Copyright ©1996, American Institute of Aeronautics and Astronautics, Inc.

AIAA Meeting Papers on Disc, July 1996A9637224, AIAA Paper 96-3103

Investigation of hydrostatic bearings operating in a turbulent, compressibleliquid

Philip C. PelfreyPratt & Whitney, West Palm Beach, FL

Vishnu M. SishtlaCarrier Corp., Syracuse, NY

AIAA, ASME, SAE, and ASEE, Joint Propulsion Conference and Exhibit, 32nd, Lake

Buena Vista, FL, July 1-3, 1996

An experimental investigation of hydrostatic bearings for use in high-pressure cryogenic rocket engine liquidhydrogen turbopumps and commercial chiller compressors is presented. A test facility was developed andfabricated using an environmentally friendly refrigerant as the working fluid. A fluid bulk compressibility of nearlyone half that of liquid hydrogen has been achieved. A hydrostatic bearing test rig was developed and fabricated tomeasure the static performance and the frequency dependent direct and cross-coupled dynamic properties. Testinghas demonstrated capabilities up to 1500 psia bearing supply pressure and applied dynamic loads up to 1000pounds force and excitations up to 1000 Hz yielding bearing coefficients with less than 20 percent uncertainty. Inaddition to obtaining bearing performance, the test program is also evaluating the effects of bearing geometry andoperating conditions on pneumatic hammer instability. These test data will enable the development and/orvalidation of a comprehensive pneumatic hammer instability criteria for bearings operating in a compressibleliquid. (Author)

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AIAA-96-3103-

INVESTIGATION OF HYDROSTATIC BEARINGSOPERATING IN A TURBULENT, COMPRESSIBLE LIQUID

Philip C. Pelfrey*Pratt & Whitney

Government Engines & Space PropulsionWest Palm Beach, FL 33410-9600

Vishnu M. Sishtlat

Carrier CorporationSyracuse, NY 13039

An experimental investigation of Ahydrostatic bearings for use in high-pressure Bcryogenic rocket engine liquid hydrogen Cturbopumps and commercial chiller compressors is Dpresented. A test facility was developed and Gfabricated using an environmentally friendly Hrefrigerant as the working fluid. A fluid bulk Kcompressibility of nearly one half that of liquid Lhydrogen has been achieved. A hydrostatic Mbearing test rig was developed and fabricated to Nmeasure the static performance (leakage, power Ploss, and load capacity) and the frequency Vdependent direct and cross coupled dynamic Xproperties (stiffness and damping). Testing has Ydemonstrated capabilities up to 1,500 psia bearing hsupply pressure and applied dynamic loads up to k1,000 pounds force and excitations up to 1,000 Hz nyielding bearing coefficients with less than 20 tpercent uncertainty. In addition to obtaining Pbearing performance, the test program is also pevaluating the effects of bearing geometry andoperating conditions on pneumatic hammer subscriptsinstability. This test data will enable thedevelopment and/or validation of a comprehensive a ambientpneumatic hammer instability criteria for bearings f filmoperating in a compressible liquid. Testing will r recessalso investigate the effects of instrumentation s supplycapacitance on the bearing's dynamic x x-directionperformance. y y-direction

Nomenclature

area or response matrixexcitation vectordampingbearing bore diameterflow ratestability response functionstiffnessbearing lengthapparent massrotor speedpressurevolumex-directiony-directiondepthfluid bulk compressibilitynumbertime1/kdensity

'Project Engineer, Rocket Component Design TechnologyMember AIAAf Senior Stall Engineer, Centrifugal Chiller DevelopmentCopyright © 1996 American Institute of Aeronauticsand Astronautics, Inc. All rights reserved.

Introduction

Externally fed, orifice compensated,hydrostatic journal and thrust bearings are beingimplemented for use as the primary rotor supportsystem for rocket engine liquid hydrogenturbopumps. They are also being considered foruse in direct drive, oil-less compressors forcommercial chillers. These bearings offer manyadvantages over conventional rolling elementbearings. Fluid film bearings, unlike rollingelement bearings, have no DN (bore diameter inmm X rotor speed in rpm) limitations and are notprone to stress-corrosion cracking, rolling-contactfatigue, or bulk-ring fractures due to material ormanufacturing defects. Free of these constraints,the rotor's speed can be increased for higheroperating efficiency. Increased operating speedalso helps to reduce the size and weight of theturbomachinery. Additional desirablecharacteristics of fluid film bearings includemechanical simplicity , accuracy of rotor position,high or low stiffness, lack of rotating-to-static partcontact during steady-state operation, andexcellent damping characteristics for improvedrotor dynamic performance. These key featuresenable significant reductions in part count andcomplexity, the use of unshrouded impellers, andsmaller seal clearances resulting in increasedperformance and longer life at a reduced cost. Forcommercial chillers, they also enable theelimination of the oil-lubricated gearbox andbearings for improved mechanical and heattransfer efficiency.

