy equation (A-5) or (A-8). Since flow in the cavity is
60-70 .79
70-80 .78 turbulent, the analysis uses (A-7) and (A-8). The difference
between the laminar and turbulent momentum equations is the
SPECIFIC LOAD, LBF/IN2 viscosity. The laminar analysis viscosity is a function of tempera­
0-200 .69 ture while the turbulent analysis uses an effective viscosity that is
200-400 .73 the sum of laminar and turbulent viscosities. The energy equation
400-600 .66 .28
differs in that turbulent kinetic energy is included in the total
600-800 .61
800-1000 .60 enthalpy for the turbulent case. The conduction term of the
1000-1200 .45 turbulent energy equation includes the partial derivative of the
enthalpy and a shear heating term is added that includes terms from
the turbulent k-e model.
The mixing of the fluid in the cavity was modelled as mass
fractions of the fluid entering at inlet one and two. The solution of
plate can be used in low to medium loaded thrust pads. Large pads, the momentum transport Equation (A-9) with the mass fraction
especially that start under load, use a ring contact. Raimondi found scalar shows mixing as a tracer dye from inlet two. The Spaulding
that a support ring radius to pad equivalent radius of one-third was and Launder model (k-e) was used for determination of kinetic
appropriate when considering elastic deflections. This design has energy and kinetic energy dissipation rate in the turbulent fluid.
an intermediate disc with ring contact at the pad and point contact B oth the kinetic energy and the dissipation rate of the energy are
loading the linl{. Ratios greater than 0.5 6 cause poor lift because solved as scalar quantities in the momentum transport Equation
the pad surface is concave. Ratios near zero cause excessive pad (A- 10) and (A- 1 1 ). This allows us to predicate the fluid shear
edge deflection. A ratio of one-third is optimum for considering heating. The major component of heating is the viscous heating
both pad edge deflection and concavity. At operating conditions, represented by Equation (A- 1 2).
the addition of thermal deflections result in a completely convex The fluids finite element solution is three dimensional and a
surface. fully coupled solution of the Navier-Stokes and energy equation at
Pads with radial line contact extending the width of the pad each node. Each fluids model has approximately 1 00,000 nodes
increase bearing capacity. The line contact reduces pad edge which form six sided hexahedral (brick) elements. S olution
deflection compared to the point and ring contact. This causes includes centrifugal effects and swirl effects of the runner on the
capacity to increase. For a center pivot bearing with a pad ratio of fluid and viscous shearing effects in the fluid. At each node, the
one, Raimondi, B oyd, and Kaufman found the maximum capacity velocity in three mutually perpendicular directions, k-E turbulence,
at the optimum crown to minimum film thickness ratio to be 30 pressure, and temperature were calculated. The laminar viscosity
percent greater for the same film thickness. of the fluid was modelled as a function of temperature. The
The pad pivot can be located by different criteria. The circum­ boundary layer near walls is viewed as a near wall laminar flow
ferential offset of the pivot from the pad entrance is a ratio with the (Law of the Wall) and a sublayer turbulent flow (Log Law of
mean circumferential length of the pad. As the offset is increased Wall). Selected plots are from a group of plots that show
from 0.5 0 for center pivot, the maximum temperature drops and the temperature, mixing, and velocity in the two orthogonal directions
film thickness increases. The minimum peak babbitt temperature longitudinally and three planes in the transverse direction.
occurs at pivot offsets between 75 percent and 85 percent when
determined experimentally. The range accounts for different Pad Entrance Temperature Problem
designs. At this offset, the local film pressures are greater and the In the two cases reviewed, a comparison of a control inlet pocket
minimum film thickness is somewhat lower. bearing at high speed but laminar, Case A, was made with the same
The radial offset tilts the pad to the outside or inside depending controlled inlet bearing running under turbulent conditions, Case
on its location. The pivot radial offset ratio is defined as the radial B. The film Reynolds number is 6 1 0 for the laminar and 1 5 1 0 for
distance from the babbitt bore to the pivot location divided by the the turbulent case. The 10- 1 /2 in, eight pad bearing is lubricated by
230 PROCEEDINGS OF THE TWENTY-FIFTH TURBOMACHINERY SYMPOSIUM

ISO VG-32 oil. The OD/ID ratio is two and the pad length to width
ratio is one. The bearing has copper pads with a 68 percent pivot
offset. Static performance data for the two cases are shown in Vector Scale

Table 4 E(
3. 7111JE-1211
PPS
Table 4. Comparison of a Control Inlet 10-112 in 8 Pad Thrust
Bearing Under Laminar and Turbulent Conditions.

