SPACE SHUTTLE MAIN ENGINE 

(SSME) LOX TURBOPUMP PUMP-END 

BEARING ANALYSIS 

I 

FINAL REPORT 

I 

-178746) SPACE SFiC.II1E MAIN ENGINE 186-22633 

(ssaE) LOP TU.BECPBEIP EUBP-ENC EEABING 

hNALYSIS Fidal Regoxi (Spectra Besearch 

Systems, Inc+) 60 p BC ~04;Cl~ 801 CSCL 21H Unclas 
SPACE SHUTTLE MAIN ENGINE 

(SSME) LOX TURBOPUMP PUMP-END 

BEARING ANALYSIS 

FINAL REPORT 

JANUARY 6, 1986 

SUBMITTED BY: SRS Technologies 

555 Sparkman Drive, Suite 1406 

Huntsville, AL 35805 

PREPARED FOR: Mr. Fred J. Dolan 

Materials and Processes Laboratory 

Engineering Physics Division 

George C. Marshall Space Flight Center 

Marshall Space Flight Center 

CONTRACT NUMBER: NAS8-36183 Modification No. 6 

TECHNOLOGIES 

SYSTEMS TECHNOLOGY DIVISION 

555 SPARKMAN DRIVEISUITE 1406 

HUNTSVILLE, ALABAMA 35805 
SRS Technologies 

TABLE OF CONTENTS 

FOREWORD. ............................. ii 

TABLE OF CONTENTS ......................... i ii 

LIST OF FIGURES .......................... i v 

LIST OF TABLES. .......................... v 

1.0 INTRODUCTION. ........................... 1 

2.0 SUMMARY .............................. 2 

3.0 SHAFTIBEARING MODEL DESCRIPTION .................. 4 

3.1 SHABERTH ShaftIBearing Model ................. 4 

3.2 SI 

3.3 Mo 

4.0 ANALYSIS OBJECTIVES AND APPROACH. ................. 9 

5.0 ANALYSIS RESULTS. ......................... 13 

5.1 Mechanical Characteristics .................. 13 

5.2 Thermal Analysis Resul ts ................... 15 

6.0 CONCLUSIONS AND RECOMMENDATIONS .................. 52 
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LIST OF TABLES 

TABLE 

NUMBER PAGE 

5.1.1 Bearing Operating Characteristics with Uniform Temp. Profile . 14 

5.1.2 Bearing Operating Characteris tics Considering Thermal Effects. 14 

5.2.1 Component Temperatures for Different Inlet Coolant 

Temperature at Low Flow Rate (3.6 lbs/sec) . . . . . . . . . . 16 

5.2.2 Component Temperatures for Different Inlet Coolant 

Temperature at High Flow Rate (7.0 lbs/sec). . . . . . . . . . 17 

5.2.3 Component Temperatures for Different Contact Friction Factors 

with 245% Increase in Heat Transfer Coefficient. . . . . . . . 18 

5.2.4 ComponentTemperaturesforDifferentContactFrictionFactors 

with 343% Increase in Heat Transfer Coefficient. . . . , . . . 19 

5.2.5 Effect of Outer Race to Isolator Clearances. . . . . . . . . . 47 

5.2.6 Effects of Thermal Isolation of Bearing Isolator (45 mm 

Bearing). . . . . . . . . . . . . . . . . . . . . . . . . . . 49 
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1.0 INTRODUCTION 

This report describes the work accomplished under the Space Shuttle Main 

Engine (SSME) LOX Turbopump Bearing Analysis effort. The objectives of this 

activity were to model the SSME LOX turbopump bearings and shaft, and to 

evaluate the sensitivity of the pump-end bearing thermal and operating 

characteristics to a broad range of imposed conditions. 
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2.0 SUMMARY 

A simulation of the shaft/bearing system of the SSME LOX turbopump has 

been developed. The simulation model a1 lows the thermal and mechanical 

characteristics to interact a1 lowing a real istic simulation of the bearing 

operating characteristics. The model was used to investigate the sensitivity 

of bearing operating characteristics to variations in parameters such as 

contact friction, preloads, heat transfer coefficients, coolant flow, etc. 

Development of the system model involved two major modeling efforts; a pump 

shaft/bearing modeled on the SHABERTH computer code, and a detailed nodal 

model of the 45 mm pump bearing modeled on the SINDA thermal code. The 

SHABERTH model includes the shaft, pump-end bearing, and turbine-end bearing 

pairs. The inboard (Number 2) pump bearing is modeled in detail to allow 

determination of component temperature profiles and average temperatures. 

wo-phase coolant conditions , and includ 

the heat generation from bearing friction and fluid stirring. An executive 

program controls the iteration between the mechanical and thermal models to 

arrive at solutions satisfying both models. Solutions that satisfy both 

models are considered to be realistic and are designated as converged solutions. When the models are not simultaneously satisfied, the solution 

diverges and a realistic operating condition is assumed not to exist. 

Using the simulation model, parametric analyses were performed on the 45 

mm pump-end bearing to investigate the sensitivity of bearing characteristics 

to contact friction, axial preload, coolant flow rate, cool ant inlet temperature and qual i ty, heat transfer coefficients, outer race clearance and 

misalignment, and the effects of thermally isolating the outer race from the 

isolator. The shaft speed was 30,000 RPM and a radial load profile was used 

that produced reactions measured in pump build 2606 RI. 