This technology is relatively mature forbearings operating in an incompressible fluid;however, data is lacking for bearings operating in acompressible liquid. This data is necessary tovalidate design codes for bearings operating inliquid hydrogen and liquid HFC-134a, in whichcompressibility effects are significant. There isalso a need for a validated, comprehensivepneumatic hammer instability criteria in order todesign stable high-stiffness, high-load-capacitybearings operating in a compressible liquid.

A self-contained fluid film bearing and sealtest rig was developed to evaluate advanced liquidrocket engine rotor support systems1. The test rigwas designed to accommodate a variety of fluid-film devices including hydrostatic bearings, foil

bearings, and damper seals. Other test articleconfigurations include the ability to test magneticbearings, ball bearings, and roller bearings. Thehydrostatic bearing configuration of the tester isshown schematically in Figure 1. The figureillustrates the test rig with both thrust bearing andjournal bearing test articles installed.

Figure 1. Bearing Test Rig

Radial rotor loads are transferred toground by the two outboard slave journal bearingswhich are mounted with interference fits within themain housing. Axial rotor loads are reacted by thehydrostatic thrust bearing or by a pneumatic thrustbearing located at the turbine end of the shaftwhich is activated only during journal bearingtesting. The rotor is driven by a Terry turbinewhich is integrally machined into the one piecerotor. The turbine is capable of providing 100 hpwhen supplied with gaseous nitrogen at 70 °F and1,000 psia supply pressure. The liquid and gascavities of the test rig are separated by a labyrinthbuffer seal/mixing cavity system.

The test journal bearing is installed in a"squirrel cage" assembly which supports thebearing axially yet is radially soft. The bearing issupported radially by four loaders — two pneumaticand two hydraulic. The pneumatic loaders arecapable of applying up to 3,000 pounds of static

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load each. They are oriented orthogonally and areused to set the eccentricity of the test article in theX and Y directions. The hydraulic loaders aremounted opposite the pneumatic loaders and arecapable of applying dynamic loads up to 1,000pounds peak at 1,000 Hz. The hydraulic shakersare used to generate independent shaft/bearingorbits at prescribed frequencies for calculation ofthe bearing frequency dependent dynamiccoefficients.

The test rig includes 51 sensors which areused for performance measurements, calculations,and health monitoring. The 51 sensors arecomprised of 26 steady state pressuremeasurements, 15 steady state temperaturemeasurements, four high frequency radialproximity probes, two high frequency load cells,two high frequency accelerometers, one axialproximity probe, and one proximity probeinterfaced to a tachometer driver which is used asa speed sensor. Fifteen of the sensors are locatedinternally to the test article - thirteen pressuresensors provide axial and circumferential pressuregradients within the pockets and on the lands whiletwo thermocouples provide the fluid temperature inthe pocket and on the land. All instrumentation iscalibrated for magnitude and the dynamicinstrumentation is calibrated for magnitude andphase as a function of frequency.

The rig is rated for a maximum allowableworking pressure of 5,000 psia and a maximumspeed of 83,500 rpm (when used with a 3 inchdiameter shaft). Bearings and seals with borediameters as small as one inch and as large as sixinches can be tested. The entire rig is made ofAMS 5664, a dimensionally stable heat treat ofInconel 718. The rig materials were selected toensure compatibility with the refrigerant HFC-134a,water, liquid nitrogen, liquid oxygen, and liquidhydrogen. By making the rig of one material, fitsand clearance changes caused by variousoperating temperatures are minimized.

Test Facility

All performance testing is being carried outat Pratt & Whitney's $200 million space propulsiontest facilities located on 7,000 acres in West PalmBeach, Florida. The test stand has beendesignated E-32. The facility consists of the testrig, a closed-loop, self-contained HFC-134arefrigerant system, a separator, and the data

acquisition/control system. A schematic of the testfacility is shown in Figure 2. Table 1 summarizesthe fluid capabilities of the test facility.

