NBS

Publications

Reference

AlllDb SbODfiD

81-2287

Laboratory Tests of A Residential

Unitary Water-Source Heat Pump

U S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Washington, DC 20234

November 1982

i

Sponsored by:

Department of Energy

^

hington, DC 20585

. U56

81-2237

ldc>2

NBSIR 81-2287

LABORATORY TESTS OF A RESIDENTIAL

UNITARY WATER-SOURCE HEAT PUMP

, ATIONAL BOTEA'T OF stand;.- 1 LIBRARY DEC 1 0 1982

Oof a c-c . - VCef

Gd \OG

21 - ^3^7

1 ^

9^

William J. Mulroy

George E. Kelly

U.S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Washington, DC 20234

November 1982

Sponsored by:

U.S. Department of Energy

Washington, DC 20585

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

NATIONAL BUREAU OF STANDARDS, Ernest Ambler, Director
: .Hr H

•

-

‘ r .

'. / .

r ., Hi -

;i:- *j3*

-

, . .....

.

. - > i

-

'' . . ' ...

'
NBSIR 81-2287

LABORATORY TESTS OF A RESIDENTIAL

UNITARY WATER-SOURCE HEAT PUMP

*

William J. Mulroy

George E. Kelly

U S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Washington, DC 20234

November 1982

Sponsored by:

U.S. Department of Energy

Washington, DC 20585

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

NATIONAL BUREAU OF STANDARDS, Ernest Ambler, Director
• ' ,

'

.

;

, • . .

. V
Abstract

The performance of a residential heat pump was measured in the laboratory over

a broad range of source water temperatures (40*F to 90°F). Tests were performed

in both heating and cooling operational modes and for both steady-state and

cyclic operation.

For both heating and cooling operation, the unit capacity and COP were found

to be linear functions of the average of the unit source and outlet water

temperatures. In heating, the unit capacity, COP, and part load performance

increased with increasing water temperature. In cooling, the unit capacity, COP,

and part load performance decreased with increasing water temperature. The

measured part load degradation coefficients ranged from 0.09 to 0.21 for heating

and from 0.10 to 0.18 for cooling. An appendix is included in which the effect

of the degradation coefficient and of supplemental resistance heat on the unit

heating and cooling seasonal performance factors is calculated.

Key Words: Central heating equipment; cooling; heat pumps; heating; heating seasonal performance; heating seasonal performance factor* test method; water source heat pumps.