Bearing component temperatures (average and peak) were found to be very 

sensitive to contact friction, axial preload, and boundary heat transfer 

conductance. The effects of coolant flow rate and inlet conditions (temperature and qual ity) are not as dramatic. However, at the more severe 

conditions (i.e., high friction and loads, they can be an important factor in 

determining the thermal stabil i ty of the shaft/bearing system. The effect on 

component temperatures of varying outer race-to-isolator clearance is small, 
SRS Technologies 

over the range of clearance investigated. Outer race clearance was lost, for 

the Number 2 bearing, for th'e majority of the cases investigated. Since the 

analysis only considered radial loads, the consequence of the loss in outer 

race clearance was not fully determined. The loss of clearance could prevent 

axial movement of the outer race, with resultant high axial loads especially 

during engine power level changes. 

A maximum surface temperature of about 1800°F was estimated for the most 

severe case that converged. This case involved an outer race misalignment of 

~29 minutes. The highest surface temperature with perfect outer race a1 ignment was ~700°F. The large increases in temperature with small errors in 

outer race a1 ignment show how critical proper installation of the outer race 

is to bearing operation. 

Load sharing of the Number 1 and 2 pump bearings is not equal. Initial - 

7y this is primarily due 

radial load to the Number 2 bearing. Due to the heavier load, the Number 2 

bearing heats up, looses internal clearance, and becomes stiffer in the 

radial direction than the Number 1 bearing. This causes the Number 2 bearing 

to share an even greater part of the load with increasing temperatures and 

further loss of internal clearance. Analysis indicates that an internal 

clearance of at least 3.9 mils is necessary to maintain thermal stability for 

the conditions evaluated. If conditions (load, friction, etc.) become too 

severe, the bearing can experience severe thermal distress which can lead to 

failure. 

Discoloration of bearing components indicates surface temperature in 

excess of the %700°F maximum predicted for stable operating conditions. 

There are two possible reasons for these differences in hardware observations 

and model predictions. The high temperatures indicated by the hardware could 

result from transient excursions which do not dwell long enough for the 

system to become thermal ly unstable. These conditions could produce high 

surface temperatures, but not dwell long enough for the average temperature 

to increase sufficiently to loose critical internal clearance. Ball milling 

or rapid ball wear experienced with the 45 mm bearing in effect increases the 

internal clearance, and could allow the bearing to survive when it would 

otherwise fail due to over temperature and loss of critical internal 

clearance. 
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3.0 SHAFT/BEARING MODEL DESCRIPTION 

A shaft/bearing sys tem model was developed to represent the Space 

Shuttle Main Engine (SSME) Liquid Oxygen (LOX) turbopump shaft and bearing 

configuration using the SHABERTH computer program. A thermal model was 

developed specifically for the 45 mm pump-end bearing using the Systems 

Improved Numerical Differencing Analyzer (SINDA) program. 

3.1 SHABERTH Shaft/Beari ng Model 

The SHABERTH model for the SSME LOX turbopump consists of the turbopump 

shaft, turbine-end and pump-end bearings. The major components of the 

shaft/bearing model are shown in Figure 3.1.1. The complete shaft/bearing 

system was modeled so that the effects of preload and shaft deflection on 

bearing load sharing could be studied. Variations in outer race clearance, 

model. The model used a shaft speed of 30,000 RPM and shaft loading conditions presented in Figure 3.1.1. These radial loads were used to produce 

the reactions reported for pump build number 2606R1. The SHABERTH model 

predicts the frictional heat generation at the contact points in the bearing 

and the load distribution for each bearing set. 

3.2 SINDA Thermal Model of 45 rnm Bearing 

The SINDA thermal model is a detailed nodal division of the SSME LOX 

turbopump Number 2 bearing. The SINDA thermal model, using the nodal representation of the bearing components (inner race, ball, outer race, etc.), is 

able to predict component temperature profiles and average component temperatures. The thermal model uses the energy conservation equation to obtain the 

temperature distribution in the bearing components. The model solves the 

conservation equation using the bearing frictional heat generation, predicted 

by the SHABERTH model, and fluid stirring heat generation. The thermal model 

also accounts for the heat generated by pump-end bearing Number 1. 

The SINDA thermal model is able to simulate different coolant flow rates 

and inlet coolant temperatures. The model is capable of simulating saturated 

coolant entering bearing Number 2. Also, thermal isolation of specific 

components in the bearing can be performed with this model. 

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3.3 Model Iteration Process 

The SHABERTH shaft/bearing model and the SlNDA thermal model are used in 

an interactive iteration process to determine the steady state operating 

conditions for a specific case. The iteration process can also predict if a 

case is thermally unstable. 

The iteration process is totally automated. The user supplies an 

initial guess for the bearing component temperatures and the models iterate 

until a solution is determined. The models are first set up to simulate a 

specific case. The user then supplies an initial guess for the component 

temperatures to the SHABERTH shaft/bearing model . The SHABERTH model then 

analyses the shaft/bearing system and calculates the frictional heat 

generation rates of the contact points in the bearing. These heat generation 

rates are placed in the SINDA thermal model. SINDA then solves the energy 

components. The average component temperatures are calculated from the 

temperature profiles and are used in a comparison with the average component 

temperatures used by the SHABERTH model. If all the temperatures compare to 

within 2"C, the steady state solution has been reached. If the comparison 

fails, the SHABERTH model uses the new temperatures to predict new frictional 

heat rates. The thermal model uses the new heat rates to predict another set 

of average component temperatures. The temperature comparison is again made. 