Figure 2. E-32 Test Facility

FluidGN2

HFC-134aWater

Capacity3 Ibjsec @ 2,500 psiaISIb^sec @ 1,500 psia55 IbVsec @ 125 psia

Table 1. Fluid Capabilities at E-32

Within the test rig, nitrogen is used to drivethe turbine, pressurize the shaft buffer seal, andpressurize the gas thrust bearing. External to thetest rig, nitrogen is used to pressurize the steady-state pneumatic loaders, load regulators, andactuate control valves. HFC-134a is the workingfluid supplied to the bearings and water is used toreject the heat from the facility and test rig.

The closed-loop HFC-134a system wasdesigned and fabricated by Carrier Corporation.The major components of this system are thewater cooled condenser, the pump/motor/starter,and the interconnecting piping. The water cooledcondenser serves as the storage reservoir for therefrigerant and provides 85 tons maximum heat

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rejection. Sub-cooled liquid HFC-134a is fed to thepump inlet from the bottom of the condenserthrough a 25 micron filter with a filtration ratio of200. A centrifugal pump (single stage inducer/impeller) driven by a 250 hp motor through agearbox provides a continuous flow rate of 15Ibjsec at 1,500 psia and 100 °F (at nominalcooling water flow rate, higher or lowertemperatures are possible by varying the waterflow rate). The pump discharge pressure ismaintained by a back-pressuring regulator which isinstalled in a by-pass leg which returns liquid to thecondenser. The slave bearings are supplied withpump discharge pressure. The test bearing supplypressure is varied with a control valve to anypressure between zero (fully closed) and pumpdischarge (fully open). As the test bearing supplyvalve is opened or closed, the back-pressuringregulator in the by-pass leg closes or opens tomaintain the constant pump discharge pressure.

The test rig buffer seal/mixing cavityprovides positive separation between the HFC-134a bearing cavities and the gaseous nitrogenturbine drive system. However, the mixing cavitycontains a mixture of liquid HFC-134a, gaseousHFC-134a, and gaseous nitrogen. A highefficiency, microprocessor based separator systemvents the nitrogen to atmosphere and returns liquidHFC-134a back to the condenser.

A single personal computer is used toacquire all steady-state data, conduct real-timeperformance calculations, display acquired andcalculated data real-time, record acquired andcalculated data to disk, and control the entirefacility with a point and click graphical userinterface. A total of 86 data channels are acquired(51 internal rig measurements and 35 facilitymeasurements). An additional 120 headers arecalculated from the acquired data. All 206 headersare displayed real time and recorded to disk at 2Hz. The personal computer also performs all limitchecking. Alarm options include changing thecolor of a value being displayed, displaying avisual message, an audible tone, a speed trip(simultaneous closing of the turbine supply valveand venting of all external loads), and an abort(simultaneously executing a speed trip and shutdown of the HFC-134a system).

Test Objectives

The primary objective of this test programis to measure the static and dynamic performanceof a variety of externally pressurized, orificecompensated, hydrostatic journal and thrustbearings operating in a compressible liquid tovalidate design codes and develop acomprehensive pneumatic hammer instabilitycriteria. Secondary objectives include measuringthe effects of capacitance from the volume of fluidin the instrumentation lines and testing potentialstability enhancing bearing designs. The testprogram has been designed to meet the needs ofthe rocket and commercial chiller industries.

Test Matrix

A minimum of nineteen different geometricconfigurations will be tested. These include sixjournal bearings with traditional parameters varied,six journal bearings which incorporate stabilityenhancements through geometry and surfacefinishes, six thrust bearings which also testtraditional parameters, stability enhancinggeometry, and surface finishes, and one journalbearing which will test a P&W proprietary stabilityenhancement. Each bearing is tested at threeeccentricity ratios (0.0, 0.25, and 0.5) for journalbearings or three clearances (0.0005, 0.002, and0.004 inches) for thrust bearings, three supplypressures (1,000, 1,250, and 1,500 psia for rocketbearings and 25, 75, and 150 psid for chillerbearings), and three speeds (16, 28, and 40 krpmfor rocket bearings and 10, 13, and 16 krpm forchiller bearings).