iii
Table of Contents

Page

Abstract ill

List of Figures v

List of Tables vi

SI Conversion Factors vii

1. INTRODUCTION 1

2. DESCRIPTION OF TEST SPECIMEN 2

3. DESCRIPTION OF TEST APPARATUS 3

4. TEST PROCEDURE 5

5. STEADY-STATE TEST RESULTS 8

5.1 Steady-State Heating Test Results 8

5.2 Steady-State Cooling Test Results 8

6. CYCLIC TEST RESULTS 9

7. ACCURACY OF RESULTS 10

8. DISCUSSION AND CONCLUSIONS 12

9. REFERENCES 14

APPENDIX A - SEASONAL PERFORMANCE CALCULATIONS

iv
List of Figures

’ Page

Figure 1. Heat pump installed in test apparatus 15

Figure 2. Heat pump installed in test apparatus 16

Figure 3. Air flow measurement tunnel 17

Figure 4. Variation of heating capacity with entering water 20

temperature for different inlet-to-outlet water temperature

differentials

Figure 5. Variation of heating capacity with mean water temperature 21

Figure 6. Variation of heating COP with mean water temperature 22

Figure 7. Variation of cooling capacity with mean water temperature 25

Figure 8. Variation of cooling COP with mean water temperature 26

Figure 9. High temperature heating cyclic tests 27

Figure 10. Low temperature heating cyclic tests 28

Figure 11. High temperature cooling cyclic tests 29

Figure 12. Low temperature cooling cyclic tests 30

Figure 13. Linear regression fit to data on degradation of the 31

coefficient of performance at part-load operation

Figure 14. Variation of coefficient of degradation with mean water 32

temperature

v
List of Tables

Page

Table 1. Steady-State Heating Test Results 18

Table 2. Steady-State Heating Test Conditions 19

Table 3. Steady-State Cooling Test Results 23

Table 4. Steady-State Cooling Test Conditions 24

Table Al. Design Conditions and Calculated Ifciit Performance for A9

Selected Cities

Table A2 . Comparison of Seasonal Performance Factors to Steady- A10

State Coefficients of Performance for the Unit Sized

at the Minimum Design Heating and Cooling Loads

Table A3. Comparison of Cooling Seasonal Performance Factor to All

Steady-State Coefficients of Performance for the Unit

Sized at the Maximum Design Cooling Loads

Table A4. Comparison of Heating Seasonal Performance Factors to A12

Steady-State Coefficients of Performance for the Unit

Sized at the Maximum Design Heating Load

Table A5. Sample Bin Method Calculation, Region IV, Maximum A13

Design Load

vi
SI Conversion Factors

Multiply *L To Obtain

Btu, IT 1.055 kJ

Btu/h, Btuh 0.293 W

Btu/lb°F, c, specific heat 4.19 kJ/kg°(

°F °C = ( °F-32)/l .8

ft 0.3048 m

ft/min, fpm 0.00508 m/s

f

t

3 /lb 0.0623 m

3 /kg

ft/min, cfm 0.472 1/ s

gpm (US) 0.0631 1/

s

inch 25.4 mm

inch of water 3.38 kPa

kBtu 1055 kJ

lb/h 0.126 g/

s

vii
'
1 . INTRODUCTION

Traditionally, residential water-source unitary heat pumps have been applied

primarily in those areas of the Southeastern United States which have abundant,

high temperature ground water. This was reflected in ARI Standard 320-76, which

specified a temperature of 70°F (21.1*0) for water entering the ref rigerant-towater heat exchanger as the heating rating point.

In the search for more efficient methods of residential heating, it has become

apparent that the water-source heat pump could provide significant energy

savings. Some of the new applications that expand the use of this device beyond

its traditional role and require new methods of rating are:

1) Heating with low temperature ground water - Although water-source heat

pumps have traditionally been employed in areas of the country where the

ground water is comparatively warm, when they are designed to operate on relatively cold (down to 40°F (4.4°C)) ground water their geographic range of

application can be expanded considerably. With proper design, a ground-watersource heat pump should operate with a higher seasonal efficiency than

an air-to-air heat pump in the same climate*

2) Solar-assisted heat pump operation - In order to transfer heat

effectively, the fluid temperature of a solar collector should be approximately 30 # F (-1.1°C) warmer than the indoor air temperature of the home which it

is heating. Under this condition, a solar collector would be ineffective

when its fluid temperature dropped below 100°F (70°F + 30°F). Solar heated

fluid at temperatures below 100°F (37.8°C), although unsuitable for direct heat

transfer space heating would, however, be a suitable source of energy for a water-source heat pump.

3) Intrastructure heat transfer - Frequently large commercial structures

have interior zones which require year-round cooling due to high internal

loads (e.g. lights, people, computers, equipment, etc.). A water-to-air heat

pump is ideal for this application because the energy extracted in the process

of cooling the interior zones can be used to heat exterior zones during the

heating season. Similar residential applications may be possible.

The primary purpose of the tests reported herein is to provide sufficient

data on the performance of a typical residential unitary water-source heat

pump to enable development of new test and rating methods. These methods

should provide sufficient information on water-source heat pump performance

and cost of operation so that consumers can purchase cost-effective energy

conserving heat-pump systems. It is important in the creation of test

and rating methods to develop procedures to measure part-load performance.

Including part load performance in calculation of heat pump system seasonal

energy usage allows credit to be given to innovations in heat pump design

and control which improve part load as well as full load, steady-state

efficiency

.

1
2. DESCRIPTION OF TEST SPECIMEN

The heat pump tested was of commercially available unitary design with an air

handling compartment above another compartment containing the compressor and the

water-to-refrigerant heat exchanger. The two compartments were separated by an insulated panel. The manufacturer’s rated heating and cooling capacities for the

unit were 72,000 Btuh (21.1 kw) and 60,000 Btuh (17.6 kw) , respectively.