The iteration process is continued in this manner until the temperatures 

used by the SHABERTH model are within 2°C of the temperatures predicted by 

the thermal model or the predicted temperatures exceed an upper limit of 

2000°F. 2000°F was chosen as the approximate ignition temperature of 440C in 

liquid oxygen. A case is considered thermally unstable if an average 

component temperature exceeds the 2000°F limit. Figure 3.3.1 illustrates the 

iteration process between the SHABERTH shaftlbearing model and the SlNDA 

thermal model for a converged case. In this case the friction heat from 

SHABERTH and the average component temperature from SINDA increase to a 

converged point. Figure 3.3.2 shows a diverged (thermally unstable) case 

which has no converged point. 
FIGURE 3.3.1 ITERATION PROCE: 

SHABERTH AND SIb 

e RESULTS FROM SHABERTH 

x RESULTS FROM SINDA 

e Y 

INITIAL GUESS 

FOR TEMPERATURE 

AVERAGE BALL TEMPERATURI 

BETWEEN 

A 

8501 b . AXIAL PREL

-230°F INLET TEMP

0.5 FRICTION FACT

7.0 Ibs/sec. COOL

343% INCREASED HE

I CONVERGED 

I TEMPERATURE 

I 

I 

I 

I 

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4.0 ANALYSIS OBJECTIVES AND APPROACH 

The objectives of the SSME LOX turbopump pump-end bearing analysis were 

to investigate the sensitivity of bearing operating characteristics to 

variations in operating parameters. The operating parameters that were 

considered are 1 isted below with the values investigated. 

Coolant Fl ow Rate (I bs/sec) 3.6, 7.0 

Contact Friction Factor 0.2, 0.3, 0.5 

Inlet Cool ant Temperature to -240, -230, -218 

Bearing #1 (OF) 

Axial Preload (lbs) 350, 480, 850 

Outer Race Clearance (mi 1 s ) 2.6, 1.7, 1.0 

Outer Race Ti1 t (minutes) 0 through 42 

Heat Transfer Between Outer With, Without 

The flow rate, friction factor, inlet temperature, and preload were 

investigated in a1 1 combinations whi le holding the remaining parameters at 

their initial values. This resulted in 54 initial cases to be simulated. 

These cases are represented in the parameter data tree of Figure 4.1. This 

figure shows that 42 of the 54 cases were thermally unstable (diverged). It 

was not necessary to run computer simulations for all the cases to determine 

thermal instability. For example, if a case with an inlet coolant temperature of -218OF, flow rate of 7.0 Ibs/sec, preload of 850 Ibs, and friction 

factor of 0.2 was unstable, then cases for friction factor of 0.3 and 0.5 

with the other parameters the same would obviously not be stable. 

The effect of changes in coolant flow rate and inlet coolant temperature 

were evaluated using the 12 converged cases. However, more converged cases, 

over the full range of parameter values, were needed to properly evaluate 

changes in the remaining operating parameters. Thus, it was decided to 

increase the boundary heat transfer coefficient to obtain more cases that are 

thermal ly stable. 

Changing the boundary heat transfer coefficient introduced heat transfer 

as a new parameter into the sensitivity analysis. A second parameter data 

tree was developed for the increased heat transfer coefficients and uneval uated parameters. Figure 4.2 shows this parameter data tree, which has 18 
FIGURE 4.1 PARAMETER DATA TREE WITH NOfliNAL HEAT TRANSFER COEFFICIENT 

I I 

350 CONVERGEO 

-2l8(SAri*f 1480 CONVERGED 1 I 

850 OIVERGEO 

350 OIVERGEO 

-24O'f 

350 OIVERGEO 

-230'F 480 OIVERGEO 

COOL ANT 

F1O;'RI:I 

3.6 LBSISEC 

850 DIVERGED 

350 OIVERGEO 

-2I8(SAT)'F 480 DIVERGE0 

850 OIVERGEO 

350 OIVERGEO 

-240°F 1480 DIVERGED 

I -218(SAT)'F 480 DIVERGED 

8 WTER RACE CLEARANCE = 2.6 ntis 

8 OUTER RACE TILT = 0 nlNUTES 

8 UITH HEAT TWSFER BEWEEN OUTER RACE AN0 ISOLATOR 

I 

350 COHVERGEO 

-240'F 1480 CONVERGED 

850 OIVERGEO 

350 CONVERGEO 

I 480 CONVERGEO t I 

I I 850 OIVERGED 

350 CONVERGE0 

350 OIVERGEO 

-240'f 1480 OIVERGED I 

850 OIVERGEO 

I 

850 DIVERGEC 

350 OIVERGEG 

-230°F 480 DIVERGEG 

850 OIVERGEO 

350 DIVERGED 

-218(SAT)'F 480 OIVERGED 

850 OlVERGiO 

350 OIVERGEO 

-240°F / 480 DIVERGE0 
FIGURE 4.2 PARAMETER DATA TREE WITH ENHANCED HEAT TRANSFER COEFFICIENT 

I) COOLANT 

FLOWRATE 

7.0 LBS/SEC 

I) INLET 

COOLANT TEMP. 