BULK
SPEED POWER OIL TEMP
RPM LOAD PRESSURE MINFILM LOSS REQD RISE
CASE DESCRIPTION (FPS) LB PSI INCH HP GPM 'F

A CO TI,�����LEl 10000 35000 635 .00079 75 19 50

(343)

B CONTROL INLE 17000 35000 635 .0013 223 56 50

TURBULENT (563)

The controlled inlet bearing pocket CFD analysis between pad

geometry is: a precavity to direct hot oil carryover radially out of
the bearing, a hot oil director wall to shed some boundary oil, a land
to wipe some of the boundary oil off the runner and, of course, the
cool oil pocket to add cool oil such that the mixed temperature of
Figure 12. Controlled Inlet Pad Cavity Axial Mid Plane Laminar
the pad inlet is lower than a flooded bearing. The advantage of the
Velocity Field (.188 in from Runner) .
controlled inlet in general is that inlet oil at 1 20'F, rather than a bulk
temperature oil equal to the outlet oil at 1 50'F, is mixing with the
boundary oil. The offset pivot drops the pad front edge or inclines the velocity distribution where back flow velocity is set to zero.
the pad more so that the ratio of pocket oil to boundary oil is higher, Inlet two is for the fresh oil. This is at a temperature of l 20'F and
the flow is the required for hydrodynamic lubrication. The-pressure

b
allowing for a lower resultant or mixed pad inlet temperature.
is assumed to be zero psig, which is the most conservative assump­
Tl tion producing the highest pad inlet temperatures. Higher pressures
will drop the inlet temperature.

OUTLET I t At outlet one, the gap is determined by pad tilt and thermal and
elastic deflections. The pressure is obtained from the pad hydrody­
namic analysis. Recirculation of the flow occurs. The CFD analysis
-
determines the gap field temperatures, proportion of recirculation,
OUTLET I GAPo
PA!l/R\HER and flow velocity distribution to the pad. Outlet two is for excess
flow bleed off.

Pad Entrance Temperature Solution

The CFD model for laminar Case A had the following fluid
INLET 2 properties: swirl at the mid plane in a plane parallel to the runner
was occurring at inlet two (Figure 1 2). Three vortices were
occurring near outlet two. Average flow velocities were approxi­
mately 1 1 percent of runner velocity with a maximum velocity of
26 percent of runner velocity. The radial mid plane, perpendicular
to the runner, velocity field was toward the runner at inlet two and
two vortices were present at outlet two (Figure 1 3). The vortex near
the outlet was feeding the outlet while the adj acent vortex was of
opposite rotation and lower in the pocket. The average field
-r -
velocity was near seven percent of runner velocity. The transverse
1 velocity field at the optimized pocket section was centered at the
pocket center and swirl was adding cool oil to the moving
Figure 11. CFD Cavity Model Including Inlet/Outlet Locations. boundary layer feeding the pad (Figure 1 4). The core rotation
factor is 4.2 percent of the runner velocity.
The controlled inlet has four boundaries, Figure 1 1 and Figure Temperatures for the pocket in the three planes were greater than
1 2 , where the flow (velocity field), temperature, and pressure are the inlet two temperatures (Figures 1 5 , 1 6, and 1 7). Entrance zone
specified. Inlet one sees hot boundary oil from the previous pad temperatures were near the 1 20'F inlet two fluid. They extended
and recirculation. The temperature and gap between the pocket for the middle one-half of the pocket in axial and radial mid planes.
front edge and runner vary with radius. The temperature can be The outer transverse section was warmer than the mean and inner
partially determined from the pad thermoelastohydrodynarnic transverse sections:
analysis. The pad outlet temperature and velocity boundary layers Tracer dye injected at inlet two is fully mixed at smaller radii
are used. The boundary layer temperatures are used with back flow (Figure 1 8). At larger radii, mixing increases from 75/25, 65/35,
temperatures set to the pre-cavity average temperatures. The to 50150 in the axial plane. This is also the case on the radial plane
thermal and elastic deflection along with pad tilt set the gap. B ack (Figure 1 9). At the optimum transverse plan mixing to the 50/50 is
flow is also occurring. Couette flow and runner speed determines occurring near the boundary layer (Figure 20).
 DESIGN CONSIDERATIONS FOR THRUST BEARING APPLICATIONS 231

Vector Scale
-E------1
4. 152E -,01
OU'lLBT ...
2 llUIINBK

on
T
IIILET 2
ID

Figure 13. Controlled Inlet Pad Cavity Radial Mid Plane Laminar
Velocity Field.

OUTLET 1

Vector Scale
«E------1
2.647E-01
...

I ' Figure 17. Controlled Inlet Pad Cavity Transverse Plane

Temperature Distribution for Laminar Flow.
§ \\'
\ \

CAVITY 80TTtll

Figure 14. Controlled Inlet Pad Cavity Transverse Plane Laminar Ll!GBIID:

Velocity Field (Radius 4.54 in).= MIN= 0.02

liAX = 0.98
INCIII!IIBIIT = • 096

LEGIIIID:
MIN= 1.0
liAX = 1.2
INCUIIEIIT = 11.01

Figure IB. Controlled Inlet Pad Cavity Axial Mid Plane Mixing
Fieldfor Laminar Flow.

Figure I5. Controlled Inlet Pad Cavity Axial M id Plane

Temperature Distribution for Laminar Flow.

Figure I9. Controlled Inlet Pad Cavity Radial Mid Plane Mixing
Field for Laminar Flow.