The air handling system was rated to deliver 2200 cfm (1.04 m^/5) at 0.15 inches

of water C37 Pa) external static pressure. The blower was belt-driven by a 3/4

horse-power (560W) motor equipped with an adjustable pulley. Included in the

air handling system were automatically controlled dampers designed to restrict

air flow and increase refrigerant head pressure during low temperature startup

in the heating mode. Manual control was substituted for automatic control of

these dampers for the tests discussed in this report.

Refrigerant flow in both heating and cooling modes was controlled by automatic

expansion valves. The use of automatic (instead of thermostatic) expansion

valves made adjustment necessary between tests. The expansion valves were

manually adjusted to provide a nominal 15°F (8.3°C) of superheat for all tests.

The refrigerant charge recommended by the manufacturer was 14 pounds (6.4 kg)

of refrigerant 22. The unit was tested as charged by the manufacturer.

The water-to-refrigerant heat exchanger was of co-axial tube-in-tube design.

Included in the water circuit were valves designed to modulate water flow to

maintain constant refrigerant conditions in the heat exchanger during both

heating and cooling mode operation. These valves were completely opened for

the test series described herein and the water flow was controlled manually.

The heat exchanger was not equipped with a water-circulating pump and, in these

laboratory tests, water supply line pressure was used to move water through

the water-to-refrigerant heat exchanger.

The heat pump was equipped with a separate ref rigerant-to-water heat

exchanger on the compressor discharge for the purpose of supplying domestic

hot water. This heat exchanger was disabled by draining and disconnecting

its integral pump prior to the start of the test series.

2
3. DESCRIPTION OF TEST APPARATUS

The test apparatus was built in conformance with ASHRAE Standard 37-78

using the Air-Enthalpy Method, Indoor Side, as the primary test method, and

the Outdoor Water Coil Method as the secondary test method. The unit with

the test duct installed is shown in figures 1, 2 and 3.

Air flow rates were determined by measuring the pressure drop across a 10-inch

(254 mm) diameter nozzle. The pressure drop across the nozzle was measured

using four interconnected upstream pressure taps at the nozzle entrance and

four downstream taps at the exit which were centered laterally on each side of

the air flow measurement tunnel. The upstream pressure taps were located

approximately 9 inches (203 mm) before the nozzle inlet and the downstream taps

were set even with the nozzle outlet. A pitot tube was installed, as a check,

centered 1 inch (25 mm) above the nozzle outlet. Figure 3 is a schematic of

the air flow measuring apparatus.

Average air temperatures entering and leaving the unit and at the nozzle

entrance were measured using two sets of 16 thermocouples connected in parallel. The temperature difference between the air entering the unit and the

air entering the nozzle was measured by a 32-junction (16 junctions on each

side) thermopile. An additional 32-junction thermopile was used to measure

the difference between the air temperatures entering and leaving the unit.

Wet-bulb temperatures of air entering the unit and the nozzle were measured

by single thermocouples covered with wetted wicks. Single dry-bulb thermocouples near the wet-bulb thermocouples were used to allow calculation of

humidity.

The water flow rate was measured by a turbine-type flowmeter. A second

turbine flowmeter was installed in series as a backup but was not used. In

addition to measuring flow rates, an integrating recorder was used to measure

integrated water flow.

The inlet and outlet water temperatures were measured by single thermocouples

in wells installed at the unit inlet and outlet fittings. Two 32-junction

thermopiles were installed in these same wells to measure the water

inlet-to-outlet temperature difference.

Thermocouples were attached to the unit refrigerant lines entering and leaving

the water-to-refrigerant and air-to-refrigerant heat exchangers.

Pressure gages were installed to measure the suction and discharge pressures

at the compressor.

The unit's total power and fan power were each recorded separately from

watt-hour meters having a 10 watt-hour resolution.

Temperature and millivolt signals were recorded by an automatic data

acquisition system having the capability of a programmable time interval

between data scans.

3
During the cyclic tests, an electronic integrator was used to record the

air-side thermopile signal for the period between compressor-on and

compressor-off. Identical instrumentation was used for recording the signal

from the water-side thermopile.

