Design Procedure of a Turbopump Test BenchJulian D. Pauw1*, Lucrezia Veggi1, Bernd Wagner2, Joydip Mondal1,3, Maximilian Klotz1,Oskar J. Haidn1

SYM

POSI

A

ON ROTATING MACHIN

ERY

ISROMAC 2017

InternationalSymposium on

Transport Phenomenaand

Dynamics of RotatingMachinery

Maui, Hawaii

December 16-21, 2017

Abstract�e high complexity of turbopumps for liquid rocket engines and their demanding requirementsnecessitate that their design process is accompanied by extensive experimental investigations andvalidation tests. �is paper presents the design procedure for a rocket turbopump test bench, wherewater is used as a surrogate for the cryogenic �uids usually used in rocket engines. Scaling methods,that allow for a comparison of tests under varying conditions, are reviewed from literature andapplied to derive the necessary dimensions of the test bench. �e resulting test bench design isshown in detail and its capabilities to support the turbopump design process are assessed. Further,the operational envelope of the derived test bench design is evaluated with respect to later tests ofdi�erent pumps.Keywordsturbopump — liquid oxygen scaling — test bench design

1Technical University of Munich, Department of Mechnical Engineering, Chair of Turbomachinery and Flight Propulsion,Divison Space Propulsion, Munich, Germany2Institute of Space Propulsion Lampoldshausen, German Aerospace Center (DLR), Hardthausen, Germany3Cryogenic Engineering Centre, Indian Institute of Technology Kharagpur, India*Corresponding author: julian.pauw@ltf.mw.tum.de

1. INTRODUCTION�e turbopumps of a liquid rocket engine (LRE) supply thecombustion devices with fuel and oxidizer at high mass �owrates and at a high pressure level. Typically, fuel and oxidizerare pumped by separate pumps. �e turbopumps increasethe pressure from a low pressure level in the tank to a highpressure level needed for the combustion process. Conse-quently, the turbopumps are a substantial part of the rocketengine. �e dimension and the performance of the turbo-pumps depend highly on the desired engine operation point,the engine cycle and the requirements of the combustionchamber.�e work performed at the Division Space Propulsion of theTechnical University of Munich (TUM) is embedded into theresearch project KonRAT, i.e. rocket propulsion engine com-ponents for applications in aerospace transportation systems.�e project aims at establishing competences in the devel-opment of turbopumps. [1, 2, 3, 4, 5]. Key objective at theDivision Space Propulsion, TUM, is the investigation of thedesign process of turbopumps and the investigation of �uidphenomena. Both parts are carried out numerically as wellas experimentally.�e Division Space Propulsion, TUM, has established a designprocess for rocket engine turbo pumps based on well-knownliterature and commercially available so�ware tools [1]. Inorder to fully understand all design parameters and to ensurethat those parameters desired can be reached, experimentalvalidation of those design parameters is an indispensablestep in the design loop of every rocket engine turbopump.�e fuel and the oxidizer of cryogenic liquid rocket engines

are stored in liquid state under boiling conditions on-boardthe launcher and thus ground tests with cryogenic �uids areof very complex nature. Besides the di�cult generation andstorage of the cryogenic �uids, the demands on the turbo-pump under development are very high as the design needsto withstand high temperature loads and it needs to takeinto consideration material choices for cryogenic tempera-tures for every tests. Consequently, tests with �uids that areliquid at ambient temperature are highly desirable. Further,it would be bene�cial to perform tests at lower rotationalspeeds without loosing information on the �ow behaviour in-side the pump. Test benches operated with water at ambienttemperature have been established as a very good solutionto this problem. A drawback of tests at ambient temperaturewith water is that the cavitational behaviour of turbopumpscannot be captured. Especially, the so called thermodynamicsuppression head (TSH), an e�ect that can be observed attests with cryogenic �uids, is not present at ambient temper-ature tests with water. A solution is provided by the �ndingsof several research groups which show that similarity of thecavitational performance can be reached by heating the waterup [6, 7, 8, 9, 10, 11].

In literature, many di�erent scaling and similarity methodsare present. �is paper will present a set of those methodsbased on a literature survey and makes use of the methodsin order to de�ne the operating conditions of a test bench fora liquid oxygen turbo pump developed at TUM. �e derivedtest bench will be shown in detail and its properties will beevaluated for later usage. �is is mostly done by making useof a numerical test bench representation in the commercially

Design Procedure of a Turbopump Test Bench — 2/11

available so�ware tool EcosimPro®. In order to be able to usethe test bench for other con�gurations as well, an outlook onthe generalized operational envelope of the test bench willcomplete this paper, so that the possible use for future testcampaigns with di�erent con�gurations can be estimated.

