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PUBLICATIONS

NB

A 1 1 1 2 b 3 m 0 ^

NATL INST OF STANDARDS & TECH R.I.C.

All 102631 409

Mulroy, William J/The performance of a c QC100 .U56 NO. 86-3422 1986 V19 C.1 NBS-P The Performance of A

Conventional Residential Sized

Heat Pump Operating With A

Nonazeotropic Binary Refrigerant

Mixture

William Mulroy

David Didion

U.S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Gaithersburg, MD 20899

October 1986

Sponsored By:

Office of Buildings and Community Systems " ° nepartment of Energy through

1 3idge National Laboratory under

100 act DE-AC05-84OR21 400 with

. IJ 5 B n Marietta Energy Systems, Inc.

86-3422

NBSIR 86-3422

THE PERFORMANCE OF A

CONVENTIONAL RESIDENTIAL SIZED

HEAT PUMP OPERATING WITH A

NONAZEOTROPIC BINARY REFRIGERANT

MIXTURE

UBS

RESEARCH

INFORMATION CENTER

Gttjoo

i

l42(.

William Mulroy

David Didion

U.S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Gaithersburg, MD 20899

October 1986

Sponsored By:

Office of Buildings and Community Systems

U.S. Department of Energy through

Oak Ridge National Laboratory under

contract DE-AC05-840R21400 with

Martin Marietta Energy Systems, Inc.

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

NATIONAL BUREAU OF STANDARDS. Ernest Ambler, Director

Executive Summary

As part of the Department of Energy (DoE)/Oak Ridge National Laboratory (ORNL)

Building Equipment Research (BER) program, the National Bureau of Standards

(NBS) performed laboratory tests of a relatively unmodified residential heat

pump designed for R22 when charged with a nonazeotropic refrigerant mixture

(NARM). Performance of the heat pump charged with a binary NARM of 6 5%

R13Bl/3 5% Rl52a was compared to R22. The advantages claimed for NARMs in this

application are improved thermodynamic efficiency as a result of gliding

refrigerant temperatures in the evaporator and condenser, and capacity

modulation through composition shifting.

Four test series (at the DoE rating conditions) were conducted: 1) heat pump

unmodified, 2) charge and expansion device area optimization tests, 3) tests

to evaluate potential performance with wide shifts in NARM composition, and 4)

tests with a receiver added to evaluate its ability to increase the NARM

composition shift.

Efficiency improvement from the gliding temperature effect was found to be

unlikely in an air-to-air heat pump or air conditioner because of the cross

flow design of most current heat exchangers and, for heat pumps, because of

the refrigerant flow reversal when switching from cooling mode to heating mode

operation. The gliding temperature effect can; however, be successfully used

in water-to-water applications.

The expansion device size should be changed when the refrigerant is changed

for one of different properties (liquid versus vapor density); so as to ensure

that proper subcooling will be maintained.

i i i
Capacity modulation through composition shift is a successful control strategy

that will increase seasonal performance in northern applications.

It was shown that if a control strategy was used which maintained subcooling

and composition could be shifted to 86% R13B1, heating capacity at 17 °F could

be increased 18% above that with R22 at comparable conditions.

Substitution of this particular NARM in the unmodified (except for expansion

device optimization) heat pump resulted in improved low temperature (17 °F)

heating performance (up 6% in capacity, no change in efficiency) at the

expense of reduced high-temperature heating (down 3% in efficiency) and

cooling (down 1 2% in capacity, 11% in efficiency) performance.

Adding a receiver resulted in a greater composition shift at low air

temperature and greater capacity enhancement relative to R22 (14% capacity

increase, 2% efficiency increase).

No potential reliability or safety problems as a result of substitution of the

NARM for R22 were observed.

It is recommended that future work on residential air conditioner and heat

pump applications of NARM's be directed toward equipment redesign to better

use their characteristics and to analysis of the performance characteristics

and potential of other NARM's than the one tested for this report.

IV
ABSTRACT

This report presents laboratory performance measurements of a relatively

unmodified residential heat pump designed for R22 when charged with a

nonazeotropic, binary mixture of R13B1 and R152a. Results are presented for

various sizes of fixed expansion devices. The effect of gliding temperature

within the saturation zone was found to be small. This experimental

investigation confirmed that flash di sti 1 lation within the accumulator would

improve low temperature heating performance. The measured performance was

approximately 11% lower in both efficiency and capacity than R22 for air

conditioning. The high temperature heating efficiency was 3% lower than R22.

The low temperature heating capacity was 14% higher and efficiency 2% higher

than R22. These results show a substantial improvement over R22 for heating

applications at the expense of reduced cooling mode performance. Further

performance enhancement for this or other mixtures is expected through various

system modifications which remain to be studied.

Key Words: air conditioning; heat pumps; non-azeotropic refrigerants;

refrigerant mixtures; refrigerants; refrigeration.