Viscous heating is occurring in the boundary layer or near it and

at the pocket pad wall in a zone that is not large (Figure 2 1 ). The
Figure I6. Controlled Inlet Pad Cavity Radial M id Plane horsepower generated by the pocket is 0.0 1 6 per pad, which raises
Temperature Distribution for Laminar Flow. the oil temperature 0. 1 OF.
232 PROCEEDINGS OF THE TWENTY-FIFTH TURBOMACHINERY SYMPOSIUM

The pocket/pad feed plane temperature distribution shows a

central area of cool oil 35 percent of pad width and 80 percent of
the gap at 1 .07 times the bearing supply oil (Figure 23). Zones on
both sides are at 1 . 1 7 times supply oil, the inner zone is 10 percent
wide and 80 percent of the gap while the outer zone is 40 percent
wide and 80 percent high. The boundary layer temperature is 1.2
times supplied temperature, the length of the pad for 12 percent of
the gap. The eight percent of the gap nearest the pad for the width
of the pad is supplied by cool oil at approximately supply oil tem­
perature. Fluid mixing distribution at the pocket/pad feed plane
shows a central region that has a mixture of 75 percent of the cool
supply oil (Figure 24). Zones to each side of this have been com­
pletely mixed, 50 percent of the cool supply oil and 50 percent of
the boundary oil. Zones at the extreme ends of the pad show 90
percent supply oil and 10 percent boundary oil due to geometry and
speed effects.

LEGEND:

MIN= LO
IIAJ[ = 1.2
INCREKBNT= 0.01

Figure 20. Controlled Inlet Pad Cavity Transverse Plane Mixing

Field for Laminar Flow. c
Figure 23. Pad with Pocket Inlet Temperature for Laminar Flow.

LEGEND:

MIN= 0.02
IIAJ[= 0.98
INCREMENT= 0.096

Figure 24. Pad with Pocket Inlet Mixing Field for Laminar Flow.

v The CFD model for the turbulent Case B showed three

vortices near outlet two with tighter radii compressed into a
shorter radial distance than Case A (Figure 25). Velocity vectors
are aligned radially to the outside of inlet two. Swirl is
occurring at inlet two and two vortices are present towards the
pad inside diameter that were absent in Case A. On the radial
midplane, at inlet two, velocities are directed towards the runner
for the full cavity depth (Figure 26). This is diverted to radial
tangential flow near the runner to a greater extent to the outside
of inlet t wo. Low velocity occurs at the cavity bottom. The outer
Figure 2I. Controlled Inlet Pad Cavity Transverse Plane Viscous vortex near the pad outside diameter is compressed and the
Heating for Laminar Flow. adjacent vortex has started to dissipate. The transverse velocity
field at the optimized pocket section was similar to Case A
The pocket/pad feed interface plane, outlet one, dominant (Figure 27).
velocity is at approximately 40 percent of runner velocity (Figure The turbulent Case B showed the larger temperature range
22). The hot boundary layer has been reduced to five percent of the (Figure 28). Outlet two flow was greater than Case A, because
gap at the pad bore and 10 percent of the gap at the pad O.D. Back of the higher speed, and the cool oil region of the cavity was
flow averaging 1 5 percent of runner velocity occurs to 50 percent 25 percent greater than Case A in the axial planes (Figure 29).
of the gap at the pad bore due to low tangential velocity and 50 The main body of fluid in the optimized section transverse plan
percent of the gap near the pad O.D. due to outlet two. Back flow was also at approximately the same temperature as Case A
is less than 1 2 percent for 65 percent pad width centrally located. (Figure 3 0).
The mixing plots show Case B inlet two fluid extending to the
LEGEND: runner boundary fluid at inlet two. In Case A, the fluid entering at
KIN=..0.52 the feed is 70 percent of inlet two fluid directly above inlet two in
MAX=+1.31
INCR.EMBNT= 0.37 the radial plane plots. Both axial and transverse planes show the 70
IIUIIIIER
percent fluid inlet two Case A being double the volume in Case B
(Figure 3 1 and 32). Fluid at the cavity inside diameter and outside
f}l\
OD
4

BOTTOil
:J diameter are at five percent to ten percent more fluid from inlet
two. The transverse plots at the inner, mean, outer radius show less
CAVITY ID

mixing for Case B. The outer plane is 83 percent, fluid inlet two at
Figure 22. Pad with Pocket Inlet Velocity Field for Laminar Flow the bulk core for the high speed case vs 70 percent fluid inlet two
(Pad Entrance Gap, Outlet 1). of Case A (Figure 33).
 DESIGN CONSIDERATIONS FOR THRUST BEARING APPLICATIONS 233

LEGEND:
MIN = 1.0
Vector Scale MAX = 1.4

< INCREMENT = 0.02

3.710E-01
FPS
PAD

Figure 28. Controlled Inlet Pad Cavity Axial Mid Plane

Temperature Distribution for Turbulent Flow.
Figure 25. Controlled Inlet Pad Cavity Axial Mid Plane Turbulent
Velocity Field ( 188 in from Runner).
.

Vector Scale
<-----1
OUTLET 2
RUNNER 1i. 1 52E-01
FPS

Figure 29. Controlled Inlet Pad Cavity Radial Mid Plane

Temperature Distribution for Turbulent Flow.

on w
1NLET 2

Figure 26. Controlled Inlet Pad Cavity Radial Mid Plane

Turbulent Velocity Field.

