NBSIR 73-316

(AFAPL-TR-73-72)

LIQUID HELIUM PUMPS

Philip M. McConnell

Cryogenics Division

Institute for Basic Standards

National Bureau of Standards

Boulder, Colorado 80302

Interim Report

June 1973

Prepared for

Air Force Aero Propulsion Laboratory

Air Force Systems Command

Wright-Patterson Air Force Base; Ohio 45433

NBSIR 73-316

(AFAPLTR-73-72)

LIQUID HELIUM PUMPS

Philip M. McConnell

Cryogenics Division

Institute for Basic Standards

National Bureau of Standards

Boulder, Colorado 80302

Interim Report

June 1973

Prepared for

Air Force Aero Propulsion Laboratory

Air Force Systems Command

Wright-Patterson Air Force Base; Ohio 45433

U.S. DEPARTMENT OF COMMERCE, Frederick B. Dent, Secretary

NATIONAL BUREAU OF STANDARDS. Richard W Roberts. Director
NOTICE

When Government drawings, specifications, or other data are used for any

purpose other than in connection with a definitely related Government

procurement operation, the United States Government thereby incurs no

responsibility nor any obligation whatsoever; and the fact that the

government may have formulated, furnished, cr in any way supplied the said

drawings, specifications, or other data, is not to be regarded by implication

or otherwise as in any manner licensing the holder or any other person

or corporation, or conveying any rights or permission to manufacture, use,

or sell any patented invention that may in any way be related thereto.

Copies of this report should not be returned unless return is required by

security considerations, contractual obligations, or notice on a specific

document.

ii
FOREWORD

This program was completed under sponsorship of the Department

of the Air Force, Wright -Patter son Air Force Base, Ohio, and was

funded with FY 1972 AF Aero Propulsion Laboratory Director's Funds.

This Technical Report has been reviewed and is approved for

publication.

The joint contributions of Lou Schafer of the Sundstrand Corporation, Rockford, Illinois, who provided the submersible motor and Jim

Lobach of the Barber -Nichols Engineering Company, Denver, Colorado,

who designed the pump (under contract), are gratefully acknowledged by

the author. * Thanks is given to K. Nuss, who machined the pump parts.

The author is grateful to V. Arp for his review of the manuscript and

his helpful suggestions throughout the course of this work. Special

thanks is also given to J. Hord for his consultative contributions.

Philip M. McConnell's present affiliation is the Todd Shipbuilding

Company, Research and Development Division, Galveston, Texas 77550.

FOR THE COMMANDER

DR. P/E. STOVER

GS-15

Chief, Power Distribution Branch

Aerospace Power Division

ACKNOWLEDGMENTS

iii
Nomenclature

area of orifice

impeller width at discharge

height of the centrifugal impeller vanes

meridional velocity at inlet tip

discharge coefficient

constants in temperature -

resistance calibration correlation

centrifugal impeller diameter

inducer hub diameter

^D^ (1 2 v )/2, mean effective,

inducer diameter

inducer tip diameter

elastic modulus

mechanical energy losses per unit mass

per unit time

local acceleration of gravity

total dynamic head

thermal strain

rate of mass flow at pump discharge

pump rotative speed

(9. 7961 m/s2

)

iv
normal boiling point

2

Pj/ Pj + /2g - P net positive

specific speed, see eq (6)

suction specific speed, see eq (7)

absolute pressure

P^ - Pj, differential pressure across

pump

differential pressure across venturi

m/

9

> volumetric rate of flow of pump

discharge

bearing cooling flow

1/2 D

m

resistance of thermometers

Reynolds Number of the discharge tube

temperature

2. 172 K, \-point temperature

torque

tangential velocity at

tangential velocity at

average stream velocity at point of

static pressure tap
w = relative velocity

W = total average 3-phase power input to motor

W = power per unit mass

= effective power imparted to the fluid by the pump,

defined by eq (4 ) W , = T' U) , shaft power

sh

Z = vertical distance above datum plane

Greek Symbols

a - angle between the tangential velocity

and the absolute velocity, see figure 9.

(3 - vane angle of impeller

Y = specific weight

T\ = efficiency

v = D^/D^, hub ratio of inducer

p = fluid density

0 = c /u , flow coefficient ml 1

cp, = c /u^ , inlet flow coefficient

1 ml 12

f = 2gH/Uj, head rise coefficient, at inlet

ou = angular velocity

vi
inlet

discharge

bearing cooling passage

critical

cavitating

hydraulic

liquid

motor

normal fluid

non- cavitating

pump

superfluid

total

ullage

vapor, volumetric

saturated vapor at inlet

vii
CONTENTS

Page

1. Introduction . 1

2. Pump Characterization 4

2. 1 Mechanical Energy Balance 4

2. 2 Pump Performance and Characteristics .... 7

2. 3 Affinity Laws 10

2. 4 Stability 11

2. 5 Helium Pump Cavitation 12

2. 6 Pump Efficiency 17

2.7 Reliability 19

3. Survey Results 20

3. 1 Literature Review 20

3. 2 Commercially Available Pumps ....... 28

4. Pump Tests . 34

4. 1 Description of the Tested Pump 34

4. 2 Test Loop and Instrumentation ....... 40

4. 3 Results of Pump Test ......... 46

4. 3. 1 LHe I, 2. 3 K < T < 4. 3 K 50

4.3.2 Supercritical Helium, 4. 7 K < T < 5. 6 K . 59

4.3.3 LHe II, 1. 8 K < T < T. ....... 61

A. 4. 3. 4 Comparison of the H-Q Data 66

4. 3. 5 Reliability ........... 66

5„ Conclusions ................ 68

6. Recommendations for Future Work 70

7. References 72

viii
LIST OF FIGURES

Page

4

Figure 1. Phase diagram for He 3

Figure 2. Schematic diagram of a liquid helium pumping

system 5

Figure 3. Typical steady state performance characteristics

of a vaneaxial blower ........... 12

Figure 4. Predicted LHe I cavitation performance of a

67. 8 mm inlet diameter, centrifugal pump;

rotative speed 25,000 rpm; inlet flow coefficient,

0 = c /U = 0. 225 16

1 rrij 1 Figure 5. Axial flow inducer; brass vanes are silver

soldered into cuts in the hub 35

Figure 6. Centrifugal impeller for helium pump, 6061 -T6

aluminum, anodized finish 36

Figure 7. Photograph of helium pump parts 37

Figure 8. Assembly drawing of helium pump and

submersible motor „ 38

Figure 9. Inlet and discharge velocity triangles based on

the design criteria 39

Figure 10. Schematic diagram of He - pump test loop and

instrumentation 41

Figure 11. Venturi flow meter used in helium pump tests . . 47

Figure 12. Typical performance characteristics of

He - pump operating in LHe I; T = 4. 15 K; n = 631 rad/s (6030 rpm); 0. 19 m £ NPSH < 0. 7 m . 53

Figure 13. Total dynamic head-capacity performance

curves in LHe I; 3. 18 K < T < 4. 15 K 55

ix
LIST OF FIGURES (Continued)

Page

Figure 14. Pump cavitation performance in LHe 1 58

Figure 15. Pump performance in supercritical helium ; n = 630 rad/s (6011 rpm) 60

Figure 16. Pump performance in LHe I and LHe II; n = 340 rad/s (3245 rpm) 63

Figure 17. Performance characteristics of He - pump

operating in LHe II; T = 1. 89 K; n = 339 rad/s

(3237 rpm); 0. 105 m £ NPSH ^ 0. 204 m .... 65

Figure 18. H-Q data normalized to the design speed,

628 rad/s (6000 rpm) 67

Figure 19. Typical reliability test cycle 69

LIST OF TABLES

Page

Table 1. Summary of values of the saturated density ratio

at the normal boiling point for several fluids. . . 14

Table 2. PumP Perf° rmance reported by Wright ... 24

Table 3. Estimated performance of helium centrifugal

pump reported by Mark and Pierce 27

Table 4. Helium pump performance test results. .... 51

x
ABSTRACT

This report summarizes studies of pump characteristics and performance in supercritical, normally

boiling, and superfluid helium, and also presents results

on a survey of commercially available pumps for helium

service. Experimental measurements were made on a centrifugal pump which produced a maximum head of about

-4 3

15 meters, and a maximum flow of about 2. 5x10 m/s„

Performance agreed approximately with classical affinity

laws, but cavitation appeared to provide less of a performance limitation than expected. The survey turned up

several pumps which have been used in helium, though

relevant performance data is lacking.

Key words: Cavitation; helium; pump performance; pumps;

superfluid.

xi

Liquid Helium Pumps*

Philip M. McConnell t

1. Introduction

Pumps are presently being used [1-4], ^ or considered [5], to

provide forced helium cooling of superconductive magnets and power

transmission lines. However, there exists little experience in pumping

helium, and no pump performance characteristics with low temperature

helium (2 K < T < 8 K) have previously been reported in the literature.

Therefore, design and performance characteristics have been largely

speculative.

A basic question which we wish to examine in this study is: What

unique features are involved in the pumping of helium? The low latent

heat of vaporization (about 22 J/g at 4K), low viscosity, and superfluidity are some of the unusual properties of helium (2 K < T < 8 K)

which could give rise to unpredictable pump performance. On the other

hand, if the helium is single phase and Newtonian, helium pumping

should be similar to the pumping of single phase, low temperature

hydrogen (or other cryogenic fluids); in this case the considerable work

which has already been done [2, 6, 7, 8, 9, 10] could be used to advantage. Actual helium data are obviously needed to verify the feasibility

and reliability of using conventional pumps to transfer the various fluid

phases (e. g, , LHe II, LHe I, two-phase helium, and supercritical helium),

This program was funded with FY-1972 AF Aero Propulsion Laboratory

Director's funds, Wright-Patterson AFB, Ohio,

t Present affiliation: Todd Shipbuilding Company, Research and

and Development Division, Galveston, Texas 77550.

Figures in brackets indicate the literature references at the end of

this paper.
to verify the pump affinity laws [11], and to check the validity of accurately predicting helium pump performance using less expensive fluids e The phase diagram for helium is shown in figure 1 (not to scale).

Systems currently under study require intermittent cooling flow rates

-3 3 which are greater than 1.9 x 10 m /s (30 gal/min) at relatively low

differential pressures. With this in mind, the study has been rather

arbitrarily limited to include pumps whose design capacities are less

-3 3

than 6. 3 x 10 m /s (100 gal/min) with relatively low total head, and

to helium inlet temperatures between 2 and 8 K. The helium thermodynamic and transport properties, necessary for any pump analysis,

are well known for these flow conditions [12].

The characterization of pumps for helium service is discussed

in general terms in section 2. , wherein we summarize the origin of

some of the important pump parameters and their applications. Although

many of these parameters were developed from work solely on impeller

pumps, they can be meaningful when applied to other types of pumps,

e. g. , ejector pumps [13]. More comprehensive treatments of impeller

pump design, performance, and similarity parameters are given by

Wislicenus [14], Stepanoff [11], and Kovats [15], and Lazarkiewicz and

Troskolanski [16], Other pump types such as reciprocating pumps and

gear pumps were also of interest, but few of these types are found for

low temperature use. Also included are practical considerations involved

in pumping helium- -e. g. , cavitation, stability of the pump, bearings,

types of motor drives, etc.

The results of a survey of helium pumps are given in section 3.

It includes results of an extensive review of the literature of helium

pumps, and results of personal communication with various pump manufacturers whose products may be applicable for helium service.

2
\

3
In section 4 we describe our own laboratory tests of a small cen-

-4 3

trifugal helium pump designed for flows up to Z. 5 x 10 m /s (4 gpm)

with a total dynamic head of 15 meters, characteristics which may be

useful for various superconductor applications where forced flow cooling

may be required.

2. Pump Characterization

2. 1 Mechanical Energy Balance

A typical liquid helium pumping system is shown in figure 2. The

pump motor may be operated at ambient temperature, location a, or submerged in the fluid, position b. Both of the configurations have

their advantages and disadvantages , which will be discussed later. The

superconducting system may be located in the pump dewar or in a remotely located dewar, depending on the particular application.