v
TABLE OF CONTENTS

Page

EXECUTIVE SUMMARY iii

ABSTRACT v

LIST OF FIGURES vii

LIST OF TABLES viii

CONVERSION FACTORS FROM METRIC (SI) TO ENGLISH UNITS ix

1. INTRODUCTION 1

2. BACKGROUND. 3

3. TEST HEAT PUMP AND FACILITY 10

4. COOLING MODE TESTS 16

5. HEATING MODE TESTS 27

6. DISCUSSION 43

7. CONCLUSIONS 50

8. REFERENCES 52

APPENDIX. OBSERVATIONS ON EXPERIMENTAL TECHNIQUE 54

v 1
LIST OF FIGURES

Figure 1. Dew and bubble temperatures for mixtures of R13B1

and Rl 52a at a pressure of 90 psia 5

Figure 2. Energy savings as a result of increased low temperature

compression system capacity reducing the need for

"make-up" resistance heat 7

Figure 3. Heat pump indoor unit installed in test apparatus 11

Figure 4. Heat pump outdoor unit 12

Figure 5. Summarized cooling mode test results 21

Figure 6. Comparison of latent capacity as a function of total

capacity 22

Figure 7. Evaporator (indoor coil) return bend temperatures

during a typical 95°F, DoE "A", cooling test 24

Figure 8. Summarized 47 °F heating mode test results 30

Figure 9. Summarized 17 °F heating mode test results 31

Figure 10. Condenser enthalpy change of R13B1/R152a binary

mixture in heating mode of 17 °F outdoor temperature 32

Figure 11. Evaporator (outdoor coil) return bend temperatures

during a typical 17 °F heating test 34

Figure 12. Condenser (indoor coil) return bend temperatures

during a typical 17 °F heating test 35

Figure 13. Comparison of binary refrigerant energy performance

to R22 46
LIST OF TABLES

Table 1. Summarized Results for Cooling Mode Tests Performed

with a 65%/35% Mixture of Rl3Bl/152a 17

Table 2. Summarized Results for Cooling Mode Tests Performed

with R22 18

Table 3. Comparison of System Pressures at the DoE 95°F,

"A", Test Condition 26

Table A. Summarized Results for Heating Mode Tests Performed

with a Mixture of R13B1/Rl52a Charged to a 65%/35%

Original Composition 28

Table 5. Summarized Results for Heating Mode Tests Performed

v/it*1 R22 29

Table 6. Heating with Subcooling and Manually Adjustable Needle

Expansion Valve 38

Table 7. Summarized Results for Binary Mixture Heating Tests

with Receiver and Comparable Tests with R22 41

v i i i
SI CONVERSION FACTORS

MULTIPLY SI TO OBTAIN

Btu/h, Btoh 0.2 93 W

Btu/lbm°F, [C

p

, specific hest] 4.19 kj/kg°i

°F •C - (°F - 32J/1.8

ft 0.3048 m

ft/min, fpm 0.00508 mJ i

ft3

/ lbm 0.0623 m

3 /kg

ft 3

/ min, CFM 0.472 m

3

/ s

gpm (US) 0.0631 L/S

inch 25.4 mm

inch of water 3 .38 kPa

LB tu/h 0.2931 kW

lbm/ h 0.000126 kg/s

ton of refrigeration capacity 3516 W

IX

1. INTRODUCTION

The original goal of this study was to determine experimentally the energy

performance benefits of a nonazeotropic binary refrigerant mixture in an

unmodified residential heat pump. The single component refrigerant, R22, for

which the unit tested was originally designed, was selected as a comparison

basis. The mixture used in this study, R13B1/R152a, was chosen because it was

one of several touted in the literature L1,2,3J as having potential for

improving performance, and also because it was readily available from the

manufacturer. This mixture may be considered as an example of any of a number

of mixture possibilities and the quantitative values of this study are

significant in that they indicate the relative merits of the different

characteristics a mixture may offer (i.e., variable temperature phase change,

variable capacity, bilevel temperature condensation/ evaporation, etc.). Other

mixtures may offer greater or lesser quantitative values depending on their

properties. The determination of the optimum mixture for heat pumps was not a

subject of this study.

As this study progressed it became apparent that a simple substitution of a

refrigerant mixture for R22 was not a fair evaluation of the relative benefits

of the mixture. The machinery components (i.e., heat exchangers, expansion

device, compressor, etc.) had been originally sized for a particular

refrigerant. The substitution of another refrigerant would cause the same

machinery to perform differently not only because of thermodynamic properties

differences, but also because parts of the machinery had become over- or

undersized. Therefore the second part of this study focused on charge

optimization and expansion device size determination.

1
A third series of tests was performed to evaluate the potential performance

with greater composition shift than had been observed in previous tests.

A fourth series of tests was performed with a receiver added to evaluate the

ability of this hardware addition to increase composition shift.

The heat pump performance criteria of prime consideration are capacity and

efficiency. Of the two, capacity is often the more important because, for

many applications (i.e., northern climates), an increased capacity at lower

temperatures will reduce the resistance backup heat demand and thus

significantly improve the seasonal efficiency. Performance criteria relating

to reliability are also important and thus discharge pressure and temperature,

suction pressure, pressure ratio, and pressure difference were measured and

reported as well.

2
2 . BACKGROUND

Nonazeotropic mixtures are those which exhibit a temperature change at

constant pressure as they condense or evaporate. Binary mixtures consist of

two components. This report deals exclusively with a binary nonazeotropic

mixture of refrigerants R13B1 and R152a applied to a residential heat pump.

Several advantages have been claimed for nonazeotropic mixtures in refrigeration systems. Among these are improved thermodynamic efficiency (as a result

of gliding temperatures in the two-phase zone, i.e. Lorenz vs. Carnot cycle}

L4J , mul ti-temperature level evaporators or condensers (useful for frozen and

produce compartments in refrigerators and for domestic hot water in heat pumps

and air conditioners) L5J , and refrigerant composition change which alters

the working media’s capacity allowing it to match load requirements [6]. The

last of these is the principal concern of this study.

The retrigerant mixture used in this study is one that has been recommended

for heat pump applications [1,2,3] . In comparison to refrigerant 22,

retrigerant 13B1 is more volatile, operating at higher pressures for

equivalent saturation temperatures. It produces more capacity in a typical

refrigeration system as a result of having a vapor density great enough to

more than compensate for a smaller latent enthalpy change. However in a

typical cycle [7], it is less efficient. Refrigerant 152a, in comparison, is

of lower volatility, has a lower operating pressure, and is a lower capacity

(lower vapor density) refrigerant with a relatively large latent enthalpy and

high efficiency at higher evaporator temperatures. The mixture of the two, 65

percent by weight of R13B1, was chosen by DuPont to give capacity and

efficiency similar to R22 during air conditioning operation and higher

3
capacity than R22 for heating operation after composition shift had taken

place.

The greater the volatility difference between the two refrigerants, the easier

it is to shift the composition of the circulating refrigerant from its initial

value to one with a higher concentration of the higher boiling point

component. This shifting may be effected passively by flash distillation in a

conventional accumulator or actively with an added distillation system. The

shifting process for the simple heat pump described within this report was a

result of liquid flooding into the suction line accumulator during low ambient

temperature of the heating mode only. This caused an increase of R13B1 in the

composition of the circulating fluid.

When evaporation or condensation of a nonazeotrope takes place, the concentrations of the two components is different in the vapor and liquid phases.

In particular, the vapor is always richer in the more volatile component and

the liquid is always richer in the less volatile component in comparison to

the initial composition. The extreme composition differences for this initial

composition occur at the beginning and end of the phase change process. For

example, as shown in figure 1, when a liquid mixture of 65% 13Bl/35% 15 2a

(x„ in figure 1) at 90 psia, first reaches its bubble temperature, it will be

<c» a in equilibrium with vapor at an approximate composition of 82% 13B1/18% 152a

(x ). An intermediate point approximately half way through the two phase

region is shown in figure 1 at b to result in equilibrium between liquid 5UX

13Bl/50% 152a (x^ and vapor 74% l3Bl/26% 152a (x v

^). The final example of

4
TEMPERATURE

(°F)

100% R152a x ViC 0% R152a

COMPOSITION (mass percent)

Figure 1. Dew and bubble temperatures for mixtures of R13B1 and Rl52a at

a pressure of 90 psia.

5
figure 1 shows nearly completed evaporation at the dew line, c, where a liquid

of 36% 13Bl/64% 152a (x^

c

) is in equilibrium with vapor at the original 6 5%

13Bl/35% 152a (x ) composition.

Thus, refrigerant liquid in an accumulator will be quite rich in the less

volatile component resulting in the remaining refrigerant being enriched in

the more volatile refrigerant. In the case of the mixture used in these

tests, flooding into the accumulator would increase the percentage of R13B1

circulating in the system and consequently increase low temperature heating

capacity over that of R22 resulting in a substantial system efficiency

increase by the reduction of need for electric resistance heat. This is shown

in figure 2.

In addition to mixture characteristics, it is also useful to understand the

system characteristics of a fixed expansion device heat pump that bring about

flooding into the accumulator with consequent flash distillation at low

outdoor temperatures. To understand this, two simplifications should be made.

First, that mass flow through the fixed expansion device is choked, which is

normally the case [8], and that, therefore, the expansion device mass flow

will be only a function of high side refrigerant properties. Second, that the

mass flow through the compressor is primarily a function of suction vapor

density, that is the low side refrigerant properties.

In the heating mode, when the outdoor temperature decreases, the mass flow

rate that the compressor pumps decreases because the suction vapor density

decreases. The amount of liquid refrigerant that the expansion device can

pass does not change because the indoor air temperature does not change and

6
THERMAL

ENERGY

OUTDOOR TEMPERATURE

Figure 2. Energy savings 86 a result of increased low temperature

compression system capacity reducing the need for "make-up"

resistance heat.

7
hence the high side saturation state does not change significantly. When the

outdoor temperature drops to such a level that the liquid mass flow that the

expansion device can pass exceeds the vapor mass flow which the compressor

is capable of pumping , liquid refrigerant will leave the condenser at a

greater rate than it is replenished until subcooling is lost and two phase

instead of liquid flow is passing through the expansion device. When steady

state is reached, this two phase flow will be equal to the compressor mass

flow.

In the cooling mode, as the outdoor temperature rises, the amount of liquid

refrigerant that the expansion device can pass increases as a result of the

increased saturation pressure in the condenser. The compressor capacity stays

about the same since the indoor air temperature does not change. As before,

expansion device capacity exceeding compressor capacity results ultimately in

flooding through the evaporator into the accumulator. Conversely, if the

outdoor temperature decreases, compressor capacity exceeding expansion device

capacity will cause liquid refrigerant to back up in the condenser and cause

the accumulator to be dry and subcooling to exist.