Condensate was collected for all cooling tests. The amount of condensate

collected was measured by weighing on a balance which had a resolution of

0.02 pounds (0.01 kg). For the steady-state tests, condensate collection was

begun at the beginning of the data period and terminated at its end. In the

cyclic tests, the condensate collection was allowed to continue 3 to 10 minutes

after shutdown to insure complete drainage from the coil. The collected con- densate was used to calculate the latent portion of the total cooling capacity.

Air flow, integrated water flow, barometric pressure, suction and discharge

pressures, power consumption, and condensate collection data were taken

manually

.

During all tests, crucial temperature and millivolt values were continuously

monitored by a 24-point printing recorder in addition to the data acquisition

system.

A range of inlet water temperatures from 40°F (4.4°C) to 90°F (32.2°C) was provided by tapping into chilled water or service water lines in the building

where tests were conducted and mixing the water from the lines.

Dry-bulb and wet-bulb air temperatures entering the unit were controlled by

the test room air conditioning equipment. In order to maintain conditions

additional heat and humidity were required, and were supplied by a gas furnace

and by portable room humidifiers.

3 3

The unit air flow rate ranged from 1960 (0.92 m /s) to 2060^(0.97 m /s) standard

cfm during steady-state heating tests and from 2085 (0.98 m /s) to 2100 (0.99m /s)

standard cfm during steady-state cooling tests. The air flow was not manually

adjusted between tests, but allowed to vary as a function of air density. The

minimum and maximum steady-state air flow rates recorded resulted in air velocities through the 10-inch nozzle of 3770 fpm (19.2 m/s) and 4020 fpm (20.4 m/s),

respectively. The above velocities are within the range of 3000 to 7000 fpm

(15.2 to 35.6 m/s) specified by ASHRAE Standard 37-78.

4
4. TEST PROCEDURE

For steady-state tests, the unit was allowed to run for 1 hour at steady

conditions prior to taking data. Data were recorded in 10 minute intervals for

a period of 1 hour. For all tests the refrigerant expansion valve of the unit was set to produce approximately 15°F (8.3°C) of superheat in the refrigerant

vapor leaving the evaporator.

Cyclic tests were performed at two sets of cycling rates.

These thermostat cycling rate sets were calculated by rounding values

calculated from the equation:

C = 4NX(1-X) (1)

where C = the cycling rate in cycles per hour.

N = the cycling rate at 50% on-time and was set equal to

either 2 or 5 for this study.

X = fractional on-time of the unit.

As an example, for the fast set of cycling times employed in this study, it was assumed that the unit would cycle five times per hour at 50 percent ontime. Substituting an on-time of 20 percent in the above equation results in

C = 4 x 5 x 0.2(0. 8) = 3.2 cycles per hour or 60/3.2 = 18.8 minutes per cycle

with an on-time of 0.2 x 18.8 = 3.8 minutes. This value (3.8 minutes) was rounded to 4 minutes to simplify performance of the cyclic test. The other

cycling times for the fast cycling rate set employed to determine the maximum

cycling effect were 4 minutes on/15 minutes off, six on/six off and 15 on/four

off.

A somewhat slower set of cyclic rates, corresponding to two cycles /hour at

50 percent on-time, consisting of nine on/38 off, 15 on/15 off, 38 on/nine

off, was used to find the system performance under what was felt to be more

normal cycling conditions.

The data acquisition system was used to record data at intervals of one to

three minutes during cyclic testing, depending on cycle length.

During cyclic tests, air flow rate and unit suction and discharge pressure

measurements were read once during each on-period and were used with the

temperatures existing at the time the reading was taken to determine the

air flow rate and refrigerant superheat. These measurements were recorded

approximately at the midpoint of the on-period.

For cyclic tests, integrated quantities (watthours, water flow, thermopile

millivolts, time) were read when the compressor started or stopped.

For cyclic tests, the air dampers on the unit were shut during the off-period

to reduce heat transfer by convection. Off-period convection flow was felt

to be an installation characteristic as opposed to a unit characteristic.

Water was prevented from circulating through the unit during the off-period

of the cyclic tests by diverting the flow through a bypass to maintain a constant water temperature.