The steady state mechanical energy balance [17], applied to the

pump between the inlet and discharge (planes 1 and 2, respectively),

can be written per unit mass per unit time as

2 2

AV f dP /\ /\

+ g A Z + ~ - "W + E = 0 , (1)

J 1 2

j P v

where the numbers 1 1 ' and '2' denote the measured inlet and outlet

/\ stream conditions respectively, W is the specific power input to the

pump, and E is the rate at which mechanical energy is irreversibly

converted to thermal energy, i. e. , friction losses. For an isentropic

process, E is equal to the rate of heat removal through the pump

casing between 1 and 2. Assumptions made in the derivation of ( 1

)

are that the flow is turbulent and that the gravitational acceleration,

g, remains constant throughout the system. The first three terms

in (1) represent the fluid or useful power per unit mass, W., imparted

to the fluid by the pump. The pump efficiency defined as

4
Vent

Pu.T

LHe

a

Delivery

Valve

. fJU ^To Superconducting

System

q (Heat Leak)

1

AZ

• 2

Figure 2. Schematic diagram of a liquid helium pumping system.

W„ W - E

1 = 7C = < 2 » w w

is determined from measurements of all the terms in (1) other than

/\

E . Since continuity of mass requires that rh - m = m, T| = W /W , v 1 2 f sh

where is the fluid power and W ^

is the shaft power to the pump.

It is sometimes convenient to divide (1) by g. In this case, the first

three terms are referred to as the total dynamic head (H), where

2 2

AV 1 r dP H =^~+ A Z + — i — . (3)

2 g g

J

j P

5
The first two terms in eq (3) are usually much less than the third. The

H has dimensions of length (usually expressed in units of the working

fluid as characterized by the inlet stream conditions), and for a given

pump operating at a constant speed and volume flow rate is (approximately) independent of the fluid being pumped. The fluid power can be

expressed in terms of H as W

f = mg (H) = Yl Q 1 (H) = Y

2 Q2

(H) (4)

where m and Q are the rates of mass and volumetric flow, respectively,

through the pump, and y is the specific weight of the fluid.

In order to evaluate the integral in eqs (1) and (3) along a representative streamline, one must know the equation of state, p = (P, T),

how T changes with P along the streamline, and the thermodynamic

path of the process from 1 to 2. For the special cases of an isothermal system, p = p (P) , and isentropic flow of an ideal gas, the integral

may be easily evaluated. In the case of isothermal flow of an incompressible fluid

,

2 dP _ P

2

-P

t Jl P P

The assumption of incompressibility applies when the change in pressure is only a small fraction (a few percent) of the absolute fluid

pressure and is usually a good assumption for most liquids and gases

at low Mach numbers. Variations in density due to changes in pressure

can be described by the equation

o o

6
where E is a "modulus of elasticity" of the volume change AV [18].

The compressibility of most liquids is small, e„ g. , for water

9 2

E = 1, 9 x 10 N/m (19, 000 atm) which means that a pressure increase

5 2

of 1 x 10 N/m (1 atm) causes a relative change in density of about

0. 005 percent. Liquid helium, however, is considerably more com6 2

pressible, e. g. , at 4, 0 K, E = 2. 5 x 10 N/m (25 atm) giving a 5,2 relative change in density of about 4 percent for AP = 10 N/m (1 atm).

In supercritical helium near the critical point, relative changes

5 2

in density of several hundred percent are possible with AP = 10 N/m , depending on the temperature. Thus, the assumption of incompressibility must be carefully examined when pumping helium.

In general, the steady state performance of a pump running at a constant speed is described by its response to both changes in capacity,

Q (non-cavitating performance), and the net positive suction head, NPSH

(cavitating performance). The latter type of performance is only meaningful in the pumping of liquids. The net positive suction head, NPSH,

sometimes referred to as the anti cavitation pressure margin, is defined

as

where subscript 1 refers to the pump inlet and the quantity in parenthesis is the stagnation inlet head. As shown in figure 2 , the inlet

pressure is the sum of the tank pressure (ullage pressure) and the pressure due to the weight of the liquid above the inlet plane, i. e. ,

2. 2 Pump Performance and Characteristics

(5)

7
If the liquid is in thermal equilibrium with its vapor and has a relativelylow inlet velocity, then

P = P , and NPSH - Z . u vl t

If this is not the case, accurate measurements of NPSH are much more

difficult, depending on the magnitude of the slope, dP/dT, of the saturation curve [19,20],

Non-cavitating performance : During a typical non-cavitating pump test, the shaft speed, n,

is maintained constant while the independent variable, Q, is altered by

throttling the delivery valve, see figure 2 . For each valve setting, the

flow rate, total dynamic head, and the shaft power are simultaneously

/ measured. The shaft power, W , = Too is obtained from either a torque

sh

measurement, T, or from measurements of the power consumption of

the driving motor of known efficiency.

In this way, the conventional, experimental non-cavitating performance curves H = f (Q), W = f (Q), and Tl = f (Q) are obtained.

sh

The difficulty of accurately determining H = f (Q) from theory lies in

accurately calculating the losses occurring in the pump. The shape

of the above curves depends strongly on the pump geometry (impeller

and casing) and a similarity parameter known as the specific speed [11,

15, 16 ].

The specific speed,

N =nQ1/2 /H

3/4

, (6,

s

8
2

is usually computed at the best efficiency point (b. e. p. ) and is useful

in characterizing types of impeller pumps; e„ g. , centrifugal pumps

have values of N between about 500 and 3500.

s Pumps of the same value of will, regardless of differences

in size and shape, tend to have similar performance characteristics.

A designer uses this fact to good advantage in that testing of one pump

will allow the performance of other pumps of the same to be accurately

predicted.

Cavitating Performance:

The cavitation performance of a pump is obtained experimentally

by running a pump with a constant liquid inlet temperature at a constant

capacity and speed, and observing the H, W ^, and T| as the NPSH is

reduced. The H = f (NPSH) curve is most commonly reported in the

literature [11,21]. The above three parameters should be independent

of the NPSH reductions down to a point, below which the curves typically

drop rather sharply. Since the parameters are not equally sensitive to

the effects of cavitation, their break-away points are usually slightly

different. Motion pictures taken of pump cavitation show that some

cavitation can occur in the pump without influencing the performance

characteristics. At the head break-away point, however, the cavitation

is sufficiently far advanced to cause partial blockage of the flow passages; this phenomenon is similar to the choking of gas flow in a convergent-divergent nozzle [15]. The shape of the H -NPSH drop-off

curves for an impeller pump depends strongly on the fluid properties

Values of N are dependent on the units used. In the U. S. , it is

s customary to use N (rpm), Q (gal/min), H (ft). This dependence is also

true for the suction specific speed, N . In this survey, the U. S.

s , cav

units are used. To convert N and N to SI units (rpm, m Is, m)

,

s s, cav

multiply the U. S. values by 0. 0194.

9
and the impeller inlet geometry. A parameter which presumably

describes flow similarity of geometrically similar pumps, operating

in the cavitation regime, is the cavitational specific speed or suction

specific speed, defined as

The suction specific speed can also be formulated as a function of

several of the basic pump design parameters [22], which can include

the thermodynamic effects of the fluid. High values of N are an

S j C cLV

indication of high speed pumps which perform well at low values of

NPSH, even though some cavitation is present, e. g. , aircraft and

rocket fuel pumps. Below N of 5,000, single -suction centrifugal

S j C 3-

pumps have poor cavitation characteristics; above values of 11,000 the

reverse is true. Meaningful values of N result when NPSH > 0. s, cav

Values of NPSH ^ 0 are an indication of boiling or two phase flow before

the liquid enters the pump.

2. 3 Affinity Laws

For a family of geometrically similar impeller pumps, neglecting

viscous effects,

10
The subscript, m, in the above equations denotes the model or test

pump and T] , T], , and 7] are the volumetric, hydraulic, and total v h

pump efficiencies, respectively. Thus, from model pump test results,

we may compute the diameter, d, and speed, n, of a prototype pump

required to produce a certain Q and H. These laws can also be applied

to the model (or test) pump to predict values of Q, H, and W ., at speed

sh

different than the test speeds (in this case d/d =1). m

2. 4 Stability

The two major requirements of a pump are the volumetric flow

discharged, Q, and the total head, between the inlet and outlet. If H

is not a single valued function of Q there is a danger of surging, i. e. , the throughput hunting back and forth between two values of Q and

generating dangerous or undesirable vibrations. Sudden changes in

pump speed can also cause unstable pump performance. One of the

advantages of polyphase A. C. motors is that the instantaneous power

is constant, resulting in a nearly constant shaft power output free

from torque pulsations.

The shape of the steady state H-Q curve is an indication of pump

stability and depends on certain design parameters. For example,

pumps which have a steep H-Q characteristic give smaller fluctuations

of the discharge, Q, for considerable variations in the total head.

Means of improving stability in a centrifugal pump are:

1) Reduction of the outlet angle, (3^> °^ ^ne impeller vanes.

2) Reduction of the impeller outlet width, b , which increases

the width ratio, d/b .

3) Reduction of the number of impeller vanes giving more

slip.

11
In general, increasing the stability involves a reduction in pump efficiency

and higher blade loading and hence, higher NPSH requirements. Pumps

with unstable H-Q curves can sometimes be used, provided the operating

point is far from the multi- valued region of the curve. For example,

a typical performance curve of a vaneaxial blower [11, 23] is shown in

figure 3 , with the recommended operating range to the right of the unstable

region. One of the disadvantages of reciprocating pumps is their tendency

to produce flow pulsations [6],

A

m Input Power

\Head

\

« Unstable

t Region

Stable

Region

Figure 3, Typical steady state performance characteristics of a vaneaxial blower.

2. 5 Helium Pump Cavitation

Tensile strengths of liquids, i. e. , pressures required to produce

incipient cavitation, are typically orders of magnitude lower than those

predicted from theory [24], The traditional explanation of this discrepancy between theory and experiment is that all real liquids contain submicroscopic nuclei or weak spots which relieve the applied tension.

In liquid helium, the existence of many of the conventional forms of

nuclei are automatically eliminated, e. g. , stabilized pockets of

12
dissolved gas. Therefore, it was originally thought that it might be

relatively difficult to induce cavitation in liquid helium. This, however,

was not the case, as various researchers using different techniques

3 2

found that extremely small pressure drops (of the order of 10 dynes /cm ) were sufficient to cavitate helium [2 5]. At these incipient pressures, the

observation of visible cavitation [49] coincides with audible cavitation noise

in LHel (T > T ) generated from collapsing cavities, whereas in LHe II

A. (T <T ) only audible noise is detectable. The generation of visible caviA. tation in LHe II is made difficult by the efficient heat transfer mechanism,

internal convection. As mentioned earlier, the small vapor formations

which occur at incipience have an insignificant effect on pump performance. Incipient cavitation is generated at local low pressure regions

along the surface of the impeller vanes, usually at the suction side of

the vanes close to the leading edge. To minimize cavitation effects, the

fluid should pass through the vanes with minimum disturbance, i. e. , with zero angle of incidence.

As the pump speed is increased or the NPSH is reduced, cavitation continues to progress. Eventually, a point is reached where the

pump passage ways become partially blocked with vapor resulting in a drop- off in the head. Pumps especially designed with an inducer stage

(high suction specific speed pumps) can handle fluids which contain

relatively large volume fractions of vapor without surging, damage, or head drop-off. An inducer is an axial flow impeller having a few long

vanes rather than many short ones. In effect, the fluid residence time

is relatively high, giving vapor formations time to recondense, leaving

sufficient vane surface to impart the required head rise to the liquid

before it enters the main stage impeller.

13
Cavitation data, obtained with a single pump operating in several

different liquids with widely differing fluid properties, have been used

to determine the relative importance of the fluid properties. One

parameter which only qualitatively describes the ability of the fluid

to block the flow passages in a given pump is the density ratio p /p [22],

t v Values of this ratio, computed for several liquids, are summarized in

table 1. Low values of p„/p indicate that blockage effects will be less

I v

severe.

Table 1. Summary of values of the saturated density ratio

p /p at the NBP for several fluids.

V

Fluid

Water 1603

L°2 256

LN

2

175

LH

2

53

LHe 8

Extrapolation of non-cavitating pump performance from fluid to

fluid and from pump to pump is reasonably well advanced. However,

predicting pump cavitation performance is still not well understood

and is a matter of concern for the designer of high suction specific

speed pumps. The ultimate goal is to be able to predict the cavitation

performance, H = f (NPSH), solely from knowledge of the pump geometry

and the fluid properties at the inlet. The predictive method developed by

Ruggeri, Moore, and Gelder [21] allows the cavitation performance of a particular helium pump to be accurately determined, provided sufficient cavitation data has been taken on the same pump operating in

another fluid, e. g. , liquid nitrogen. Unfortunately, there are no LHe

14
cavitation data to verify the extrapolation of this predictive method

to LHe. Figure 4 shows the predicted helium cavitation performance

of a high suction specific speed, centrifugal pump using data obtained

in liquid hydrogen [21], The predicted helium curve shows that blockage effects occur at much lower values of NPSH than those for liquid

hydrogen, i. e. , at a value of the ratio of head rise coefficients of 0, 9

(equivalent to a head drop-off of 10 percent) the pump in helium requires

only about 6 meters of NPSH compared to 29 for hydrogen. This result

is of practical importance in that for certain applications, pumps of

high N (which tend to be expensive) may not be necessary for

s , c av

satisfactory operation in liquid helium.