OUTLET l

Fig u re 30. Controlled Inlet Pad Cavity Transverse Plane

CAVITY BOTTOM
Temperature Distribution for Turbulent Flow.

Figure 27. Controlled Inlet Pad Cavity Transverse Plane

Turbulent Velocity Field (Radius 4. 54 in).
=
B ackflow is occurring at the cavity outside diameter for two
percent of the pad radial length and the full gap height below the
runner boundary layer. B ackflow is not occurring at the pad gap
Viscous heating for the turbulent model is occurring at the inside diameter as in the laminar case.
runner boundary and pad wall (Figure 34). The zone is approxi­ The pad inlet gap temperature distribution range is to 1 .4 times
mately double the laminar model. At the inlet pad corner a larger the inlet temperature (Figure 36). The lower temperatures are near
zone of energy dissipation is occurring. The horsepower generated the pad and the highest temperatures are at the gap outside
by the pocket is 0.043 per pad, which raises the oil temperature less diameter and inside diameter near the runner. At inlet two in the
than one-tenth of 1 'F. gap and within 30 percent of the pad, the temperature is 1 . 1 0 times
The pocket/pad feed interface plane outlet one, dominant the inlet temperature for one-fourth of the pad radial length. The
velocity is about 30 percent of the runner mean velocity (Figure center one-third of the gap height at 1 . 1 6 times and the gap near the
35). The hot boundary layer is approximately nine percent of gap. runner is at 1 .4 times the feed temperature.
234 PROCEEDINGS OF THE TWENTY-FIFTH TURBOMACHINERY SYMPOSIUM

LEGEND :
MIN m 0 . 02
MAX= 0 . 98
INCREMEIIT = • 096

Figure 34. Controlled Inlet Pad Cavity Transverse Plane Viscous

Heating for Turbulent Flow.

Figure 31. Controlled Inlet Pad Cavity Axial Mid Plane Mixing
LEGEND:
Field for Turbulent Flow.
M I N = -0 . 40
MAX = + 1 . 3 1
I R CRBNENT = 0 . 3 4

RUIOO!R

OD CAVITY BOTTOH ID
Figure 32. Controlled Inlet Pad Cavity Radial Mid Plane Mixing
Field for Turbulent Flow.
Figure 35. Pad with Pocket Inlet Velocity Field for Turbulent Flow
(Pad Entrance Gap, Outlet 1).

LEGEND:

IIIR = 1 . 0
MAX = 1 . 4
INCREIIEI'fT = 0 . 02

Figure 36. Pad with Pocket Inlet Temperature for Turbulent

Flow.

LEGEND:

HlN = 0 . 02
HAX = 0 . 98
INCREMENT = 0 . 096

Figure 33. Controlled Inlet Pad Cavity Transverse Plane Mixing

Field for Turbulent Flow. Figure 37. Pad with Pocket Inlet Mixing Field for Turbulent
Flow.

The fluid mixing maps show a similar distribution as tempera­

ture maps at the pad feed plane (Figure 37). Fluid is 75 percent inlet oil being fed at 1 20oF. For the turbulent model running 1 .7
cavity feed fluid near the pad at cavity inlet two. The mixing near times faster, the hot oil carryover temperature was reduced by 5 6
the runner is 50 percent of inlet two fluid. Towards the pad percent.
extremes the fluid is 30 percent of cavity feed fluid. The increased Some of the data for the laminar and turbulent models are sum­
speed of runner B has resulted in less mixing, especially at the marized in Table 5 . The turbulent viscosity to oil viscosity ratio at
central pad area in the gap. Case B shows stratified temperature the pad entrance and over the cavity domain are shown. Cavity
and mixing plots. swirl velocities are mass averaged values on a horizontal mid plane
For the laminar flow model, the average hot oil carryover temper­ 0. 1 in wide at the transverse plane at a radius of 4.54 in.
ature was reduced by 53 percent when mixed with the controlled Centrifugal fluid velocities are also shown.
 DESIGN CONSIDERATIONS FOR THRUST BEARING APPLICATIONS 235