2. SCALING METHODSIn order to test a pump in a test facility at conditions thatdi�er from the normal operating conditions, the similarity ofthe �ow passing through the pumps needs to be establishedboth at the test bench and at the operational conditions. Ac-cording to [12, 13, 14], this similarity can be reached if fourdi�erent parameters are comparable: (1) the geometry un-der investigation, (2) a comparable establishment of velocitytriangles at pump inlet and outlet, (3) a similar dynamic be-haviour of the pump and (4) comparable thermodynamicproperties of the �uids.

Following Sigloch [13, 14], geometric comparability of twoturbo machines is ful�lled if the dimensions of the partsconducting the �ow follow a certain ratio in all spatial di-mensions. �is o�ers not only comparable interaction ofthe �uid with the static and dynamic parts of the machine,but also yields similar velocity triangles at every correspond-ing point for friction-free conditions because the velocityvectors are then governed by passage dimensions and therotational speed only. It is important to make sure that notonly the rotor and blade geometries are comparable, but alsothe gap geometries and clearance distances between rotorand housing.

λgeometry =dimension of prototype

dimension of model=

Xm

Xa(1)

λvelocity =velocities of prototype

velocities of model=

vm

va(2)

In the case of single-phase �ows, i.e. in the case of non-cavitating �ows, the forces acting on the pump are the fol-lowing: inertia forces, pressure forces, viscous forces and thegravitational force. Additionally, forces resulting from theelasticity of the mechanical system, especially the sha�, canact on the pump components. �e �ow through two di�erentpumps can only be considered similar if the ratio of the forcesacting on these pumps is constant in all spatial directions.

λForce =forces at prototype

forces at model=

Fm

Fa(3)

�e forces acting in pumps with single-phase �ows can berelated by the Froude number, the Reynolds number and theEuler number. �ese characteristic numbers should be thesame for the model and the prototype.�e Froude number is de�ned as the ratio of inertia forces togravitational forces. It is commonly used in hydro-mechanics

to model �ows with free surfaces where gravitational forceshave a large impact.

Fr =inertia force

gravitational force=

v2

gLc(4)

�e Reynolds number relates between the acting inertiaforces and the viscous forces. Equivalent Reynolds num-bers predict a similar development of boundary layers insidetwo compared pumps. �e generation of boundary layersis closely linked to the surface roughness of the parts thatconduct the �ow. Especially for geometrically scaled modelsit is di�cult to generate a comparable surface roughness ofall pump components. For the purpose of comparing di�er-ent radial machines, the Reynolds number is o�en de�ned asthe product of peripheral velocity U3 at the impeller outletand the corresponding tip radius R3 divided by the kinematicviscosity υ. When comparing the scaled measurements indetail, it is strongly recommended to calculate the Reynoldsnumber locally to guarantee comparable boundary layers allover the pump [14, 13, 15, 16].

Re =inertia forceviscous force

=U3Rt,3

υ=ωR2

t,3

υ(5)

�e Euler number describes the ratio between pressure forcesand inertia forces. For the scaling of pumps, it is inevitableto keep the Euler number the same for model and prototype.

Eu =pressure forceinertia force

=∆pρv2 (6)

Further, the thermal properties of the �uids, i.e. the spe-ci�c heat, the enthalpy and the thermal di�usivity, largelyin�uence the �ow characteristics. �is is especially true formachines that exhibit a large pressure di�erence. It is advis-able to make use of �uids with comparable properties and toclosely track property changes.

�e above scaling methods need to be incorporated in orderto design a test bench that can be used to test a LOX turbop-ump with water. For ground tests of liquid rocket engineturbopumps, the gravitational e�ects are negligible. �e dif-ferences in height and the resulting e�ects are comparablysmall. �e Froude scaling is thus irrelevant. �e Reynoldsscaling is of high importance. At least for this application, itis su�cient to achieve fully turbulent conditions in the modeland the prototype because the boundary layers that developat turbulent conditions are assumed to be comparable. Fullyturbulent conditions can be considered for Ret,3 > 106. Dueto the high rotational speeds of liquid rocket engine turbop-umps, this requirement can be ful�lled without any problems.Figure 1 sums up the properties that need to be scaled forliquid rocket engine turbopumps.Obeying the above given similarity conditions, the operatingproperties can be transferred between model and prototype

Design Procedure of a Turbopump Test Bench — 3/11

Fluid-dynamic scaling of

flow

Geometric scaling of

rotating parts

Comparable boundary layers

Thermodynamic scaling of fluid

Similar velocity triangles

Figure 1. Comparable properties for �uid-dynamicsimilarity

with the help of well-known scaling laws for pumps[14].