The orientation of a submerged He-pump can influence its suction

performance. For maximum NPSH the pump should be mounted vertically with the inlet near the lowest point in the dewar. When a pump

is mounted horizontally a short distance below the liquid level, vapor

filled vortices might be drawn into the pump when the liquid is sucked

downward. This could destroy the pump performance. If the pump

must be mounted horizontally, a 90 degree elbow section, with the inlet

facing the dewar bottom, provides a much better inlet condition [19].

In LHe II, pump cavitation is not anticipated to be a problem

because the high effective thermal conductivity of the liquid makes it

difficult for cavities to grow to visible size [25],

15

2. 6 Pump Efficiency

In a helium pump, inefficiencies are not only expensive in the

additional power required but also produce costly losses of fluid. This

is especially true with liquid helium because of its extremely low latent

heat of vaporization. The economic situation will dictate the need for

gas recovery systems. An analysis by Jacobs [26] of pump losses

involved in the transfer of liquid helium shows that for high discharge

to inlet pressure ratios, P^/P^ , large quantities of helium can be

wasted, even though the pump may be highly efficient. The limitation

of the present survey to relatively low head pumps is consistent with

this analysis. Large diameter transmission lines tend to minimize

P /P . If the fluid pumped is supercritical helium, the thermal effects

caused by pump inefficiency can produce relatively large increases in

fluid temperature and hence decreases in the density. The importance

of pump efficiency in a particular system also depends on whether the

pump is to operate continuously or intermittently and whether the motor

is submerged or not.

When the motor is submerged (location b, figure 1) the losses

are (1-71 ), where T| = Tl Tl , and T] and T| are the pump and motor

t t m m

efficiencies, respectively. Little information is available on the performance characteristics of cryogenic motors. Efficiencies of small

induction motors are typically higher (by about 10 percent) at cryogenic

temperatures than at room temperature but they decrease with motor

size [8]. By using long conduction paths, the heat leak through the

motor electrical leads can be made small. The main attractions of a submersible motor close-coupled to the pump are compactness, elimination of a rotating seal, elimination of heat leak down the shaft, and minimal mechanical problems due to the short shaft. These features are

17
favorable for flight type systems. A superconducting motor, submerged

in liquid helium, might be used to drive a liquid helium circulating pump.

Small, superconducting motors have successfully been operated at speeds

up to 1200 rpm with all parts flooded with liquid helium [27]. However,

the performance characteristics of these motors are not reported. Since

superconducting motors are still in the development stage, their reliability may be poor.

For a pump operated with a room-temperature motor (location a),

the only work which does not reach the pump is that dissipated by the

shaft motion in the fluid and bearing friction. The amount of heat generated will depend on the surface speed and the contact area. Compared

to the shaft work and the heat leak into the cryostat these friction losses

are small and can usually be neglected. This configuration requires a relatively long pump shaft passing through a rotating seal to the motor.

A tubular shaft made of stainless steel rather than a solid shaft greatly

reduces the heat leak from the room-temperature motor to the pump.

Several commercially available room-temperature, rotary seals are

currently available. These seals, guaranteed to be leak tight for vacuum systems, are designed for speeds up to 10, 000 rpm.

Since the density and viscosity of liquid helium is low, it can

readily pass through small apertures. Because of this, a piston type

pump would need close-fitting parts which in turn might produce high

friction losses.

Some efficiency design criteria have evolved from the large

accumulation of pump data. A few of these criteria are:

1) Pump efficiency depends on the specific speed;

e. g. , the highest centrifugal pump efficiencies

occur in the specific speed range of 1500 to 2000.

18
2) A wide centrifugal type impeller of small diameter

gives higher efficiencies than a larger thinner impeller;

the latter shape, however, gives a steeper, more stable

H - Q curve.

3) Pump efficiency increases with Reynolds Number, Re.

4

At Re > 2 x 10 , pump efficiencies are independent of Re.

4) The efficiency of axial flow pumps is more nearly constant

with variations in head rise, compared to other pump types.

5) A vane discharge angle, (3 , of about 22-1/2 degrees should

be used, if possible.

6) All parts which take part in the generation of head should

be streamlined. All fluid passages should be smoothed

by polishing or with special coatings.

2„ 7 Reliability

The reliability of a helium pump depends primarily on the bearing

life and the operation of the motor. Hydrodynamic bearings should be

avoided because of complexity. Since the small surface contact area

of a radial-contact ball bearing will result in a minimum of heat generation, they are recommended for impeller pumps. At helium tempera

tures, bearing heating is not expected to be a problem. Modern cryogenic ball bearings use a glass-reinforced PTFE retainer. The life

of these bearings with small loads over wide ranges of speed are of the

order of thousands of hours. Submerged centrifugal pumps usually

require that one or more bearings are lubricated by the helium, which

has poor lubricating qualities. Redmond and Bott [28] have used a retainer impregnated with MoS^ to provide lubrication of the balls

(440 C stainless steel) at helium temperatures. If the ball bearings

are prevented from skidding during rapid acceleration from rest,

19
intermittent operation should not effect the bearing life. Contaminates,

such as ice crystals, can be abrasive and thus a threat to long bearing

life and pump reliability.

Since positive displacement type pumps (gear pumps) have

relatively high weight and wear rates [8], their reliability may be

poor.

Polyphase cryogenic induction motors are reliable because there

are no brushes to wear out. However, D. C. brush type motors have

been successfully operated in cryogenic liquids and gases [8], At

1 to 10 watts design input, the motor reliability is lower because of

low torque and interference problems due to possible contamination in the

bearings. To minimize the starting torque requirements, stable

pumps of low specific speed should be started with the discharge valve

closed. The design and construction details of small induction motors

which operate at helium temperatures are given by Redmond and Bott [28].

These motors have operated for hundreds of hours at temperatures down

to 20 K and therefore are quite reliable.

3. Survey Results

3. 1 Literature Review

An extensive review of the available literature has been made to

assist in assessing the state of the art of helium pumping. Although no

complete cryogenic helium performance characteristics are reported,

supercritical helium has been pumped in several closed loop heat transfer systems and liquid helium has been pumped by a centrifugal pump

and by a small reciprocating pump. Attempts to pump liquid helium

below the X-point with a mechanical pump have not been reported.

20
Johnston (1946 - 1947)

In his pioneering work on pumping liquid cryogens, Johnston

[29, 30 ] tested a variety of commercially available pumps —gear pumps,

diaphragm pumps, and centrifugal pumps. Although he did not do any

pumping of helium, many of the problems he experienced with these

commercial pumps operating in LN^ should be applicable to the heuristic

problem of using an off-the-shelf water pump in liquid helium. Of the

above pump types tested, it was concluded by Johnston that a centrifugal

pump, with a straight- vane impeller and carefully designed diffusion

passages for outflow, worked the best.

Gottzmann (1959)

In his study of high-pressure liquid hydrogen and liquid helium

pumps, Gottzmann [31] used a single stage, centrifugal forepump to

provide adequate head to the liquid entering a reciprocating, highpressure pump. This work is the first reported attempt to pump

liquid helium and was apparently quite successful. The centrifugal

pump was driven by a close-coupled, canned electric motor which was

immersed in the liquid helium. The head rise and the NPSH required

at the design conditions (presumably for water) was 24 m (80 ft) and

0. 61 m (2 ft), respectively. The forepump apparently worked troublefree during the course of the tests. Unfortunately, neither the details

of the centrifugal pump nor its helium performance data were reported.

Yonemitsu and Okada (1964)

The authors £32] describe the use of a plunger type pump to transfer liquid nitrogen to cool an infrared detector. The reciprocating

plunger is driven by a room-temperature crank mechanism via a stainless steel connecting rod which passes through a warm O-ring

21
seal. The pump output is controlled by varying the speed of a room

temperature motor0 The piston and cylinder are made from a hardsurface material with a very fine finish in order to use the liquid as

a lubricant. Details of the pump dimensions and performance characteristics are not reported,

Parrel and Schoch (1965)

A small liquid helium superconducting pump capable of trans

-6 3

ferring 6. 94 x 10 m /s (0. 11 gpm) with a discharge head of 0. 92 m

-6 3

(3 ft) and 3. 3 x 10 m /s (0. 053 gpm) at a 1.83 m (6 ft) head is

reported by Barrel and Schoch [33]. The pump is 7/8 inches in

diameter and 2 inches long. A superconducting niobium piston,

suspended by a thin-wall nickel bellows, is forced to oscillate at a resonant frequency of 120 Hz by 60 cycle current flowing through a superconducting coil, located below the piston. The flow rate could

be adjusted by varying the amplitude of the coil current. The mass

of the piston (armature) and the bellows spring constant were selected

to produce resonance at this frequency. When the piston moves upward,

hydrodynamic forces cause the suction valve to close and the discharge

valve to open, forcing liquid through the outlet tube. These valves are

disk-type made of 0. 015 inch thick nylon. In reciprocating pumps, the

geometry and spring design of disk-type inlet valves can affect the cavitation performance.

The reliability of this pump depends primarily on the life of the

bellows, which is not well known. The leakage rate past the valves and

hence the volumetric efficiency is not reported. These could be severe

problems in pumps of larger size. Considerably more testing of this

type of pump is needed before a true evaluation of its merits can be

given.

22
Kolm, Leupold, and Hay (1965)

A small reciprocating pump using a metal bellows has been

designed for circulation of supercritical helium by Kolm, et al. [34].

The pump armature is driven by a superconducting coil surrounding the

outside of the pump housing. Two gravity-loaded ball valves provided

unidirectional flow. Both the bellows displacement and cycling rate

could be continuously controlled. A second reciprocating pump which

used a double-acting stainless steel piston was also made by the authors

for the circulation of supercritical helium. When operating at two

6 3

strokes /sec, this second pump produced a flow rate of 6. 4x10 m/s

(0. 1 gpm). The complete performance characteristics of these two

pumps were not reported. The use of ball check valves in liquid have

been known to cause cavitation difficulties [29].

Caine and Pradhan (1967)

The design criteria of small, submersible de stratification fans

(axial flow pumps) are described by Caine and Pradhan [8], These fans

are designed for liquid and gaseous hydrogen but should work approximately the same in helium. A typical fan of 48 mm (1. 9 inch) tip

diameter rotating at 3500 rpm produces a flow of 35 gpm and a head

of about 0. 67 m (2. 2 ft). Typical characteristic curves and model

data at the b. e. p. are given for a 0. 3m (1 ft) diameter axial flow pump,

with N = 16, 000. These axial flow fans are commercially available,

s Wright (1967)

The construction and performance of a double -entry centrifugal

pump for transferring LN^ are discussed [22], Quiet and very reliable

performance are reported. At the time this paper was published, the

pump was commercially available. The author mentions that the pump

should work in helium above the X-point; however, no helium tests were

23
performed. The outside diameter of the pump is 48 mm. It is vertically mounted and driven by a warm, 10-watt, electrical motor with

a long, stainless steel shaft. The impeller (plastic or metal) has six

unshrouded backward-curved blades. Two bearings are used: 1) to

support the impeller, and 2) to prevent whipping of the shaft. The

3

estimated bearing life is 10 hours. The fluid enters the pump from

both the top and bottom of the pump casing through 10 mm diameter

holes. The driving shaft also enters through the top hole. The use of

an unshrouded, double entry design gives low impeller bearing loads

and overcomes problems of vapor accumulation (i. e. , gravity forces

allow vapor formations in the eye of the pump to escape vertically

through the top hole in the casing). Table 2 is a summary of the

reported performance obtained in LN^ .

Table 2. LN pump performance reported by Wright

N Q Head

(rpm) (m3

/s) (gpm) (m) (ft)

2, 850 0. 0 0. 0 1. 73 5.7

3. 3 x

10" 5

0. 53 0. 0 0. 0

10, 000 0. 0 0. 0 15. 7 51. 5

83. 5 x 10' 513. 2 0. 0 0. 0

Mulder (1968)

A small electrically driven centrifugal pump (N = 700) was used

s by Mulder [3 5] to circulate high pressure supercritical helium (25 atm).