5
In addition to the steady-state and cyclic operating tests, a heat loss

test was performed to measure the thermal loss factor of the compressor compartment. This heat loss test was performed by heating the compartment to a steady-state temperature with a fixed electric resistance heater and dividing

the heater power input by the resulting steady -state compartment temperature

less the room air temperature. This loss factor, in conjunction with com- partment interior and ambient temperatures, was used to determine the unit

jacket loss from this compartment. Determining the jacket losses allowed

a more accurate comparison of results calculated using the air-side and

water-side measured data.

The air side heating capacity during a test was calculated from the

equation

( 2 )

the air side cooling capacity was calculated from the equation:

^tci ^sci +^lci (3)

(4)

q,H . = 1060 W

lci c (5)

the water side heating capacity was calculated from the equation:

( 6 )

and the water side cooling capacity was calculated by the equation:

(7)

In the above equations, the following nomenclature has been employed:

C

pa = specific heat of air, Btu per (lb)(°F) of dry air

C

pw

= specific heat of water, Btu per (lb)(°F)

E = power input, total watts

Qmi = air flow, indoor, measured, cfm

q lci = latent cooling capacity, (indoor side data), Btuh

6
q sci

q tci

q tco

q thi

q tho

fc a

fc al

t a2

fc cc

fc w3

t w4

UA

cc

v'nW

W

W

w

sensible cooling capacity, (indoor side data), Btuh

total cooling capacity, (indoor side data), Btuh

total cooling capacity, (outdoor side data) Btuh

total heating capacity, (indoor side data), Btuh

total heating capacity, (outdoor side data), Btuh

temperature, ambient, °F

temperature, air entering indoor side, dry bulb, °F

temperature, air leaving indoor side, dry bulb, °F

temperature inside compressor compartment, °F

temperature, water entering outdoor side, °F

temperature, water leaving outdoor side, °F

thermal transmittance of compressor compartment, Btu/hr °F

specific volume of air at nozzle, cu-ft per lb of air-water mixture

flow rate, indoor coil condensate, lb per hr

humidity ratio, at nozzle, lbs moisture per lb dry air

flow rate, water, lb per hr.

7
5. STEADY-STATE TEST RESULTS

5.1 STEADY-STATE HEATING TEST RESULTS

The steady-state tests were run at conditions chosen to generate a set of

heating capacity and COP curves as a function of entering water temperature

for inlet-to-outlet water temperature differences of 10°F, 15°F and 20°F

(5.6°C, 8.3°C and 11.1°C). For the steady-state tests, the entering air temperature was held constant at 70°F (21.1°C) and the evaporator superheat at

15°F (8. 3°C) .

The results and test conditions for these tests are listed in tables 1 and 2.

The temperature limit on the range of water temperatures used in the tests was set by ice formation in the water-to-refrigerant heat exchanger which

occurred when the leaving water temperature was below approximately 35°F (1.7°C)

Although the unit was provided with reverse cycle defrosting this feature was

not used since its operation would not have produced results comparable to the

results of the tests at the higher water temperatures. At the highest inlet

water temperatures used in this test series (90°F, 32.2°C), the unit capacity

was limited by the size of the expansion valve, since at or above this temperature the evaporator superheat exceeded 15°F (8.3°C) with the valve fully opened.

The three curves of capacity as a function of inlet water temperature are shown in figure 4. In figure 5, the same capacity variation data are plotted

against the average of the inlet and outlet water temperatures (mean water

temperature). Use of the mean water temperature instead of the entering water

temperature as the independent variable results in the three inlet-to-outlet

temperature differential curves becoming superimposed, eliminating temperature

differential as an independent variable.

In figure 6, the heating coefficient of performance may also be observed

to be a linear function of mean water temperature.

5.2 STEADY-STATE COOLING TESTS

Steady-state cooling tests were performed with a range of inlet water

temperatures from 42°F to 95°F (5.6°C to 35.0°C) and a range of inlet-to-outlet

temperatures differentials from 10°F to 40°F (5.6°C to 22.2°C).

The data and results for these tests are summarized in tables 3 and 4 and

figures 7 and 8.

Although data were taken that would allow plotting of capacity and COP as functions of temperature differential at different entering water temperatures,

it was found that these cooling data fell along a single line when plotted

against mean water temperature, as had the heating data. Thus the results in

figures 7 and 8 are shown as a function of the mean water temperature.

8
6. CYCLIC TEST RESULTS

Cyclic tests were performed at high and low entering water temperatures in both

heating and cooling modes.

