NAT'L INST. OF STAND & TECH

AlllDb 2bD2fl4

NBSIR 83-2756

A Laboratory Investigation of

Refrigerant Migration in A Split

Unit Air Conditioner

U.S. DEPARTMENT OF COMMERCE

National Bureau of Standards

Center for Building Technology

Washington, DC 20234

April 1983

Issued August 1 983

uCiou

, 1)56

65-2756

1963

Prepared for: U.S. Department of Energy

ffice of Conservation and

Renewable Energy

ashington, DC 20585

national cyrxAU

Liurjir./ CEP 1 5 1983

NBSIR 83-2756 f

A LABORATORY INVESTIGATION OF

REFRIGERANT MIGRATION IN A SPLIT

UNIT AIR CONDITIONER

William J. Mulroy

David A. Didion

U S. DEPARTMENT OF COMMERCE

National Bureau of Standards

Center for Building Technology

Washington, DC 20234

April 1983

Issued August 1983

Prepared for: U.S. Department of Energy

Office of Conservation and

Renewable Energy

Washington, DC 20585

U.S. DEPARTMENT OF COMMERCE, Malcolm Baldrige, Secretary

NATIONAL BUREAU OF STANDARDS. Ernest Ambler. Director

ABSTRACT

The relationship between cyclic refrigerant migration and cyclic loss for a residential, split-system air conditioner has been investigated. The cyclic

refrigerant migration was measured at different points in the operating cycle

by simultaneously shutting five pneumatic valves which isolated the refrigerant

in the major system components. The refrigerant was then removed, weighed,

and returned to the system. The unit tested was found to have a high initial

capacity as migrated refrigerant was removed from the evaporator and then a low, slowly increasing capacity as trapped refrigerant was returned to the

system from the accumulator. The unit performance was also compared to single

and double time constant regressive approximations and to the time constant

calculated from the evaporator mass and heat transfer coefficient. Although

relationships between migrated refrigerant and cyclic capacity were observed, no practical refrigerant migration test method that would be less burdensome

than the cyclic tests of ASHRAE Std. 116 appears possible at this time.

Key Words: Central air conditioners; cyclic testing; heat pumps; refrigerant

migration

iii

TABLE OF CONTENTS

List of Tables vii

List of Figures ix

Metric Conversion Factors xi

1. INTRODUCTION 1

2. TEST SPECIMEN AND FACILITY 3

3. TEST PROCEDURE 9

4. DETERMINATION OF COOLING CAPACITY, PART LOAD FACTOR,

AND REFRIGERANT MASS IN ISOLATED SECTIONS 14

5. DISCUSSION OF RESULTS 16

6. CONCLUSIONS 33

7. REFERENCES 35

APPENDIX A: Comparison of Cyclic Performance to Reduced

Charge Steady-State Performance 36

APPENDIX B: Refrigerant Migration Measurement Techniques 40

v

LIST OF TABLES

TABLE 1. COMPONENT VOLUME MEASUREMENTS 5

TABLE 2. COMPONENT MASS MEASUREMENTS 5

TABLE 3. SUMMARIZED STEADY-STATE TEST RESULTS 17

TABLE 4. SUMMARIZED CYCLIC TEST RESULTS 18

TABLE 5. REFRIGERANT CONTENT OF ISOLATED SECTIONS 19

TABLE 6. POUNDS OF LIQUID AND VAPOR IN ISOLATED SECTIONS

(CALCULATED FROM LIQUID AND VAPOR DENSITIES,

ISOLATED SECTION VOLUMES AND DATA OF TABLE 3) 20

vii

LIST OF FIGURES

Figure 1. Schematic of NBS Heat Pump Test Apparatus 4

Figure 2. Refrigerant Migration During an "On" Cycle 21

Figure 3. Comparison of Time for Capacity to Reach Steady-State

to Time Required to Empty Accumulator 24

Figure 4. Comparison of Transient Air-Side Capacity to

Coincident Transient Refrigerant Side Capacity 25

Figure 5. Comparison of Refrigerant Side Capacity to

Refrigerant Mass in Evaporator 27

Figure 6. Regressive Fit to Experimental Transient Normalized

Capacity Data 29

Figure 7. Capacity Decay after Compressor Shut-off with

Continuous Fan Operation 31

Figure Al. Unit Capacity as a Function of Charge 37

Figure A2. Comparison of Refrigerant Side Capacity to Capacity

Calculated from Accumulator Content and Reduced

Charge Steady-State Data 38

Figure A3. Comparison of Air Side Capacity Calculated from

Accumulator Liquid Charge and Reduced Charge

Steady-State Data to Experimental Air Side Data 39

METRIC CONVERSION FACTORS

Length 1 inch (in) = 25.4 millimeters (mm)

1 foot (ft) = 0.3048 meters (m)

Area 1 ft

2 = 0.092903 m2

Volume 1 ft

3 = 0.028317 m

3

Temperature F = 9/5 C + 32

Temperature

Interval 1°F = 9/5°C

Mass 1 pound (lb) = 0.453592 kilograms (kg)

Mass Per Unit Volume 1 lb/ft 3 = 16.0185 kg/m3

Energy 1 Btu = 1.05506 kilojoules (kj)

Specific Heat 1 Btu/ [ (lb) (°F) ] = 4.1868 kj/I(kg) (°K)

]

Gallon 1 gallon = 0.0037854 m

3

xi

1 . INTRODUCTION

Because of the concerns about test cost and testing time requirements of the current evaluation procedure for central air conditioners, the Department of Energy

(DoE) requested the National Bureau of Standards (NBS) to investigate several

alternative approaches for determining cyclic degradation that might result in

a simpler, cheaper test procedure without sacrifice of repeatability or of

comparability to results obtained currently.

Laboratory and field tests have shown that units which do not allow off-cycle

refrigerant migration (i.e., units with thermostatic expansion valves without

bleed ports) have less performance degradation when cycling than do units

which allow off-cycle pressure equalization and consequent refrigerant migration. As a result the existing cyclic test procedures were conceived to allow

for credit to be given for those designs that might minimize cyclic losses.

Modeling efforts which might be used to mitigate some of the testing requirements were thwarted since the amount of migration varies from one design to

another and is also function of operating conditions. Therefore, it was not

possible to characterize an air conditioner by a simple time constant as was

done for the furnace /boiler evaluation procedure. Theoretical approaches

to cyclic performance analysis have been presented which use refrigerant migration during the off-cycle as an input [1] [2], However, no adequate

empirical data has existed, in the public domain, to quantify the refrigerant

flow during the dynamic periods of a typical air conditioner operation.