A stainless steel shaft was used to minimize the heat leak from the

room-temperature motor to the pump. The centrifugal impeller, consisting of 12 radial blades, was run at speeds ranging from 12,000 to

18, 000 rpm. The performance characteristics of the pump are given;

24
however, these were presumably not measured in cryogenic helium.

The pump appears to be stable and has a maximum efficiency of 50

percent. At the b. e. p. the head and flow are about 77 m and

-4 3 4x10 m /s (6. 4 gpm). Operating problems (if any) are not discussed

by Mulder. A model of this pump is not commercially available [2],

Morpurgo (1968)

A double action piston pump was designed and used by Morpurgo

[ 1 ] to force cool a superconducting system with supercritical helium.

The pump is driven by a warm electrical motor and a long, thin rod and

crank mechanism. The pump speed is variable from 5 to 200 turns per

minute. The maximum speed corresponds to a helium flow of about

-4 3

1. 17 x 10 m /s (1. 9 gpm) at a low differential pressure.

Sixsmith and Giarratano (1970)

The authors [2] describe a miniature centrifugal pump, designed

2

to circulate supercritical helium at high pressures (10 MN/m or 100 atm).

Head- capacity data, obtained with cold nitrogen gas at several speeds,

are included. The pump is stable at all speeds tested and at 30, 000 rpm

produces a maximum head rise at shut-off of about 67 m and a maximum

-4 3

flow of 2. 5 x 10 m /s (4 gpm). This pump is driven by a room temperature motor connected by a stainless steel shaft, about 140 mm in

length. The centrifugal impeller (25. 4 mm diameter) is doubly- shrouded

and is composed of 18 backward curving vanes. In order to eliminate

surging at low flow rates, the pump was made without a diffuse r. It has

several hundred operating hours with helium at temperatures from 4. 2

to 30 K. However, it is not presently instrumented for the acquisition

of helium performance data. Fabrication of this pump is relatively difficult and therefore expensive.

25
There are two features of this pump which might give rise to

poor cavitation performance in liquid helium:

1) The inlet section (as designed) has several sharp corners

which could induce premature cavitation.

2) The high number of vanes gives relatively small flow areas.

Therefore, blockage of the flow passage ways due to cavitation might

be considerable.

Wagner (1970)

A small electrically driven pump and stirring unit for use at

2

temperatures from 4 K to 350 K and pressures up to 10 MN/m is

described [36]. The unit was designed for circulation of gas-liquid

systems for phase equilibrium experiments. A small cryogenic D„ C.

motor was developed especially for this work. The liquid phase is

sucked by an axial flow impeller and pumped into a mixing chamber.

The gas phase is driven by a six-vane, centrifugal impeller. Both

impellers, connected to the same shaft, are driven by the motor. At

3000 rpm (0. 04 watt motor input) the pump produces a flow rate of

- 6 3 5x10 m / s (0. 08 gpm). It has been operated in nitrogen and helium

for continuous periods of 80 and 4 hours, respectively. The estimated

life of the ball bearings is 2500 hrs.

Heron and Cairns (1970)

A small centrifugal pump, used to circulate supercritical helium

(T ^ 4. 5 K) in a closed loop heat transfer experimental system, is

described [ 3 ]. The pump performed trouble-free throughout the experiments. However, the authors did experience some problems with thermal oscillation in the flow loop. The complete head-capacity characteristics could not be obtained because of poor throttling characteristics

26
of a bypass valve. The impeller, 30 mm in diameter, has 12 straight

radial vanes, 4 mm wide. The pump is driven at 18, 000 rpm by a 3-phase room- temperature motor connected to a 60 mm long stainless

steel shaft. The power into the pump was 24 watts. The maximum

5 2

flow of supercritical helium at 4 x 10 N/m (4 atm) at 6. 5 K was

-5 3 estimated to be 7x10 m /s(l.l gpm).

Mark and Pierce (1971)

The authors [37] have used axial flow fans (air blowers) to circulate liquid hydrogen through hydrogen targets. Prior to their operation in LH^, these fans are run for a short time in LHe. The blowers

are similar to those described by Caine and Pradhan [8] but are

operated at speeds far below the design speed and are therefore very

inefficient. These fans also have been arranged in series for higher

output head. Although the fans are adequate, a small centrifugal pump

better suited to the work has been developed and tested in liquid nitrogen.

At the time of the report, work was in progress to build hardware to

test the pump in hydrogen and helium. The pump should have the performance given in table 3.

Table 3. Estimated performance of helium

centrifugal pump reported by Mark and Pierce.

N Q Head

(rpm) (m 3

/s) (gpm) (m) (ft)

3,400 3. 3 x

10" 5

0. 5 3 1. 5 4. 9

10, 000 9. 7 x

10" 5

1. 56 13 42

Specific details of this centrifugal pump are not reported but it is

believed to be similar to Sixsmith and Giarratano's [2].

27
Summary:

Although much of the performance data is sketchy, the literature

review shows that supercritical helium and liquid helium above the

X-point can effectively be transferred by mechanical pumps. The impeller pumps are by far the most popular and appear to offer distinct advantages of stability, wide ranges of operation, and ease of control. Such

pumps are ideally suited for low head high flow situations. Reciprocating

pumps on the other hand offer no obvious advantages (other than their size

and shape) for producing low heads and high flows. Unfortunately, most

of the pumps described are made for specific applications, are not commercially available, and are expensive to make. Since the same types

of pumps are effective in LHe as well as supercritical helium, a single

machine should be able to pump both phases of helium, provided cavitation effects are negligible. The cavitation effects on a pump operating

in liquid helium are not presently known and pumping LHe II using a mechanical pump has apparently, here-to-fore, never been attempted.

The author has recently pumped LHe II (T = 1.9 K) with a small centrifugal pump with no apparent difficulties. The results of these tests are

presented in part 4 of this report.

In future helium experiments which require the use of a pump,

the installation of additional instrumentation to determine the performance characteristics would be of considerable help in the development

of optimum helium pumps.

3

3. 2 Commercially Available Pumps

As noted from the review of the literature, pumps for helium

service are not readily available. In this section the results of

The pumps referred to in this section are coded. A copy of the code

list may be obtained by writing to Dr. Edward L. Brady, Associate

Director for Information Programs, Institute for Basic Standards,

National Bureau of Standards, Washington, D. C. 20234

28
personal communications with various pump manufacturers are reported.

It was hoped that suitable helium pumps might only require slight modifications from the manufacturers standard line of pumps for which a complete set of water performance characteristics are usually available,,

The effort was concentrated on the established cryogenic pump manufacturers. However, many manufacturers of other types of pumps were

also contacted. Most of the commercially available cryogenic pumps are

relatively large, powerful impeller pumps and not designed for submersion or for operation at low flow rates. The helium losses for these

pumps would be considerable. Many of the pump manufacturers contacted by letter did not respond while others thought that pumping helium

would be impossible.

Pump A

Pumps of this basic type are made for LN^ service. It is a

single stage centrifugal pump. For helium service, the pump

would be vacuum jacketed and made of stainless steel.

A complete predicted performance map is available. According to the

- 3 3 manufacturer, it produces 6. 3 x 10 m /s (100 gpm) with a total dynamic head of 159 m (520 ft) and has an efficiency of 60 percent. The NPSH

at these conditions is 3 m (10 ft). It is driven at 7200 rpm by a roomtemperature, 2. 2 kW (3 hp) motor, controlled by a variable frequency

power supply. The price of the pump (including motor and power supply)

is approximately $11, 000.

Pump B

This submersible pump is part of an automatic LN level control

-5 3

system. It will discharge approximately 4„ 9 x 10 m /s (0 o 78 gpm)

against a head of 1. 5 m (5 ft). The maximum diameter of the pump is

about 50 mm. It is driven by an electric motor (10 watts) and costs

approximately $400.

29
Pump C

This is a small, centrifugal pump designed for operation in roomtemperature fluids. This model is made of 316 stainless

steel and would require modifications of the bearings to operate

in helium. It is driven by an electric motor at 10,000 rpm via

a ceramic magnetic -coupling. Use of a magnetic coupling eliminates

the need for a rotating seal. Wagner [36] found that at temperatures

less than 80 K, some ceramic magnets lose much of their residual

magnetism and thus their ability to transmit torque. An investigation

would be needed to see if this is the case with pump C. Stray magnetic

fields from the superconducting system might also affect this coupling

if not shielded. A solution to these problems is to directly couple a room-temperature motor to the pump and use a rotating seal or use

a close -coupled submersible motor. The total length of the motor and

pump is 178 mm and the maximum width is 88 mm. The reported

performance (head-capacity curve) is stable and produces about 12 m

-4 3

(40 ft) head at zero flow and 4. 4x10 m/s(7 gpm) at zero head. The

pump and motor sells for about $100.

Pump D

This is an all-PTFE pump which uses a nutating (360°

oscillating) disk to produce pumping . The pump is sealed with a bellows and can be driven with a 37 watt (1/20 hp) motor. Although this

pump is interesting and novel, materials tend to become brittle at helium

temperatures which might give rise to unsatisfactory performance at

-5 3

these temperatures. It produces 6. 7 x 10 m /s (1 gpm) with a head

of about 10 meters (30 ft). Self- lubrication of the PTFE reduces friction

and wear of parts. The pump has been used successfully with a wide

variety of fluids; however, operation in cryogenic liquids is not reported.

Because of the low viscosity of helium, leakage in this pump might be

considerable. Certainly, more testing is needed before it can be

accepted as a useful cryogenic pump.

30
Pump E

This is a small centrifugal pump designed for room

-5 3 temperature liquids. It produces about 8. 4 x 10 m /s (1. 25 gpm)

at a head of 1 3 m (44 ft). Its appearance is quite similar to pump C and

is magnetically coupled to the motor in order to eliminate a dynamic

seal. Operation of this pump in helium would require modifications

similar to those recommended for pump C.

Pump F

These submersible pumps are of low weight and small

size and are designed for blowing air. Consequently they produce high

flow at relatively low head. Driven by close coupled, submersible

motors modified for cryogenic use, these pumps have been used at

reduced speeds in liquid hydrogen experiments to provide forced circulation. However, at these low speeds, the efficiencies are low. They

have also been run for short times in liquid helium. The head- capacity

curve shows an unstable characteristic at the lower flow rates, see

figure 3 Q These motor-pumps (modified for cryogenic use) cost a few

hundred dollars, depending on the size.

Pump G

This is a submersible 100 mm (4 inch) centrifugal pump

made for pumping liquid propane. Rotating at 3560 rpm,

-3 3

it develops a head of 43 m (140 ft) at a flow of 1. 3 x 10 m /s (20 gpm).

The motor should draw about 225 watts (1/3 hp) at the above output condition when operating submerged in liquid helium. However, this motor

has never been operated at helium temperature. This submersible pump

(including motor) costs about $2600.

31
Pump H

This is a centrifugal, cryogenic pump driven by a roomtemperature electric motor. At maximum efficiency, 44 percent,

-3 3

and a speed of 1450 rpm, it will produce 5 x 10 m / s (75 gpm) with

30 m (100 ft) of head. The pump, made of 304 stainless steel, comes

with flanges at the inlet and outlet for mounting in a horizontal position,

for example, in a pipe line. The manufacturer claims that vapor binding

is prevented by having a vertical top suction. Vapor binding is interpreted as vapor accumulation around the impeller. The cavitation tests

of DiStefano and Caine [19] showed this configuration to be least desirable in a submerged pump. The pump (without motor) is priced at about

$3000.

Pump IThis pump is a submersible, LO pump used for NASA space

applications. The internal parts of the motor do not come in contact

with the oxygen. The motor, when corrected for the specific gravity

of liquid helium, should draw about 66 watts (0. 09 hp). At the design

-3 3

speed, 11,000 rpm, it produces a flow of 2 x 10 m /s (30 gpm) with

a head of 10 m (30. 4 ft). Because of the many special features incorporated for space flight, this pump would be relatively expensive.

Pump J

This is a small stainless steel centrifugal pump designed

for operation in liquids at room-temperature (and higher). It is

driven at 3100 rpm by a canned, 50 watt (1/15 hp) motor and is

similar to pump C. Its head- capacity curve (which is stable) shows

that the pump will produce about 3. 5 m of head at shut -off and at full

-3 3

capacity, 1. 1 x 10 m /s (16 gpm), about 1. 2 meters. The water

32
NPSH required at full capacity is about 5 m (15 ft). The base price

of the pump and motor is less than $100. Modifications similar to

those mentioned earlier for pump C would be needed before this pump

could be used in helium.