-230" F 

FRICTION 

FACTOR 

0.2 

AXIAL PRELOA

350 CONVERGE

480 CONVERGE

HEAT TRANSFER I 1 850 CONVERGE

COEFFICIENT 

INCREASE 

245% 

I

350 CONVERGE

350 CONVERGE

@ OUTER RACE 

CLEARANCE = 2.6MILS 

480 CONVERGE

350 CONVERGE

850 CONVERGE

1 

350 CONVERGE

@ OUTER RACE 343% 

TILT = 0 MINUTES 

350 CONVERGE

@ WITH HEAT TRANSFER 

BETWEEN OUTER RACE AND ISOLATOR 1 480 CONVERGE

0.3 

0.5 

480 CONVERGE

850 CONVERGE
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more cases. The effect of varying friction factor and axial preload were 

properly evaluated using these cases. 

The heat transfer coefficient was increased to obtain converged cases 

over the range of parameter values. A thermally stable solution was obtained 

for a coolant flow rate of 7.0 lbs/sec, inlet temperature of -230°F, friction 

factor of 0.5, and preload of 480 Ibs with a 245 percent increase in heat 

transfer coefficient. An increase in heat transfer coefficient of 343 

percent was needed for the 850 1 b preloaded case. 

The variation of outer race clearance, outer race tilt, and thermal 

isolation of the isolator were evaluated separately, holding a1 1 other 

parameters constant. A nominal case with a flow rate of 7.0 lbs/sec, inlet 

temperature of -230°F, friction factor of 0.2, and preload of 480 lbs was 

used to evaluated these parameters. The 245 percent increased heat transfer 

was nee 

The 343 percent increased heat transfer coefficient was needed to evaluate 

the effects of changing outer race tilt. 
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5.0 ANALYSIS RESULTS 

The analysis of the SSME LOX turbopump included different loading 

conditions and bearing temperature profi 1 es. A mechanical analysis of the 

shaft/bearing system was performed using the SHABERTH computer model. The 

thermal analysis of the pump-end Number 2 (inboard) bearing was performed 

using the SINDA thermal model. The use of these computer models, in an 

interactive manner, enabled the comprehensive evaluation of the bearing 

performance. 

5.1 Mechanical Characteristics 

The mechanical analysis was performed using a shaft speed of 30,000 RPM 

and radial loads on the preburner impeller, main impeller, and turbine of 

382, 3000, and -220 lbs, respectively. These radial loads, illustrated in 

2606R1. The model used an unmounted diametrical clearance of 6.3 mils, inner 

race curvature of 0.55, and outer race curvature of 0.52 for the 45 mm 

inboard (Number 2) bearing. The model also simulates the outer races sliding 

with axial load. This was done by manipulating the turbine-end bearings so 

that they transmit very small axial loads to the pump-end bearings. 

The bearing operating characteristics were determined for a case using 

480 lbs preload, 7.0 Ibs/sec coolant flow rate, and friction factor of 0.2. 

Table 5.1.1 shows the bearing operating characteristics for a uniform 

temperature profile of -230°F. Table 5.1.2 shows the same case a t i t s steady 

state (converged) temperature profile. As can be seen from these tables, the 

pump-end bearings (Numbers 1 and 2) do not share an equal amount of the load. 

This is primarily due to the shaft deflecting from the large radial load 

placed between the pump-end bearings and the turbine-end bearings (Numbers 3 

and 4). Bearing 2 received about 72 percent of the load to the pump-end 

bearings with a uniform temperature profile. The bearing received about 76 

percent of the load, considering thermal effects. Bearing 2 received a 

greater portion of the load with the steady state temperature profile because 

of thermal growth of the ball, with respect to the races. The thermal growth 

of the ball decreased the internal clearances in the bearing which increased 

the axial load. The increased axial load caused the bearing to become 

"stiffer" and able to support a larger radial load. 
TABLE 5.1.1 BEARING OPERATING CHARACTERISTICS 

WITH UNIFORM TEMP. PROFILE 

480 LBS AXIAL PRELOAD 

6.3 MILS DIAMETRICAL CLEARANCE 

2.6 MILS OUTER RACE CLEARANCE 

TABLE 5.1.2 BEARING OPERATING CHARACTERISTICS 

CONSIDERING THERMAL EFFECTS 

480 LBS AXIAL PRELOAD 

6.3 MILS DIAMETRICAL CLEARANCE 

1.6 EllLS OUTER RACE CLEARANCE 
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5.2 

The SINOA thermal model predicted temperatures for each node of the 

bearing components. The program then calculates the volume average temperature for the main components. Tables 5.2.1 through 5.2.4 show the average 

component and maximum track temperatures for the cases studying preload, 

friction factor, inlet temperature, and flow rate. 

The calculated results with nominal heat transfer coefficient are shown 

in Table 5.2.1 and 5.2.2. The effect of preload and inlet coolant temperatures are shown for a friction factor of 0.2 with two flow rates considered. 