In either case the cycle accommodates itself to the changed outdoor

temperature to achieve steady state with the mass flow through the compressor

and expansion device equal. In the case of an expansion device capacity less

than the compressor capacity, this is accomplished by a decrease in suction

pressure and increase in head pressure that slightly increases expansion

device flow and, as a result of this increased pressure difference, reduces

the cycle efficiency. In the case of an overly large expansion device

8
compared to compressor capacity, the mass flow through the expansion device

will be reduced to match the compressor capacity by passing vapor (which

contributes little to capacity) as well as liquid refrigerant. This also

reduces cycle efficiency.

In the control strategy for this binary mixture, it is not desired to have

concentration shift at high temperature cooling even though this would help

match the load line because it would result in high head pressures and reduced

efficiency which would be undesirable for summer peaking utility applications.

Hence the cooling expansion device orifice should be chosen to give enough

superheat at the 82°F rating point to prevent flooding at higher outdoor

ambient temperatures.

In heating the expansion device orifice should most likely be chosen to give

best efficiency (i.e., a small amount of subcooling) at 47°F and to accept

whatever degree of composition shift occurs at lower temperatures, although

these criteria are less clear than those for cooling.

It should be noted that flooding into the accumulator is dependent on loss of

subcooling in the condenser. More refrigerant in the accumulator means higher

quality and void fraction in the condenser. Adding a receiver to the system

as discussed for the last series of tests in this report increases system

design flexibility by allowing varying (but equal in sum) amounts of

refrigerant to be in the receiver and accumulator while maintaining a constant

high side quality defined by the requirement that the receiver be at

saturation conditions if it is partially full.

9
3. TEST HEAT PUMP AND FACILITY

A nominal three ton, split system, air-to-air unitary heat pump was used for

this investigation. 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 adjoining environmental

chambers. The unit was tested in the environmental chambers and apparatus

described in Reference [9].

The indoor section of the unit attached to the air flow measurement tunnel is

shown in figure 3. The outdoor section of the unit with the cover removed to

show the coil return bends is shown in figure 4.

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

cooling capacities was the indoor air enthalpy method [10]. 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 the

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.

10
mmM

Figure 3. Heat pump indoor unit installed in test apparatus

11
1 2

Figure 4 Heat pump outdoor unit
A thirty-junction thermopile was used 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 relative humidity of the air entering and leaving the indoor unit was

measured using an aspirated psychrometer. An aspirated psychrometer was also

used to measure the relative humidity of the air leaving the indoor unit

during cooling mode tests and the relative humidity of the air entering the

outdoor unit during heating mode tests.

The volatile refrigerant flow method [10] was used as the secondary test

method. The liquid refrigerant 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. Adequate subcooling

to prevent flashing at the turbine meter was necessary for accurate

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

possible to use this secondary method for some tests. The refrigerant flow

was determined by the turbine meter and an electronic totalizer. The density

of the refrigerant was determined from the temperature and pressure readings

and the mixture equations of state 111, 12J.

Watthour meters were installed to measure the electrical energy consumption of

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

13
and printers were connected to the watt-hour 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 gauges were installed at various locations and

temperatures and pressures were recorded to provide additional information on

the performance of the unit and as a check for consistent operation.

At the conclusion of the initial test series the unit was modified by the

addition of valved circuits in parallel with the capillary tubes to allow use

of interchangeable orifices to facilitate expansion device optimization. A

single orifice with a tee was used to replace the two capillary tubes on the

indoor "A" coil. A manifold allowing selection between three orifices was

used in place of the single capillary tube on the outdoor unit.

At this same time additional thermocouples were added to measure coil

temperatures. Individual insulated thermocouples were soldered to all the

return bends on one end of the outdoor coil and on one end of one half of the

indoor coil.

Through-the-wal 1 thermocouples were also installed through 13 return bends on

the indoor coil and 9 return bends on the outdoor coil. In their most

successful form these were created by silver-6oldering an 18 gage hypodermic

needle into a small hole drilled through the tube wall, inserting a

14
thermocouple, filling the base of the hypodermic with epoxy glue, evacuating

the system sufficiently to draw the epoxy into the hypodermic needle, and

finally crimping the back of the hypodermic needle. These through-the-wal

1

thermocouples gave trouble throughout the project with leakage, particularly

in the coil which happened to be operating as a condenser. For this reason,

several of them were eliminated toward the end of the project.

Refrigerant composition for the binary-mixture tests was measured by

withdrawing a small sample into a separate sample bottle which was later

analyzed with a gas chromatograph.

A receiver installed in the liquid line near the indoor unit was used in some

tests. An inverted oil separator with the oil return line blocked was used

for this purpose. The internal volume of this oil separator shell used as a

receiver was 140 cubic inches.
4. COOLING MODE TESTS

Cooling mode test results are summarized in Table 1 for tests performed using

the binary refrigerant mixture and in Table 2 for tests performed using R22.

As indicated, these tests were performed at either the DoE Test "A" (95°F

outdoor ambient) or at the DoE Test "B" (82°F outdoor ambient) rating

condition [13]. In both cases, the specified indoor entering condition is

80°F dry bulb and 67 °F wet bulb. Direct comparison between the capillary tube

and orifice results should not be made since the expansion device change

involved some change in refrigerant distribution, and the superheat criteria

were different. Also, the two test series were separated by a nine month

period during which this heat pump was used for another test series which

could have resulted in some wear on the equipment.

The charging criteria for the capillary tube tests was 14°F of superheat

leaving the evaporator at the 95°F outdoor ambient cooling te6t condition.

For the orifice tests this was reduced to 10°F. This amount of superheat was

desired for this test series for two reasons. First, it was desired to

quantify cooling mode performance of the nonazeotropic refrigerant without the

composition shift that would occur with flooding into the accumulator.

Second, it was desired to use the refrigerant-side mas6 flow multiplied by the

enthalpy change as a secondary capacity measurement method to establish

confidence in the tabulated refrigerant mixture properties. This capacity

measurement method as specified in ASHRAE Standard 37-78 requires 3°F of

subcooling entering and 10°F of superheat leaving the indoor coil to ensi.i <

its accuracy. Since these tests were performed, ASHRAE Standard 116-83 has

been adopted which allows use of the refrigerant enthalpy flow method with V'F

of superheat leaving the indoor coil.

16
Table

1.

Summarized

Results

for

Cooling

Mode

Tests

Performed

with

a

65%/

35%

mixture

of

R13Bl/152a

CO i—< CO

<n co cn

N N CO

• • m co n* O' vO CM «H \D

m in oo co in O'* cm m on m

CO CM CM m m in m

m O' r>. n h m

m m cm m m cm cn

• moo cm cm m

cl m m cm m cn cm cm cm m

U U o

0) C/3 C/3

ro u

0) o U 03

in U H IU 2 S H •o

Cl <D 0) a > -o

«rH -H 3

00 —

CL U

CO 03 C O -H M

17

Includes

control

circuit

power

of

approximately

40

W.
Table

2.

Summarized

Results

for

Cooling

Mode

Tests

Performed

with

R22

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18
For both binary and conventional refrigerant fixed-area-expansion-device heat

pumps, this level of subcooling (3°F to 5°F) is typical, but the level of

superheat is high. Typically at the 95°F rating point superheat would be 5°

or less. Superheat levels of 5°F to 10°F would be typical of thermostatic

expansion valves systems operating at this rating point. Although the design

or simulation of an expansion valve system was not included as part of this

project, the comparative results are probably quite close to those which would

be obtained with a suitable cross-charged expansion valve. It is more likely,

however, that an expansion device for a binary refrigerant system would employ

an electric valve sensing both outdoor temperature and liquid level in a

receiver or accumulator, in addition to temperature and pressure leaving the

evaporator.