For the volume �ow rate:

Qm

NmDt,m3 =

Qa

NaDt,a3 (7)

For the head rise:

Hm

Nm2Dt,m

2 =Ha

Na2Dt,a

2 (8)

For the input power:

Pm

Nm3Dt,m

5ρm=

Pa

Na3Dt,a

5ρa(9)

2.1 Scaling of Cavitation Phenomena�e scaling methodology above is valid for ideal �uids, i.e.non-cavitating single-phase �ows. In operational regimeswhere cavitation at the pump blades occurs, a two-phase �owis present as a phase change is triggered by a pressure dropwhere the local �uid pressure is below the vapour pressureof the �uid. Usually, a certain amount of cavitation can beaccepted for every pump, but excessive cavitation results ina head break down of the whole pump and de�nes an oper-ational limit of the pump when the local static pressure ismuch smaller than the vapor pressure of the pumped liquid.In rocket engine turbopumps, cavitation usually �rst occursat the leading-edge of the inducer and it is for most appli-cations best described by heterogeneous nucleation due to�uid impurities [17]. By de�nition, the net positive suctionhead (NPSH) calculates to the di�erence between the totalpressure and the �uid’s vapor pressure at the inlet dividedby the local density of the �uid. [18].

NPSH =Pst − Pv

ρg(10)

�e required net positive suction head (NPSHR) is commonlyde�ned as the point where operation of the pump is still

feasible. �e minimal required NPSHR has to be smaller thanthe available NPSHA.

NPSHR ≤ NPSHA (11)

In order to show comparable behaviour with water as a sur-rogate compared to cryogenic �uids in terms of cavitation, itis necessary to match the same �ow coe�cient ϕ as well asthe cavitation number σ [14, 15].

ϕ =2Q

ADtω(12)

σ =(pst − pv)2ρlD2

tω2 (13)

However, in experiments with several �uids, especially withcryogenic �uids, that operate close to their critical point likeliquid hydrogen and liquid oxygen, it was possible to observethat the inducers of pumps are operating without cavitationat NPSH values smaller than the NPSHR which would beexpected for an ideal �uid. �is e�ect is commonly calledthermal suppression head (TSH). By de�nition, the TSH cal-culates to the di�erence between the available NPSHA andthe NPSHideal f luid that would be expected for an ideal �uid.NPSHideal f luid equals the NPSHtank minus friction lossesin the inlet tubing [18].

TSH = NPSHavailable − NPSHideal f luid (14)

�e mechanism of bubble formation, growth and collapse incavitation depends largely on the instantaneous heat trans-fer between the bubble and the surrounding �uid, the sizeof the pump and the speed of the pump. �us, additionalscaling methods are needed to allow for a prediction of themechanisms related to thermodynamic e�ects. Several pa-rameters and models have been suggested in order to de-scribe and predict the occurrence of cavitation in this regard[19, 12, 20, 21, 22, 23].Brennen [24], by simply looking at a heat balance betweenthe vapor and liquid phases, established a dimensional param-eter that, as he claimed, should be identical for replicatingidentical cavitation behavior in two liquids. He assumed thatthe heat-transfer between the bubble and its surrounding isof conductive nature. �e developed parameter Σ is widelyused. �ereby, he tried to show the di�erent ranges withinwhich the cavitation behavior of di�erent �uids resembledthe cavitation behaviour of water.

Σ =ρ2vh2

f g

ρ2l

√αlCp,lTl

(15)

Ehrlich and Murdock [25] further developed this parameterto a non-dimensional thermal scaling parameter called Di-mensionless Bubble (DB) parameter by considering bubble

Design Procedure of a Turbopump Test Bench — 4/11

growth over a time-varying pressure �eld. �e resulting DBparameter is similar to the formulation suggested by Ruggeriand Moore [20] and is considered very convenient for esti-mating the thermal operational boundaries of the test benchat an early stage in the test bench design process, as it isbased solely on thermodynamic properties of the bulk �uidrather than empirical correlations. For the detailed investiga-tion of a speci�c pump, it is of high importance to take intoaccount all available prediction models, especially those forwhich be�er validation data is available in literature.

DB =Rtω

3/2Cp,lTlρ2l

√αl

h2f gρ2v

(16)

3. TEST OBJECTOne main goal of the test bench under construction is toestablish a result validation loop within the numerical designprocess for turbopumps at TUM. �e pump to be developedis a liquid oxygen turbopump designed for a liquid hydrogenand liquid oxygen expander cycle engine. �e desired opera-tional parameters are given in Table 1. �e resulting thrustlevel is in the order of magnitude of the VINCI upper stageengine.