No stable solution was achieved for 850 lbs preload. Therefore, it was 

necessary to increase the heat transfer coefficient to obtain more stable 

cases. Table 5.2.3 presents the results for three friction factors and three 

preloads for cases where the heat transfer coefficient has been increased to 

245% of its nominal valu 

thermally stable solution for all cases except the case with 0.5 friction 

factor and 850 lbs preload. To obtain a thermally stable solution for this 

case, the heat transfer coefficient was increased by 343%. The results of 

the cases studied with 343% increase in heat transfer coefficient have been 

reported in Table 5.2.4. The maximum track temperature has been found to be 

719°F with a friction factor of 0.5 and a preload of 850 lbs. It should be 

observed that the higher track temperatures for stable cases occur with the 

higher heat transfer coefficients and friction coefficients. These conditions allow high local heating while providing the capability to remove 

sufficient heat to prevent thermal instability. 

Effect of Axial Preload 

The effect of varying axial preload on the 45 mm inboard bearing was 

- determined for friction factors of 0.2, 0.3, and 0.5. Figures 5.2.1 through 

5.2.3 show the effect of axial preload using the increased heat transfer 

coefficient of 245%. Figures 5.2.4 through 5.2.6 show the effect of preload 

using the 343% increased heat transfer coefficient. The figures show that 

axial preload does have a significant effect on operating temperatures. The 

effect of preload is increased as the friction factor is increased. At a 

friction factor of 0.2, the average ball temperature increases by 31°F with 
TABLE 5.2.1 COMPONENT TEMPERATURES FOR 

TEMPERATURES AT LOW FLOW A 

45 nn BEARING 

DIFFERENT IKET COOLAMT 

rTE (3.6 LBSISEC) 

COOLANT FLOWRATE = 3.6 LBSISEC 

;QICTION FACTOR = 0.2 

* THERMALLY UNSTABLE 
TABLE 5.2.2 COMPONENT TEMPERATURES FOR DIFFERENT INLET 

COOLANT TEMPERATURES AT HIGH FLOW RATE (7.0 LBS/SEC) 

45 w BEAR I NG 

COOLANT FLOWRATE = 7.0 LBS/SEC 

FRICTION FACTOR = 0.2 

* THERMALLY UN5TABLE 
TABLE 5.2.3 COMPONENT TEMPERATURES FOR D 

FACTORS WITH 245% INCREASE I1 

45 WI BEARING 

FFERENT CONTACT FRICTION 

HEAT TRANSFER COEFFICIENT 

FLOWRATE = 7.0 LBS/SEC 

INLET COOLANT TEMPERATURE -230°F 

INCREASE IN HEAT TRANSFER COEFFICIENT OF 245% 

* THERMALLY UNSTABLE 
TABLE 5.2.4 COMPONENT TEMPERATURES FOR DI FFERENCT CONTACT FRICTION FACTORS WI

INCREASE IY HCki TRANSFER COEFFICIENT 

45 HH BEARING 

FLOWRATE = 7.0 LBS/SEC 

INLET COOLANT TEMPERATURE = -230 F 

iNCREASE IN HEAT TRANSFER COEFFICIENT OF 343' 
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FIGURE 5.2.6 45mm BEARING OPERATING TEMPERATURES 

VS. AXIAL PRELOAD 

o 0.5 FRICTION FACTOR 

0 7.0 LBS/SEC COOLANT FLOW 

1 r -230°F INLET COOLANT TEMPERATURE 

INNER R

o INCREASED HEAT TRANSFER COEFFICIENT (343%) MAX IMUM 

TRACK T

AVERAGE TEMPER

OUTER R

INNER RA

300 400 500 600 700 800 9 00 

AXIAL PRELOAD (LB) 
SRS Technologies 

an increase of preload from 350 to 850 lbs. At a friction of 0.5, the 

average ball temperature increased by 226°F with an increase in preload from 

350 to 850 I bs. The dashed line in Figure 5.2.3 indicates that the 850 Ib 

preload case did not have a thermally stable solution. 

Effect of Friction Factor 

The effect of increasing the friction factor for preloads of 350, 480, 

and 850 lbs and the two increased heat transfer coefficients are shown in 

Figures 5.2.7 through 5.2.12. The results show that contact friction has a 

large effect on bearing operating temperatures. An increase in friction 

factor from 0.2 to 0.5 increased the average ball temperature by 123°F for a 

preload of 350 1 bs and 318°F with a preload of 850 lbs. 

The more severe cases involving the higher friction factors and preloads 

were thermal ly unstable using the nominal heat transfer coefficient. Thus, 

the heat transfer coefficient was increased to obtain stable solutions over 

the range of friction factors and prel oads considered. Figures 5.2.13 

through 5.2.15 show the results of increasing the heat transfer coefficient. 

The temperatures decrease with increasing heat transfer coefficient. The 

heat transfer coefficient had a larger effect on the average ball temperature 

than it did on the average inner and outer race temperatures. This was 

because the ball has a larger surface to volume ratio exposed to the fluid 

than do the inner and outer races. Thus, the ball had a larger heat loss 

increase than did the inner or outer race as the heat transfer coefficient 

for surface to fluid increased. 