Table 5. CFD Controlled Inlet Fluid Parameters. The material type also can reduce local babbitt temperatures . A
drop in temperature of about 25.F is shown in Figure 38 by using
.:... a strengthened copper with one percent chromium in place of a
o(
Cl
::1 steel for the pad backing material. The temperature reduction is
'
w
�
'...I
It:
�.... (.) w w due to the two effects. The higher thermal conductivity of copper,
z (.) (.)
z
i:i z z 190 BTU/hr-ft-·F as compared to steel 3 1 BTU/hr-ft-·F, reduces the
� � � � .... w w
...1 ...1 temperature differential from the top to the bottom of the pad. This
u z
� � u w ::1
Gl 0
i= � ::1
Gl 0
> g > g It: reduces the thermal crowning of the pad, which reduces the local
o(
(.)
w
>
% RUNNER
� �
% RU N N E R
� � i:i � ::1
(.) .... � babbitt temperature and pressure along with increasing the film
CASE FLOW IPS SPEED IPS SPEED DIM. DIM.
thickness. A one dimensional thermal model also shows local
babbitt temperatures being reduced directly. The temperature
reduction changes with speed and load. It can be as great as 54•F
A LAMINAR 1 72 4.2 84 2.0 0.6 1 .0
at a load of 700 psi to no decrease at a load of 1 00 psi. Speed
effects the temperature reduction, 36•F below 1 0,000 rpm and 1 8 .F
B TURBULENT 235 5.7 1 52 3.7 1 .5 2.0
above. Copper does have an elastic modulus of 1 6 . 1 E6 psi, about
one-half the modulus for steel. This allows for more crowning of
the pad due to pad elasticity. When copper pads are at high loads,
the pad crowning should be studied for loss of load capacity and
* MASS WEIGHTED AVERAGE
film thickness. Pads can be redesigned with greater depth or a
!J. R = 4 . 54" different support system to compensate for the loss of modulus .
The cool oil pocket features prepad mixing o f boundary layer
and bearing inlet oil. A standard bearing will mix the boundary
DESIGN FEATURES AND PERFORMANCE OF A layer with lubricant that is at a bulk oil temperature above the
CONTROLLED INLET BEARING bearing inlet oil. The bulk temperature used is usually the bearing
Controlled Inl�t, Offset Pivot, and Copper Pads outlet temperature assuming a flooded bearing. This higher bulk
temperature occurs because of the parasitic losses both through
Thrust bearings can feature offs et pivot, chromium copper pads, flow and surface area drag and mixing of the oil leaving the sides
or pad cool oil pockets individually or in combinations. Pad pivot of the pad. Many bearings have a bulk temperature rise of 30.F
offsets traditionally were under 60 percent, because various isovis­ with a bearing inlet temperature of 120·F. Thus, the standard
cous analysis showed that optimum loading of tilt pads was flooded bearing mixes the hot boundary layer oil from the previous
occurring near an inlet to outlet film ratio of two. Since bearing
pad with 1 50•F oil; whereas the pocket or controlled inlet bearing
failure is based on local babbitt temperature and film temperatures mixes bearing inlet oil at 1 20•F with the boundary layer to
drop at higher pad inclinations, i.e., pivot offsets, by changing the
determine pad inlet temperatures . 1\vo other features of the con­
criteria, greater pad offs ets have become popular. A plot is shown
trolled inlet bearing help reduce the pad inlet temperature: the hot
in Figure 38 of local babbitt temperature vs pivot position at 500
oil director and wiper. The director moves previous pad side
psi loading and 6000 rpm. For the steel pads, babbitt temperature
leakage oil to the collar area to be spun out of the bearing instead
decreases to pivot offsets as high as 80 percent. Many bearings
of raising the pad inlet oil temperature. The wiper, or controlled
today have offs ets between 65 percent and 70 percent.
inlet land, will shed a portion of the boundary layer into the
260 director stream. Since the wiper's distance to the collar is below

u..
250

2 40
� j -
-
Cr-Cu Pad
S teei Pad
� but near the pad inlet film thickness, one to two ten thousands of
an inch, a major portion of the boundary layer still enters the cool
oil pocket to be mixed.
� 230
� The offset pivot, copper material, and pad controlled inlet all
.3 appear to reduce pad babbitt temperature which is the prime
2
QJ 220
� � indicator of bearing capacity. Using two or all three of the features
a.
E
QJ 21 0
� � does not mean that the temperature reduction of the film can be
f-- --- estimated by adding the features directly; since, they are related
�
�
200
,.., ......._ nonlinearly with interaction effects.
:.0 .......
.0
0
CD 1 90
.......
-.... A Controlled Inlet Thrust Bearing
1 80
The controlled inlet bearing (Figure 39) has design features that
1 70 increase load capacity along with reduce flow requirements and
50 60 70 80
power consumption. The load capacity is augmented by three
Pivot Position, 7. o f Pad Arc features : the controlled inlet (Figure 40) which consists of a hot oil
Figure 38. Pad Temperature V s Pad Pivot Offset. director, wiper, and cool oil pocket; the offset pivot which tips the
shoe to a greater inclination allowing more cool oil from the pocket
As pivot offset increases, the inclination of the pad to the to mix with the boundary layer hot oil; and the pad line contact
rotating collar increases. This allows a higher volume of bearing support reduces radial pad deflection that button type supports
feed oil to enter the pad compared to the carryover boundary layer inherently have. The reduced deflection translates into greater load
oil. The shear rate of the lubricant is reduced, thus dropping the capacity. The capacity can be as high as 750 psi unit load.
film temperature. The tradeoff is with babbitt creep at local sites on The flow requirements are reduced by the controlled inlet cool
the pad. Design charts showing the babbitt yield strength vs tem­ oil pocket that places and holds the cool oil at the pad front edge
perature with a design safety factor allow the local babbitt pressure and the controlled inlet lubricant feed system that delivers the cool
and temperature to be reviewed for different placements of the oil to the pocket. The oil containment by the controlled inlet and
pivot. If the babbitt stress is approaching the yield limit of babbitt feed system reduce parasitic loss. Horsepower loss reduction is
or creep initiation the percent of pivot offset can be reduced. surface drag and through flow losses. Less oil is heated by shear at
236 PROCEEDINGS OF THE TWENTY-FIFTH TURBOMACHINERY SYMPOSIUM

Figure 4I. Controlled Inlet Thrust Retainer Showing Assembly of

Lower Link Pin and Links.
Figure 39. Controlled Inlet Thrust Bearing.

retainer in the housing, locating the lower link circurnferentially,

securing it against rotation, and holding the assembly together for
shipping and turbomachine assembly.

