NATIONAL BUREAU OF STANDARDS REPORT

3341

CALORIMETER TESTS OF A PROTOTYPE 1/2 HP

FIVE»CYLINDER RADIAL COMPRESSOR

by

Minoru Fujii

Co Wo Phillips

P« Ro Achenbach

Report to

Office of Ihe Quartermaster General

Department of the Army

U. S. DEPARTMENT OF COMMERCE

NATIONAL BUREAU OF STANDARDS
U. S. DEPARTMENT OF COMMERCE

Sinclair Weeks, Secretary NATIONAL BUREAU OF STANDARDS

A. V. Aslin, Director

THE NATIONAL BUREAU OF STANDARDS

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^Office of Basic Instrumentation ^Office of Weights and Measures.
NATIONAL BUREAU OF STANDARDS REPORT

NBS PROJECT NBS REPORT

1003-20-4^32 June 11, 1954 3341

Calorimeter Tests of a Prototype 1/2 HP

Five-Cylinder Radial Compressor

by

Minoru Fujii

Co ¥o Phillips

Po Ro Achenbach

Heating and Air Conditioning Section

Building Technology Division

to

Office of The Quartermaster General

Department of the Army

U. S. DEPMT^ENT OF COMMERCE

NATIONAL BUREAU OF STANDARDS

T». pubiicii..,.. Approved for public release by the

Director of the National Institute of

25,D.C. Suchper

, ^

..II, It, I Standards and Technology (NIST)

on October 9, 2015.

In part, is prohibited

andards, Washington

port has been specifijport for Its own use,

CALORIMETER TESTS OF A PROTOTYPE 1/2 HP

FIVE-CYLINDER RADIAL COMPRESSOR

by

Minoru Fujii, C. W. Phillips, and P. R. Achenbach

Abstract

Calorimeter tests were made of a prototype fivecylinder radial compressor submitted by Longstreth

Enterprises to determine the suitability of this

design to military applications. Special features

of the compressor included a magnesium housing,

five-cylinder radial construction, overhung crankshaft, special gasketing, countersunk bolts with

Allen heads for the housing, and continuous circulation of oil through the compressor without

the use of a sump. The calorimetric measurements

indicated a linear change of capacity with speed

and practically constant volumetric efficiency

in the speed range up to 2300 rpm. The capacity

of the compressor was one-half ton at a suction

pressure of 14 psig, a discharge pressure of 117

psig and a speed of 1750 rpm. The horsepower per

ton of refrigeration was found to be somewhat

higher and the volumetric efficiency somewhat

lower than for slow-speed compressors on similar

duty, but these deviations might not be prohibitive

in view of possible savings in space and weight.

The magnesium housing became perforated with many

small holes during the course of the tests, probably

the result of corrosion of the magnesium by dichlorodifluoromethane. Means were not provided for

mounting the compressor and vibration was excessive

at speeds above 2300 rpm. Although the specimen

might have potential value for military applications

because of its small size and light weight, further

development would be required to make it practical.

I. INTRODUCTION

At the request of the Office of The Quartermaster General,

study was made to determine the performance characteristics of a
prototype refrigeration compressor submitted by Longstreth Enterprises of La Jolla, California. The prototype compressor incorporated a number erf nonv. conventional features including a magnesium

housing, five-.cylinder radial construction, overhung crankshaft,

special gasketing,'^ and*^countersunk bolts with Allen heads for the

housing. The purpose of this investigation was to determine the

suitability of this design to military purposes.

II . Description of Test Specimen

The specimen compressor was cylindrical in shape without fins.

It was made of magnesium and weighed h\^ve pounds. An oil separator

was provided as a separate accessory for the compressor. Fig. 1

shows an exterior view of the assembled unit. Fittings A, B, and

C are the suction gas inlet, discharge gas outlet and oil return

from the oil separator respectively. Hub D is the bearing housing

and E is the shaft.

Fig. 2 shows an interior view of the compressor with the

round cover plate opposite the hub removed from the compressor

housing. This plate was fastened to the compressor housing with

5 Allen head bolts screwed into tapped blind holes. Synthetic

rubber gaskets of the 0-ring type were used at A and B to seal the

casing and to separate the high pressure side and the low pressure

side of the compressor. The five cylinders discharged into a common

annular space just inside the periphery of the housing and the suction

gas entered the crankcase at the center of the compressor. The lower

end of one of the five pistons, which were arranged radially in one

plane, is visible at D. The master connecting rod is shown at G

and the other four link connecting rods were attached to it with

link rod pins as typified at F. The end of the crankshaft is shown

at E. Refrigerant gas was admitted to each cylinder from the crankcase through the port indicated at C and 20 holes around each cylinder

sleeve which were uncovered at the bottom of the piston stroke. Three

of these holes are visible through the port at C. .