The Dittus Boel ter equation was used to determine the heat transfer 

coefficients for the inner and outer races. The Katsnellson equation was 

used for the ball. The properties were evaluated at the film temperature 

which is the average of the surface and saturation temperature. 
FIGURE 5.2.7 45llM BEARING OPERATING TEMPERATURES VS. FRICTION FACTOR 

r 350 LB AXIAL PRELOAD 

r 7.0 LBSISEC COOLANT FLOW 

o -230°F INLET COOLANT TEMP. 

r INCREASED HEAT TRANSFER 

COEFFICIENT (245%) 

OUT

MAX 

TRA

AVE

BAL

OUT

INN

FR I CT ION FACTOR 
FIGURE 5.2.8 45MM BEARING OPERATING TEMPERATURES VS. FRICTION FACTOR 

6 480 LB AXIAL PRELOAD 

o 7.0 LBSISEC COOLANT FLOW 

e -230°F INLET COOLANT TEMP 

e INCREASED HEAT TRANSFER 

COEFFICIENT (245%) 
FIGURE 5.2.9 45mm BEARING OPERATING TEMPERATURES VS. FRICTION FACTOR 

r 850 LB AXIAL PRELOAD 

B 7.0 LBS/SEC COOLANT FLOW 

B -230 "F INLET COOLANT TEMPERATURE 

e INCREASED HEAT TRANSFER COEFFICIENT (245%) 

INNER RACE 

/ 

MAX I MUM 

0 TRACK TEMP. 

/ 

/ 

/ 

/ 

I 

I 

- , --OUTER Rk 

A/-- C - --C 

(f = 0.5 

THERMALLY 1 UNSTAEIL E ) I 

I 

I 

1 1 I I 

0.2 0.3 0.4 0.5 

FR:CTION FACTOR 

r l GURE 5.2.11 45MM BEARING OPERATING TEh!PERATURES VS. FRICTION FACTOR 

480 LB AXIAL PRELOAD 

7.0 LBS/SEC COOLAND FLOW 

-230°F INLET COOLANT TEtfP 

INCREASED HEAT TRANSFER 

COEFFICIENT (343%) 

OU

RA

MA

TR

TE

TE

BA

OU

RA

IN

RA

FRICTION FACTOR 
F I GURE 5.2.12 4.5HM BEARING OPERATING TEMPERATURES VS . FRICTION FACTOR 

800 0 850 LB AXIAL PRELOAD 

7.0 LBSISEC COOLANT FLOW 

0 -230 OF INLET COOLANT TEMP. 

0 INCRASED HEAT TRAFSFER 

COEFFICIENT (343%) 

600 

h 

LL 

0 

V 

v, 

400 

3 

I2 

W 

% 

2? 

I200 AVERAGE 

0 

-200 

FRICTION FACTOR 
FIGURE 5.2.13 45 tltl BEARING OPERATING !TEMPERATURES VS. HEAT TRANSFER COEFFICIENT 

e 350 LB AXIAL PRELO

,a -230°F INLET COOLA

e 7.0 LBSISEC COOLAN

e 0.2 FRICTION FACTO

i 

--'"--- 

NOMINAL 245% 

INCREASE 

343

INCREA

HEAT TRANFER COEFFICIENT 
FIGURE 5.2.14 

45MM BEARING OPERATING TEMPERATURES US. HEAl 

I 

NCMINAL 

TRANSFER COEFFICIENT 

[IAL PRELOAD 

ILET COOLANT TEMP. 

;EC COOLANT FLOW 

'ION FACTOR 

O

F

T

A

B

O

I

HEAT TRANSFER COEFFICIENT 
FIGURE 5.2.15 

45MM BEARING OPERATING TEMPERATURES VS. HEAT TRANSFER COEFFICIENT 

(NOMINAL 

CASE 

THERMALLY 

UNSTABLE) 

NOMINAL 

850 LB AXIAL PRELOAD 

-230°F INLET COOLANT TEMP. 

7.0 LBSISEC COOLANT FLOW 

0.2 FRICTION FACTOR 

-OM

T

A

B

O

245% 343% 

INCREASE INCREA

HEAT TRANSFER COEFFICIENT 
SRS Technologies 

Effect of Coolant Flow Rate 

Figures 5.2.16 through 5.2.19 show the effect of changing the coolant 

flow rate. The operating temperatures do not change much over the range of 

flow rates studied. However, an increase in coolant flow rate could prevent 

a marginal case from diverging. 

Effect of Inlet Coolant Temperature 

The effects of varying inlet temperatures are shown in Figures 5.2.20 

through 5.2.23. Two subcooled cases (-240°F and -230°F) and a saturated case 

(Tsat = -214°F) were considered. The coolant was introduced at Bearing 1 and 

was allowed to increase in temperature while passing through the bearing. 

For the saturated case, the coolant was introduced at -218°F and was allowed 

to increase to the saturation temperature of -214°F when entering Bearing 2. 

nlet coolant temperature did not have a very significant effect on 

the bearing operating temperatures. But, lowering the inlet temperature 

could cause a case that would be thermally unstable to become stable. 

Effect of Outer Race Misalignment 

Angular misalignments of the outer race up to 31.5 minutes were studied 

to determine the effect on bearing operating temperatures. The outer race 

was tilted so as to place the heaviest loaded ball in the "pinch point". 

Consequently, the misalignment contributed to the radial loading of the ball. 

Figure 5.2.24 shows the effect of outer race misalignment for a case using 

480 lbs preload, 7.0 Ibs/sec flow rate, 0.2 friction factor, -230°F inlet 

temperature, and 343% increased heat transfer coefficient. The maximum outer 

race tilt, that produced a stable solution, was about 29 minutes. This tilt 

caused the highest stable operating temperatures observed in this analysis. 