Pad
The radial rib with line contact has replaced the spherical pivot
with point contact which reduces radial elastic pad deflection, thus
increasing load capacity or film thickness and results in lower
babbitt temperatures . The pad and other components are retained
by a screw lock that prevents parts falling out at assembly.

Controlled Inlet
The controlled inlet feeds each pad front edge (Figure 42). The
fluid flows radially in and axially up to the cool oil pocket. In the
pocket, the cool oil mixes circurnferentially by swirl and radially
by centrifugal forces with the boundary layer hot oil carryover. The
pocket cross section has optimized mixing geometry at the outer
one-third of the pad front edge since this section requires the most
flow and develops the highest local temperatures downstream, near
Figure 40. Controlled Inlet Showing Hot Oil Director Wiper and the trailing edge of the pad.
Cool Oil Pocket.

the shaft/retainer annulus and the collar housing annulus because

of lower fluid velocities. Shoe side shear is still occurring at the
front and rear of shoe. The entrance, between pad, and discharge
entrance loses are less due to lower flows and fluid velocities.

Thin Axial Profile Bearing

The bearing assemble has a thin axi al profile because the
retainer (Figure 4 1 ) has an optional bottomless feature. Axial
height is reduced 3 1 percent for the feature. This can reduce rotor
length on new machines and will allow the bearing to be fit into all
upgrade or rerate designs. The bottomless retainer produces a
tighter stack height tolerance which allows the links to give a
greater proportion of their adjustment range to misalignment duties
as compared to manufacturing tolerance issues. By omitting the
retainer trepan bottom with parallel, flatness, and tolerance issues
which are occurring on the largest assembly part, and transferring
tolerances to small parts, the stack height tolerance is reduced. The
retainer bottom tolerance has been replaced by a dowel pin
tolerance. Thus, the thin thrust bearing assemble is more accurate­
ly made and functions better for misalignment.
The retainer is not part of the axial load system with this bearing,
it functions to hold bearing parts in the assembly. The retainer is
replaced by the lower link dowel pin in the axial load system. The
pin functions include transferring the axial load, positioning the Figure 42. Pad with Controlled Inlet and Lube Feed Piston.
 DESIGN CONSIDERATIONS FOR THRUST BEARING APPLICATIONS 237

The controlled inlet consists of a precavity and hot oil director 1 05

- BASE LD
that move some of the hot oil from the previous pad radially out of - LOW FLOW
the bearing in the precavity. The land or wiper also reduces the - H I G H LD
carryover oil to the pocket by shedding some of the boundary fluid
at the wiper front edge to the precavity. The reduced boundary
layer is then mixed in the pocket cavity with cool inlet oil to enter
the pad.
The controlled inlet is made of babbitt so that the runner "sees"
only soft babbitt as the pad surface. The controlled inlet is detach­ I"'"
able so that a flooded or controlled inlet bearing differ only by
adding the controlled inlet and pistons. This allows for easy
upgrades in the field, if desired. 1 .5

Piston
1 04
1 03 1 .5 2.5 3 7 B 1 04
The controlled inlet pocket feed system includes the retainer cir­
SPEED,RPM
cumferential annulus and an axial half cylinder with large comer
radii for reduced pressure losses to feed the radial floating bronze
hollow piston. The piston is located in the retainer and moves Figure 44. Load V s Speed, 10-112 in Thrust Bearing with Pocket
radially by differential hydraulic forces to seal against the con­ Inlet, Offset Pivot, and Steel Pads.
trolled inlet. This design is easier to assemble and has fewer parts
or higher assembly reliability. The large bore diameter of the 1 02
piston, half cylinder ports in the retainer, and large turning radius - BASE LD
in the controlled inlet have reduce feed system pressure drop. 6 - LOW FLOW
5 - H I G H LD
4
Links
The links (Figure 43) are designed to provide a misalignment :::;;
capacity. Contact stresses are lower where a pair member is made Q_
(.!) �.-..
1 01 -�
of a soft material, at the pad/upper link interface and the lower ";i
0
link/link pin/housing interface. Higher contact stresses exist by ..J
u_
�
design where the pair members are both hard, that is upper - -
link/lower link contact. At current thrust bearing catalog loads, 500 _.....
psi, the pad/upper link contact stress is 50 percent of the spherical �
pivot contact stress. The lower link support pin and support
pin/housing or shim contact stress is 70 percent and 10 percent of 1 00
the lower link/retainer, cylinder on plate contact stress. 1 03 1 .5 2.5 3 7 B 1 04
SPEED, RPM

Figure 45. Flow V s Speed, 10-112 in Thrust Bearing with Pocket,

Offse t Pivot, and Steel Pads.