Table 1. LOX turbopump nominal operating conditions

Property Value Unit

Rotational speed N 20000 rpmNominal mass �ow rate Ûm 25 kg/sTotal pressure at pump inlet pt,1 2.5 barTemperature at pump inlet T1 90 KTotal pressure at pump outlet pt,3 70 bar

A radial-type impeller was selected based on the suctionspeci�c speed NSS value of the pump. Additionally, in orderto avoid cavitation in the radial impeller stage, a high headinducer with cylindrical tip shape was positioned in frontof the impeller. [1, 5] A summary of the design details atnominal operating conditions with liquid oxygen is given inTable 2.A preliminary CAD sketch of the pump assembly with in-ducer, impeller and turbine is shown in Figure 2. �e testbench is designed in such a way as to investigate the pumpdetached from the turbine system. �e parts of the pumpassembly that are relevant for the test bench design are high-lighted.

4. TEST BENCH DESIGN�e main properties of the test bench can be selected basedon the dimensions of the pump under development and thegeneral scaling methods for pumps, which have both beenintroduced in the sections before. �e following section

Table 2. LOX turbopump detailed design parameters

Property Value Unit

Pump Characteristics

Speci�c speed NS 24.34 rpm · m2/√sDesign head coe�cient ψst,pump 0.56 -Hydraulic e�ciency ηhyd 0.88 -Pump overall e�ciency ηtot 0.75 -Sha� power Psha f t −197 kW

Inducer Characteristics

Design suction speci�c speed NSS 421.42 rpm · m2/√sDesign head coe�cient ψtot,inducer 0.11 -Design �ow coe�cient ϕinducer 0.089 -Number of blades Zinducer 2 -Hub-to-tip ratio, inlet Dh,1/Dt,1 0.40 -Hub-to-tip ratio, outlet Dh,2/Dt,2 0.64 -Tip diameter Dt,1 70.83 mm

Impeller Characteristics

Design head coe�cient ψtot,impeller 0.50 -Design �ow coe�cient ϕimpeller 0.10 -Number of blades Zimpeller 6 -Tip-to-outlet ratio Dt2/Dt3 0.72 -Hub-to-outlet ratio Dh2/Dt3 0.46 -Tip diameter Dt3 100.284 mm

Figure 2. Preliminary CAD view of the developedturbopump at TUM without housing and volute. �e pumpcomponents are highlighted.

identi�es the test objectives which are of interest and theimplementation in the test bench.

4.1 Test ObjectivesOne of the key objectives of the test facility is to provideperformance validation data for the turbopump developmentat the Division Space Propulsion, TUM. �is can be yieldedby measuring the pump’s head rise at varied volume �ow

Design Procedure of a Turbopump Test Bench — 5/11

F

Main Tank Inducer & Impeller under Investigation

Main Throttle Valve

Bearing

Unit

Massflowmeter

F

Massflowmeter

Pre

ssu

re

Tem

per

atu

re Bea

ring

D

ata

Pre

ssu

re

Tem

per

atu

re

Pressure

Temperature

Torq

ue

Motor

Water Pipe

Pump Shaft

Pressure Control System

Auxiliary Pump

ElectricalWater Heater

Optional Water Heating System

Spe

ed

Figure 3. Drawing of the test bench layout

rates and rotational speeds in order to generate a pump per-formance chart. �erefore, it is necessary to acquire thefollowing data: the pressure di�erence over the investigatedpump, the mass �ow rate, the sha� torque and the rotationalspeed of the sha�. �e torque measurements give informa-tion about the e�ciency of the pump under investigation.A second goal is to experimentally observe the cavitationbehaviour of the inducer and the pump with water tests inorder to predict the cavitation behaviour under liquid oxygenconditions. �erefore, as evident from the cavitation scalingtheory, the water needs to be heated up in order to operatethe facility with a liquid which is close to its boiling point.

4.2 Test Facility DimensioningFor an operation of the pump in non-cavitating conditions,the scaling methods of equation 7, equation 8 and equation9 have to be applied. �e maximal rotational speed of themotor that drives the pump has been chosen to Nmax =

5500 rpm. �e following considerations will approve thischoice. Further, it is in good agreement with the rotationalspeeds of comparable test facilities. Consequently, sub-scaletests will be performed. �e geometry stays the same asin the LOX hardware. �e hardware designed for the LOXturbopump will be used on the test bench without any scaling:λgeom = 1. �e e�ciency of the pump on the test bench andthe e�ciency of the original LOX pump are assumed to beequal: ηm = ηa .With the properties of the pump under development at TUM,given in Table 1 and Table 2, the maximal mass �ow rate atNmax based on the design point of the original LOX hardwarecalculates to

Qm,Nmax ≈QaNmax

Na= 24.75 m3/h (17)