The test bearing designs are expected tocover a range from stable, to marginally stable, tounstable. With the marginally stable designs, it isexpected that stable or unstable operation can beachieved by varying the fluid compressibility bycontrolling the supply pressure and temperature.These tests, along with the capacitance testing,will result in some geometries being tested morethan once. One test point consists of obtaining thestatic and frequency dependent dynamic bearingperformance at a given supply pressure andtemperature, speed, and eccentricity or clearance.The entire test matrix consists of 624 test points.

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Pneumatic Hammer Instability

Self-excited instability or "pneumatichammer" is a major concern in the design ofhydrostatic bearings which use a compressibleliquid as the working fluid. Pneumatic hammer is aself-sustained fluctuation in the bearing recesspressure. Since the recess pressure is partiallyresponsible for supporting the static and dynamicloads of the shaft, any fluctuations in the pressurewill translate into shaft movements which can leadto internal damage of the turbopump orcompressor. Pneumatic hammer is alsocharacterized by a loss of dynamic damping withinthe bearing2.

The factors governing the onset of self-excited instability are the recess area, recessdepth, film thickness, lubricant compressibility,supply pressure, and ambient pressure around thebearing. In most cases, the film thickness is set toachieve a desired flow rate and is limited bymachining and assembly tolerances. Ambientpressure around the bearing is often dictated bythe secondary flow path limitations of themachinery. Static and dynamic loading of thebearing will require a minimum level of supplypressure, while the maximum supply pressure isset by the pump or compressor output. Giventhese restrictions, the bearing geometry is usuallymodified to ensure stable operation of the bearing.

Past stability criteria have been based on"rule of thumb" experience from gas-lubricatedbearings3:

(1)

The product of the first two terms represents thedimensionless compressibility which is 1.0 forperfect gases, approximately 0.1 for liquidhydrogen and 0.025 for liquid oxygen. This criteriahas been extended to replace the "Z-Factor"compressibility with the fluid bulk compressibility":

{nr(hrAr + Vs)/.(nDLhf )](PS - Pfl )|3 «0.02(2)

Table 2 provides some typical values of bulkcompressibility for several fluids of interest basedon:

dV (3)

FluidLH2

HFC-134aLN2LOXH2O

k(lb/in2)9,82217,08035,06062,830299,047

Table 2. Bulk Compressibilities

The stability criterias presented inequations (1) and (2) both use the ratio of recessvolume to land film volume as an indication ofstability for operation at any given fluid conditions.Therefore, according to these criterias, bearingswith various recess areas and depths would havethe same stability if the ratio of recess volume toland film volume was the same.

Pratt & Whitney has developed a thirdcriteria based on achieving an overdamped flowsystem:

H<f(Gf}

(4)

The criteria is based on the ratio of the change inmass flow through the orifice and lands as afunction of pocket pressure to the change in massflow across the lands due to a change in operatingclearance compared with a fluid compressibilityresponse function within the recess.

As part of this test program, a variety ofbearings will be tested which have unique recessareas and depths but which maintain a constantratio of recess volume to land film volume.Additionally, the effects of orifice geometry andorientation on stability will also be evaluated. Theresults of these tests will enable the generationand/or validation of a comprehensive pneumatichammer instability criteria.

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Steady State Performance

Steady state performance measurementsinclude load capacity, static stiffness, total bearingleakage, individual pocket leakages, and dragtorque in the laminar and turbulent flow regime.The maximum Reynolds number tested was100,000.

Load capacity is measured at twoeccentricity ratios for journal bearings (0.25 and0.5) and three clearances for thrust bearings(0.0005, 0.002, and 0.004 inches). Radial steadystate loads are applied with pneumatic actuatorswhich are operated in tension to pull the testjournal bearing to the desired eccentricity ratio.The applied load is calculated based on theactuator pressure times the piston area (less thepiston rod). Axial loads are applied by pressurizingthe top cavity of the test rig. The relative bearing-to-shaft clearance is measured with KamanKDM7200-1UEP thermally compensated proximityprobes with a calibrated resolution of 0.05 mils.This data is also used to calculate the staticstiffness.

Total bearing leakage is calculated using acalibrated sharp-edged orifice located external tothe test rig in the bearing supply line. Fluidpressure, temperature, and pressure drop acrossthe orifice are measured. The test bearing flowrate and the slave bearing flow rates are allcalculated using this method. Additionally, a flowmeter measures the total flow rate provided to allthree bearings. The total flow to all three bearings,as measured by the flow meter, is equal to the sumof the three bearing legs as calculated by thesharp-edged orifices.