Fig. 3 shows an interior view when the cylinder block was taken

out of the housing. A. The ”0”-ring gasket B and suction ports C

on the hub side are shown in this view. The crank pin is shown at

E„ The discharge valves were reed valves of the cantilever type,

one of which is visible at D„ In this figure F and G show the location of one of the link connecting rods and the master connecting

rod, respectively.

Fig. 4 shows the journal bearing for the crankshaft at A,

retainer C, and thrust ball bearing B. One rubber ”0”-ring and

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one graphite sleeve gasket combined with a spring-loaded "O^’-ring

were used for sealing the shaft. These parts are not shown in

Fig. 4.

Forced-feed lubrication was provided utilizing the pressure

differential between the oil separator and the crankcase to return

the oil from the discharge line. The oil was fed to the journal

bearing from the oil return, thence through a small hole in the

shaft to the crank pin and finally to each cylinder through the

five connecting rods.

No means of mounting was provided for the test specimen. For

test purposes the compressor was mounted on a wooden block which

was cut with a semi-cylindrical surface to fit the bottom half of

the compressor. Steel straps were used to secure the compressor

to the wood block.

Drawings and other data submitted by Longstreth Enterprises

gave the principal dimensions and other physical characteristics

which are listed as follows:
~ 4

"

Table 1

Principal Dimensions and Characteristics of Specimen Compressor

Compressor manufacturer

Type

Number of cylinders

Bore*

Stroke*

Displacement*

Method of lubrication

Oil

Type of valves (discharge)

Manufacturer’s model

Dimensions

:

Total length of compressor

Diameter

Capacity

Power

Direction of rotation

Design internal pressure

Upper temperature limit for

discharge gas

Upper and lower temperature limit

for suction gas

Upper and lower pressure limit

for suction gas

Longstreth Enterprises

La Jolla, California

Reciprocating, single stage,

radial

Five

lo25 in»

Oo75 in.

4o601 cUo in, per revolution

Force feed

12 fluid ounces, 300 Viscocity

refrigc oil

Flapper

S^L-201

4 1/2 in,

7 in,

^ 6000 Btu per hour

1/2 Horsepower

Clockwise or counter clockwise

300 lb, per sq, in,

200°F

°F, -20°F

32 psig, 25” vac. Hg

*The measured bore and stroke were 1,375 and

respectively, corresponding to a displacement of

5,6^1 cu, in, per revolution.

III, Testing Method and Procedure

A secondary refrigerant calorimeter was used to determine the

capacities of the compressor. Heat load was imposed on the secondary

refrigerant calorimeter by electric heaters. Suitable controls were

used for maintaining the desired suction pressure, head pressure,

suction temperature and other conditions. Calibrated pressure gages

and watthour meters were used for observing pressures and energy

consumption, respectively. Thermocouples were attached to the refrigerant tubing and compressor at selected locations for observing

temperatures. The copper tubing was covered with hairfelt insulation

for a distance of approximately 6” on each side of the stations of
temperature measurement. Thermocouples enclosed in stainless steel

wells were inserted in the suction line 6" from the compressor

shell and 6” from the expansion valve to measure refrigerant temperatures. A water-cooled condenser was used for condensing the refrigerant .

A shunt-wound, direct-current potor was used to drive the

compressor for most of the tests. Pulleys and belts of different

sizes and a bank of rheostats in the armature circuit of this motor

were used for a coarse adjustment of the motor speed, A governor,

connected to the motor shaft, actuated a slide-type rheostat in

the shunt field circuit of the motor to obtain more precise control

of the compressor speed. Although the current flow through the

armature circuit and voltage at the motor terminals were measured

the power consumption of the compressor was not determined since

the, efficiency of the D.C. motor at various speeds was not measured.

Absolute power consumption at a speed of about 1700 rpm was determined by means of a 1/2 horsepower A.C. motor as described later.

Three kinds of tests were made with the specimen compressor.

For one series of tests, the compressor speed, discharge pressure

and suction temperature were maintained constant, while the suction

pressure was varied. For the second series of tests, suction pressure

head pressure and suction temperature were maintained constant while

the compressor speed was varied, A special test was made with a 1/2

H.P. A.C, motor directly connected to the specimen compressor. The

motor had previously been calibrated by means of a dynamometer.