Pump K

Submersible axial flow pumps described by Caine and Pradhan

[8] are commercially available. These are high flow - low head

machines of proven reliability in various cryogenic environments.

33
4. Pump Tests

4. 1 Description of the Tested Pump

The test pump, which was designed and fabricated in a few week's

time, consists of an axial flow inducer, a centrifugal impeller, and a two-part pump housing which includes a vaneless diffuser. These parts

are shown in detail in figures 5, 6, and 7, respectively. Figure 8 is

an assembly drawing of these parts. The pump was designed to meet

the following specifications in liquid helium at 4. 2 K with a rotative

speed of 628. 3 rad/s (6000 rpm):

-4 3

a) capacity = 2, 52 x 10 m /s (4 gal/min)

b) total dynamic head = 15.2 m (50 ft)

c) efficiency > 50 percent

d) stable operation from shut-off to full flow capacity

e) cavitation free operation with NPSH ^ 0. 305 m (1 ft).

The inlet and discharge velocity triangles, based on the above design

criteria, are shown in figure 9. The selection of |3 =60° was based

on consideration of the desired steepness of the head-capacity curve and

the required output for a given speed and impeller diameter. The choice

of 3^ is important, since all the non-cavitating design constants depend

on this value. To meet the stability requirement, the centrifugal impeller

was made with a low number of backward curving vanes (six) and a relatively high breadth ratio, D^/B^ . The centrifugal impeller and

inducer vanes were machined with a numerically controlled mill. This

allowed duplicate (or scaled) pieces to be made quickly at low cost.

These rotating parts were statically balanced before final assembly. The

clearances in the pump had to account for differential expansion, AL/L,

between brass and aluminum over the temperature range of 1. 8 to 300 K.

_4 The magnitude of this difference is about 3x10 m/m.

The above pump parts were designed to mate to an existing 3-phase

submersible, cryogenic induction motor rated at 35 watts, as shown

34
0.653"

(16.59 mmT

VIEW A-A

TYPICAL BRASS VANE ( 6 eOCh )

Figure 5. Axial flow inducer; brass vanes are silver soldered into

cuts in the hub.

35
1.965

2"(49.9! mm!

u ! " (19.05mm)

.260 ( (6.6 mm)

.12" R

(3.05 mm)

B2

.035"

(.89 mm)

.031" .134'

(.79 mm) (3.40 mm)

B,

Figure 6. Centrifugal impeller for helium pump, 6061 -T6 aluminum,

anodized finish.

36
uaOu

Gfl Oo

Oh

OX)

37
Stainless Steel Tubing

Helical Compression

Spring

^6-32 NC Self -locking Nut Low - temperature

Bearings with Glass impregnated, PTFE Retainer

Submersible 3-Phase

Induction Motor,

400 Hz, 208 VAC L-L

120 VAC L-N

Power max 3 5 W

3 Holes, .088"dia. (2.24 mm)

for Bearing Cooling

Groove for PTFE O-Ring

035'

(89 mm)

(95.25 mm)

Figure 8. Assembly drawing of helium pump and submersible motor.

38
,= 20.1°

u, = r u) = 397 cm/s

a. Inlet velocity triangle at mean effective inducer diameter,

D = 12. 6 mm (0. 497 in ); n = 628 rad/s (6000 rpm); m _4 3 Q = 0. + Q = 3. 1.6 x 10 m /s; zero prerotation.

u, = 1 568 cm/s T ft

b. Velocity triangle at centrifugal impeller discharge, ; n = 628 rad/s (6000 rpm); Q = Q + Q = 3, I6x 10"'4 m

3 /s.

Figure 9. Inlet and discharge velocity triangles based on the design

criteria.

39
in figure 8. The motor speed is controlled with a three phase, variable

frequency (300 - 500 H ), variable voltage (0 - 120 V L-N) power supply.

This motor, originally made for NASA to drive miniature de stratification

fans [28, 8], is designed to operate over a wide range of temperatures

(4 K < T < 300 K), and since it has been operated for hundreds of hours

in air, gaseous oxygen, liquid nitrogen (LN^) „ and liquid hydrogen

(LH^), is quite reliable. In the present studies, the motor efficiency,

when running submerged in helium is not accurately known but is estimated at 35 to 50 percent [8], Efficiencies of small cryogenic induction

motors typically decrease with the size of the motor.

4. 2 Test Loop and Instrumentation

The pump test loop and instrumentation are schematically shown

in figure 10. The vertically mounted pump is suspended a few centimeters from the bottom of a 1. 22 m deep, stainless steel, cylindrical

dewar by two stainless steel support tubes. The electrical leads for

the submersible motor, liquid level meter, speed sensor, and resistance

thermometers are fed through these support tubes. The fluid is pumped

vertically upward through a venturi flow meter, then passes through a 3/8-inch cryogenic throttling valve (flow coefficient = 1. 1) and is finally

discharged at low velocities through a conical diffuser. The leak rate

of the throttling valve when closed and pressurized with helium gas was

checked in liquid nitrogen by observing extraneous bubbles. Since at

pressures twice those expected in the pump tests, only a few bubbles

were occasionally emitted through the stem, the valve was considered

to be sufficiently tight to determine the head rise of the pump at zero

flow.

The pump speed, n, is determined from the frequency of an

alternating voltage, generated from an electro-magnetic pickup, whose

magnetic field is periodically perturbed by a 20-tooth ferrous gear.

This gear is rigidly attached to the top of the pump shaft. The frequency

is measured byJ a counter. /in
Pressure Lines

He Environment

OSCILLATOR

100 Hz

CAPACITANCE

BRIDGE

LIQUID LEVEL

20 fjk CONSTANT

CURRENT

SUPPLY

SWITCH

Ri R2 R3

RESISTANCE THERMOMETER CIRCUIT

FREQUENCY COUNTER SHAFT

3 PHASE, VAIIAILE VOLTAGE, VARIABLE FREQUENCY POWER

SUPPLY

POLYPHASE WATT

METER

SPEED

MOTOR

POWER

5.7 cm

Centrifugal impeller

Axial flow inducer

,aP.

Extended valve stem

Support tubes

Discharge control valve

Venturi meter

Concentric cylinder

capacitance level gage

R

2 resistance thermometer

at venturi inlet

Conical diffuser

Magnetic speed sensor 20-Tooth, steel spur gear

R3 resistance thermometer

in bearing coolant passage

3-Phase induction motor

Pi

Stagnation pressure tap

at pump inlet

R1 resistance thermometer

at pump inlet

Figure 10. Schematic diagram of He-pump test loop and instrumentation.

41 IB- 103798~|
The fluid discharged from the diffuser mixes the fluid surrounding

the pump and helps maintain a constant bath temperature, determined

by the vapor pressure above the liquid surface. Consequently, when

pumping liquid helium, the height of liquid above the pump inlet is equal

to the NPSH. A calibrated capacitance level gage, shown in figure 10,

is used to measure the liquid level in the dewar and hence the NPSH at

the pump inlet. The sensing capacitor, one leg of the capacitance

bridge, consists of two concentric stainless steel tubes running the full

length of the dewar. Only the very top and bottom of the outer tube are

open to the bath. The bridge output voltage, measured with a DVM, was

somewhat noisy, varying as much as 2. 4 percent during a measurement

interval of 1 minute. The precision of the liquid level measurements,

as typified by the overall standard deviation of the calibration data, is

40 mm.

The fluid temperatures at the pump and venturi inlet , T^ and

T , respectively, are measured with calibrated, 1/8 watt carbon

resistors, R,12and R . The resistor R„ is positioned halfway inside

2 r 7 the discharge tube and tightly sealed from the helium bath. A third

resistor, R , located inside the motor and near the top, is used to

detect any significant temperature increase due to possible bearing

heating; none was observed. These thermometers, powered by a constant

D. C. current of 20 \s A ± 0. 01 |j,A, were calibrated in place against the

saturation temperature of LHe (in equilibrium with its vapor) over the

range 1, 75 K<T< 5, 2K. To extend the calibration range to 20 K, the

three thermometers were immersed in liquid hydrogen. The voltage

drops across R^, R , and R^ were measured on the same DVM mentioned

earlier. Steady background thermal emf1 s were subtracted out by taking

two readings, one with reversed current. During the calibration, the

42
system pressure, maintained constant with, a manostat to within

2

930 N/m (7 mmHg), was measured by directly observing one of three

manometers, depending on the pressure range. Systematic errors in

the T-R calibration data can occur due to thermal stratification in the

fluid and slight changes in resistance with time. Errors in temperature

can be as high as 2 K in some liquid helium systems with severe stratification. Past experience using carbon resistor thermometers in helium

indicate that they are reproducible to within 0. 010 K over a period of a few months.

Over the temperature range 1. 7 K < T < 20 K the calibration data

fit a logarithmic expansion of the form

6

£n T = C + Y C. ((nR)

1

, o . , i

i = i where C , C , . . . C. are constants. With measurements of R , o 1 6 1 R , and R , this function is used to compute T , T , and T in the

analysis of pump performance. With these curves, the total rms error

of the calibration data is 0. 86 percent, 0. 55 percent, and 0. 44 percent,

respectively for T^, T^, and . The differential pressure across the pump,AP , and across the

P

venturi , AP , are directly measured with differential, bellows -type,

pressure gages. These measurements are used in the calculations of

the total dynamic head and the rate of volumetric flow discharged,

respectively. The pump discharge static pressure, P^ , and the ullage

pressure, P , are measured with manometers. These manometer

u

measurements are used to check the direct measurements of AP and

P

to compute the fluid densities at the pump inlet and discharge. At

system pressures equal to or greater than ambient, the manometer

43
measurements of and P are gauge pressures. With system pressures less than ambient, P is measured with an absolute manometer

u

and P is determined from P = P + AP , where P is the sum of

2 Z 1 p 1 pressure due to the hydrostatic fluid head above the pump inlet and P

When pumping liquid helium, the unknown location of the liquid

vapor interface in the pressure lines is a source of both random and

systematic error in these pressure measurements. This problem can

be eliminated by using vacuum insulated pressure lines with a heater

attached at a known location on the inner lines. This step, which greatly

complicates the system and gives additional heat leak, was not required

in the present work. In the differential pressure measurements, it is

assumed that this interface occurs at the same elevation in each of the

two lines since temperature differences between the different pressure

lines were minimized by wrapping them with copper wire and soldering

them altogether at several points.

Both differential pressure gages, sized for the pump design criteria,

were calibrated over their full range against a differential manometer

filled with distilled, room-temperature water. The precision of the

AP gage (range 0-20 inches of H O) , estimated from the standard

deviation of the calibration data, is about 0. 8 percent of full scale. The

absolute error in AP is estimated to be less than 1. 8 percent at the v

high flow rates and increases to about 8 percent at the lowest recorded

flow rates. Similarily, the precision of the AP gage (range 0-100 inche

P

of H O) is 0. 4 percent of full scale. The absolute error in AP is

2 p

estimated to be less than 4. 0 percent over the range of AP's typically

produced by the pump. The response time of both AP and AP systems

v p

(i. e. the pressure lines and gages) to sudden changes of differential

pressure in the dewar is less than 5 seconds.

44
In order to accurately measure the ullage pressure, P , in

5 2

U

the dewar over the entire range from 0 to 3 x 10 N/m , three manometers were used which covered the following ranges of absolute

pressure:

5 2

0 < P < 0, 6 x 10 N/m , Maximum, estimated error of

u

pressures measured in this

range is 8. 3 percent.

5 2

0„ 8 < P ^1.4x10 N/m , Maximum, estimated error of u pressures measured in this

range is 0„ 5 percent.

5 5 2 1.4x10 <P < 3. 0x10 N/m , Maximum, estimated error of

u

pressures measured in this

range is 0. 1 percent.

Significant thermal-acoustic pressure oscillations were not

observed in any of the pressure gages during the course of these experiments. This is a fortuitous result, since in general, thermal oscillations

and instabilities have been known to be problems of major importance in

helium systems [40], The elimination of such oscillations may partially

be due to the construction of the stainless steel pressure lines, which

typically consist of a length of a small diameter tube (1 0 4 mm i. d. ), originating from the pressure source, followed by a larger tube

(2. 9 mm i. d. ) which leaves the low temperature environment. The

different sizes of line represent resonant cavities of different volumes.

Therefore, the possibility of system resonance to oscillations of any one

frequency are presumably reduced.