250
- BASE LD
2 40 - LOW FLOW
- HIGH LD
u_ 230
.0
" 220
Figure 43. Controlled Inlet Thrust Bearing Upper and Lower Link -.....
l() -
" 21 0
Assembly. v r-
w -
0:: 200
::::> -
�
0::
1 90
Performance of a Controlled Inlet Bearing w
�
Q_ 1 80 ..... 1---
A controlled inlet bearing with offs et pivot will have higher :::;; I--
w
I- 1 70
.,...
capacity, lower power losses, or a combination of both. The ,......
following discussion tries to show the interplay of load, speed, and 1 60
flow in the bearing performance parameters, horsepower loss, and
1 50
babbitt local temperature. Figures 44 through 47 should be viewed 1 000 2000 3000 4000 5000 6000 7000 8000 9000 1 0000

simultaneously. That is, for a given speed, load, and flow (Figure SPEED, RPM
44 and Figure 45) , the resulting babbitt temperature and horse­
power loss (Figures 46 and 47) may be determined. Most published Figure 46. Pad Temperature, 10-112 in Thrust Bearing with
thrust bearing data does not show flow effecting local temperature. Pocket, Offset Pivot, and Steel Pads.
The load/speed map is shown in Figure 44 for a 10- 1 /2 in thrust
bearing running at base conditions, 30 percent more load, and 5 0
percent l e s s flow. Th e speed range is 2000 t o 8000 rpm and loading The babbitt temperature plot (Figure 46), which is a good indi­
is to 3 9000 pounds at the higher speed. The required flow is shown cation of bearing capacity, shows an average temperature
in Figure 45 for the three cases: base condition, high load, low difference at 2000 rpm of 22·F when comparing either the high
flow. load or low flow bearing to base conditions. At a higher speed of
238 PROCEEDINGS OF THE TWENTY-FIFTH TURBOMACHINERY SYMPOSI UM

300
f:=
� === E
1 02
- BASE L D P O C K ET: D I R E C T E D , O FF S ET P I VO T , S T L PAD
6 - LOW FLOW 280 S T D : F L O , CTR P IVOT, STL PAD r--
5 - r--
HIGH LD /v
u.. 260
.......: 0,....
(!) w
(!) 0:: 240
0
-'
v-::::� :::>
-�--

0::
� � � 220
w 0::
/ w '
:s: 1 01 (L 200
0 :::;;: .......
(L w
w .......
f- 1 80
(!)
0:: L()
0 /'-
I "'-.. 1 60
L()
/'-
1 40 3 0 0 P S I U N I T LOAD

1 20 8000 RPM

1 00 1 00
1 03 1 .5 2.5 7 8 1 04 0 1 0 20 30 40 50 60 70 80
SPEED, R P M FLOW, GPM

Figure 47. Horsepower Loss Vs Speed, 10-1/2 in Thrust Bearing Figure 49. Pad Temperature Vs Flow, 10-112 in Thrust Bearing,
with Pocket, Offset Pivot, and Steel Pads. Flooded V s Pocket.

CONCLUSIONS
8000 rpm, the average difference between the high load or low
Tilt Pad Thrust Bearing Selection
flow and the base bearing is 2TF. In a more general plot (Figure
49), the local temperature increases with reduction in bearing flow • With Just in Time manufacturing, manufacturing cells, short lead
requirement. A standard lug type bearing with center pivot running time NC machined bearing components, and engineering tools like
flooded has a higher local pad temperature and cannot be operated CFD, thrust bearings can be designed for the application rather
to as low a flow as a controlled inlet bearing. The latter bearing than having to use a catalog item.
with both the advantage of pad pivot offset that allows more cool
• Establish what parameters control bearing capacity by plotting
oil to enter the front of the pad, because that edge has dropped
bearing on the load speed map and using Table 1 .
farther from the runner, and the control of cool inlet oil at the front
of the pad by the pocket allow for lower measured temperature. • Ask bearing supplier what type of analysis was used. As the
Parasitic horsepower loss is correlated with speed and flow. B y solution includes more variables, it emulates run conditions closer.
reducing the flow, the bearing horsepower loss is less. Assuming a Isoviscous approximation by closed form solutions for f ilm,
•

bulk temperature rise of 40'F to 50'F, which is common for con­ i.e., short or long bearing theory
trolled inlet bearings, reductions in flow can reduce bearing
horsepower loss, provided of course, flow reductions do not go • Isoviscous hydrodynamic film solution
below the minimum flow requirement for hydrodynamic lubrica­ • Thermohydrodynarnic film solution
tion. The horsepower loss, shown in Figure 47, for base and low
flow bearings being approximately the same at 2000 rpm, since at
• Thermohydrodynarnic film solution with turbulence
low speeds parasitic losses are not significant when compared to Thermohydrodynamic film solution with turbulence and
•

the hydrodynamic pad shearing loss. At 8000 rpm, the horsepower pad thermal effects
was reduced by 1 8 percent for the low flow bearing. Larger
Thermohydrodynamic film solution with turbulence and
•
reduction will occur at higher speeds. Horsepower loss vs bearing
pad elastic effects
supply flow is shown in Figure 48. Here, horsepower loss drops by
40 percent as the flow is reduced from 40 gpm to 7 - 1 12 gpm, at a Thermohydrodynarnic film solution with turbulence and
•