The maximum outer race track temperature was found to be 1858°F. The tilt 

caused the outer race to heat and expand more than the inner race. Thus, the 

bearing is able to maintain a sufficient operating clearance for higher ball 

temperatures. 

Effect of Outer Race to Isolator Clearance 

Clearances of 2.6, 1.7, and 1.0 mils between the outer race and isolator 

were investigated. The heat transfer coefficient was increased 245% of the 
FIGURE 5.2.16 

45MM BEARING OPERATING TEMPERATURES VS COOLANT FLOWRATE 

r '350 LB AXIAL. PRELOAD 

r -240°F INLET COOLANT TEMP. 

r 0.2 FRICTION FACTOR 

NOMINAL HEAT TRANSFER 

COEFFICIENT 

AVERA

\I I I I I I 1 

1.0 2.0 3 .O 4.0 5.0 6.0 7.

COOLANT FLOW (LBS/SEC) 
FIGURE 5.2.17 45MFI BEARING OPERATING TEMPERATURES VS COOLANT FLOWRATE 

e 350 LB AXIAL PRELOAD 

e -230 OF INLET COOLANT TEMP. 

e 0.2 FRICTION FACTOR 

e NOMINAL HEAT TRANSFER 

COEFFICIENT 

O

M

T

AVERAGE 
(j,) 3dnlVH3dW31 
FIGURE 5.2.19 45MM BEARING OPERATING TEMPERATURES YS INLET COOLANT FLOWRATE 

r 480 LB AXIAL PRELOAD 

r -230°F INLET COOLANT TEMP. 

r 0.2 FRICTION FACTOR 

r NOMINAL HEAT TRANSFER 

COEFFICIENT 

AVERAG
FIGURE 5.2.20 45Ml BEARING OPERATING TEMPERATURES VS. INLET COOLANT TEMPERATURE 

a 350 LB AXIAL PRELOAD 

a 7.0 LBS/SEC COOLANT FLOW 

r 0.2 FRICTION FACTOR 

r NOMINAL HEAT TRANSFER 

COEFFICIENT OUTER R

AVERAGE 

INLET COOLANT TEMPERATURE (OF) 
FIGURE 5.2.21 45MM BEARING OPERATING TEMPERATURES VS . INLET COOLANT TEMPERATURE 

0 480 LB AXIAL PRELOAD 

7.0 LBSISEC COOLANT FLOW 

0 0.2 FRICTION FACTOR 

NOMINAL HEAT TRANSFER 

COEFFICIENT 

I_________ 

BALL 

MAX I MU

TRACK 

AVERAG

0 

AVERAGE TEMP. 

-200 

INLET COOLANT TEMPERATURE (" F) 
FIGURE 5.2.22 45MM BEARING OPERATING TEMPERATURES VS. INLET COOLANT TEMPERATURE 

3 00 e 350 LB AXIAL PRELOAD 

r 3.6 LBS/SEC COOLANT FLOW 

e 0.2 FRICTION FACTOR 

r NOMINAL HEAT TRANSFER 

COEFFICIENT 

O

100 

- 

L. 

0 

- AV

W 

CZ 

3 

t-- 

Q 0 CY 

W 

a 

Z 

W 

+ 

- 100 

-200 

- 3 00 

-240 -220 

INLET COOLANT TEMPERATURE ( 'F) 
FIGURE 5.2.23 45~~ BEARING OPERATING TEMPERATURES VS. INLET COOLANT TEMPERATURE 

300 e 480 LB AXIAL PRELOAD 

r 3.6' LBS/SEC COOLANT FLOW 

e 0.2 FRICTION FACTOR 

r NOMINAL HEAT TRANSFER COEFFICIENT 

BALL 

AVERAGE 

OUTER 

RACE 

' INNER 

RACE 

INLET COOLANT TEMPERATURES (OF) 
o 480 LBS PRELOAD 

o 7 LBSISEC COOLANT FLOW 

o 0.2 FRICTION FACTOR 

o -230 F INLET COOLANT 

o 3432 INCREASE IN HEAT 

TRANSFER COEFFICIENT. 

1200 

800 

AVE

OUTER RACE MISALIGNMENT (MINUTES) 
n~minal value to obtain stable solutions over the range of parameters 

studied. The results are shown in Table 5.2.5 and Figure 5.2.25. The 

.clearance, over the range studied, had a small effect on the bearing operating temperatures. The average ball temperature increased 46OF for a decrease 

in clearance of 2.6 to 1.0 mi 1s at a preload of 850 I bs. However, these are 

radial loaded conditions and the effects could be more serve under axial 

load. The bearing outer race could be restricted from moving in the axial 

direction since the operating clearance between the outer race and isolator 

is zero for these conditions. 

Effect of Thermal Isolation of the Outer Race 

The effect of thermally isolating the outer race from the isolator was 

determined. The results are shown in Table 5.2.6. The temperature increase 

of the heat generated is transferred to the coolant and very 1 ittle is transferred through the 

bearing isolator. 

Cool ant Qua1 i ty 

The coolant qua1 i ty profile was determined for the saturated cases using 

flow rates of 7.0 and 3.6 lbs/sec. The profiles are shown in Figures 5.2.26 

and 5.2.27, The quality was calculated as the mass of vapor per total mass 

of coolant. The mass of vapor was determined from the amount of heat 

transferred to each fluid node. The fluid quality was decreased by 

approximately 70% by increasing the flow rate from 3.6 to 7.0 lbs/sec. 