The average 3-phase power into the motor was measured with a two element, polyphase wattmeter. The power and frequency ranges

of this meter are 0-60 watts and 2 5 to 800 Hz, respectively. The

estimated accuracy of the power measurements is 0. 75 percent.

45
The venturi meter used to measure the helium flow rate is shown

in detail in figure 11. Since the venturi is not calibrated, its dimensions,

which are based on standard forms, had to be machined accurately in

order to use the recommended [41] discharge coefficient, which is a

5 function of the tube Reynold's number, Re. When Re > 2 x 10 , the

discharge coefficient is constant and equal to 0. 984. It has been found

experimentally that the discharge coefficient is almost the same for

compressible fluids and incompressible fluids for well designed venturi

meters [17], The internal dimensions of the venturi were obtained by

pouring molds of the inside contours with dental plaster, whose setting

contraction is 0. 05 percent. The external mold dimensions were then

measured with an optical comparator. The errors in fluid density and

in AP both contribute to the total error in the volumetric flow rate, v estimated to be 4„ 1 percent at the lowest values of flow and 1. 1 percent

at the highest. Flow measurements in LN^ and LH^ have also been

made in previous work, using similar venturi meters [42,43],

4. 3 Results of Pump Test

During a typical (non- cavitating) performance test, the pump,

started with the discharge throttling valve closed, is slowly brought

up to the desired speed, n, which is maintained constant to within

± . 3 percent. The measured head rise across the pump typically

4

reaches a steady value within 20 to 30 seconds. This comparatively

long time appears to be associated with the creation and decay of

perturbed temperature distributions in the pressure lines due to

movements of cold fluid into the initially warmer lines. At this

shut-off condition, AP = 0, indicating zero flow through the venturi

meter. There is, however, some flow through the bearing cooling

passages, shown in figure 8. The magnitude of this coolant flow, Q ,

~*

! !

Similar waiting times were also characteristic of the measured head

rise when the pump was started with the discharge valve fully open.

46
(tutu 8*;

(UJW98 -QlT

47
is not accurately known, but can be estimated from the relationship which

describes orifice flow,

where C is the discharge coefficient, A is the orifice area (11. 8 mm ), and AP/p is the static head loss across the orifice. At the pump design

speed, Q is estimated to be about 25 percent of the design flow capacity.

When the pump tests were finished, it was concluded that bearing heating

was insignificant and that this forced cooling of the bearings was

unnecessary.

The pump flow capacity, Q, is altered by throttling the discharge

valve. For each valve setting, the static pressure rise, AP , the

P

power consumption to the motor, W, and the helium flow rate through

the venturi, are simultaneously measured. In this way, the pump

characteristics, H = f (Q), W = f (Q), and total efficiency = f (Q) are

determined. Data are taken for both increasing and decreasing values

of Q. The total dynamic head, generally defined for steady state

conditions, is

(11)

Z

(12)

1 and is directly related to the power, W^, expended by the pump on

the fluid, where

W

f = mg (H) . (13)

In the application of (13) to the present study, rh is the rate of mass

flow through the discharge tube, measured by the venturi meter. However,

48
the pump is also expending a certain amount of useful work in pumping

fluid through the bearing cooling passages; e„ g. , at 628 rad/s (6000 rpm)

the magnitude of this rate of work is estimated to be about 0. 8 watts.

This contribution is neglected in calculations of efficiency.

If the fluid is isothermal and incompressible, the integral in

(2) is simply (P - P )/gp. The first two terms in (2) are small compared to the third. Therefore, the percent error involved in calculating

the total dynamic head is approximately the same as the error in

AP measuring ^- , which is believed to be less than 5. 3 percent. Since

P typical measured temperature differences (T - T ) are only a few

hundreths of a degree, the isothermal approximation is valid. However,

the incompressible assumption for pumping helium must be examined

closely. In the present tests, the values of (P - P ) are at the most

4 2 . . 1. 5 x 10 N/m (0. 15 atm), and hence, the error in the incompressible

approximation is small when pumping liquid helium. For the tests with

supercritical helium, this assumption is not valid and the integral must

be evaluated numerically.

The pump efficiency, T) , is the fluid power divided by the shaft

P

power, i. e. ,

fl = WJW .

, p f sh

where W , = T 'ud. In the present tests, the motor is close -coupled to

sh

the pump and submerged in helium. Therefore, conventional torque

measurements could not be made to determine W , . The motor efficienc

sh

was not accurately known either. Because of these difficulties, the

total efficiency is measured and is the product of the motor efficiency

and the pump efficiency,

71 = I] Tl = W./W . t m p f

49
The error in W is estimated to be less than 9. 4 percent at low values

f of the mass flow rate and 5. 2 percent at the high values.

Effects of cavitation on pump performance are determined by

maintaining a constant liquid inlet temperature, a constant pump speed,

and a given valve setting while the NPSH decreased due to helium boiloff.

The total dynamic head is independent of reductions in NPSH until a critical NPSH is reached. At this point the pump passages become partially

filled with vapor, resulting in a substantial drop-off in the head. Cavitation effects could also be observed by inducing two-phase flow in the

pump inlet as follows: The pump is run at a constant speed, head, and

flow rate at values of NPSH greater than critical with the dewar vent

valve closed. The saturated liquid then self-pressurizes to a certain

value. With the pump operating, the vent valve is suddenly opened. This

sudden depre ssurization results in superheated liquid which then boils

violently with appreciable entrained vapor at the pump inlet. Equilibrium

and steady-state performance are restored after about 30 seconds. In

this test, the pump cavitation performance was similar to the low NPSH

experiments.

The pump data are summarized in table 1. The accumulated test

time in liquid helium is about 8-1/3 hours and in supercritical helium

about 2/3 hours.

4. 3. 1 LHe I, 2. 3 K < T < 4. 3 K

Typical performance characteristics of the pump operating in LHe I

at a nominal speed of 631 rad/sec (6030 rpm) are shown in figure 12. At

-4 3 flow rates near the design capacity (2. 52 x 10 m /s) the total dynamic

head is only about 1/3 of the design head (15. 2 m). The reasons for

this discrepancy are not known at this time. In general, head losses

can be caused by cavitation, entrance losses at the leading edges of

the impeller blades, friction and hydraulic losses in the impeller

50
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52
ft

Total Efficiency {%)

a H-Q

' DESIGN

POINT

Total Average Power

(Watts)

at

M

Total Dynamic Head(m)

Fluid Power (Watts)

^7

I I I I I I I I I I I I

1

I I I I

1 I I I I1 I I I I 0.5 1.0 1.5 2.0

Q .CAPACITY, m

3/s

2.5 3.0x10

Figure 12. Typical performance characteristics of He-pump operating

in LHel; T = 4. 15 K; n = 631 rad/s (630 rpm);

0 o 19m < NPSH £ 0. 7 m.

53
passages, and losses in the diffuser. The latter three factors are

associated with the hydraulic efficiency, T) . In these tests, cavitation

is ruled out as a possible explanation because the developed head was

independent of the NPSH when NPSH > NPSHC

. At the design point,

the specific speed, N

g = n (rpm) Q(m 3 /s)/TDH(m) 3 / 4

, is 12. 3,

(638 [ft, gpm]). For small, single stage centrifugal impellers, the

hydraulic losses, 1 -- T| , increase rapidly at low specific speeds and

might be as high as 0. 35 for the test pump [12],

The maximum total efficiency, about 16 percent, occurs at a -4 3

flow of about 1. 5 x 10 m /s. The specific speed at this point is

14. 8 (765 [ft, gpm]). Using the typical values of the motor efficiency [8]

(about 35 percent at 10 watts) the maximum pump efficiency, T| is

P

estimated to be about 46 percent. The H= f (Q) characteristic is seen

to be a monotonically decreasing function which is an indication of good

stability. The shape of this curve is similar to that typically found

with low specific speed pumps [39],

The pump was tested in LHe I at several different speeds and

with different bath temperatures. The data points produce smooth

TDH = f (Q) as shown in figure 13. Constant speed data taken at different temperatures is shown in curve 5. The affinity laws for a given pump are:

54
3.0x10

-4

Q CAPACITY, m

3

/s

Figure 13. Total dynamic head- capacity performance curves in

LHel; 3. 1 8 K £ T < 4. 15 K.

55
The subscript "1" in the above equations denotes known test conditions

and Tj , T| , and T] are the volumetric, hydraulic, and total pump

v h p

efficiencies, respectively. These laws can be applied to a test pump

to predict values of Q, H, and W , at speeds different than the test

sh

speeds. It is generally assumed that the efficiency ratios are equal to

unity, in which case the head-capacity curves are parallel [44], e. g. , in figure 13, curve 3 is predicted from curve 2. However, curves 4, 5,

and 6 are slightly steeper than the low speed curves, 2 and 3; i. e. , values of H, and Q predicted for the high speeds from the low speed

data, are typically higher than the actual data points. Such effects have

been reported by others [2] and indicate that the volumetric and hydraulic

efficiencies are slightly less at the higher speeds. From (4), the higher

rotative speeds give higher heads at the impeller tip and hence, higher

leakage losses and rates of bearing coolant flow. Such "losses 11 tend to

shift head-capacity curves to the left [ 1 ]. Hence, plugging of the bearing cooling holes would probably give data which are in closer agreement

with the affinity laws.

Liquid nitrogen data are given in curves (1) and (2) of figure 13.

The nitrogen runs were restricted to lower speeds because of the motor

power limit. At the same speed, curve 2, the nitrogen data are not

significantly different from the LHe I data. This result is important

from a practical standpoint, since the cost of liquid helium is much

higher than LN^ . In these tests, the pump Reynolds number at the

inlet, about one order of magnitixde higher with LHe than with LN^,

has no apparent effect on the performance. Typical values for LHe I 5 4 and LN^ are 4 x 10 and 5x10 , respectively. Reynolds number

effects in centrifugal pumps are usually secondary and are negligible

4

above 2 x 10 [45] .

56
The effects of cavitation on the H-Q performance were investigated at two different speeds, the design speed, 628 rad/s, and the

maximum speed produced by the power supply, 749 rad/s. The helium

bath temperature was nominally 4. 1 K. The data shown in figure 14,

were taken with the throttling valve fully open. The limiting or critical

NPSH (NPSH ) is only about 0. 09 m ± 0 o 04 m ( 3. 5 in ± 1. 6 in) . At

c values of NPSH >NPSH , the head is essentially independent of the

c NPSH. However, when NPSH < NPSH , cavitation causes the head to

c drop off rapidly. In this regime, the rotative speed of the induction

motor tends to increase by several percent and the pressure drop

across the venturi meter decreases, presumably because of reductions

in fluid density. Since the percent error in the NPSH measurements

is relatively high at these low values, the shapes of the cavitation curves

are not precisely known. The limiting NPSH is significantly lower than

expected. This is an important result. It suggests that a pump of

simple design can be used to pump liquid helium with relatively little

liquid cover. The tendency of cavitation to block the impeller flow

passages is probably less severe with LHe I compared to other liquids

because of the exceptionally low density ratio, p / p % 8 at the normal

t v

boiling point [22] . Because this ratio is temperature sensitive, the

limiting helium NPSH is also expected to be temperature dependent.

Variations in the pump NPSH with inlet temperature have previously

been reported by Ruggeri and Moore [21] for liquid hydrogen. The

above explanation is only qualitative and is certainly not the whole story.

Additional helium cavitation data are needed to give insight into the

general problem of accurately predicting cavitating pump performance.

By replacing the existing inlet nut, shown in figure 8, with one which

is hydrodynamically smooth, even better cavitation performance might

be possible.

57
o 1.0

OS

o

cr

0.6

z

o

Ll. 0.4

u_

UJ

o

u

UJ 02

crQ

<

UJ 0

X

TO

T = 4.15 K

n= 628 rad/s (6000 rpm)

0,= 0.35

\// = 6.556

M,NC

T = 4.09 K

n = 749 rad/s (7154 rpm)

0.34

W/= 6.596

rl,NC

0.4

NPSH ,m

0.6 0.8

Figure 14. Pump cavitation performance in LHel.

58
4. 3 P 2 Supercritical Helium, 4.7K<T<5. 6K

The pump was given one test run near the design speed in super5 2 5 critical helium. During this test, 2. 69 x 10 N/m <s P > 2.77 x 10 N/m

and 4. 80 K ^ T^ ^ 5. 51 K. Without the stabilizing influence of bath liquidvapor equilibrium, the pump temperature was difficult to control and

tended to increase with time. In the derivation of the total dynamic

head, eq (2), a steady-state process is assumed. In this test with

supercritical helium, the process is really one of unsteady- state.