This report presents results of dry coil steady-state and comparable cyclic tests

performed at the DoE/NBS test procedure [3] rating point (27.8°C or 82°F outdoor
temperature) using the recently introduced continuous indoor fan option

for the cyclic tests. The refrigerant mass in five major areas (compressor

and accumulator, liquid line, vapor line, condenser, evaporator) was determined by shutting quick closing pneumatic ball valves isolating these areas and

then extracting and weighing the refrigerant contained in each. Refrigerant

mass distribution data were taken for steady-state operation and at the

beginning, end (6-minutes) and at several intermediate times during an "on"

cycle performed as an extension of 6 minute "on" and 24 minute "off" cyclic

tests performed in accordance with the DoE/NBS test procedure [3]

No adequately confirmed theoretical model for calculating the cyclic

degradation factor from this experimental data or recommended alternate test

procedure is presented at this time. This empirical study indicates that

such a procedure may be possible; although the representative function will

be more than a simple time constant and the method for obtaining the test

data may not be less burdensome than the current test methods.

2
2. TEST SPECIMEN AND FACILITY

A nominal three ton, split system, air-to-air unitary heat pump operated in the

cooling mode was used for this investigation. The unit, being a heat pump,

differed from a typical split system air conditioner in its coil sizing and in

possessing a four way valve and an accumulator. These design changes generally

result in a heat pump having a lower steady-state efficiency and a greater

cyclic degradation coefficient than that, of a central air conditioner of similar cooling capacity. It was believed that this expected increase in the variable

to be studied, the cyclic degradation factor, would increase the measurement

accuracy. The unit was equipped with a crankcase heater which was automatically

energized during all "off" cycles. The outdoor unit contained the compressor,

outdoor coil, fan defrost controller and associated equipment. The indoor and

outdoor sections of the heat pump were installed in separate but adjoining

environmental chambers

.

System measurements not generally considered in a performance test, but of

particular interest in this study were the volumes and masses of various system components some of which are summarized in Tables 1 and 2. The volume

measurements in Table 1 listed as experimental were determined by vapor

charging the system component volumes isolated by pneumatically operated ball

valves (Fig. 1) with refrigerant 22. The system volumes were then calculated

from the specific volumes of the thermally stabilized refrigerant in each section (taken from the R22 superheat tables at the measured temperatures and

pressures) and the measured refrigerant mass introduced by taking before and

after charging cylinder weights (correcting for charging hose volume) . As a

check, comparable volumes listed as dimensional in Table 1 were calculated from

3
PNEUMATICALLY

OPERATED

BALL

CO3

4-1 CO

y<

coO.a

co

<u

)-4 3

bO •H

(j-i

4

•

Temperature

measurement

point

a

Pressure

measurement

point
tubing lengths, outside diameters and presumed wall thicknesses for all components except the compressor—accumulator section which was too complex to

allow a reasonably accurate volume estimate from the available dimensional data.

Table 1

Component Volume Measurements

Outdoor Coil

Dimensional Experimental

4310 mL (263 in

3

) 5020 mL (306 in

3

)

Compressor and Accumulator — 10,000 mL (610 in 3

)

Vapor Line 1000 mL (61 in

3

) 870 mL (53 in 3

)

Liquid Line 1030 mL (63 in3

) 1280 mL (78 in 3

)

Indoor Coil 5150 mL (314 in

3

) 5760 mL (346 in

3

)

The component mass measurements given in Table 2 were calculated using the

dimensional data used in preparing Table 1 with the addition of measurements

of the coil fin dimensions.

Table 2

Component Mass Measurements

Outdoor Coil 12.4 kg (27.4 pounds) of copper

11.4 kg (25.1 pounds) of aluminum

Indoor Coil 16.7 kg (36.7 pounds) of copper

8.9 kg (19.6 pounds) of aluminum

Vapor Line 3.0 kg (6.5 pounds) of copper

The unit was tested in the environmental chambers and apparatus described in [5].

5
The principal method used to measure the heat pump heating and sensible

cooling capacities was the indoor air enthalpy method. An air flowrate

measuring apparatus was connected to an insulated duct attached to the discharge side of the indoor section of the heat pump. This apparatus consisted

of a receiving chamber and discharge chamber separated by a partition containing a nozzle. An exhaust fan was attached to the duct leaving the discharge

chamber so that the static pressure of the air leaving the indoor section of

the heat pump could be adjusted to give air flow rates consistent with the

manufacturer's performance data.

Manometers accurate to within 1% of the reading were used to measure static

pressure across the nozzle and pressures obtained by a pitot tube placed at

the nozzle exit. A thermocouple was installed to determine the air temperature at the nozzle exit,

A thirty-j unction thermopile was constructed to measure the air temperature

difference entering and leaving the indoor section of the heat pump. One side

of the thermopile was placed in the entering duct while the other was placed

in the insulated discharge duct leading to the receiving chamber of the air

flow measuring apparatus. The junctions of the thermopile were spaced at

centers of equal area across the inlet and discharge ducts. The millivolt

signal generated by the thermopile was integrated by an electronic integrator

in order to obtain time-temperature difference data of the air passing over

the indoor coil. A duplicate thermopile was read as point data for integration by use of Simpon's rule and for preparation of the graphical presentations of capacity as a function of t^ime in this report.

6
The relative humidity of the air entering the indoor unit was measured by a

motorized psychrometer . A motorized psychrometer was also used to measure the

relative humidity of the air leaving the indoor unit.

The volatile refrigerant flow method [6] was used as the secondary test to

determine the full-load, steady-state cooling capacities. The liquid refriggerant line at the indoor unit was modified to incorporate a sightglass, turbine meter and four valves. The valves served to properly direct the flow of

refrigerant through the turbine meter when the heat pump operated in either

the heating or cooling mode. All tests discussed in this report were performed

in the cooling mode.

Adequate subcooling at the turbine meter was necessary for accurate

determination of the refrigerant flowrate. For this reason, it was not

possible to use this secondary method for part load tests. The refrigerant

flow was determined by the turbine meter and electronic totalizer. The density of the refrigerant was determined from the temperature and pressure readings.

Watthour meters were installed to measure the electrical energy consumption of

the compressor and outdoor fan, and the indoor fan. Electronic totalizers and

printers were connected to the watthour meters to provide a record of the

energy consumption.

The dry-bulb and dew-point temperatures in both the indoor and outdoor environmental chambers were continuously recorded. The average temperatures of the

air entering the outdoor unit was measured using four thermocouples connected

in parallel. In addition to the preceding instrumentation, thermocouples and

pressure gages were installed at various locations and temperatures and

7
pressures were recorded to provide additional information on the performance

of the unit and as a check for consistent operation.

Electric resistance heaters which cycled "on" or "off" as the unit cycled were

installed in the test chamber to stabilize the load on the test chamber

environmental control systems.