�e scaled design head rise at Nmax equals

Hm,Nmax ≈HaNm

2

Na2 = 33.30 m (18)

At the same operating point, the sha� power calculates to

Pm,Nmax ≈PaNm

3ρm

Na3ρa

= 3.29 kW (19)

�e dimensions of the new test facility have to be chosenin such a way that those parameters can be satis�ed. Awider range of operation above those limits is desirable forpotentially subsequent expansions of the test facility.In order to test the cavitation performance of the pump in-ducer, it is necessary that the mass �ow coe�cient ϕ, the cav-itation number σ and the thermodynamic properties matchthe ones of the LOX application. Further, the turbopumphas to be operated in regions where Re > 106. As for non-cavitating conditions, the geometry of the LOX hardware isused without dimensional scaling.�e minimal necessary rotational speed for which Re = 106

is satis�ed can be described, according to Equation 5, by

ω ≥ υReR2t,3= 400.55 rad/s or N ≥ 3825.00 rpm (20)

�e thermodynamic properties are assumed to be compara-ble if the Dimensionless Bubble (DB) parameter, as denotedin Equation 16, is equal or close to equal. For the LOX tur-bopump, with the operational parameters given in Table 1and the dimensions of the impeller given in Table 2, thisparameter calculates to DBLOX,re f = 0.306 for the impellerof the TUM design. �e plot in Figure 4 shows the LOXDB reference value DBLOX,re f as a constant line. Togetherwith this constant value, the DB value calculated for di�erent

Design Procedure of a Turbopump Test Bench — 6/11

rotational speeds is shown in dependence of the water tem-perature. It is clearly visible that for each rotational speedof the turbopump on the test bench, a speci�c water tem-perature has to be set in order to satisfy the equality of theDimensionless Bubble (DB) parameter. With increasing ro-tational speed, the required water temperature rises. Forthe maximal chosen rotational speed of the motor Nmax , thewater temperature iteratively calculates to TNmax = 368 K . Ithas to be pointed out that this value for the bulk water tem-perature at the pump inlet only serves as design constraintto the water heating system at this early point in the testbench design procedure.

Water Temperature [K]280 300 320 340 360

DB

[-]

10-2

100

102

104

1500rpm 3100rpm 5500rpm LOX-Reference

Figure 4. Dimenionless Bubble (DB) parameter at di�erentrotational speeds. �e DB reference value has beencalculated for the a LOX pump as described in Table 1. �eLOX geometry is used without any changes for the watertests. �e DB values are calculated for pt,1 = 1 bar .

4.3 Detailed Test Bench DesignWater is selected as the test medium. �e use of water o�ersseveral advantages. �e most important advantage is the easyhandling. �is allows for a comparably cheap operation ofthe test facility. Compared to tests with cryogenic �uids, thesafety of tests is also improved. Further, only very few mate-rial incompatibilities are known. �e system is designed as aclosed circuit. �is is especially advantageous for tests withheated water: a�er preheating the water up to the desiredtemperature, the heater only needs to keep the temperatureat the desired level. �is can improve the temperature con-trol accuracy. �e use of a closed water loop is also in goodagreement with other test facilities for pumps at di�erentinstitutes and national standards [6, 7, 8, 26, 27, 9, 10, 11, 28].As the water is supposed to stay within the circuit for theduration of multiple test campaigns, deionized water is used.�is is bene�cial as depositions on water circuit componentsare limited. Especially for heated water, the sedimentarydeposition of limescale is reduced signi�cantly [29]. It hasto be taken into account that the water needs a minimalconductivity greater 20 µS/cm to guarantee the operation of

inductive mass �ow meters. �e decay of the water qualityis slowed down as well due to the fact that no permanentaccess of light is present in the system. A major drawback ofa closed-loop system is that, in case of occurring cavitation,vapour bubbles can persist in the system and be sucked intothe inlet again.�e drawing in Figure 3 shows an overview of the circuit.Necessary sensor positions, actuators and their positionswithin the loop are shown as well.�e water reservoir is realized by a stainless steel tank witha volume of Vtank = 2000 l. It is designed to withstand themechanical and thermal loads of water at ptank,max = 4 barand Ttank,max = 100◦C. �e tank is equipped with a EPDMmembrane �lled with pressurized air. �is membrane has avolume of Vmembrane = 500 l. �e air pressure within thismembrane and therefore also within the tank is variable andcan be controlled and regulated electronically. In operationalmodes where cavitation occurs, it is very likely that bubblesare transported into the tank and might disturb the pumpmeasurements if they are sucked into the pump again. Inorder to signi�cantly reduce this e�ect, it is desirable tomaximize the residence time of the water in the reservoir.�is is a�ained by redirecting the inlet �ow in circumferentialdirection. In addition, the amount of dissolved oxygen canbe measured in order to a�ain a good repeatability of thetest conditions. Further, the tank can be depressurized up toa negative pressure of pt,tank = 0.9 bar to remove dissolvedoxygen. �is is especially important for tests with heatedwater. For safety reasons, the tank is also equipped with anover-pressure valve. �e controllable static inlet pressureallows to perform NPSHR evaluations on the test pump.�e inlet to the pump from the main reservoir and the outletpiping from the pump to the main tank is created by stainlesssteel tubes with circular cross-section and standard �angeconnectors. �e tube dimension has been chosen to equal thestandard dimension DN80. �is results in an internal pipediameter of all tubes of Dtube = 80.8 mm. It is favourableto have a fully turbulent �ow within the tubes present atall times. �is is feasible with Dtube for mass �ow ratesfrom Ûmmin = 0.6 kg