Figure 5.2.26 shows that no vapor was generated in Bearing 1 for the higher 

flow rate. 
TABLE 5.2.5 EFFECT OF OUTER RACE TO ISOLATOR CLEARANCES (45mm BEARING

Flowrate = 7.0 Ibslsec 

Friction Factor = 0.2 

245:! Increase in Heat Transfer Coefficient 

Inlet Coolant Temperature = -230°F 
-J W U a 

3 

at04: 

4w 

2 a 

051 

cnw 

-IotI-Z W WZ 

UO 

an 

ar z 

0 

av 

W 
TABLE 5.2.6 EFFECTS OF THERMAL ISOLATION ( i BEARING ISOLATOR (45mm BEA

Flowrate = 7.0 1 bs/sec 

Friction Factor = 0.2 

245% Increase in Heat Transfer Coefficient 

Inlet Coolant Temperature = 230°F 
FIGURE 5.2.26 FLUID QUALITY THR 

7.0 LBS/SEC COOLANT FLOWRATE 

480 LBS AXIAL PRELOAD 

JGH BEARING #2 

BEARING 2 

I 

FLUID NODE LOCATION

FLUID NODES 
FIGURE 5.2.27 FLUID OUALITY THROUGH BEARING #2 

3.6 LBS/STC C30LANT FLOW RATE 

480 LBS AXIAL PRELOAD 

BEARI

FLUID NODE 

FLUID NODES 
SRS Technologies 

6.0 CONCLUSIONS AND RECOMMENDATIONS 

The turbopump bearing component temperatures are very sensitive to load, 

contact friction, and fluid/boundary heat transfer coefficients. Alignment 

of the outer race was found to be critical as bearing temperatures are very 

sensitive to this parameter. For the majority of the conditions evaluated, 

the operating clearance between the outer race and isolator was lost. 

Although, for the radially loaded conditions, the effect of isolator 

clearance on bearing temperatures was nominal, the loss of clearance would be 

considerably more detrimental for axial load transients. While not as 

sensitive to coolant flow (according to analysis) and coolant quality, these 

parameters can become important for marginal operating conditions. According 

to the analysis, the internal operating clearance of the 45 mm bearing must 

be at least 53.9 mils to maintain a thermally stable operating condition. 

coefficients, further work should be done to investigate the possible effects 

of the internal flow field characteristics or fluid/boundary heat transfer-. 

In addition, recent tests of the Bearing and Seal Materials Tester showed a 

higher dependence of bearing temperature on cool ant f 1 ow than was predicted 

by analysis. This inconsistency in the analysis should be investigated. 

Improved methods for cooling the bearings, i .e., under race cooling, 

introducing the coolant between the bearings, etc., should be modeled and 

eval uated for improved efficiency in cool i ng the LOX turbopump bearings. 
555 SPARKMAN DRlVElSUlTE 1406 

HUNTSVILLE, ALABAMA 35816.3425 

(205) 830-0375 

SYSVEMS TECHNOLOGY DIVISION 

January 6, 1986 

OL86- 1249 

TO: DISTRIBUTION 

SUBJECT: Final Report for Space Shuttle Main Engine (SSME) LOX Turbopump 

Pump-End Bearing Analysis 

PREPARED FOR: Mr. Fred Dolan 

Materials and Processes Laboratory 

Engineering Physics Division 

George C. Marshall Space Flight Center 

Marshall Space Flight Center, AL 35812 

SUBMITTED BY: SRS Technol ogies 

echnol ogy Division 

man Drive; Suite 140 

Huntsvi 1 le, AL 35805 

CONTRACT NO. : NAS8-36183 Modification No. 6. 

DATE OF PUBLICATION: January 6, 1986 

The enclosed R&D Final Report provides a description of the work performed 

under the subject contract modif ications. 

Sincerely, 

SRS TECHNOLOGIES 

Systems Technol ogy Division '\ 

Joseph C. Cody 

Enclosures : as stated 

project ~ngineer 
DISTRIBUTION: George C. Marshal l Space Fl ight Center 

Mr. Fred J. Dolan, EH14 (6 Copies + Repro.) 

Mr. Schwinghamer, EHOl 

Mr. Riggs, EP23 

Mr. McCarty, EP21 

Mr. Geotz, EE51 

Mr. Lombardo, SA53 

Mr. G. Smith, SA51 

Mr. Lovingood, SA51 

AT0 1 

AS240 (3 Copies) 

EHll 

AP29-F 

EM13B-21 

BF30 

CCOl/Wof ford 

NASA Scientific and Technical Information Center 

ATTN: Accessioning Department (1 Copy + Repro.) 

DCASMA 
SRS Technologies 

FOREWORD 

This report was prepared by SRS Technologies under Contract NAS8- 

36183 Modification Number 6 for the George C. Marshall Space Flight Center 

of the National Aeronautics and Space Administration. The work was 

administered under the technical direction of the Engineering Physics 

Division of the Materials and Processes Laboratory with Mr. Fred F. Dolan as 

Project Manager. This report describes the work accompl ished in accordance 

with Contract Modification Number 6. Mr. Joseph C. Cody was the SRS Project 

Engineer. A list of the key contributors is shown below. 

Mr. David Marty 

Dr. Alok Majumdar 

Mr. Bruce K. Tiller