Pneumatically-operated ball valves operated by a solenoid pilot were installed

in the locations shown schematically in figure 1 so that portions of the system could be isolated and their refrigerant content determined by transferring

the refrigerant to a canister and determining the change in refrigerant container weight

8
3 . TEST PROCEDURE

Prior to each day's testing, the unit was vapor charged with the same amount of

refrigerant that was removed after the previous day's test.

In later tests it was found to be simplest and most accurate to charge the

unit to a repeatable steady-state performance after test room conditions had

been obtained. In this case the refrigerant flowrate was the observed quantity. When this method was used, the refrigerant cylinder was weighed before

and after charging to compare the amount of refrigerant removed at the end of

the test.

Vapor charging, because of the reduced charging line tare, was found to be

considerably more accurate than liquid charging. However, the best measure of

the weight of refrigerant added was that observed when the cylinder was

detached from the charging manifold. The best step-by-step procedure for

charging by weight was as follows:

1) The desired amount of refrigerant to be added was determined.

2) The refrigerant cylinder was weighed.

3) The upright refrigerant cylinder was placed in a bucket of slightlybelow-room-temperature water to maintain a constant pressure as

refrigerant was evaporated into the system. The cylinder and the

water filled bucket were mounted on a platform scale.

9
4) Using a refrigerant charging manifold, the cylinder was connected to

the low pressure side of the unit and to a vacuum pump.

5) After the manifold was evacuated, the vacuum pump was isolated and

disconnected

6) The cylinder valve was opened and the manifold and tubing were pressurized to a value near the expected steady-state suction pressure.

7) The manifold to cylinder hose was adequately supported to prevent

indicated weight changes as a result of tank or hose motion.

8) The platform scale was balanced.

9) The scale arm was set to desired final weight.

10) The unit valve was opened until tank pressure was reached.

11) The heat pump was started and the valves were closed when scale arm

fell.

12) The refrigerant cylinder was detached from the unit, removed from

the water bath, the outside dried, and weighed.

10
The most accurate measurement of the amount of refrigerant added is

the difference between the cylinder weights taken in steps 2 and 12. A

consistent bias error of approximately 23 g (0.05 pounds) between the refrigerant

added and the amount removed was observed during preliminary tests and was used in the calculation of the charge to be added when this method was used.

It was found that the unit performance for equal calculated total charges

would be different if charged from a vacuum as opposed to charging from

a residual vapor charge. In this case the unit charged from an evacuated

state would be undercharged because of refrigerant absorption by the crankcase oil. For all tests in this report the unit was charged from a residual

vapor charge that would result in a quantity of refrigerant absorbed in the

oil near the steady-state value. The steady-state suction pressure was approximately 550 kPa (65 psig) . The refrigerant was extracted with a cylinder

packed in water ice (0°C [32°F] = 465 kPa [52.8 psig] saturation pressure).

When steady-state operation of the unit and test chamber reached the 27.8°C

(82 F) dry coil test condition of the DoE/NBS test procedure, data were taken

for a minimum period of 1/2 hour. For the three tests reported as QC 3,

QC 5, and QC 6 in Table 3, the unit was shut off and the pneumatically operated

valves simultaneously closed at the end of the steady-state data period. For the

11
other reported tests, manually-operated cyclic operation was begun at the

end of the steady-state data period. These cyclic tests were performed following the guidelines of the DoE/NBS test procedure using the 6 minute "on" cycle

length and 24 minute "off" cycle length. All tests were performed with continuous indoor fan operation (i.e., during both the "on" and "off" cycles).

Three complete cycles were performed to allow conditions to become repeatable

in the cyclic mode. A fourth cycle was then performed during which DoE/NBS

test procedure test data were taken. Following this "data" cycle an additional

24 minute "off" period followed by a 6 minute "on" period was performed. The

pneumatically operated valves were closed at various times between cut-on and 6

minute of operation during this "on" period. The compressor was wired to stop

simultaneously with pneumatically operated valve closure.

In all cases, closure of the pneumatic valves was immediately followed by

extraction and weighing of the refrigerant in each isolated section according

to the following procedure:

1) The cylinder to be used was weighed prior to refrigerant removal.

2) The upright refrigerant cylinder was placed in an ice bath.

3) One hose of a charging manifold was connected to the cylinder

,

another to the unit volume from which the refrigerant was removed,

and the third to a vacuum pump.

12
4) The manifold was evacuated and the isolation valve was closed. Then

the vacuum pump was shut off and disconnected.

5) The valves on tank and unit were opened.

6) The pressure was observed, when pressure reached approximately 60 psig

(510 kPa) the tank and unit valves were shut.

7) The tank was disconnected from the manifold hose, removed from the

ice bath, excess water was wiped off and the tank was weighed.

8) Steps 1 through 7 were repeated for each isolated area.

9) Approximately 1 hour after removal the pressures and temperatures of

all isolated areas were checked to verify that they contained superheated vapor.

10)

After temperatures were stabilized (preferably the next morning) the

temperature and pressure of each isolated area was recorded so that

the mass of vapor remaining in each section could be calculated.

13
4. DETERMINATION OF COOLING CAPACITY, PART LOAD FACTOR, AND REFRIGERANT MASS

IN ISOLATED SECTIONS

As previously noted, the test facility was designed for determining the

sensible cooling capacity by the indoor air enthalpy method. As a check

on the measurement of the cooling capacity during steady-state, full-load operation, the volatile refrigerant flow method was used. The equations used for

determining the cooling capacity are given in [6]. Since all reported tests

were performed with a dry indoor coil (no latent capacity) the total cooling

capacity was equal to the sensible cooling capacity.

The equations used for calculating the part load factor, PLF, and cyclic

degradation factor C^, are given in [3]. The test procedure for this report

differed from [3] in having continuous operation of the indoor fan (i.e.

,

during both "on" and "off" cycles) as proposed in the ASHRAE Standard 116P.

The cooling done during cyclic operation was defined as that done during the

6 minute "on" cycle. No credit was given for the cooling that occurred during

the first few minutes of the "off" cycle nor debit taken for the "off" cycle

fan or crankcase heat. The unit "on" cycle power including "on" cycle fan power

was used as the total cyclic power in calculating the coefficient of performance and cyclic performance degradation coefficient, C^. The cyclic and comparable steady-state cooling capacities were calculated from both data taken by an

electronic integrator reading and a 30 junction thermopile across the evaporator

and by numerical integration of point data taken from a duplicate thermopile.