A value of 0 is listed in Tables 1 and 2 for subcooling at the expansion

device when subcooling was lost as indicated by a lack of agreement between

the refrigerant and air side capacities.

The binary test shown in column 6 of Table 1 has minimal superheat. This test

was conducted since many likely control strategies would result in less

superheat than was used for these tests. The binary test shown in column 8

has an increased percentage of R13B1 to show the effect of this variable in

the cooling mode.

The expansion device and charge are listed only for the 95°F test. The 82 °F

test tabulated immediately below each 95°F was performed with the same

expansion device and charge on the same or the following day.

19
The capacity and efficiency test results are also summarised in figure 5. The

NBS steady-state heat pump simulation model, HPSIM [8], was used to verify the

shape of these curves for the E22 teste and confirmed the observation that

as the expansion device becomes overly restrictive the 82°F and 95°F capacity

approach each other and eventually cross. The shape of these graphs for the

binary refrigerant are quite similar, showing the above described narrowing of

the 95 e F and 82°F capacity difference for small orifices and showing

simultaneous peaks in capacity and efficiency.

The orifice giving the highest capacity at 95°F has a 30% smaller area for the

R22 tests and a 6% greater area for the binary tests than the orifice area

giving the greatest capacity at 82 °F. For general comparison, the optimum

expansion device and charge were chosen as those giving the best efficiency at

the 82°F rating point. These orifices were the 0.0781 in I.D. for R22 and the

0.0820 in I.D. for the binary mixture. Best performance for both refrigerants

was obtained with minimal high-side pressure while still maintaining some

subcooling entering the expansion device.

That the percent latent cooling for a given capacity is not strongly affected

by the different refrigerants and their associated different temperature

patterns through the evaporator is shown in figure 6 derived from Table 1 and

2. This suggests that the decrease in total capacity of the binary mixture,

also included a decrease in latent capacity due to an average increase in

evaporator temperature. The apparent continuity of this plot for the two

refrigerants as shown by the near coincidence of their respective least aquare

linear regression lines indicates that there i6 no other substantial

difference between them.

20
EER

(Btuh/w)

CAPACITY

(Btu/hr)

ORIFICE I.D. (in)

Figure 5. Summarized cooling mode test results.

21
PERCENT

LATENT

CAPACITY

TOTAL CAPACITY,

SENSIBLE PLUS LATENT (kBtu/hr)

Figure 6. Comparison of latent capacity as a function of total capacity.

22
For the mixture to have the same percent latent cooling as the R22 case, the

air flow rate would have to be reduced approximately by the ratio of their

respective capacities. This air flow reduction would result in a lowering of

the evaporator saturation temperature and in a small additional loss of

capacity and efficiency.

Plots of return bend temperatures through the evaporator for a typical test

are shown in figure 7. The saturation, bubble and dew lines shown are calculated from pressures at the evaporator inlet and outlet, assuming a straight

line gradient through the coil. The R22 temperatures can be seen to follow

the saturation temperature line predicted from these pressures until the

refrigerant reaches a quality of unity, after which point it rapidly reaches

the imposed conditions of 10 °F of superheat. The binary refrigerant

temperature gradient results from a tradeoff between the rate of heat gain and

pressure drop. The most rapid period of temperature rise occurs when the coil

circuitry splits in half, reducing the mass flow in each circuit accordingly,

and simultaneously is routed to the front of the coil where the highest heat

gain rate occurs. The temperature rise through the binary evaporator, is not

as great as the difference between the dew and bubble temperatures at this

pressure (18°F) . This is due to the fact that while the changing liquid

composition is causing an increase in fluid temperature, the flow pressure

drop is causing a decrease in saturation temperature of the fluid. The actual

temperature rise through the evaporator for the binary refrigerant was

approximately 14°F and that of the R22 approximately 1°F. This is, however, a

result of the 10°F superheat criteria reinforcing the temperature gradient for

the binary refrigerant and cancelling that of the R22. If the refrigerant

left the evaporator as saturated vapor, the temperature difference, inlet-tooutlet, assuming the same pressure drop, for the binary refrigerant would be

23
DISTANCE THROUGH EVAPORATOR

Figure 7. Evaporator (indoor coil) return bend temperatures during

a typical 95°F, DoE "A", cooling test.

24
only 3 °F and that of the R22 would be -8°F. The actual temperature gradient

for the binary refrigerant while it is in saturation is indicated in figure 7

as 9°F. The temperature difference (gliding temperature effect) of this

mixture through the evaporator tends to be of the same order of magnitude as

that for R22. In general the value for a binary mixture would be greater or

less than that for R22 depending on the temperature difference between its

bubble and dew lines, the pressure gradient, and the specified amount of

leaving superheat. Return bend temperatures are shown and discussed for both

a typical condensing operation and for a flooding evaporator in the following

section on heating test results.

The refrigerant charge was not measured for most tests. This was because the

unit was charged from separate bottles of the pure components and because as

the charge was adjusted to obtain the required superheat and composition

frequent addition or removal of small amounts of refrigerant was necessary.

This resulted in considerable uncertainty as to the final remaining charge

because of the cumulative small errors.

The refrigerant mass flow rates listed for tests in which the liquid refrigerant was subcooled were directly measured with the turbine flow meter. For

those tests in which subcooling was lost at the turbine meter, an approximation

was made at either test condition "A" or "B" by assuming a constant volumetric flow rate at the compressor suction line and using the measured pressures and

temperatures to calculate the refrigerant specific volume.

A principal reason for performing these expansion device optimization tests

was the concern about the high pressures observed during the original binary

25
refrigerant tests which used the capillary tube sized for R22. The

improvement in observed pressure at the

conditions, 95°F, is summarized in Table

most severe of

3.

the cooling test

Table 3. Comparison of System Pressures at the DoE 95 °F, "A", Test Condition

Refrigerant R22 R22 Binary Binary

Expansion Device (cap. tube

or orifice diameter, in.)

Cap .078 Cap .082

Discharge Pressure, psia 27 8 27 9 317 2 97

Discharge Temperature, °F 23 2 242 221 208

Suction Pressure, psia 82.9 91.1 82.8 90.9

Pressure Ratio 3.35 3.06 3.83 3.27

Pressure Difference, psi 195 188 234 206

The increased suction pressure for the orifice test6 over the capillary tube

tests for R22 is a reflection primarily of the change in test plan to a lower

value for the refrigerant superheat leaving the evaporator since the expansion

device was essentially optimized in both cases. Optimization of the expansion

device did reduce the binary pressure ratio to a value between the listed R22

values. The pressure difference remains slightly higher than for R22.

Compressor discharge temperature, also listed in Table 3, is seen to be

consistently better (i.e., lower) for the binary mixture than for R22.

26
5. HEATING MODE TESTS

Heating mode test results are summarized in Table 4 for those tests performed

with the binary refrigerant mixture and in Table 5 for those tests performed

with R22. As indicated in Tables 4 and 5, these tests were performed at both

the DoE 47 °F and 17 °F outdoor ambient test conditions. In both cases, the

specified indoor entering air condition was 70°F. The refrigerant charge used

for these tests was that which had produced the best efficiency at the 82°F

cooling rating condition as shown in Tables 1 and 2. For the binary tests,

this resulted in a charge of 9.5 lbs with the 0.0820 in. I.D. orifice

installed in the indoor coil for cooling mode operation. For the R22 tests, a

charge of 7.17 lbs was used with the 0.0780 in. I.D. orifice installed in the

indoor coil for cooling mode operation. Caution should be exercised in making

direct comparison between the capillary tube and orifice results since the

expansion device change involved changes in distribution and in superheat

criteria. Also, these two test series were separated by a nine month period

during which this heat pump was used for a different series of tests. This

could have resulted in some wear which might bias the results.