s on. Based on the head rise scaling con-siderations in Equation 18, all tubes have been designed inthe pressure class DN80-PN25. Due to restrictions of thesensors and auxiliary equipment in the loop, the static pres-sure in the inlet section, including the tank, is limited top1,max = 4 bar . �e pressure in the outlet section is limitedto p3,max = 10 bar . In order to avoid possible sources ofcavitation in the tubing, all transitions and redirections of�ow are manufactured as smooth as possible. Further, allsources of �ow disturbance are avoided in all tubes in frontof sensors and in the tube in front of the pump inlet at a tubelength of Lmin = 10 Dtube.�e head rise of the pump needs to be reduced to the pressurelevel of the tank. �is is done by a thro�le valve con�gura-tion in-between the pump and the tank. In order to reducethe risk of cavitation at the thro�ling system, two identi-cal valves are arranged in a daisy-chain con�guration. Both

Design Procedure of a Turbopump Test Bench — 7/11

valves are equipped with an electronically controllable ac-tuation unit. �is makes it possible to set the pressure dropacross each valve independently. Further, the mass �ow ratecan be controlled. �is feature is needed in order to do aperformance mapping of the pump under investigation andit is a crucial requirement for cavitation scaling.�e mass �ow rate is sensed at two locations. One inductivemass �ow meter is placed in the inlet tubing directly in frontof the pump inlet. A second inductive mass �ow meter isplaced a�er the pump. At the same position, a measurementori�ce is placed. �is is where the pressure drop across ade�ned through-�ow area is measured. �us, for a given �uiddensity, the mass �ow rate can be calculated in a second, inde-pendent way. �is allows for a comparison of measurementresults and the improvement of the measurement accuracyin all measurement ranges.For the controlled operation of the test bench, pressures andtemperatures are captured at di�erent locations of the watercircuit. �e static pressure is monitored, as depicted in Figure3, in the pump inlet section, in the pump outlet section anddirectly a�er the main thro�le valve con�guration. At allthree locations, the pressure measurements are averaged overthe circumference of the horizontally placed pipes in orderto compare for gravity e�ects. �e measured static pressuredi�erence between pump inlet and pump outlet can be usedto calculate the head rise of the pump. �e static pressuredi�erence over the thro�ling valve con�guration allows fora safe operation of the same. In addition, the tank pressureis also monitored. Additionally, temperatures are detected atall mentioned pressure sensor locations. Especially for testswith heated water, these temperature readings yield valuableinformation for the thermal control system.�e pump is driven by a three-phase alternating current (AC)electric motor with a maximal power of Psha f t,max = 12 kW .�e motor reaches its maximal torque Tsha f t,max = 27 Nmat its design speed Nre f = 3100 rpm. �e maximal speed ofthe motor, without the use of any additional transmission, isNmax = 5500 rpm. �e motor is connected to the sha� by a�exible coupling that dampens the temporary high torqueduring start-up. �e sha� is held in position by a bearingunit in overhung con�guration - the bearings are positionedbetween the pump and the drive unit. �e bearing unit is de-signed as an arrangement of a �xed bearing close to the pumpand a �oating bearing close to the drive unit. All loads onthe drive unit are closely monitored. �is includes the torqueof the sha�, the rotational speed and the axial force actingon the �xed bearing. Further, a bearing monitoring systemis established by permanent observation of the bearing racetemperatures. All bearings are run with grease lubricationand the grease quality is observed in �xed intervals. Further,the pump housing is equipped with three acceleration sen-sors for the investigation of potential instabilities. �e dataobtained from those sensors can also be used to monitor thebearing operation.As shown in section 2, in order to reproduce the cavitationbehaviour of the LOX turbopump, it is necessary to run