Individual pocket leakage rates arecalculated based on the pressure differentialacross the bearing supply orifice. Internal riginstrumentation measures the bearing supplypressure and temperature immediately upstreamof the bearing orifice. Additionally, each bearingpocket has at least one static pressuremeasurement (oriented at the same location inevery pocket). One effective orifice dischargecoefficient is selected so that the sum of all pocketleakage rates equals the total flow rate to thebearing.

Prior to bearing performance testing, thetest rig was operated without a test bearinginstalled to measure drag torque "tare" values forconcentric operation as a function of speed.Internal rig instrumentation measures the turbinesupply and vent pressures and temperatures.Additionally, the turbine mass flow rate iscalculated using an external sharp-edged orifice inthe facility supply line. However, the flow rate canalso be calculated from the internal rig supplyconditions since the turbine is driven by sonicnozzles. Using this data along with the operatingspeed, turbine power and torque are calculated.Assuming the drag torque is independent ofoperating eccentricity (for eccentricity ratios up to0.5), the "tare" value is subtracted from the testdata to determine the power loss and drag torquedue to the test article.

Dynamic Performance

This research identifies the dynamiccharacteristics (with associated uncertainties) of aparticular test bearing from measured quantities.Uncertainties in the bearing coefficients arederived from the uncertainties of the individualmeasured quantities. The test rig is modeled as atwo degree of freedom system. The dynamiccoefficients are expressed in terms of anequivalent spring and damper system with cross-coupling between the X and Y coordinates. Therelationship between the dynamic forces andbearing motions for this model is given by:

01 1X0[MV \ + [C]\

(5)

where fx(t) and fy(t) represent applied forces whichresult in relative motion between the bearing andshaft. M, C, and K correspond to the fluid filminertia, damping, and stiffness matrices,respectively.

If journal motion is restricted to smalldisplacements about a steady operating point inthe bearing, the above equation can be linearizedby neglecting local variations in the coefficients.The dynamic coefficients cannot be directlymeasured; they are derived. The sinusoidalexcitation method was chosen due to thestraightforwardness of its implementation, its

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repeatability, and because comprehensiveuncertainty analyses indicate its ability to providelow uncertainty results. This method permits theuse of direct data averaging in the time domainover many cycles of operation, thus leading to areduction in the effects of noise and other phaseindependent disturbances.

Two linearly independent response orbitsare required to solve for the coefficients. Theassociated uncertainties are generated in part bythe manner in which these orbits combine. Orbitpairs that are not linearly independent will combinein a manner which will yield a nearly singularresponse matrix. A balance must be struck insetting the orbit amplitude. The response shouldbe maximized in order to minimize uncertainty yetminimized so as not to violate the assumption oflinearity.

The dynamic characteristics arerepresented by eight linearized stiffness anddamping coefficients. The stiffness coefficientsare equivalent stiffness. Because the fluid iscompressible, the apparent mass due to inertiaeffects cannot be separated. A total of eightequations is required to solve for the eightunknowns. The typical response to the knownforce input is an elliptical displacement orbit.Because an ellipse is uniquely defined by fourpieces of information, two independent orbits arerequired. A second orbit is created by a change inthe external dynamic excitation (magnitude, phase,or both) at the same frequency and staticoperating conditions. Assembling the equationsinto matrix form yields and 8-by-8 set of coupledordinary differential equations. By algebraicmanipulation, the equations are separated into twosets of 4-by-4 matrix equations. The equations areof the general form A * X = B where A is themeasured response matrix, X represents thecoefficient vector, and 6 is the measured excitationvector. The coefficients are then found by matrixinversion.

A second personal computer isresponsible for controlling and acquiring the highfrequency data at a maximum sample rate of 100kHz per channel simultaneously. A signalgenerated within the PC is sent to the hydraulicshaker master controller which activates theapplication of a dynamic load of the prescribedmagnitude at the frequency and phase of thereference signal. The X and Y shakers are

controlled independently; however, the phase leador lag, if any, is controlled by the personalcomputer. Zonic hydraulic shakers excite the testbearing at frequencies up to 1,000 Hz and loadsup to 1,000 pounds peak. The input load andacceleration are measured by Zonic dynamic loadcells and Endevco high frequency accelerometers.The local bearing-to-shaft clearance response ismeasured by the Kaman KDM7200-1UEPproximity probes (A/C coupled). Allinstrumentation is calibrated for magnitude andphase as a function of frequency. Figure 3 showsthe results of the frequency dependent, directdamping coefficients for a journal bearing testarticle with angled injection.