The capacity and efficiency results listed in Tables 4 and 5 are summarized in

figures 8 and 9. The outstanding observation from these figures is the

insensitivity of binary capacity to orifice size. An exception was the

smallest orifice at 47 °F which presumably failed to flood the evaporator. It

appears that the capacity gain from concentration shift is being offset by the

capacity loss resulting from the incomplete condensation leaving the

condenser. This point is illustrated in figure 10 in which the enthalpy

change through the condenser is compared to the value that would have resulted

had complete condensation taken place.

27
Table

4.

Summarized

Results

for

Heating

Mode

Tests

Performed

with

a

Mixture

of

R13B1/R152a

Charged

to

a

65X/35Z

Original

Composition

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28
Table 5. Summarized Results for Heating Mode Tests Performed with R22

Expansion Device, I.D., In., ^ Cap

47°F DoE Rating Condition

0.0630 0.0650 0.0670

Air Side Capacity, Btu/hr 36,200 34,628 35,340 35,220

COP (Total System) 2.79 2.69 2.74 2.73

Refrigerant Mass Flow Rate, lb/hr 383 388a 395 398

Quality Leaving Condenser, % 0 1.7 0 0

Compressor Suction, psia 64.7 68.7 68.6 68.8

Compressor Suction, °F 51 31.2 42.0 40.7

Compressor Discharge, psia 222 216 219 218

Compressor Discharge, °F 208 190 196 194

0 Compressor Power, W 3327 3300 3308 3306

0 Indoor Fan Power, W 473 473 473 473

Expansion Device, I.D., In.,

^ Cap

17 °F DoE Rating Condition

0.0630 0.0650 0.0670

Air Side Capacity, Btu/hr 20,510 19,090 20,040 19,740

COP 1.92 1.80 1.89 1.87

Refrigerant Mass Flow Rate, lb/hr 222

a 205

3 215

a 215a

Quality Leaving Condenser, % 5.8 6.8 6.7 8.2

Compressor Suction, psia 40.4 43.3 43.5 43.7

Compressor Suction, °F 11 17.4 17.4 17.1

Compressor Discharge, psia 177 175 174 174 Compressor Discharge, ®F 203 206 206 205

Compressor Power, 2642 2621 2627 2610

Indoor Fan Power, W

e 484 484 484 479

aRefrigerant mass flow and quality calculated from air side capacity and volumetric

flow meter.

^Cap implies a single capillary tube, 0.09 in. I.D. by 37.5 in. long. Otherwise the

I.D. for a single 0.5 in. long orifice is given. In both cases the expansion device

directly fed the first pass of the evaporator as shown in Figure 11.

Q Includes outdoor fan power of approximately 320 W.

^Includes indoor fan power of approximately 340 W.

6 Includes control circuit power of approximately 40 W.

29
COP

CAPACITY

(Btu/hr)

Figure 8. Summarized 47 °F heating mode test results.

30
COP

CAPACITY

(Btu/hr)

22000

21000

20000

19000 - O Binary

A R22

Figure 9. Summarized 17 °F heating mode test results.

3 1
ENTHALPY

CHANGE

MASS

FLOW

RATE

TIMES

(Btu/lb)

ENTHALPY

CHANGE

(Btu/hr)

ORIFICE I.D. (in)

Figure 10. Condenser Enthalpy Change of R13B1/R152a Binary Mixture

in Heating Mode at 17°F Outdoor Temperature

32
The enthalpy used in figure 10 and the tabulated mass flov for most of these

tests (as indicated by a superscript in Tables 4 and 5) was calculated by

varying the assumed quality at the measured pressure at the turbine flowmeter

to find a value which was consistent with the volumetric flow measured by the

turbine meter and with the measured air side capacity (see Appendix).

Although the capacity was insensitive to orifice size, the efficiency showed a

clear peak for 47 °F operation since the shift to concentrations richer in the

more dense R13B1, helped to offset capacity loss but not efficiency loss. A

similar peak for 17 °F operation would be expected if orifices small enough to

cause evaporator starving at that temperature had been used.

The binary control strategy, as originally conceived, called for a dry

accumulator during 47 °F operation and flooding causing composition shift

during low temperature operation. It appears that starving the evaporator

causes severe efficiency loss and that the best control strategy is to barely

flood at 47 °F, which still results in composition shift at lower temperatures

because the degree of flooding still increases as outdoor temperature drops.

As was observed in the cooling mode tests, the binary discharge pressure is

higher and the discharge temperature lower than would be the case for R22.

In figures 11 and 12, return bend temperatures through the evaporator and

condenser for R22 and the mixture are compared for 17 °F flooded evaporator

operation. In this case the refrigerant mixture gives temperatures within the

evaporator (figure 11) that rise or fall depending on the relative effects of

pressure drop and heat transfer with a leaving temperature approximately 1°F

33
cdo

•Ha

;>> jj

cd

a>u

flEc •H

U34

DISTANCE

THROUGH

EVAPORATOR
RETURN

BEND

TEMPERATURES

(°F)

110

I

I

I

I

I

I

80

Outlet Inlet

DISTANCE THROUGH CONDENSER

Figure 12. Condenser (indoor coil) return bend temperatures

during a typical 17°F heating test.

35
higher than its entering value, whereas the R22 temperatures steadily fell

leaving the evaporator 5°F lower. The heat transfer rate to this coil was

more uniform than the indoor coil in that its circuitry zigzagged from front

to back instead of proceeding row-by-row. At each branch point the mass flow

through each circuit is halved thus resulting in an approximate halving of the

flow-pressure-drop rate. Since heat transfer is primarily limited by the air

side heat transfer coefficient and surface area, an approximate doubling of

the heat transfer rate per pound was observed.

The binary mixture can be seen to go through approximately a 9°F change in the

condenser two-phase zone (where the pressure drop and the saturation temperature gradient reinforce) in comparison to an approximately 3°F phase temperature gradient due to pressure drop alone for R22 (figure 12). The pressure

gradient was not measurable with the high-side pressure gauges used on the

condenser. Hence horizontal bubble, dew, and saturation lines are shown in

figure 12. The 3°F saturation temperature difference observed for the R22

would correspond to a saturation pressure difference of approximately 7.5 psi.

It should be noted that compared to an evaporator, condenser pressure

gradients are less because the lower average void fraction of the condensing

refrigerant results in lower velocity and, consequently, lower pressure drop.

Also, for this mixture, the difference between the dew and bubble lines is

reduced at the high pressure of a condenser in comparison to evaporator

pressure levels. For a 65X/35X mixture of R13B1/R152a, this dew-bubble line

temperature difference is 1A.7°F at 250 psia and 23.5°F at 60 psia.

After it had been observed that the subcooling loss that was occurring

concurrently with composition shift was adversely affecting the performance of

36
the binary mixture, it was decided to perform comparative tests with

subcooling at the 17 °F heating test condition to evaluate the potential

performance obtainable by composition shifting. These tests were performed

with an adjustable needle valve in place of the interchangeable orifice

expansion devices used for the other reported tests because sufficiently small

orifices were not available. The results of these tests are shown in Table 6.

A 47 °F test with a high concentration of R13B1 is also shown in Table 6 for

the purpose of comparison to the data shown in Tables 3 and 4.

The high concentrations of R13B1 shown in Table 6 were obtained by charging

the unit with the desired mix of refrigerants rather than as the result of any

control strategy affecting a base mixture at 6 5% R13B1 as was the case in the

tests recorded in Table 3. The expansion device was adjusted for best

capacity which in all cases was at the point of marginal subcooling.
Table 6. Heating Tests with Rubcooling and Manually Adjustable Needle

Expansion Valve.