300 psi load, and a speed of 8000 rpm. pad thermal and elastic effects
Thermohydrodynamic film solution with turbulence and
•

80
pad and runner thermal and elastic effects
CFD coupled thermohydrodynamic film solutions with
•

300 PSI turbulence and pad and runner thermal and elastic effects
70
8000 RPM
(!) • Review design features with bearing suppliers to determine
(!)
0
-' f.- bearing capacity, features as pad ratio and pad spacing. Discuss
n::: 60 _,..,.... pivot position strategy, pad crowning criteria, etc. Establish the
w
:s:
0
kC: special features you need for your bearing application, such as no
(L
w
(!) 50
1/ load axial vibration control or a particular film thickness ratio for
n:::
0
I
/ a process fluids application.

40
/ Pad Entrance Temperatures
"'
• The two dimensional array, temperature field, at the pad entrance
is determined by a CFD analysis of the controlled inlet pocket
1 0 20 30 40 50 flows. Previous procedures to determine pad entrance temperature
FLOW, G P M
have produced an average value of temperature. Bearing capacity
is partially determined by an accurate pad inlet temperature.
Figure 48. Horsepower Loss Vs Flow, 10-1/2 i n Thrust Bearing • Using an inlet temperature distribution rather than an average
with Pocket, Offset Pivot, and Steel Pads. inlet temperature should predict babbitt temperatures more
 DESIGN CONSIDERATIONS FOR THRUST BEARING APPLICATIONS 239

accurately. With the ability to view the pad entrance gap as a two T l - Ts
dimensional field of temperatures, it appears that the boundary Tr - Ts
layer near the runner can be above the babbitt limit temperature
k Turbulent kinetic energy
because the fluid near the pad babbitt is at a lower temperature for
highly loaded bearings. Mass at inlet i
Turbulent Prandtl number
• CFD analysis can be used to design an optimum cavity for a Pressure
specific load/speed envelope. Flow to pad
• The cavity transverse swirl is optimum at the pad entrance radius Flow from pad
where the highest pad pressure occurs downstream. A swirl rota­ Recirculation flow
tional velocity reduces pad inlet temperatures by interacting and Bearing supply flow
shedding a portion the boundary layer. A higher degree of mixing Source terms
occurs at the optimized plane compared to the mean and inner Temperature
transverse plane. Fluid temperature at pad entrance, op
Fluid temperature at pad exit, °F
• The controlled inlet cavity cool oil entrance is optimally placed Runner temperature, op
so the lowest pad inlet temperature field is where the pad will see Fluid supply temperature, °F
the highest local temperatures and pressures downstream.
• The cavity velocity maps were similar qualitatively for the
laminar and turbulent cases. 3 dimensional orthogonal velocities of a spacial position
in the fluid
•The turbulent case with increased feed due to the speed had filled
more of the cavity volume with cool oil.
Fluctuating velocity component
• Higher inlet flows will cause less mixing resulting in a tempera­ Mean fluctuating velocity component
ture stratification at the pad entrance. This is occurring in the
circumferencial and axial plane of the cavity.
• CFD bearing analysis can be used to estimate internal bearing 3 dimensional orthogonal velocities of a spacial position
in the fluid
parasitic power losses.

Controlled Inlet Thrust Bearing Features r Diffusion coefficient, K/cp

dT Temperature rise along pad streamline, •p
• Higher capacity by design, 50 percent more. Higher bearing e Thrbulent kinetic energy dissipation rate
capacity due to offset pivot pad and controlled inlet design, 750 A. Mixing constant
psi. Viscosity
Jl
• Low stack height due to retainer construction, 31 percent less. Jle Effective viscosity
J..4 Thrbulent viscosity
• Lower feed flow requirement due to controlled inlet, 40 percent
p Fluid density
less .
crk K-e Schmidt number
• Lower horsepower loss due t o reduced flows and lower surface Oe K-e Schmidt number
drag and through flow power loss, 20 percent less. 't Fluid stress
• Lower Hertz contact stress in links due to conformable surfaces tj> Mass fraction
or line contacts replacing spherical pivot contacts in soft pairs. APPENDIX
• Modularity of controlled inlet allows for field retrofits for Computational fluid dynamics equations for laminar and
increased load or reduced horsepower losses. turbulent steady state flow are shown below. The terms in the
• The hydraulic feed piston and controlled inlet oil pocket supply equations are in the NOMENCLATURE section.
cool oil to the front pad edge with fewer parts and a simpler
assembly. Reynolds Equation
• The controlled inlet functions as a collar wiper, director for hot
oil carryover, and cool oil supply cavity for the pad.
A- 1
NOMENCLATURE
Body force
For Laminar Flow
Constant
Conservation of Mass
Constant
Constant
Specific heat
Groove energy sink A-2
Pad energy sink
Runner energy sink
Side leakage energy sink
Shear energy source Conservation of Momentum
Gravitational force
Total enthalpy
Enthalpy A-3
Conductivity
242 PROCEEDINGS OF THE TWENTY-FIFTH TURBOMACHINERY SYMPOSIUM