From this calibration the brake horsepower of the compressor

was determined for a limited range of suction pressure. Special

tests were made at a compression ratio of 9«29 using the D.C. motor

drive and at a compression ratio of 11,16 using the 1/2 hp A.C, motor.

The compressor capacity for each test was computed by means

of the following formula, taken from ASRE Standard No. 23-R, entitled

"Methods of Rating and Testing Refrigerant Compressors":

Q = I’fl (Qh + Qe) (1)

"g2 -

where h^^^ = heat content of refrigerant liquid at a temperature

corresponding to the pressure of refrigerant vapor

leaving compressor, Btu per lb.
hf2 = heat content of refrigerant liquid entering the

expansion valve, Btu per Ibo

h_]^ = heat content of refrigerant vapor entering compressor,

Btu per Ibo

h „2 = heat content of refrigerant vapor leaving calorimeter,

Btu per Ibo

Q = compressor capacity, Btu perhro

= calorimeter heat input, Btu per hr.

= heat leakage into calorimeter, Btu per hro

A secondary method was employed to substantiate the compressor

capacities computed by the above formula from the calorimeter data.

This method consisted of measuring the volume flow of refrigerant

through the calorimeter circuit per unit time. This was accomplished

by using two receiving tanks in the refrigerant circuit and alternately filling one tank with liquid from the condenser, while emptying

the other tank into the evaporator. The two tanks were calibrated

volumetrically, and the time required to fill a certain fraction of

the volume of one tank was observed. From this elapsed time the

volume-flow per hour was computed. The following formula, taken

from ASRE Standard No. 23-R, was then used to compute the compressor

capacity for each test:

Q = VD (1 » 0.01^ wd) (h 1

- OokU w (

£»2 “ ) where

d = density of liquid refrigerant, for temperature

at which flow is measured, lb per cu ft

hfi = heat content of refrigerant liquid at a

saturation temperature corresponding to the

pressure of refrigerant vapor leaving compressor,

Btu per lb.

hgl = heat content of refrigerant vapor entering

compressor, Btu per lb.

Q = compressor capacity, Btu per hr.
- 7

-

tfi = saturation temperature corresponding to pressure

of refrigerant vapor leaving compressor, deg F

V

w

temperature of refrigerant vapor entering compressor

deg F

flow of mixture of liquid refrigerant and oil, cu ft

per hr.

weight of oil per unit weight of mixture of liquid

refrigerant and oil, lb per lb.

Tables in ASRE Circular NOo 12 on the thermodynamic properties

of dichlorodifluororaethaae (F-12) were used to determine the heat

contents called for in the above formulae A value of 0,02 for w

in the formula was used, based on a single determination of the oil

content of the refrigerant.

In preparation for each test the compressor was run for a period

of more than two hours before the test period was started, in order

to allow the calorimeter system to come to steady state conditions.

Observations were made every twenty minutes for at least two hours

under steady conditions of temperature, pressure, heat input and

speed.

Tests were made with only one direction of rotation of the compressor which was counter clockwise as viewed from the hub side of

the compressor,

IV, Test Results

The results obtained during the calorimeter tests of the compressor specimen are summarized in Tables 2 and 3 and various relationships are shown graphically in Fig. 5 to 10, inclusive.

Fig. 5 shows the relationship between compressor capacity and

speed for the range from ^00 to 2300 rpm. During this series of

tests the suction pressure was maintained at psig and the discharge pressure at 117 psig corresponding to a compression ratio of

6,^6. The suction gas temperature was maintained at 65°F, The

volumetric efficiency was constant for all practical purposes through

out this range of speed and the straight line drawn through the observed values, if extrapolated, shows that the capacity becomes zero

almost simultaneously with the speed. This indicates that the

effectiveness of the valve action was equally good at all speeds

up to 2300 rpm.
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Fig 6 shows that the relationship between compressor capacity

and suction pressure was linear for a constant discharge pressure*

This figure reveals that the capacity of 6000 Btu/hr claimed by

the manufacturer was attained at a suction pressure of 14 » 3 psig

corresponding to an evaporator temperature of about 10°F, If the

capacity curve in Fig 6 were extrapolated as a straight line the

indicated capacity would be zero at a suction pressure of 10 in. Hg

vacuum.