Outdoor Ambient Air Temp., °F 17.4 16.9 46 .4

Ref r igerant , (R22 or % R13B1

)

R22 85.9% 85.7%

Air Side Capacity, Btu/hr 21,500 25,46 0 39,130

COP (Total Sy stem) 1.98 2.00 2.41

Ref rigerant Mass Flow Rate, lb/hr 216 483 a

7 90

Compressor Suction, psia 42.9 61 .6 96 .5

Compr e s sor Suction, °F 20 1

1

33.8

Compr e s sor Discharge, psia 180 259 308

Compr e ssor Discharge, °F 212 178 172

Compr e ssor Power, W 27 1 2 b

3 2 53 b 4291

Indoor Fan Power, W 47 4

d 473 d 46 9

(a) Estimated from compressor suction vapor density.

(b) Includes outdoor fan power of approximately 340W.

(c) Includes outdoor fan power of approximately 320W.

(d) Includes control circuit power of approximately 40W.

38
It was observed during the low temperature heating tests that large amounts of

composition shift could not be obtained without serious loss of subcooling.

The simplest solution to this problem and the one tested here is to install a

receiver in the liquid line which will remain filled with refrigerant at the

circulating composition during both cooling and high temperature heating when

subcooling exists in the liquid line. During low temperature operation when

subcooling is lost such a receiver would trap bubbles, filling itself with

vapor and sending its liquid refrigerant to the accumulator where it would

effect a composition shift. This mechanism does require that subcooling be

lost, but only by a small amount which should result in a large composition

shift. Receiver emptying will not occur completely when subcooling is first

lost with the original circulating composition because the composition shift

to a greater amount of R13B1 will result in a maintaining of saturation

conditions in the accumulator and stable, steady-state operating condition

with a partly emptied receiver. This is because the shift to a greater

concentration of R13B1 increases vapor density more than liquid density thus

increasing compressor capacity relative to expansion device capacity. Liquid

level in such a receiver could also be controlled by level sensors operating

an electric expansion valve as described by Young [14] instead of the fixed

orifice system described in this report.

The actual receiver used in these tests was an inverted oil separator with the

oil return line blocked. An inverted oil separator was used instead of a

normal receiver so that both inlet and outlet would be on the bottom resulting

in insensitivity to direction of flow. The volume was 140 cubic inches which

would result in the filled receiver holding approximately 5.8 pounds of either

R22 or the 6 5% R13B1, 35% R152 binary mixture if completely filled. This
extra amount of refrigerant in the receiver during steady-state operations

resulted in a total system charge that would have overfilled the accumulator

during cyclic operation. Since cyclic testing was not part of this study, no

additional system modifications (e.g., larger accumulator or liquid line

solenoid) were employed to correct this.

Since a 14 month period had passed since performance of the R22 baseline

tests, R22 tests were rerun to reestablish the baseline for comparison to the

receiver installed tests. For all tests listed in Table 7, the unit was

charged in the cooling mode to 10°F superheat leaving the evaporator at the

95°F test condition. The cooling mode orifice used in charging with R22

was .0780 in I.D. The cooling orifice used for charging for the mixture tests

was .0820 in I.D. with the exception of the heating test performed with

the .0670 in I.D. orifice. This test was charged at 95°F using a .0807 in

I.D. orifice to insure that no flooding was occurring. The charging

composition was 65% Rl3Bl/35% Rl52a. As is shown in Table 7, the performance

with R22, particularly at low temperature, had degraded in comparison to the

performance listed in Table 5. This is attributed to wear resulting from

particularly abusive testing during the 14 month hiatu6 in the binary

refrigerant research program. The most abusive of these tests involved cyclic

operation with the accumulator removed which would result in severe oil

dilution by refrigerant and consequent cylinder wear. Cylinder wear would be

expected to be mo6t damaging to performance at the high compression ratios

associated with low temperature heating.

40
Table

7.

Summarized

Results

for

Binary

Mixture

Heating

Tests

with

Receiver

and

Comparable

Tests

with

R22

41

Includes

control

circuit

power

of

approximately

40W.
At the 47 °F operating condition (for the mixture tests), the receiver

prevented the unit from losing subcooling (indicated by the refrigerant

flovrate being listed as directly measured) by partially emptying into the

accumulator so as to enrich the circulating refrigerant with R13B1. The

smaller orifices resulted in severe performance loss at this temperature by

starving the evaporator. This loss with smaller orifices interferes with the

concept of choosing an orifice size such that most composition shift would

occur at a typical building balance point.

At 17 °F, composition shift greater than that which occurred without the

receiver was observed and this resulted in improved capacity in comparison to

R22.

4 2
6. DISCUSSION

The most highly publicized attribute of nonazeotropic mixtures, that of

variable temperature phase change and its associated minimal heat transfer

irreversibilities in a counter flow heat exchanger with a sensible heat sink

or source, seems to have rather limited application for direct working fluid

substitution in R22 air-to-air heat pump hardware. Test results in this

study confirmed our model predictions that the saturation temperature decrease

due to the refrigerant flow's static pressure decrease virtually offsets the

increase in composition shift temperature in the evaporator. Thus, the bulk

temperature profile of the mixture in the evaporator is in fact more constant

than with R22, which has only the saturation temperature decrease phenomenon

occurring. Of course, in the condenser the mixture's composition temperature

and saturation temperature shifts are additive and a rather significant

refrigerant phase change temperature decrease results. However, current outdoor coil designs are typically single row wraparound resulting in almost pure

crossflow heat exchange. Furthermore, the reversing action of the

heating/cooling heat pump confounds the design problem of the heat exchangers

to be effective with mixtures.

Mixtures should be evaluated on a reliability basis as well as energy

performance. For the vapor compression cycle this usually means consideration

of the discharge pressure and temperature levels as well as the pressure

difference and ratio. For a heat pump, the high-stress condition resulting

from high pressures and temperatures will occur during the highest condenser

operating conditions, which corresponds to the cooling mode.

43
Table 3 summarizes some system temperatures and pressures for the cooling

mode. The 6 5% / 3 5% mixture can be seen to produce suction pressures

substantially the same as R22. For this mixture the discharge pressure is

approximately 18 psia higher than R22 (and consequently the suction to

discharge pressure ratios are different). The pressure ratio was, on average,

7 percent higher than R22. The discharge temperature was, on average, 19°F

lower than with R22. Thus, when using a superheat leaving the evaporator as a

charging criterion, this mixture does not produce pressures significantly

higher than R22 and its lower discharge temperatures would likely more than

compensate for the slightly higher pressures in affecting system life.

The energy performance of the 6 5% / 3 5% (R13B1/R152a) mixture, charged in the

cooling mode to give a similar superheat as R22, was evaluated. The system

operated such that the accumulator was dry throughout the cooling mode which

meant no shift in composition was occuring. Cooling efficiency was 10Z worse

than R22 at the 95°F test condition and 11Z worse than R22 at the 82 °F test

condition.

In the heating mode, it was observed that flooding into the accumulator caused

an increase in the mass concentration of R13B1 in the circulating refrigerant

of up to 11 percentage points with the largest orifice used. As a part of

normal fixed orifice system operation, the subcooling was decreased with each

increase in orifice size until completely lost thus offsetting this potential

gain in capacity and efficiency at lower temperatures. A possible solution

for this loss in subcooling is an additional volume (e.g. , a receiver) on the

high pressure side which could transfer refrigerant to the accumulator without

starving the condenser.