The refrigerant content of each isolated section, for the quick closing valve

tests, was calculated as the weight removed plus the mass of the vapor

remaining calculated from the section volume and the density of superheated

14
vapor at the final pressure and temperature recorded after sufficient time

had been allowed for these quantities to stabilize. The quality (mass fraction

of vapor) in each isolated section was calculated from the above determined

total refrigerant mass, the volume of each section, and from refrigerant vapor

and liquid densities calculated from the pressures and temperatures occurring

at the time of valve closure. In this case the vapor was assumed to be superheated at the average temperature of each section and the liquid was assumed to

be saturated at the lowest temperature measured by any of the component

thermocouples on that section, or if lower, the section saturation temperature

corresponding to the measured pressure.

Only one test, QC 11, was performed with the unit indoor fan cycling with

the compressor. This test was terminated at the end of a cut-off period with

pneumatic isolation valve closure at that time. As can be seen in Tables

4 and 5 the off-cycle refrigerant migration for this cyclic fan test was not

significantly different from the continuous fan test data.

15
5. DISCUSSION AND RESULTS

The test results for the steady-state and cyclic periods preceeding pneumaticallyoperated valve closure are given in Tables 3 and 4.

The refrigerant migration data taken with the quick closing pneumaticallyoperated valve system are listed in Tables 5 and 6 and are plotted as a function of time in figure 2. Some comments which can be made about this data are:

1) Most of the refrigerant was on the high pressure side during steady-state

operation. Approximately 46% of the total charge was in the condenser and 37%

in the liquid line for a total of 83%, the remaining 17% being on the low

pressure side.

2) Most of the refrigerant was on the low pressure side at the end of the

off-cycle. Approximately 56% of the total charge was in the evaporator, 2% in

the vapor line, and 22% in the compressor and accumulator for a total of 80%,

the remaining 20% being on the high pressure side.

Refrigerant migration to the evaporator during the off-cycle was initially the

result of the pressure gradient between the condenser and evaporator. After

these pressures equalized, which was typically in about 5 minutes for this unit,

migration continued as a result of the temperature gradient between the condenser and evaporator. The refrigerant migration to the accumulator during the

off -cycle was presumably by gravity flow from the evaporator since the accumulator in the compressor compartment, was considerably warmer than the evaporator

during the off-cycle. At the end of a typical off^cycle, temperatures of 34.5°C

(94.1°F) and 28.1°C (82.6°F)were recorded from insulated thermocouples soldered

to the top and bottom of the accumulator respectively. This bottom temperature

16
TABLE

3.

SUMMARIZED

STEADY-STATE

TEST

RESULTS

HS W CO 3Eh w

CJ

gX Sg CTv

=Ss ov VO LO o m r**

=x 2>»= rm CM cn =T*= cn o m ^ 2 St vD vO i"- vO m sf m m vO vO vO H £32 00 00 00 00 00 00 co 00 00 oo 00 00 • 00 00 w

PiwH £ mhj: 1 CJ 3 o o o o o o O o o o O

O o o CO CTi m m rH CM as X MT CM O CM o 00

Q <3 4-»

m ST MT Os r>» CM 1—1 OV CT> cn <r LO SS" O' o

cn o O o o o o Os Ch Oi Oi cr< CT\ OV

CNI CO CO CO CO CO CO CM CM CM CM CM CM CM

HXM

CO

OH

G

O

OV o o o o o o O o O

O

O

O o rH o o r-. co m CM 00 O <r 00 CM m l. <3 x LO M3 CM <r CO 00 r-' St vO X m VO 00 Pi P-j PQ 00 OS os os C\ Os 00 00 00 00 00 00 00 cn

l-j <J

<3u

CM CM CM CM CM CM CM CM CM CM CM CM CM CM

a s 1 % Ph fx Ph Cn Ph PH

X Ph

XX Ph Ph Ph —I o o o o o o O O o O O o o o O H 2 o <3 «

sr 00 o rH MT vO

m

O 00 CM

m

m vO 2 § o rH CM rH CM rH rH O O £co £^

rH rH rH rH rH rH rH rH rH rH rH rH rH rH

•K/—S

gj

H 00 00 00 00 oo 00 00 00 00 00 oo 00 00 00 P3 P5 X X •H •H •H •H •H •H X •H •H

Q

O <3 CO CO CO CO CO CO CD CD CO CO CO CO CO CO <3 cp O W W pq (i O* a CX ex a X CX X X X X X X

ts

w S

Pi o

CM <r CO vO vO

m

mm

m

m <r

m CM CM

P-4 |H o o o o O o o O o O o a o o

P-4 CM CM CM CM CM CM CM CM CM CM CM CM CM CM O

EhS3 o o W 5 ^

Ph PH Ph Ph Ph pH Ph

X

X

X

X Ph Ph

X

Sag “ s & 0 o o e 0 e o « 0 e 0 0 o 0 St• <• o•

(H.• CO Os• CM• •<r• m• in

O• vO• CM

m

rH oc Os 00 00 Os

O Ov ov o av

O cn

CM rH rH rH rH rH rH CM rH rH CM rH CM rH

rvi P^ gin 00 00 00 00 00 00 00 oo 00 00 00 00 00 oo •H •H •H •H •H •H •H •H •H •H •H •H •H •H |gS COa

COO,COPu

CO(X

COaCO(X

COcxCO(XCOa

U3X

COX

COX

COX

COX

S3g w

in

m o o o St st

O rH

O o O

O

m j < • • • • • • • • • • • • cx M 5 St m m VO vO m m m in in m un LO m w

vO so vO VO vO VO vO vO vO vO vO VO vO vO o • Ph Cn Ph IX Ph Ph X Ph Ph Ph Ph PH X Ph o o o o o o O o o o 0 O O o § § 1—1

ON vO 00 CM o Os r^. 00 00 r'. in CM CM s s • • • • • • • • • e • # gM p 00o

r*- as OS OS

r-~ o o Os OV cr> OV OV o o 00 00 r-. r- 00 00

Pio • Ph Ph Ph Ph IX Ph

X Ph Ph Ph Ph

X Ph Ph

O P-i o o o o O o O o o o o 0 o o o s

LO CM o <3- CM o rH <r rH o co VO 00

H

W * * • * • • • • • • • • • £d Eh CM CM CM CM CM CM CM CM CM CM rH CM o 00 00 00 00 00 00 00 00 00 00 00 00 00 00

•K

•K co Eh o rH vO 00 CTV o m vO r-. oo as rH rH CM CM

CPwH =*>= =fe =>$= ;>>G ***= =>>s cj C_) o CJ o u a o o u o CJ CJ CJ O' O' O' O' O' O' O' O' O' O' O' O' O' O'

(Uc CO X cO3 X G X G 0 X

cr X

X

X

Oo

a)