16001400-

•E-1200

iooo80

J" 600 -I

E 400 •raQ 200 -I-t-

0 100 200 300 400 500Excitation Frequency (Hz)

Figure 3. Dynamic Damping vs. Frequency

As shown in the figure, there is a significantdecrease in direct, dynamic damping as excitationfrequency is increased from 100 to 500 Hz. Thisfigure is typical of a bearing at the onset ofpneumatic hammer instability. The vertical lines ateach data point represent the magnitude of theuncertainty which, for this case, is approximately14 percent.

Instrumentation Capacitance

The bearing recess pocket pressure is ahighly desirable parameter to measure during atest. However, in a compressible liquid, thecompressibility of the fluid can lead to responseproblems which can effect the dynamiccharacteristics of the bearing. Previousresearchers investigated the effects of recessaccumulators on the performance of bearings5.Their findings indicated that the resistance andvolume of the accumulator could significantly

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reduce the stiffness and damping of a hydrostaticbearing. The volume of fluid in the hypotubingwhich connects the pocket to a strain gagetransducer may act as an accumulator and alterthe bearing dynamic performance.

The magnitude of the effects due toinstrumentation capacitance can be mitigated twoways: 1) reduce the volume and 2) increase theresistance. The volume was reduced by usingsmall diameter hypotubing (0.032 inch outsidediameter). Additionally, the length of hypo tubingwas minimized by locating the transducer at thetest rig. The resistance was increased with a0.007 inch diameter separation hole which isolatesthe volume of fluid in the pocket from the volume offluid in the hypotubing. Analysis has predicted adecrease in dynamic damping of five percent foran instrumentation hypotubing length of 7 inches6.The effects of instrumentation capacitance will beevaluated experimentally by comparing thedynamic coefficients of identical test points withhypotubing lengths of 10 feet and 10 inches.

Conclusions

A test rig and facility were designed andfabricated to measure the static and dynamicperformance of externally fed, orificecompensated, hydrostatic journal and thrustbearings operating in a turbulent, compressibleliquid. Testing has demonstrated the ability tomeasure and/or extract the static (leakage, loadcapacity, and drag torque) and dynamic (eightfrequency dependent, direct and cross-coupledstiffness and damping coefficients) at bearingsupply pressures up to 1,500 psia, static loads upto 3,000 Ibs, dynamic loads up to 1,000 Ibs,frequencies up to 1,000 Hz, and Reynoldsnumbers up to 100,000. Bearing dynamiccoefficients have been derived using sinusoidalexcitation to achieve multiple independent orbitsyielding coefficients with less than 20 percentuncertainty. Completion of the extensive testmatrix will enable the generation and/or validationof a comprehensive pneumatic hammer instabilitycriteria, evaluate the effects of instrumentationcapacitance on the dynamic performance, and testa variety of stability enhancing bearingconfigurations.

Acknowledgments

A portion of this work was funded by theAdvanced Research Projects Agency as part ofthe Technology Reinvestment Project.

Test results published with permission ofthe Dual-Use Hydrostatic Bearing Consortiumwhich consists of Pratt & Whitney (Lead), CarrierCorporation, and the United States Air ForcePhillips Laboratory.

References

1 Pelfrey, P.C., Turbopump Design andFabrication Support", PL-TR-95-3026, Volume 1

2 San Andres, L.A., "Effects of FluidCompressibility on the Dynamic Response ofHydrostatic Journal Bearings", Wear, Volume 146,pp. 269-283, 1991

3 Reddecliff, J.M., and Vohr, J.H.,"Hydrostatic Bearings for Cryogenic Rocket EngineTurbopumps", Journal of Lubrication Technology,pp. 557-575,1969

4 San Andres, L..A., "hydrosealt User'sManual", December, 1993

5 Goodwin, M.P., Hooke, C.J., and Penny,J.E.T., "Controlling the Dynamic Characteristics ofa Hydrostatic Bearing by Using a Pocket-Connected Accumulator", IMECHE vol 197C, pp.255-258, 1983

6 Hibbs, R.I., Scharrer, J.K., Pelfrey, P.C.,and Justak, J., "Pressure Tap Effect on theDynamic Characteristics of a CryogenicHydrostatic Journal Bearing", AIAA 95-2965

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