44
In spite of the above control problem, which reduced the performance of the

binary refrigerant below expectations, a substantial improvement in low

temperature capacity in comparison to R22 was observed. Without the receiver

the binary refrigerant exhibited 6.3% more capacity at the 17 °F rating point

than R22. With the receiver, the capacity was 14.2% better than R22. In both

cases this improvement was the result of composition shift brought about by

the normal control by the expansion device (orifice). An additional

comparison with slight subcooling leaving the condenser and the binary

refrigerant artifically adjusted to give an 86% concentration of R13B1 (in

comparison to 73% without the receiver and 76% with it installed) resulted in

an 18.4% capacity improvement over R22.

The performance of the mixture (i.e., capacity and COP) is shown in figure 13

relative to R22. All of the tests had an initial charge composition of 6 5%

R13Bl/35% R152a, which was the operating composition at 47 °F. The composition

shifted to 76% Rl3Bl/24% R152a at 17 °F. The performance results with the

receiver as shown in figure 13 have been increased before plotting by the

amount their comparable R22 tests had degraded during the previously mentioned

interuption in the test program. This composition shift caused less

reduction in low temperature performance than occurred with R22 at low

temperatures; that is, as the evaporator temperature is reduced, there is more

available mixture capacity relative to R22. The empirical results also show

a relative improvement in mixture efficiency performance in comparison to

4 5

R22.
40

36

32

28

24

20

3.0

2.6

2.2

1.8

85% R13B1

87.5% R13B1

76% R13B1 ' 65% R13B1

65%

R13B1

(WNReglon of likely

operation for

R13B1/R152a

binary

A R22

O Binary • Computer

simulation

V Binary with

receiver

10 20 30 40 50

OUTDOOR TEMPERATURE (°F)

i.8on of binary refrigerant energy performance to R22.

46
The improvement in comparison to R22 is understandable when the basic

properties of the refrigerants are considered. A review of the p-h diagram

[11] for the mixture indicates that while the latent heat decreases with

increasing percentage of R13B1, the vapor density of the mixture increases

enough to more than compensate for this effect due to the higher vapor

pressure of R13B1 relative to Rl52a. Thus, an increase in R13B1 should result

in an increase in the product of vapor density times latent enthalpy to which

capacity is approximately proportional for a compressor with a constant

volumetric flow rate (i.e., positive displacement machines). Also, as the

outdoor temperature decreases from 47 °F to 17 °F, the suction pressure does not

have to decrease as much to attain the necessary evaporator temperature

because of the composition increase in R13B1. The discharge pressure also

increases with shifting composition but at a slower rate than the suction

pressure so that the (P^/Pg) term, of which work is a function, does not

increase as much as does capacity when composition shift occurs; thus

efficiency also begins to increase relative to the value which would occur

without composition shifting.

Also illustrated on figure 13 are the results of the NBS mixtures/heat pump

simulation model L 1 2] . The crosshatched band represents the region of likely

operation for this mixture. The capacity region is bounded on the bottom by a

65% R13Bl/35% Rl52a circulating composition with a 0°F superheat and a 2°F

subcooled condition for both the 47°F and 17°F. The top bound is an 85%

R13B1/15% Rl52a composition with the same operating conditions. These bounds

represent the optimum capacity this heat pump could produce with the

circulating compositions shown. In order to proceed through this region on

any path it is necessary to have full control of the expansion device, the

47
composition and the charge. For example, the path between the 85% R13B1 point

at 47 W F and the 87.5% R13B1 point at 17 °F represents the fixed orifice and

charge case; that is, the current heat pump operation. In this case, these

parameters (i.e., expansion device and composition) are not fully controlled

resulting in part load performance degraded from the optimum. For this case,

there is an associated loss of subcooling when liquid storage occurs in the

accumulator, which might well be mitigated with an increase in initial charge.

In summary then, there appears to be an inherent trade-off when operating an

unmodified heat pump with R13B1/Rl52a mixtures so as to increase capacity at

lower temperatures. The trade-off is that as the evaporator pressure

decreases and begins to flood through to the accumulator and allows the

composition shift to occur, the subcooling on the high pressure side also

decreases and causes a decrease in the available latent enthalpy difference.

A possible solution to this problem might be to increase the amount of initial

charge. This would cause an increase in head pressure, which could be a

problem in the cooling mode. Another possible solution is to incorporate a

receiver xn the system in the liquid line between the expansion devices.

The primary areas for future work with binary mixtures are the investigation

of additional mixtures, consideration of the changes necessary for conventional

heat pumps to take best advantage of the characteristics of mixtures, and,

finally, the possibility of more advanced heat pump cycles. Other refrigerant

mixtures that have been frequently suggested are R22/R114, R12/R23 and

R22/R23. Equipment changes might include water-to^air and water-to-water heat

pumps which would allow more ideal counter flow heat exchange. A control

scheme monitoring refrigerant level in a receiver as described in L14J may be

48
useful for air-to-air applications. Since there is interrelation between

application, refrigerant properties, and equipment design some parallelism

between these different paths should occur. In fact the uncertainties in

predicting mixture properties at this time require that, after an initial

screening has been made based on predicted properties, final evaluation must

be made in a realistically constructed machine operating at conditions of

application.

49
7 . CONCLUSIONS

The following conclusions can be drawn from this experimental study:

0 Substitution of this particular mixture in an unmodified heat pump,

designed for R22 operation, resulted in improved low temperature, 17 °F,

heating performance (up 6% in capacity, no change in efficiency) at the

expense of reduced high temperature heating (down 3% in efficiency) and

cooling (down 12% in capacity, 11% in efficiency) performance.

° The low temperature performance was further improved in comparison to

R22 by fitting a receiver (up 14% in capacity, 2% in efficiency) at the

17 °F rating point.

0 The expansion device size should be changed when the refrigerant mixture is

changed for one of different properties (liquid vs. vapor density)*, so as

to ensure that proper subcooling will be maintained.

0 Efficiency improvement from the gliding temperature effect (Lorenz vs.

Carnot cycle) is unlikely in an air-to-air heat pimp primarily because of

the cross flow design of most current heat exchangers. The gliding

temperature effect can be made successful in water-to-water applications.

0 Capacity modulation through composition shift is a successful control

strategy that will increase seasonal performance in northern applications.

° No potential reliability or safety problems as a result of substitution of

the mixture for R22 were observed.

50
° Future work should be directed toward equipment redesign to better use the

characteristics of nonazeotropic mixtures and to analysis of the

performance characteristics and potential of other mixtures than the one

tested for this report.

51
8 . REFERENCES

1. Cooper, W. D. , and Borchardt, H. J. , "The Use of Refrigerant Mixtures in

Air-to-Air Heat Pumps," XVth International Congress of Refrigeration,

Proceedings, Vol. IV, 1979.

2. Cooper, W. D. , "The Use of Mixed Refrigerants in Air-to-Air Heat Pumps,"

ASHRAE Transactions, V. 88, Pt. 1, American Society of Heating,

Refrigerating and Air Conditioning Engineers, Inc. , Atlanta, Georgia,

1982.

3. Freon Products Division, Freon* 13Bl/152a Refrigerant Mixture for Heat

Pumps, RT-74, E. I. DuPont de Nemours and Co., Inc., Wilmington, Delaware.

4. Carr, F. , "Power Savings in Process Refrigeration," Industrial and

Engineering Chemistry, 41, 776-7 80, 1949.

5. Kazachki, G.S., "Calculation of Two Branch Cycles with Nonazeotropic

Refrigerant," Proceedings of Meeting of Commission Bl , B2, El, E2,

International Institute of Refrigeration (pp. 153-157), Mons (Belgium),

1980.

6. Schwind, H. , "Application of Binary Refrigerant Mixtures and their

Presentation in the Enthalpy—Pressure Diagram," Kaltetechnik , 14, 98-

105, 196 2.