X

x: c X X C 0) X pGX X

a) •> X (0 c X p co o a) a X

CO

X

co 0) pGX

00 o CX X X X X cO X X X o

0) pV| X

CO X • o X o X COCO

CO CJO'

CO X

X

X oX XG X

nJ o

CO

u CO

<u • CD X X X

<D

X CX i c 3 P § o a X u x o CD o pG X x X G G

co

X pG

X

X

ci) a)X

00,G 3 X 4J X CD X 00 CO

0)

G cO

X

X pG X a.

a) a) -G x

G

x c X

0)

X

X o CO CD X X X X a • 0) Pn

CDOO

X o x CO a, CO G B X CO O o o 3 3 a x G X G X X D. CO X X x G

cO ^ o o x x cj pO 0) o 3 x O X O co X CD X X G X X X CO

0)

X

Cl)

g CD CD & i pG a H x H O

•JC **

17
TABLE 4. SUMMARIZED CYCLIC TEST RESULTS

TEST STEADY-STATE

EER

CYCLIC

EER

CYCLIC

CLF

CYCLIC

PLF

CYCLIC

CD

QC #7 7.46 4.98 .125 .668 .38

QC #8 7.46 5.06 .126 .678 .37

QC #9 7.42 5.09 .130 .686 .36

QC #10 7.34 5.00 .127 .681 .37

QC #11* 7.28 5.37 .136 .737 .30

QC #16 7.36 5.12 .129 .696 .35

QC #17 7.18 5.09 .132 .708 .34

QC #18 7.29 5.23 .133 .717 .33

QC #19 7.30 5.25 .135 .720 .32

QC #20 7.39 5.33 .135 .721 .32

QC #21 7.48 5.18 .130 .692 .35

*The indoor fan was cycled in phase with the compressor

for Test QC #11. For all other tests the indoor fan

operated continuously.

NOTE: The variation in cyclic data for this test series

is somewhat greater than that which would exist if

performed in strict accordance with [3] and [4]

because of charge variability and because of the

varying point data intervals which were selected

for other testing considerations.

18
TABLE

5.

REFRIGERANT

CONTENT

OF

ISOLATED

SECTIONS

eS CO o r-~ CO r-H

m CM CO in Ov ov VO

n- *<r vO i-«

m

m VO vO vO

m vO vO cE3H CO 00 00 00 00 00 00 00 00 00 00 00 00 00

pioHC

S'? LO <r CM vO vO in

m 1"- VO vO Ov 00 <r CO

CM i—i CO p- in CO r-H CM rH

m

mm

uu<

US

CN CM CM CO CO CM CM CM rH rH i—1

Pio

co co

co

a VO

m vO CO vO 00 <1* CM OV CM OV r-. vO

g & ov 00

O CM OV CO rH o m o o & o rH l—1 CN CO CM CM CM CM rH rH rH

PH ou

hJ CO CO

m <r Ov rH rH VO rH i^. OV

Ho S'? • o rH 00 rH Ov 00 00 OV CM CM VO vO vO

cj rH rH co CO CO CO CO CO Sf pi o CO o O rH 00 1—1 rH 00 CO 00 o m rH vO VO QL_j OV OV o CO CO CM vO VO OV O o

!b o CM CO CO CO CO CO CO co co o Ph

i-J B^

rH• o o OV rH CM rH CM co

m 00 Ov

M CO VO

O 00 00 in p- 00 00 o o O

CJ CO

m vO 1—1

rH rH

Pio

COQ rH i-". 00 C" rH o vO Ov rH rH rH OV

Q §

Ov 00

O r» CO vO p^ 00 Ov ov

ISM

O

Ph <r

m o o W

B~9

CO• CO CO CO

m

m in in VO vO vO vO

ISM

CSI CM rH rH

O o o o o o O o d

O

hJ COo O o rH i—1

VO <r

mm VOm

m CM CM rH rH o o o o o o o o o o £3 • • • • • • • • • • • • • • > o

Ph o o o o o o o o o o o w

r-» rH 00 o rH Ov

m rHm in VO o m

IS B^ oc Ov

m rH CO o CM

m vO vO r-' VO vO

rJ CM CM CM CO CO co CO CO CO co CO QM

t=> O'

COO VO Ov OV CM CM OV 1—1

p» vO vO <r <r <r

m

g

p- 00 o CO vO o rH rH rH rH rH

M

rJ

O

Ph rH CM CM CM CM CO CO CO CO CO CO S

HC

dd

d

d (1) a) <u

d

dd

d PI •H •H •H •H •u •u 4-1

MH

Ph

•H •H •H •H s B

B S cd cd cd 6 6 0 B s d d 4J u 4-1 o o o O •H *H co CO CO

<Jr

m o o un o o 0 0 l 1 1 PiO

CM CM CM CMm • • • • to >v • I' t rH rH CM vO vO T) ho 03

e» e» cd cd cd COWQ

CH 4h CM 0) 0) 0)

<4H cm CM

d p!

d

ddddd •u •u 4-)

O o o o o o o o o o o CO CO CO

rH o i—1

vO r^. 00 ov o

00 i—1

Csl CN rH rH 1—1

rH Ov t—1CO

m vO

EH

=tfc =Sfe =tfc =*>= 5**;. :>Yd o o U o o o CJ) CJ CJ CJ o o O

u

O' O' O' O' O' o o O' o O O' O'

O o

19

*The

indoor

fan

was

cycled

in

phase

with

the

compressor

for

Test

QC

#11.

For

all

other

tests

the

indoor

fan

operated

continuously.
ocCH<2

CJ

CJ<

VAPOR .90

mX .44 .33 *yCO

CNCO

r>* CO .39 .45

cOXCC/3

G

CN 2 O' CO ^y CN in CN £ 2 O X ^3- O cc O* r—

X O O • CN CN CN rH 2 2

2: 2 c

CJ

.J 2O O' O' CN >y 2 2 2 2 O — a. -y <r in >y <r -y >y >y c < • • • • • • o O > u

E- IT. C£ < W C 0 C >_<

X CN O' <r CN O' *<y 2

£5

O 2 <r CN -y O' O'

X O' CN \C

CJ ^ E- c- G c 2 2 CN CN CN CN CN CN co CO < > C 2 W 2O 2 C/3 2 2 c O' O X 00 O' r— O 2 >y