Fig 7 shows the power consumption of the Longstreth compressor

for a limited range of suction pressure as determined by using a

calibrated 1/2 hp A.C. motor as a direct drive. This curve shows

that a 1/2 hp motor would only be adequate for driving this compressor at suction pressures of zero gage or lower. Using the data

from Fig 7 and a minor extrapolation of the curve in Fig 6 it can

be shown that the brake horsepower per ton of refrigeration for the

specimen compressor was about 3.75 for a suction pressure of 2.5

psig and a discharge pressure of 117 psig. By comparison, the

brake horsepower per ton observed (;>.n a larger slow-speed compressor

was 2.S for the same compression ratio.

Fig S shows the temperatures of the bearing housing and the

compressor shell for a range of compressor speed. These temperatures

increased in a linear relationship with compressor speed at constant

ambient temperature. The scattering of the plotted values at a

speed of 1750 rpm resulted from the variation in suction pressure

selected for different tests at this speed.

The relation between volumetric efficiency and compression ratio

at constant speed is shown in Fig 9o The volumetric efficiency

ranged from about for a compression ratio of 7 to about 5^%

for a compression ratio of 2. 5. The volumetric efficiency of the

Longstreth compressor at a compression ratio of 5 was comparable

to that of another prototype compressor tested at the same speed

during an earlier investigation, but it is as much as 20% lower than

for slow-speed compressors designed for similar capacity.

Fig 10 shows the volumetric efficiency plotted against compressor

speed for a fixed compression ratio. Fig 10 corroborates the conclusion drawn from Fig 5 that the volumetric efficiency was practically

constant for the range of speed from ^00 rpm to 2300 rpm.

According to information received from the manufacturer, the

compressor was designed to operate at speeds in the range from 1000

rpm to 4000 rpm. At speeds higher than 2300 rpm, however, vibration

of the specimen was so severe that the refrigerant lines broke repeatedly and fittings developed leaks. A better means of mounting
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the compressor was being constructed when a leak test revealed

refrigerant loss through a large nujnber of pores in the housing

of the compressor. The pores were found in the cylindrical surface

of the magnesium compressor housing and could barely be observed

by a binocular which had a magnification of one hundred. The

leakage through each pore was minute, but the total amount of leakage

was considerable. It was not determined whether the pores were

present in the magnesium casting originally or whether they were

caused by corrosion,

V, Discussion and Conclusions

The prototype compressor appeared to have potential value to

the Office of The Quartermaster General because of its relatively

light weight and small size. There were, however, some disadvantages.

The power consumption per unit capacity was relatively high. This

characteristic may have been due in part to the absence of suction

valves. The suction ports were not uncovered to admit refrigerant

to the cylinders until the piston reached the bottom of the cylinder

thus increasing the amount of work done during each suction stroke.

The vibration of the prototype was excessive. The absence of any

means of mounting the compressor may have been responsible in part

for this excessive vibration, but it is believed that better balancing

of the compressor is also needed.

The Data Book of the American Society of Refrigerating Engineers

recognizes that Freon-12 is corrosive to magnesium. It is probable,

therefore, that corrosion caused the porosity of the magnesium compressor housing that was observed at the end of the tests.

The compressor housing and shaft bearing operated at relatively

high temperatures even when air was circulated over the compressor

by an auxiliary fan. The compressor was not provided with fins for

extended cooling surface. The wooden supports used for mounting the

prototype during the tests probably reduced the heat transfer from

the lower half of the housing somewhat.

An oil separator is an essential accessory to this compressor

because lubrication of all moving parts depends on a steady flow of

oil from the compressor, through the separator, and back to the oil

inlet connection on the bearing housing.

The design of the compressor was considered well-suited to

mass production because most of the parts could be made with automatic machine tools.

THE NATIONAL BUREAU OF STANDARDS

Functions and Activities

The functions of the National Bureau of Standards are set forth in the Act of Congress, March

3, 1901, as amended by Congress in Public Law 619, 1950. These include the development and

maintenance of the national standards of measurement and the provision of means and methods

for making measurements consistent with these standards; the determination of physical constants and properties of materials; the development of methods and instruments for testing materials,

devices, and structures; advisory services to Government Agencies on scientific and technical

problems; invention and development of devices to serve special needs of the Government; and the

development of standard practices, codes, and specifications. The work includes basic and applied

research, development, engineering, instrumentation, testing, evaluation, calibration services, and

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Information on tbe Bureau’s publications can be found in NBS Circular 460, Publications of

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publications should be addressed to the Office of Scientific Publications, National Bureau of

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