However, since the time involved in the acquisition of each data point

is short, the use of the total dynamic head is still considered meaningful.

The integral f dP/p is evaluated numerically for each data point using

1 J

the equation of state [12] for supercritical helium and assuming an

isothermal process. The resulting performance curves are shown in

figure 15, with the values of the average pump temperature and density

shown next to the H- Q data points. The final values of fluid density near

shut-off are about 1/2 the starting density near full capacity. That the

highest temperatures and lowest densities occurred near shut-off is a result of the test procedure and not a property of the pump.

Superposition of the H-Q curve in figure 15 over that obtained

with LHe I at near the same speed, figure 9, shows that values of H

obtained in supercritical helium are considerably higher than those of

LHe I at comparable values of flow capacity, which should not be the

case [15], This discrepancy might have been caused by some systematic

error in the pump differential pressure measurements and in the temperature measurements obtained from R . A thermodynamic analysis [46]"

5 2

of the pump conditions (5. 5 K, 2. 5 x 10 N/m , and 50 percent efficiency)

5 2 shows that values of dT/dP can be about 1 K/10 N/m . Then with

4 2 a dP ~ 10 N/m (. 1 atm), which was typically observed, dT ^ 0. 1 K.

59
4

0 .CAPACITY, m

3/s

Figure 15. Pump performance in supercritical helium, n = 630 rad/s

(6011 rpm).
This is about one order of magnitude higher than that obtained with a similar analysis for the liquid state. Such variations in temperature

could give errors in fluid density of approximately 20 percent in this

near- critical range, which could possibly account for some of the

discrepancy. The measured temperature difference (T - T ) however

is only a few hundreths of a degree, but since is located some 100 mm

from the impeller outlet and the thin wall discharge tube is not insulated,

some heat transfer between the bath and the discharge tube occurred.

Any heat transfer between the two should tend to reduce the temperature

of the discharge stream at station 2.

4. 3.3 LHe II, 1, 8 K < T < T.

A At saturated vapor pressures, a transition from LHe I to LHe II occurs when the temperature T = 2. 172 K. At this temperature, called

A. the X-point, visible bubble formations which normally occur during

vacuum pumping of the bath suddenly cease, a result of the liquid's

anomalously high thermal conductivity; the transition from LHe I to

LHe II can therefore be easily recognized. At T > T , LHe I behaves

A as an ordinary Newtonian fluid, and is referred to as "normal helium. "

In contrast to this, however, the macroscopic properties of LHe II at

T < T cannot be described by the classical laws of fluid mechanics.

A Its behavior is most simply described phenomenally by a two-fluid

model. The two fluids, termed the "normal fluid" and the "superfluid, " have temperature dependent densities such that

P + P =.P

,

n s

where p is the ordinary (observed) density of liquid helium. Since the

4

liquid at T < T is an assembly of He atoms which are all identical,

A the two-fluid model does not imply that LHe II is a mixture of two real

61
fluids. This model, which correlates a wide variety of experimental

results, postulates that the superfluid component carries zero entropy

and experiences no flow resistance,, Comprehensive treatments of

the two -fluid model, experiments done in LHe II and their interpretation

are given, for example, by Donnelly [47] and by Wilks [48].

Attempts to use a mechanical pump to transfer LHe II have not

been reported in the literature. The present experiments in LHe II are therefore believed to be the first attempts to pump superfluid helium.

Although they were not designed or intended to investigate the basic

physics of pumping superfluid, the two experiments in LHe II showed

that LHe II (with approximately 50 percent mass fraction superfluid) can

easily be pumped without significantly increasing the temperature of the

fluid.

In one experiment, the pump was run at a constant speed and

throttling valve setting while the bath temperature was reduced from

above to below the \-point. The results of this experiment are shown

in figure 16. The total dynamic heads and the volumetric rates of flow

measured in LHe II are not significantly different than those in LHe I. With these data, the main impeller tip speed and relative velocity at

are of the order of 10 m/s. Therefore, the superfluid component is

envisaged to be highly turbulent. At the lowest temperature, T % 1, 9 K,

approximately 50 percent of the liquid mass is superfluid. In these

data, the measured temperature rise across the pump, (T - T ), is

of the order of a few hundreths of a degree.

In another experiment the apparent constant speed performance

characteristics were measured. The term "apparent" is used because

in LHe II measurements of the rate of volumetric flow using a venturi

meter are ambiguous [38]; i. e. , the flow of the superfluid component

through the venturi may not contribute to the measured pressure drop.

The same argument can be applied to the differential pressure measured

across the pump. The resulting pump performance data taken at a
3.0x10

4.0

Figure 16. Pump performance in LHel and LHell; n = 340 rad/s

(3245 rpm).

63
nominal temperature of 1. 89 K, are given in figure 17. The curve

drawn through these H-Q data is obtained from the low speed pump data

obtained with LHe I (curve 2, fig 13) and use of the affinity laws. The

solid circles represent data taken several days previous to those represented by the open ones. The curves of total efficiency and total

average power as a function of Q (fig. 17) are dissimilar compared to

those typically obtained in LHe I (fig. 12) and supercritical helium

(fig. 15). As mentioned earlier, it would be more meaningful to plot

the shaft power and the pump efficiency rather than the total quantities.

Measurements of the shaft power to the pump should be incorporated in

future work.

It should be noted that the overall performance in superfluid helium

was not greatly different from that which would be scaled by classical

hydrodynamics from measurements in normal helium- -the difference

being certainly less than a factor of two. Probably this comes about

because the impeller tip velocity far exceeds the "critical" velocity of

helium (a few centimeters per second) at which vorticity is generated

in the superfluid. The tangled mass of vortices, so generated, have

been shown to move approximately in accordance with classical hydrodynamical equations in many experiments. Our results are in qualitative

accord with this fact.

During the experiments with LHe II, the time for both differential

pressure gages to reach steady values was measurably increased. The

significance of this observation is not well understood at the present

time and may partially be caused by film flow inside the pressure lines.

However, the slower response times did bring to mind critical velocity

experiments which show that the superfluid component will (sometimes)

not interact with a solid surface moving relative to the fluid because of

its frictionless like behavior. In future pump experiments in LHe II, it

would be interesting to see if critical velocity effects exist.

64
15

10

° D D

D

TOTAL EFFICIENCY {%)

w

TOTAL AVERAGE POWER

( watts

)

H

TOTAL DYNAMIC HEAD (m)

.4

0 3x10

APPARENT CAPACITY, mVs

Figure 17. Performance characteristics of He -Pump operating in

LHell; T = 1. 89 K; n = 339 rad/s (3237 rpm);

0, 105 m < NPSH 0. 204 m.

65
4. 3. 4 Comparison of the H-Q Data

Using the pump affinity laws, all of the H-Q data were normalized

to the design speed, 628 rad/s. The results, summarized in figure 18,

show that the LHe I data fall in two zones depending on the speed. The

two zones appear to be off- set by a pure translation in Q about the same

order of magnitude as the estimated "leakage" through the bearing cooling

-4 3

holes, Q ?s 0. 6 x 10 m / s. At the lower flow rates, the supercritical

data are considerably higher than all the other data. The LHe II data

are dispersed, with several points falling below the two zones. Possible

reasons for these latter deviations have been discussed earlier.

4. 3. 5 Reliability

The reliability of the helium pump depends on the life of the low

temperature ball bearings and the operation of the submersible motor.

The induction motor used in these tests has good low temperature starting

characteristics and is highly reliable because there are no brushes to

wear out. The helium environment provides adequate bearing cooling

but gives little lubrication to the balls. The ball bearings used in these

experiments, made of 440 C stainless steel, are retained with glassreinforced PTFE, impregnated with M S_, . The M provides the

o 2 o 2

lubrication. Under normal operating conditions the maximum bearing

5

life is estimated to be of the order of 10 hours. Since skidding of the

balls is prevented during rapid motor acceleration by a spring preload

(see fig. 8), the bearing life should not depend on the duty cycle. Impurities such as ice crystals can be highly abrasive and therefore are a threat to long bearing life and pump reliability. For this reason, special

care was taken to eliminate traces of water from getting into the helium

bath.

The start-up reliability of the pump was studied in LHe I at T = 4. 1 K by periodically starting and stopping the pump, while subjected

to a certain load. The nominal load condition, n = 628 rad/s (6000 rpm),

66
SC He 4.80 K < T S 5.51 K

n = 629 rod / s (6011 r/min)

LHeH 1.85 K < T < 1.95 K

n = 339 rod /s (3237 r/min)

LHe I T = 4.02 K

204 rad/s (1952 r/min) < n < 393rad/s

(3754r/min)

LHe I 2.23 K < T < 4.20 K

582 rad/s (5553 r/min) < n < 749 rad /s (7154 r/min)

.8 1.2 1.6 2.0

Q , CAPACITY, mVs

2.4 2.8 3.2x10r4

Figure 18. H-Q data normalized to the design speed, 628 rad/s (6000 rpm)

67
head rise ^5 m, and Q % 2. 5 x 10 m/s, was chosen arbitrarily. Typical time response curves of the head and capacity are shown in figure 19.

The pump is allowed to run at steady conditions for about 60 seconds.

Fifty (50) such test cycles were performed with no anomalous behavior

observed. For all 50 values of the steady state data, the maximum variation in head was 4. 0 percent and the maximum variation in capacity was

5. 5 percent. Approximately 30 additional switchings were also performed

during the course of running the standard performance tests.

5. Conclusions

The following is a summary of conclusions formulated from the

results of this work.

1. A pre-induced, single-stage impeller pump is a simple, controllable, and reliable method of generating forced helium flow.

2. A single-stage inducer-impeller pump of simple design can

be used to pump all phases of helium from 1. 85 K to 5. 5 K.

3. A simple inducer-impeller pump can operate satisfactorily in

LHe I and LHe II with little liquid cover. This result may not be

universal and needs further testing and evaluation.

4. Results of non-cavitating pump performance in liquid nitrogen

can be used to estimate non-cavitating pump performance in LHe I.

5. The pump affinity laws are applicable to the non-cavitating

helium pump performance curves provided leakage effects are

accounted for. Additional work is required to improve the design

of helium pumps so that the design point is closer to the measured

performance.

6. The non-cavitating performance of geometrically similar pumps

of different size might be estimated using the results of the present

tests and the pump scaling laws [16, 15, 11], Since the validity of

the scaling laws applied to helium pumps has yet to be tested, the

accuracy of such predictions would be unknown.

68
n =628 rad/s (6000 rpm)

J I I L 20 40 60

TIME,s

80 100

Figure 19. Typical reliability test cycle.

69
7. The helium performance of different pumps whose specific

speed at the best efficiency point is approximately 15 (780 [ft, gpm]),

might also be estimated using the results of the present tests, see

for instance Stepanoff [H], p0 176,, Such predictions are subject to

the same uncertainties as noted in 6.

6. Recommendations for Future Work

Further insight into the detailed behavior of the pump and the

relative importance of some of the parts would have been determined

from the following:

1) Determine the pump efficiency from torque measurements

or from motor efficiency,,

2) Obtain additional pump cavitation data in liquid helium and

in liquid nitrogen or hydrogen to evaluate the performance prediction method developed by Ruggeri, Moore, and Gelder [21,39]=

3) Eliminate the bearing cooling flow of the present design.

We know now that this is not necessary and it adds uncertainty

to the flow measurements.

4) Compare performance with and without the inducer.

5) Streamline the entrance hub nut for better cavitation

performance.

6) Increase the discharge vane angle of the centrifugal impeller

to give a higher head at the design flow.

7) Add pressure taps in the pump housing to determine

1) leakage flow past the inducer,

2) the head rise across the inducer and the centrifugal

impeller,

3) the diffuse r efficiency.

Such instrumentation would give insight for improving the

present design.

Of the above recommendations, items 1 and 2 are considered to be the

most crucial.

70
Systematic errors which may have occurred due to thermal

gradients in the pressure lines could possibly be eliminated in future

work by using pressure transducers designed to work at helium temperatures. This would also be a necessary step for automating the data

acquisition and would help remove the ambiguity from pressure measurements made in LHe II.

71
7. References

Morpurgo, M. , A superconducting solenoid cooled by forced

circulation of supercritical helium, CERN Report 69-25,

Geneva (1969).

Sixsmith, H„ , and Giarratano, P, , A miniature centrifugal pump,

Rev. Sci. Inst. 41 , No. 11, 1570-1573 (1970).