;z E— 2 a. ro <r CN rH rH 2 CN CN CN C CJ c < • • 2 u: u >

CJ cc

lc

a c 2 C/3 c O

m CN (N 2 OX

H

Q 2 lO co CN r-n CO '•y in

X o < 2 O2 <r H C 2 2 c 2 cr. C 00 X 'O- CO CO CO CO CO HT 2 OZ 2 C- c O 0 O O O O O O w 2 < • • 2 2 2 >

E— 2

C£

C/3 2 c C 2 2

n- u: CN

m CN rH 2 CN r— < g < c

2 O 0 O O O c O O > > 2 • x Q C ^ Z C- co 3 < cc > cj 2 C CN O' G rJ 2 2 r-i 0 0 O O O c O c h Q cq 2 < • • => z < 2 >

O' < H 2 a

-J c u. 2 G 2 O 2 2 X co CN O' 2 X in in

U. 2 O' 2 X r- 0 CO

X 0 2 2 o o < 2 O • 2 2 CN CN CN CN CO CO CO

VJ >J < 2 Q Q 2 n o c c c c C c C C a; 2 c c c 2 2 2 2 2

e- u. < 2 2 E E E E E E E E 2 O O O O O • O m O O m O O O X 2 CN CN in

E- • • rH 2 CN *y

X

V

w 2 2 2 2 c c c c C c C ca 5

a:0 0 0 0 6 c 0 © 0 2

cn E- 8 0 u u u CJ u u a 1So 0 •H •rH tH 2 •H 2 2 2 •O 2 2 rH 2 2 2 2 2 CCJ U a u a CJ 0 O u <D

>“n so So So So :o So So 2 O CJ CJ CJ CJ O CJ O inX

O' » 00 2 in * _ O' CO

4. * H w 00 CJ CJ CJ W2

cr 0 O' 2 2 2 O 0 2 X O'. X O' O 0

CN CN 2 2 2 2

cc O0

X > CJ CJ CJ O CJ CJ > > < O o- O' O' O' O' < <

20
22% r

Compressor

&

accumulator

39MVH0 1V101 JO INJOHJd

CM

21

Figure

2.

Refrigerant

migration

during

an

"on"

cycle.
was consistent with the presence of saturated liquid and the top temperature

was indicative of the high temperature in the compressor compartment indicating

that this liquid refrigerant could not have been transferred to the accumulator

by thermal gradients

Some liquid refrigerant was also found in the suction line at the end of the

off-cycle. It is presumed that during the off-cycle, liquid refrigerant

flowed by gravity from the evaporator through the suction line to the

accumulator and that refrigerant vapor was transferred in the opposite

direction by the existing temperature gradient.

There was no evidence of substantial liquid refrigerant transfer to the com~

pressor crankcase during the off-cycle or on start-up. During a typical

cyclic test, an insulated thermocouple soldered to the bottom corner of the

compressor shell indicated a steady temperature rise from 65.0*C (149*F) at

cut-on to a temperature of 76.1*C 0.69 O F) after six minutes of operation. A

sharp decrease in the temperature would have occurred had there been a

significant quantity of liquid refrigerant in the compressor crankcase.

Consequently, it was assumed that all the refrigerant in the liquid phase

in the compressor-accumulator section was in the accumulator.

3) Most of the refrigerant that was in the evaporator at the end of the "off"

cycle left within 15 seconds after start-up. The amount of refrigerant in the

evaporator reached a minimum at approximately 1.5 minutes after start-up and

then steadily increased to its steady-state value.

4) The amount of refrigerant in both the condenser and liquid line increased

steadily from their start-up to their steady-state values.

22
5) The amount of refrigerant in the compressor and accumulator section reached

its maximum recorded amount of refrigerant (38% of the total charge) at the data

point 15 seconds after start-up. The amount of refrigerant in this section slowly

decreased to its steady-state value (5% of the total charge) as the accumulator

gradually returned its liquid refrigerant to circulation. Clearly much of the

cyclic loss can be attributed to the unit being in effect "undercharged" while

the liquid refrigerant is stored in the accumulator. This point is emphasized in

figure 3 in which the unit capacity and the temperature measured by a thermocouple

soldered to the bottom comer of the accumulator minus the compressor saturated

suction temperature are plotted against time. Here it is seen that the capacity

does not reach its steady-state value until the accumulator bottom temperature

departs substantially from the compressor saturated suction temperature, an

indication that the accumulator has been emptied of liquid refrigerant.

Further comparison of the refrigerant migration data to unit performance

requires calculation of the refrigerant side capacity. Fig. 4 shows the

measured air side capacity for the first six minutes of operation and a

refrigerant side capacity which was calculated from this air side data.

The refrigerant side capacity was calculated by drawing a smooth curve through

the air side data points and then assuming that, for small time intervals

(0.1 minute), the air side capacity was exponentially approaching a constant

refrigerant side value. Expressed as an equation this would be:

5.1

23
amvA aivis-Aams jo iNJoaad

2 b

Figure

3.

Comparison

of

time

for

capacity

to

reach

steady-state

to

time

required

to

empty

accumulator.
100

90

80

70

60

50

40

30

20

10

0

Refrigerant side capacity

calculated from air side data

Air side capacity plus fan power _

TIME (min)

ire 4

.

Comparison of Transient air-side capacity to

coincident transient refrigerant side capacity.

25
where Q^p = refrigerant side capacity

= air side capacity

t l

,t

2 = t:*- mes 1 an^ 2, min *

x = evaporator time constant, MC/UA, min.

The evaporator time constant was calculated from the coil mass data given in

Table 2 and from the steady-state test value of UA (equal to the steady-state

capacity divided by the air to refrigerant saturation temperature difference)

giving a value of 0.63 minutes for the evaporator time constant. In figure 5

this calculated refrigerant side capacity is compared to the mass of refrigerant in the evaporator. The two curves follow one another in shape quite well

after 15 seconds and lead to the following interpretation for the shape of the

refrigerant side capacity curve.

These results may be explained by considering at cut-on, that there is a

sufficient amount of refrigerant in the evaporator to allow near steady-state

capacity even though the capillary tube is not feeding sufficiently to sustain

this capacity level. As the refrigerant leaves the evaporator by nucleate boiling and slug flow, liquid is carried into the accumulator. This drops capacity

to a level that can be sustained by the flow that the capillary will permit.

This "capillary-limiting" capacity value is reduced below the steady-state

value primarily because the unit is effectively "undercharged" by the amount

of liquid temporarily lying in the accumulator. As the accumulator returns

its retained liquid to the system, the condenser pressure increases causing

the capillary flowrate to increase toward its steady-state value. This

phenomena is further complicated by the lack of stable liquid seal ahead of

the capillary during this period.

26
100

90

80

70

60

50

40

30

20

10

TIME (min)

igure 5

.

Comparison of refrigerant side capacity

to refrigerant mass in evaporator.