7. "ASHRAE Handbook of Fundamentals", Chapter 16, "Refrigerants," American

Society of Heating, Refrigerating and Air Conditioning Engineers, Inc.

,

Atlanta, Georgia, 1985.

8. Domanski, P. and Didion, D. , "Computer Modeling of the Vapor Compression

Cycle with Constant Flow Area Expansion Device," NBS BSS 155, National

Bureau of Standards, Gaithersburg, Maryland, 1983.

9. 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, Inc., Atlanta, Georgia, 1977.

10. "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, Inc., Atlanta, Georgia,

196 9.

11. Domanski, P. and Didion, D. , "Equation of State Based Thermodynamic Chart*

for Non-Azeotropic Refrigerant Mixtures," ASHRAE Transactions, V. 91, Pt.

1, American Society of Heating, Refrigerating and Air Conditioning

Engineers, Inc., Atlanta, Georgia, (1985).

12. Domanski, P. , "Modeling of a Heat Pump Charged with a Non-Azeotropic

Mixture Refrigerant," NBS Technical Note 1218, National Bureau of Standards,

Gaithersburg, Maryland, 1986 .

52
13. Federal Register, Part III, "Test Procedures for Central Air Conditioners,

Including Heat Pumps," Vol. 44, No. 249, p. 76708, December 27 , 1979.

14. Young, David J. , "Development of a Nothern Climate Residential Air-Source

Heat Pump," ASHRAE Transactions, Vol. 86, Part 1, American Society of

Heating, Refrigerating, and Air Conditioning Engineers, Inc. , Atlanta,

Georgia, 1980.

15. Bergles, A.E., Collier, J.G. , Delhaye, J.M. , Hewitt, G.F., Mayinger, F. , "Two-Phase Flow and Heat Transfer in the Power and Process Industries,"

Hemisphere Publishing Corporation, McGraw-Hill Book Company, 1981.

53
APPENDIX. OBSERVATIONS ON EXPERIMENTAL TECHNIQUE

This section discusses the experimental techniques that were used for

refrigerant sampling for composition analysis, temperature measurement with

immersed thermocouples, and two-phase refrigerant mass flow rate estimation.

Two methods of refrigerant sampling were employed, liquid and vapor sampling.

Liquid sampling is appropriate for determining the contents of a cylinder or

for sampling from a subcooled liquid line of an operating heat pump. Liquid

sampling was done with a two chamber bottle consisting of a small chamber to

be filled with subcooled liquid and a large chamber separated by a valve to

allow expansion of the liquid into room temperature superheated vapor for

introduction into a gas chromatograph. This technique requires some knowledge

of the mixture composition so that these volumes may be correctly chosen to

allow complete expansion into superheated vapor. A pressure gage was found

useful to check for superheat after expansion. In some cases, if expansion

results in saturation it may be possible to revaporize the sample to a superheated vapor state in a hot water bath and then bleed the sample down to a

mass that will result in superheated vapor at room temperature. When

connecting the bottle to the sample ports, refrigerant should not be blown to

clear the line of air as this will cause composition shift in the line from

which the sample is to be drawn. Inclusion of air does not affect the ratio

of the two refrigerants as measured by the gas chromatograph.

When taking a vapor sample, vapor from a superheated line in the heat pump was

bled into the sample bottle until a pressure was reached that was adequate for

injecting the sample into the chromatograph but wa6 well below saturation.

54
In both sampling techniques, oil should not be allowed to enter the bottle

since the different refrigerants will absorb into the oil in proportions

different from their circulating composition and will be rapidly released from

the oil in the low pressure of the sample bottle. Sampling tees, therefore,

should point upward, to avoid trapping oil. If a downward facing fitting must

be used for vapor sampling, vapor should be blown shortly before sampling to

clear the line.

As was discussed in Section 4, an attempt was made to read refrigerant

temperatures with immersed, through-the-wall thermocouples sealed into

hypodermic needles with epoxy. Epoxy failures prevented this from being a

technique worth recommendation.

For cooling tests, the majority of the data points at each ambient temperature

had subcooling and, thus, a valid measurement of refrigerant mass flow rate

using the turbine meter. This was used to calculate the ratio of mass flow to

vapor density at the compressor inlet which ratio was then used to estimate

the mass flow for the remaining points at that same test condition.

The refrigerant mass flow for heating tests in which two-phase flow existed at

the flowmeter was calculated by assuming that the flowmeter correctly measured

volumetric flow and using the air side capacity to calculate the quality and

mass flow. The equation for R22 tests for calculating quality was:

Qas = V(hin ” ^ xllv

+ (1 “ x)h£))/(xv

v

+ (1 - x)v^) + QF

where: Q^g = air side capacity, Btu/hr
Qf

= fan power, Btu/hr

^in = suPer^®ated vapor enthalpy entering condenser, Btu/lb

hy = saturated vapor enthalpy leaving condenser, Btu/lb

h^ = saturated liquid enthalpy leaving condenser, Btu/lb

vv

= saturated vapor specific volume leaving condenser, cf/lb

v^ = saturated liquid specific volume leaving condenser, cf/lb

V = volumetric flowrate, cf/h

x = quality, mass % of vapor

and for calculating mass flow was:

Mass flow = V/ (xvv

+ (1 - x)V£)

For the binary tests a similar calculation was performed with the exception

that the computerized properties program returned two phase properties giving:

^AS

= ^^in “ +

Q*F anc* mass fl°w = v

*/ v

20 which were solved by

iteration.

The results given by this method were, for all reported tests, within 10% of

(and higher than) those given by ignoring the capacity contribution of the

refrigerant vapor and calculating the refrigerant mas6 flow as:

m = (<^AS " ^F)/(hin

“

A detailed discussion on the use of turbine meters in two phase flow is given

in L15J .

56
NBS-114A (REV. 2-80

U.S. DEPT. OF COMM.

BIBLIOGRAPHIC DATA

SHEET (See instructions)

1. PUBLICATION OR

REPORT NO.

NBSIR-86/3422

2. Performing Organ. Report No 3. Publication Date

JANUARY 1987

4. TITLE AND SUBTITLE

The Performance of a Conventional Residential Sized Heat Pump Operating with a Nonazeotropic Binary Refrigerant Mixture

5. AUTHOR(S)

William Mulroy, and David Didion

6. PERFORMING ORGANIZATION (If joint or other than N BS. see instructions)

NATIONAL bureau of standards

DEPARTMENT OF COMMERCE

WASHINGTON, D.C. 20234

7. Contract/Grant No.

8 . Type of Report & Period Covered

9.

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

Office of Building and Community Systems

U.S. Department of Energy

(via Oak Ridge National Laboratory)

1000 Independence Avenue, SW

Washington,. DC 20585

10.

SUPPLEMENTARY NOTES

U 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)

This report presents laboratory performance measurements of a relatively unmodified

residential heat pump designed for R22 when charged with a nonazeotropic, binary

mixture of R13B1 and Rl52a. Results are presented for various sizes of fixed

expansion devices. The effect of gliding temperature within the saturation zone

was found to be small. This experimental investigation confirmed that flash

distillation within the accumulator would improve low temperature heating performance. The measured performance was approximately 11% lower in both efficiency and

capacity than R22 for air conditioning. The high temperature heating efficiency was

3% lower than R22. The low temperature heating capacity was 14% higher and efficiency

2% higher than R22. These results show a substantial improvement over R22 for heating

applications at the expense of reduced cooling mode performance. Further performance

enhancement for this or other mixtures is expected through various system modifications

which remain to be studied.

12. KEY WORDS (Six to twelve entries; alphabetical order; capitalize only proper names; and separate key words by semicolon s) air conditioning; heat pumps; nonazeotropic refrigerants; refrigerant mixtures;

refrigerants; refrigeration

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