Heron, K. , and Cairns, D. N„ H. , Experiments on a pumped

supercritical helium test loop, Central ElectricityResearch Laboratories Report RD/LIN 190/70 (1970).

Keilin, V. E„ , and Klimenko, E„ Yu, Cryogenic problems of

forced cooled superconducting systems, Cryogenics 1

2

, No. 4, 292-296 (1972).

Snowden, D. P. , Superconductors for power transmission,

Sci. Am. _226, No. 4, 84-91 (1972).

Scott, R. B. , Cryogenic Engineering, pp. 263-267 (Van Nostrand,

New York, N. Y. , 1959).

Daney, D. E. , Ludtke, P. R. , Chelton, D. B. , and Sindt, C. F. , Slush hydrogen pumping characteristics, Nat. Bur. Stand.

(U.S.), Tech. Note 364 (1968).

Caine, G. H. , and Pradhan, A. V. , Pumps or fans for

destratification of hydrogen liquid and gas, Book, Adv. in

Cryogenic Eng. 13_ , Ed. K. D. Timmerhaus, pp. 728-738

(Plenum Press, New York, N. Y. , 1968).

Urasek, D. C. , Meng, P. R. , and Connelly, R. E. , Investigation

of two -phase hydrogen flow in pump inlet line, NASA

TN-D-5258 (1969).

Wright, B. M. , Centrifugal pumping of liquefied gases,

J. Sci. Instruments 44 , 469-470 (1967).

72
11„ Stepanoff, A. J. , Centrifugal and Axial Flow Pumps, 2 Ed0 (Wiley and Sons, New York, N. Y. , 1957).

12. McCarty, R. D„ , Thermodynamic functions for helium 4 for

temperatures from 2 to 1500 K with pressures to

2

100 MN/m (1000 atmospheres), Nat. Bur. Stand.

(U.S.), Tech. Note 631 (1972).

13. Cunningham, R. G. , Hansen, A. G. , and Na, T. Y. , Jet

pump cavitation, ASME Paper No. 69-WA/FE-29 (1969).

14. Wislicenus, G. F0 , Fluid Mechanics of Turbomachinery

(McGraw-Hill, New York, N. Y. , 1947).

15. Kovats, A. , Design and Performance of Centrifugal and Axial

Flow Pumps and Compressors (Pergamon Press, New York,

N. Y. , 1964).

16. Lazarkiewicz, S. , and Troskolanski, A. T. , Impeller Pumps

(Pergamon Press, New York, N. Y. , 1965).

17. Bird, R. , Stewart, W. E. , and Lightfoot, E. N. , Transport

Phenomena (Wiley & Sons, New York, N. Y. , I960).

18. Schlichting, H. , Boundary Layer Theory, 4th Ed. (McGraw Hill,

New York, N. Y. , I960).

19. DiStefano, J„ F. , and Caine, G. H. , Cavitating characteristics

of tank mounted cryogenic pumps and their predicted

performance under reduced gravity, PESCO report (1961).

20. Jacobs, R. B. , Martin, L. B. , Hardy, R. J. , Direct

measurement of net positive suction head, J. of Basic

Engineering 81 , 47-152 (1958).

21. Ruggeri, R. S. , and Moore, R. D. , Method for prediction of

pump cavitation performance for various liquids, liquid

temperatures and rotative speeds, NASA Tech. Note

D-5292 (1969).

73
22. Wright, M. K. , Design comments and experimental results for

cavitation-resistant inducers up to 40, 000 suction specific

speed, J. of Engineering for Power 8_6, 176-180 (1964).

23. Bulletins C-5246, C-5166, and C-5126, Globe Industries , Inc , 2275 Stanley Avenue, Dayton, Ohio 45404 (1967).

24. Apfel, Rc E. , The tensile strengths of liquids, Scientific

American 227, No. 6, 58-71 (1972).

25. Finch, R. D. , Kagiwada, R. , Barmatz, M. , and Rudnick, I. , Cavitation in liquid helium, Phys. Rev. 1 34 , No. 6 A,

1425-1428 (1964).

26. Jacobs, R. B. , Single phase transfer of liquefied gases,

NBS Circular 596 (1958).

27. Schoch, K. F. , Superconducting cryogenic motors, Book,

Advances in Cryogenic Engineering <6

, Ed. K. D„

Timmerhaus, pp. 65-72 (Plenum Press, New York,

N. Y. , 1961).

28. Redmond, J. H. , and Bott, F. W, , Design study of a. c. motors to operate in helium from liquid to gaseous

state, J. Spacecraft^, No. 9, 1065-1069 (1968).

29. Johnston, H. L. , Liquid Hydrogen: Propellant for aircraft

and rockets, Final Report, Ohio State Research

Foundation, Columbus 10, Ohio (1946).

30. Johnston, H. L. , Liquid Hydrogen: Propellant for aircraft

and rockets, Second Summary Report, Ohio State

Research Foundation, Columbus 10, Ohio (1947).

31. Gottzmann, C. F. , High pressure liquid-hydrogen and

helium pumps, Book, Advances in Cryogenic

Engineering _5 , Ed. K. D. Timmerhaus, pp. 289-298

(Plenum Press, New York, N. Y. , I960).

74
32. Yonemitsu, H„ , and Okada, J„ , Spot-cooling with a cryogenic

pump, Book: Adv. in Cryogenic Engineering 9 , Ed. K. D. Timmerhaus, pp. 507-512 (Plenum Press,

New York, N. Y. , 1964).

33. Darrel, B. , and Schoch, K. , An electrically pumped liquid

helium transfer system, Book, Adv. in Cryogenic

Engineering 11 , Ed. K„ D. Timmerhaus, pp. 607-611

(Plenum Press, New York, N. Y, , 1966).

34. Kolm, H. H. , Leupold, M. J. , and Hay, R. D. , Heat transfer

by the circulation of supercritical helium, Book,

Advances in Cryogenic Engineering 11, Ed. K. D.

Timmerhaus, 530-535 (Plenum Pre s s, New York,

N. Y. , 1966).

35. Mulder, I. J. , A cold transfer system (Proc. 2nd Int. Cryogenic Conf. Brighton, U. K. , 50, 1968).

36. Wagner , W. , An electrically driven pump and stirring unit

for use at low temperatures and high pressures,

Kaltetechnik-Klimatisierung 22 , No. 10, 342-344,

In German, English translation by A. W. Beffie (1970).

37. Mark, J. W. , and Pierce, W. B. , Hydrogen targets at SLAC,

SLAC-PUB-883, Stanford Linear Accelerator Center,

Stanford University 94305 (1971).

38. Meservey, R. , Superfluid flow in a venturi tube, Phys. of

Fluids 8, No. 7, 1209-1212(1965).

39. Gelder, T. F. , Ruggeri, R. S. , Moore, R. D. , Cavitation

similarity considerations based on measured pressure

and temperature depressions in cavitated regions of

freon-114, NASA TN D-3509 (1966).

75
40. Arp, V. , Ballinger, E. R„ , Giarratano, P. J. , and others,

Helium heat transfer, Unpublished NBS Report

(1972).

41. ASME Power Test Code, Chapter 4, Flow Measurement

(Am. Soc. of Mech. Eng., New York, N. Y. , 1959).

42. Hord, J. , Anderson, L. M, , and Hall, W. J. , Cavitation

in liquid cryogens I_, NASA CR - 2054 (1972).

43. Moore, R. D. , and Meng, P. R. , Comparison of noncavitating

and cavitation performance for 78°, 80. 6°, and 84° helical

inducers operated in hydrogen, NASA TN D-6361 (1971).

44. Pump Manual, prepared by the sub- committee on pumps of

the equipment testing procedures committee. (Am. Inst,

of Chem. Engineers, 345 E. 47th St. , New York, N. Y.

10017, I960).

45. Peck, J. F. , Design of centrifugal pumps with computer aid,

Proc. of Inst, of Mech. Eng. 183 , Part 1, No. 17,

321-351 (1968-1969).

46. Arp, V. , Forced flow, single -phase helium cooling systems,

Book, Advances in Cryogenic Engineering J_7, Ed. K.

Timmerhaus, pp. 342-351 (Plenum Press, New York,

N. Y. , 1972).

47. Donnelly, R. J. , Experimental Superfluidity (University of

Chicago Press, Chicago, 111. , 1967).

48. Wilks, J. , The Properties of Liquid and Solid Helium

(Clarendon Press, Oxford, 1967).

49. Wood, G. M. , Visual cavitation studies of mixed flow pump

impellers, J. of Basic Eng. , 85, 17-28 (1963).

76
Unclassified

Security Classification DOCUMENT CONTROL DATA -R&D

(Security classification of title, body ot abstract and indexing annotation must be entered when the overall report is classified)

1. ORIGINATING ACTIVITY (Corporate author) Cryogenics Division

National Bureau of Standards

2a. REPORT SECURITY CLASSIFICATION unclas sified

2b. GROUP

3 REPORT TITLE LIQUID HELIUM PUMPS

4. DESCRIPTIVE NOTES (Type of report and inclusive dates) Summary report covering 5/72 through 12/72.

5- Au THORiS) (First name, middle initial, last name)

Philip M. McConnell

6- REPOR T DATE June 1973

7a. TOTAL NO. OF PAGES

81

7b. NO. OF RE FS 49

8a. CONTRACT OR GRANT NO. 9a. ORIGINATOR'S REPORT NUMBER(S)

MIPR FY 14557200411 NBSIR 73-316

b. P ROJ EC T NO.

9b. OTHER REPORT NOISI (Any other numbers that may be assigned

this report) AFAPL-TR-73-72

0. DISTRIBUTION STATEMENT

Approved for public release; distribution unlimited.

II. SUPPLEMENTARY NOTES 12. SPONSORING MILITARY ACTIVITY Air Force Aero-Propulsion lab.

Air Force Systems Command

Wright-Patterson Air Force Base, Ohio

13. ABSTRACT

This report summarizes studies of pump characteristics and

performance in supercritical, normally boiling, and superfluid helium,

and also presents results on a survey of commercially available pumps

for helium service. Experimental measurements were made on a centrifugal pump which produced a maximum head of about 15 meters,

and a maximum flow of about 2. 5 x 10 m3 /s. Performance agreed

approximately with classical affinity laws, but cavitation appeared to provide less of a performance limitation than expected. The survey

turned up several pumps which have been used in helium, though

relevant performance data is lacking.

DD FORM1 NOV 65 1473 Unclassified

SecuntyClassification
Unclassified

Security Classification

1 4. KEY WORDS

LINK A LINK B LINK C ROLE W T ROLE W T ROLE W T

Cavitation; helium; pumps; pump performance

superfluid.

Unclas sified

Security Classification
FORM NBS-114A (1-71) U.S. DEPT. of comm. 1. PUBLICATION OR REPORT NO. 2. Gov't Accession

BIBLIOGRAPHIC DATA NBSIR 73-316

N°*

3. Recipient's Accession No.

4. TITLE AND SUBTITLE

LIQUID HELIUM PUMPS

5. Publication Date June 1973

6. Performing Organization Code

7. author(s) Philip M. McConnellt

t Present affiliation: Todd Research and Technical Div.

,

8. Performing Organization

9. PERFORMING ORGANIZATION NAME AND ADDRESS Caiveston, 1 exas

NATIONAL BUREAU OF STANDARDS , Boulder Labs

DEPARTMENT OF COMMERCE

Boulder, Colorado 80302

10. Project/Task/Work Unit No. 2750459

I I. Contract/ Grant No. MIPR FY14557200411

12. Sponsoring Organization Name and Address Aero-Propulsion Laboratory

Wriaht-Patte rson AFB Ohio 45433

13. Type of Report & Period Covered Interim

May - December 1972

14. Sponsoring Agency Code

15. SUPPLEMENTARY NOTES

16. ABSTRACT (A 200-word or less factual summary of most significant information. If document includes a significant

bibliography or literature survey, mention it here.) This report summarizes studies of pump characteristics and performance

in supercritical, normally boiling, and superfluid helium, and also presents

results on a survey of commercially available pumps for helium service.

Experimental measurements were made on a centrifugal pump which produced

a maximum head of about 15 meters and a maximum flow of about Z, 5 x 10"^ m^/s

Performance agreed approximately with classical affinity laws, but cavitation

appeared to provide less of a performance limitation than expected. The survey

turned up several pumps which have been used in helium, though relevant

performance data is lacking.

17. KEY WORDS (Alphabetical order, separated by semicolons)

Cavitation; helium; pump performance; pumps; superfluid.

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(THIS REPORT)

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