27
The unit sight glass was observed during several cycles and found to be 1/2

full after 30 seconds of operation. The outlet port of the sight glass was

3/4 full after 1 minute of operation. After 1 minute and 45 seconds, the outlet port was covered but the bubble in the top of the sight glass occasionally

connected to the outlet port until 4.5 minutes after start up. This indicated

that total liquid stability between the condenser outlet and the capillary tube

entrance had not yet been established. The sight glass was completely clear

5.5 minutes after start-up. The pseudo-high refrigerant capacity noted in the

first minute then, is a result of the excess of evaporator refrigerant accumu- '

lated during the off-cycle and, as this is "pumpedM out, the capillary feed

tends to starve the evaporator until, gradually, a steady flow rate is achieved.

The two sets of test results, shown in figure 6, were obtained by allowing the

unit to come to steady-state capacity after several 6 minute-on/24 minute-off

cycles. They differed in the indoor air dew point which was 37.0°F (2.8°C) for

Test QC 26 and 20.5°F (-6.4°C) for Test QC 27. The second test (at the lower

dew point) was pe-formed to determine if, when a very low saturation temperature

occurred during the starved evaporator period, there would be a significant

intra-cycle latent dooling effect. It was concluded from the closeness of these

two curves that this was not a significant concern.

The insight provided by figure 5 is useful in preparing a regressive fit to the

capacity as a function of time presented in figure 6. The experimental data

shown in figure 6 were fitted with two curves. In the first, a single time

* * —t/t1

constant exponential curve, Q/Q

gs = (1 - e ), where Tl + 3.0 minutes, is

28
8

*<

*5

AllOVdVO BiViS AQVaiS JO lN30H3d

29

Figure

6.

Regressive

fit

to

experimental

transient

normalized

capacity

data.
satisfies the 0 and infinity time conditions and matches the transient data

values fairly well after four minutes of operation. However, before that time

the indicated capacity is much too low.

_. t

/

t2 If a second time constant in the form of a term (1 + a e ) , where

a = 4.0, T2 = 0.8 minutes, is introduced to simulate the very high, but

rapidly decaying, initial capacity shown in figure 5, the resulting equation is

* t

Q = Q

ss

= (1 - e'

T/Tl

)(l + a e‘

T/T

) 5.2

This equation form, when fitted to the data, provides a very good trace of

the phenomena observed. It should be noted that the terms, Tl and t2 , in

these equations have now become regressive curve fitting constants and are

not time constants in the classical transient heat-transfer capacitance

sense of equation 5.1.

An additional test was performed to record the capacity decay after shut-down.

The results of this test are shown in figure 7. The regressive least squares

fit to this data resulted in the shown time constant of 0.61 minutes. During

this off-cycle it is not surprising that a simple decay function should simulate the transient capacity so well since none of the accumulator or capillary

complications are significant. Furthermore, it is interesting to note that for

this air-conditioning unit a similar coefficient value may be construed for the

start-up period.

30
Figure 7. Capacity decay after compressor shut-off

with continuous fan operation

31
As shown in figure 4, the unit refrigerant side capacity for the first 3/4

minute of operation is near the steady-state value. Making the assumption of

constant refrigerant-side capacity for the first 3/4 minute of operation, a bx

regressive exponential least-squares fit of the form y

= ae was made to the

experimental air-side data by rearranging the unit air-dide capacity to the form

*

' Kv

1 - Q/Q = ae , where b = -1/t.

ss

The results of this fit were a = 0.99 and T = 0.80 minutes. The value of T is

reasonably close to the value of 0.63 minutes calculated from the coil mass,

specific heat, and heat transfer rate. A somewhat higher experimental time constant could be Attributed to additional system masses in addition to the evaporator and to the measurement system. This time constant being equal to l2 in

equation 5.2 is purely a coincidence.

32
6 . CONCLUS IONS

The empirical results of this study suggest that it is possible to simulate the

dynamic operation of an air conditioner on two levels of complexity. The first

level, a single time constant characterizing the start-up and shut down operation, tends to ignore some of the actual physical phenomena that occur, however, as an approximation or a next step past steady-state towards reality

this may be adequate. Such an approach may be considered when the model is

for characterizing a "typical" or the "general class" of air conditioners.

Unfortunately a time constant value that is the best representative of the

state of the art is not possible to determine without considerable further

study of many different units. The value obtained for this unit is undoubtedly

a typical value but how wide the range or what the median is for the air

conditioning units of today is not known.

The second level of simulation is that which is of the form shown in equation

5.2. Here both facets of the cyclic degradation are accounted for and such a

form could be used to characterize the operation of any given air conditioner.

This level of simulation and its respective empirical data are essential if

the intention of the simulation is to distinguish among individual units’ characteristics or dynamic performance. That is to say, if proper credit is to

be attributed to a specific design for rating purposes. If the air side capacity of a specific unit were to be determined, as shown in Fig. 4, from a

continuous plot of the air side thermopile temperature difference, that air

conditioner’s dynamic performance would probably be simulated as accurately as

is done by the current cyclic tests. However, between the pre-measurement operations required to assure periodic repeatability and the confirming shut down

33
dynamic measurements, the experimental burden is no less than the existing

cyclic tests. It would therefore appear that the best hope for a less burdensome characterization of a given unit's dynamic performance is through sufficient study of the individual components' interaction and impact on refrigerant

migration and then an establishment of a series of enhancement (positive or

negative) factors which could be attributed to a given unit's steady-state

performance depending on its particular design.

34
7. REFERENCES

1. Chi, J. and Didion, D. , "A Simulation Model of the Transient Performance

of a Heat Pump", International Journal of Refrigeration, Volume 5, Number 3,

(International Institute of Refrigeration, Paris, France, May 1982).

2. Nguyen, H., Goldschmidt, V. W. and Tree, D. R. , "Analysis of the DoE Test

Procedures", Publication No. HL 81-11 (Ray W. Herrick Laboratories, Purdue

Unqversity, West Lafayette, Indiana 47907, March 1981).

3. Kelly, G. E. and Parken, W. H., "Method of Testing, Rating and Estimating

the Seasonal Performance of Central Air Conditioners and Heat Pumps Operating in the Cooling Mode", NBSIR 77-1271, (National Bureau of Standards,

Washington, D.C. 20234, April 1978).

4. "Methods of Testing for Seasonal Efficiency of Unitary Air-Conditioners

and Heat Pumps", ASHRAE Standard 116-83, (American Society of Heating,

Refrigerating and Air Conditioning Engineers, Atlanta, Georgia 30329, 1983).

5. Parken, W. H., Beausoliel, R. W., and Kelly, G. E., "Factors Affecting the

Performance of a Residential Air-to-Air Heat Pump", ASHRAE Transactions,

Vol. 83, Part 1, (American Society of Heating, Refrigerating, and Air

Conditioning Engineers, Atlanta, Georgia 30329, 1977).