tests with heated water. �erefore, a water heating systemis included in the test bench setup. �is heating system isdesigned as a second auxiliary circuit that can be decoupledfrom the pump circuit. �us, it is possible to heat the watercontained in the main tank to a desired temperature. �ewater is heated up by an electrical heater with a power ofPht,el = 60 kW . �e water in the second auxiliary watercircuit is driven by a separate pump. �e heating system isdesigned to control the temperature within ∆T1 = ±1 K andcan a�ain a maximal water temperature of Tmax = 100◦C.For means of �ow control in the auxiliary heating circuit, thewater temperature at the heater outlet, the static pressure atheater inlet and heater outlet as well as the mass �ow rateare closely monitored. �e mass �ow rate is detected by ameasuring ori�ce. �e bladder inside the main tank servesas a compensation reservoir for volumetric changes due tothe heating of the water. A dedicated cooling system is notimplemented, but the insulation of the tank and the tubingis designed in a way that continuous heating is necessary toa�ain a constant high temperature.Figure 5 shows a CAD plot of the main circuit of the testfacility that is currently being constructed at the DivisionSpace Propulsion, TUM.

Figure 5. CAD view of the turbopump test bench at TUM

4.4 Numerical Design MethodsIn order to enhance the test facility development describedin the previous sections, the test bench was modelled nu-merically in parallel with the physical setup. �erefore, thewater circuit including all pipes, valves and the tank havebeen modelled in the so�ware tool EcosimPro®. Especiallyall components that are foreseen to be electronically con-trolled were investigated numerically in detail. �erefore,the heating system and the tank, including the membranebladder, were represented as detailed numerical models. Forall components, the pressure drop across those components

Design Procedure of a Turbopump Test Bench — 8/11

Q [m3/h]0 5 10 15 20 25 30 35

H [

m]

0

20

40

60

80

10030%20%10% 40%

100%90%80%70%

60%

50%

Figure 6. Numerically obtained system performance chart.�e percentile values denote the level of opening of thethro�le valve con�guration.

is implemented for di�erent mass �ow rates as core func-tionality. Further, for investigations on the thermal controlof the test bench, empirical correlations for heat losses atall surfaces have been added. �e so�ware suite EcosimPro®

o�ers many tubing and piping elements readily availablein its libraries that have been adapted to the properties ofthe test bench. �ey can be combined in a modular way.�e EcosimPro® library ESPSS expands the building blocksby components for rocket engines. �is includes special tankcon�gurations as well as a generic turbopump model.

�e graph in Figure 6 shows a numerically obtained systemperformance chart for di�erent positions of the thro�le valveassembly. For the generation of this chart, the thro�le valvecon�guration was opened at a �xed percentage and the mass�ow rate was varied. �e static pressure at the pump inletand the pump outlet was measured and the resulting headof the system was calculated. Especially the performancefor the fully opened thro�le valve con�guration is of highinterest as this curve describes the minimal head that a pumphas to generate in order to be tested on the test bench.Further, a numerical evaluation of the pump operation withinthe circuit has been evaluated. �erefore, the generic pumpmodel has been initialized with a speci�c speed of NS =

24.34, a total head of Htot = 602.3 m, the design rotationalspeed of NLOX = 20000 rpm and an estimated e�ciencyof η = 0.887. �e graph in Figure 7 shows the computedpump characteristic for N = 1500 rpm, the motor designspeed Nre f = 3100 rpm and the maximal rotational speedof the motor Nmax = 5500 rpm. Additionally, the systemperformance map is partially shown. �is makes it possibleto identify the resulting operation points. �is chart hasbeen created by varying the opening level of the thro�levalve con�guration from 0% to 100%. �e rotational speed ofthe pump has been kept constant for each curve. �e headrise across the pump has been measured.

Q [m3/h]0 5 10 15 20 25 30 35

H [

m]

0

10

20

30

40

50

60

1500 rpm 3100 rpm 5500 rpm

10% 30% 40%

60%

50%

70%

80%

90%100%

20%

Figure 7. Numerically obtained pump performance chartfor N = 1500 rpm, N = 3100 rpm & N = 5500 rpm. �esystem performance chart for di�erent opening levels of thethro�le valve con�guration are plo�ed in gray.

5. SUMMARY AND OUTLOOK�e key characteristics of the developed test facility are sum-marized in Table 3. �e designed test bench provides a valu-able facility to test radial pumps with inducers in a sub-scaleenvironment. According to the presented scaling methods,the results obtained here can be used to predict the non-scaledpump performance of the original application. Additionally,the possibilities to heat up the water, to control the �ow rateand to control the inlet pressure separately, allow for theinvestigation of the occurrence of cavitation at the inducerblades.

Table 3. Operational Characteristics of the Test Facility atTUM

Property Value Unit

Rotational speed N ≤ 5500 rpmPump Power P ≤ 12 kWInducer Diameter Dt,1 ≤ 80 mmImpeller Diameter Dt,3 ≤ 150 mmFluid Temperature T ≤ 100 ◦CTotal pressure at pump inlet ptot,1 ≥ 1 barTotal pressure at pump outlet ptot,3 ≤ 10 barCavitation Number σ 0.02 . . . 0.45 -Flow coe�cient ϕ ≤ 0.11 -

�e presented test facility has been designed to meet the testcriteria for the LOX-turbopump which is currently under de-velopment here at TUM. All dimensions and characteristicsof the components have been chosen accordingly. Never-theless, the test bench can be used to investigate any pumpof similar constructive form, as long as it stays within theconstraints given in Table 3. Based on the Barber-Nichol’schart for pumps [30], the operational boundaries of the test

Design Procedure of a Turbopump Test Bench — 9/11

Specific Speed NS

10-3 10-2 10-1 100 101 102 103 104

Spec

ific

Dia

met

er D

S10-1

100

101

102

Operational envelope of the test facility at TUM Hmax

NSS,min

Pmax

Qmax

Figure 8. Operational envelope of the turbopump test facility at TUM 11 �is �gure is composed by overlaying the Barber Nichols chart for pumps [30] with the operational boundaries of the test facility at theDivision Space Propulsion, TUM.

bench have been plo�ed in a Speci�c Speed - Speci�c Diam-eter (NS − DS) chart. �e visualization of the boundaries isshown in Figure 8. For this �gure, the maximal mass �owrate has been chosen to Ûmmax = 25 kg

s . In general, the mass�ow rate of the test facility is not limited, but, accordingto the system performance chart in Figure 6, the necessaryminimal head rise of the investigated pump increases withincreasing mass �ow rate.

�e numerical implementation of all components withEcosimPro® in parallel with the physical construction of thetest facility has shown to be bene�cial. It was possible toinvestigate component dimensions at question numericallyand the a priori understanding of the test facility was greatlyimproved. Especially the dimensioning of the water heatingsystem and the development of the inlet pressure controlsystem in the tank were supported by numerical studies. Im-plementations of simple circuit components like straight andbent pipes, valves and tanks are available and the empiricalcorrelations are of good quality for the full operating regionof the test bench. �e preliminarily calculated performancechart derived from the generic pump model shows the sameorder of magnitude as the analytically scaled values. Any-how, especially the implementation of the pump, based onthe generic pump component from the ESPSS library, doesnot yield the desired amount of details. Two possibilitieshave been identi�ed in order to perform improved numer-ical analyses: (1) Experimental measurement of the pumpperformance map and implementing this information in amodule or (2) detailed modeling of the single pump compo-nents. Both approaches are currently being followed at TUM.

For the second approach, a detailed model of the inlet, theinducer, the impeller, the di�usor and the volute of the pumpdeveloped at TUM are under development.

NOMENCLATURESymbolsA areaα thermal di�usivityCp speci�c heatD diameterDB Dimensionless Bubble parameterη e�ciencyEu Euler numberF forceFr Froude numberg gravitational acceleration. g = 9, 81 m/s2

H head riseh f g heat of vaporizationλ scaling factorÛm mass �ow rateN rotational speed

[min−1]

NS speci�c speed. NS = N ·√Q

H0.75

NSS suction speci�c speed. NSS = N ·√Q

NPSH0.75

P powerp pressureψ head coe�cient. ψ = 2gH/U2

3

ϕ �ow coe�cientU peripheral velocityQ volume �ow rate

Design Procedure of a Turbopump Test Bench — 10/11

R radiusρ densityRe Reynolds numberΣ cavitation parameter (Brennen)T temperatureυ kinematic viscosityv velocityω rotational speed [rad/s]X lengthZ number of bladesSubscriptsA availablea application prototype (not scaled)c characteristicel electricalh hubht heating circuithyd hydraulicl liquidm model for testsR requiredSS suction speci�cst statict tiptot totalv vapor1 inducer inlet2 interface between inducer and impeller3 impeller outletAbbreviationsNPSH Net Positive Suction HeadAC Alternating CurrentCAD Computer Aided DesignDB Dimensionless Bubble parameterLH2 Liquid HydrogenLOX Liquid OxygenLRE Liquid Rocket EngineTSH �ermodynamic Suppression HeadTUM Technical University of Munich

ACKNOWLEDGMENTS�is project is supported by the Ludwig Bolkow Campus,funded by the Bavarian government. �e authors greatly ap-preciate the good cooperation with the consortium partners.

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