6. "Methods of Testing for Rating Unitary Air Conditioning and Heat Pump

Equipment", ASHRAE Standard 37-69, (American Society of Heating, Refrigerating, and Air Conditioning Engineers, Atlanta, Georgia 30329, 1969).

35
APPENDIX A - COMPARISON OF CYCLIC PERFORMANCE TO REDUCED CHARGE STEADY-STATE

PERFORMANCE

In an attempt to verify the "starved evaporator" concept suggested, several

steady-state tests were performed with the refrigerant charge reduced for

comparison to the transient data taken during the period in which liquid trapped in the accumulator was believed to be the major cause of cyclic loss. This

data is shown in figure A1 and capacities calculated from the indicated linear

least squares fit to the capacity data in figure A1 are compared to the transient capacity from figure 4 in figure A2 . The capacity from 1.5 minutes

after start-up until 6 minutes after start-up exceeds the value calculated

from the transient data by approximately 10%. The air side capacity which is

equivalent to the refrigerant side capacity calculated from the accumulator

contents and reduced charge steady-state capacity as shown in figure A2 is

compared to the experimental air side capacity in figure A3.

36
CHARGE (kg)

3.0 3.5 4.0

Figure Al. Unit capacity as a function of charge.

37

CAPACITY

(kW)
90

80

70

60

50

40

30

20

10

T

Capacity calculated from

accumulator charge and reduced

charge steady-state data

Refrigerant side capacity

calculated from air side

cyclic test data

1 2 3 4 5 6

TIME (min)

Comparison of refrigerant side capacity to capacity

calculated from accumulator content and reduced charge

steady-state data.

38
AllOVdVO 31V1S-AQV31S 30 lN30H3d

39

Figure

A3.

Comparison

of

air

side

capacity

calculated

from

accumulator

liquid

charge

and

reduced

charge

steady-state

data

to

experimental

air

side

data.
APPENDIX B

REFRIGERANT MIGRATION MEASUREMENT TECHNIQUES

This appendix describes unsuccessful measurement techniques which were tried in

addition to the pneumatic valve isolation technique which produced the data discussed in this report. With additional effort it may be possible to surmount

the difficulties which are described below. This effort would be worthwhile as

these unsuccessful measurement methods (weighing and accumulator surface temperature measurement) both produce data without interruption of unit operation.

Prior to installation of the pneumatically operated ball valves for isolation

of system components, a considerable amount of effort was put into attempts to

measure refrigerant migration by weighing the outdoor unit. These attempts

failed for several reasons, some of which are:

1) Stiffness of the suction and discharge lines - this reduced the scale

sensitivity and, if the unit rocked on the scale, produced a zero

shift

2) Flexure of the suction and discharge lines with pressure—as the

unit cycled the discharge pressure varied from approximately 220 psia

to 160 psia and the suction pressure from 65 psia to 160 psia.

These fluctuations in the lines resulted in forces being transmitted to the scales upon which the unit rested of approximately

+0.5 pounds. This effect was studied by pressurizing the unit with

helium and was found to depend on the initial force the lines exerted

40
on the unit (i.e., how they were supported) and to have poor repeatability. It is presumed that there was also an effect on line force

due to cyclic temperature changes of the lines, however, no attempt

was made to verify or measure such an effect.

3) The refrigerant weight is a small percentage of the total weight of

the outdoor unit. Typically the refrigerant weight to measured would

be on the order of 4 pounds and the unit weight 160 pounds (4/160 =

2.5%). A small error in unit weight would result in a large error

in presumed refrigerant weight.

After a continuing lack of success with weighing the outdoor unit, the ball

valves were installed which produced the data discussed in this report. The

ball valve data indicated substantial "off" cycle refrigerant migration to the

accumulator. This migration to the accumulator was important in evaluating

the sources of cyclic loss. Since the accumulator is normally part of the

outdoor unit along with the compressor and the condenser, weighing the outdoor

unit, since it cannot distinguish between refrigerant contained in the condenser as opposed to that in the accumulator, would not have provided adequate

data for this analysis.

An attempt was made to determine the refrigerant content of the accumulator by

monitoring insulated thermocouples soldered to the accumulator and spaced 25 mm

(one inch) apart up its side from bottom to top. These indicated a refrigerant

content in the accumulator approximately 50% greater than that measured with

the quick closing valves. The possible reasons for this are turbulence of the

refrigerant in the accumulator as a result of boiling and depression and perturbation of the refrigerant surface caused by the high velocity of the

entering refrigerant vapor.

41
NBS-714A (rev. 2-«c) U.S. DEPT. OF COMM.

BIBLIOGRAPHIC DATA

1. PUBLICATION OR

REPORT NO.

2. Performing Organ. Report No. 3. Publication Date

SHEET (See instructions) NBSIR 83-2756 August 1983

4. TITLE AND SUBTITLE

A Laboratory Investigation of Refrigerant Migration in a Split Unit Air Conditioner

5. AUTHOR(S)

William J. Mulroy and David A. Didion

6. PERFORMING ORGANIZATION (If joint or other than NBS, see instructions) 7. Contract/Grant No. NATIONAL BUREAU OF STANDARDS

DEPARTMENT OF COMMERCE t. Type of Report & Period Covered

WASHINGTON, D.C. 20234

9. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street. City. State. ZIP)

U.S. Department of Energy

Office of Conservation and Renewable Energy

1000 Independence Avenue, S.W.

Washington, D.C. 20585

10. SUPPLEMENTARY NOTES

I I Document describes a computer program; SF-185, FIPS Software Summary, is attached.

11.

ABSTRACT (A 200-word or less factual summary of most significant information. If document includes a significant

bibliography or literature survey, mention it here)

The relationship between cyclic refrigerant migration and cyclic loss for a residential, split-system air conditioner has been investigated. The cyclic refrigerant

migration was measured at different points in the operating cycle by simultaneously

shutting five pneumatic valves which isolated the refrigerant in the major system

components. The refrigerant was then removed, weighed, and returned to the system.

The unit tested was found to have a high initial capacity as migrated refrigerant was removed from the evaporator and then a low, slowly increasing capacity as trapped

refrigerant was returned to the system from the accumulator. The unit performance was also compared to single and double time constant regressive approximations and

to the time constant calculated from the evaporator mass and heat transfer coefficient. Although relationships between migrated refrigerant and cyclic capacity were

observed, no practical refrigerant migration test method that would be less burdensome than the cyclic tests of ASHRAE Std. 116 appears possible at this time.

12. KEY WORDS (Six to twelve entries; alphabetical order; capitalize only proper names; and separate key words by semicolon s) central air conditioners; cyclic testing; heat pumps; refrigerant migration

13. AVAILABILITY 14. NO. OF

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| | For Official Distribution, Do Not Release to NTIS

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47

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