Cyclic test results are shown in figures 9 through 14. Each data point plotted

is the mean of the values measured during two to four individual on/off cycles.

The dashed lines show one standard deviation away from this plotted mean. All

reported values are based upon air side data.

The percent full load COP is defined as the ratio of the COP when cycling to

the steady-state COP at the same ambient conditions. The heating (cooling) load

is defined as the ratio of the cyclic heating (cooling) done over a complete onoff cycle to the steady-state heating (cooling) that would have been done if the

unit had remained on over the entire cycle which makes L a percent rather than

a "load". This may be expressed in algebraic form as follows:

L = Qcyc

Qss(ton + toff) (8)

where L=heating or cooling load

Qcyc=heating (cooling) done over a cycle

Qss=steady-state heating (cooling) capacity

ton=on-time

tof f=of f-time

The degradation coefficient, defined as the slope of a line drawn through

the 20 percent and 100 percent load points of the part-load coefficient of

performance degradation curve (see references 1 and 2), is tabulated below:

Degradation Coefficient

Mean Water Temp./ High Cycling Rate Low Cycling Rate

Heating or Cooling Test (5 cph at 50% load) (2 cph at 50% load)

95.7°F (35.4°C) Water/Cooling

44.7°F (7.1°C) Water /Heating

54.2°F (12.3°C) Water Cooling

85.5°F (29.7°C) Water/Heating

0.189 0.163

0.200 0.135

0.112 0.105

0.101 0.084

These degradation coefficients were calculated by performing a linear

regression to the data plotted on a graph of percent full load COP versus

load. The resulting curve was then used to find the percent full load COP

at 20 percent load. With this information, the degradation coefficient was determined from the following equation:

100 - percent of full load COP at 20% load (9)

C D

“ 80

9
7. ACCURACY OF RESULTS

In the heating mode, the average difference between the air and water side capacity measurments for the 10 steady-state tests was 1.8 percent, with the

water side measurement, on the average, higher than the air side.

In the cooling mode the average difference between the air and water side

capacity measurements for the 9 steady-state tests was 1.2 percent, again with

the water side measurement higher, on average, than the air side.

Standard error and the coefficient of variation are given by the equations:

Standard Error = (yi - y ) 2

n - 2

7

Coefficient of Variation = Standard Error

y

Where = the y value for the i th data point

y = the corresponding value from the least-squares fit

y

= the average y value

n = the number of data points

The standard error between the heating capacity given by the linear regression

fit to the capacity versus mean water temperature data and the individual data

points is 2500 Btuh (732 W) , with the coefficient of variation being 3.8 percent.

Similarly for the heating COP versus mean water temperature plot, the standard

error is 0.0850 with a coefficient of variation of 2.8 percent.

For steady-state cooling the standard error between the capacity given by the

linear regression fit to the capacity versus mean water temperature data and

the individual data points of 1670 Btuh (489W), with the coefficient of variation being 3.4 percent. Similarly for the cooling COP versus mean water temperature plot, the standard error is 0.1048 with a coefficient of variation of

4.2 percent.

The standard deviation of the cyclic test data generally increased as the

cycles became shorter and the run times less. The following table lists

the maximum standard deviation as a percent of the mean observed at each

cyclic rate for air side capacity and COP:

10
Cycling Times

Maximum Standard

Deviation in Capacity,

% of mean

Minimum Standard

Deviation in COP

% of mean

4 on/15 off 2.2 2.8

6 on/6 off 1.0 2.0

15 on/4 off 0.4 1.1

9 on/38 off 1.1 2.0

15 on/15 off 0.8 1.6

38 on/ 9 off 0.4 1.5

11
8. DISCUSSION AND CONCLUSIONS

For both heating and cooling, the capacity and COP of the heat pump under test

could both be considered linear functions of the average (mean) temperature of

the water in and out of the unit for inlet water temperatures ranging from 40°F

to 90°F (4.4°C to 32 . 2°C)

.

The steady-state heating capacity and COP increased with increasing mean water

temperature, as shown in figures 5 and 6 . A linear regression of the capacity

and COP data versus mean water temperature was found to correlate the data

well. The correlation indicates that, increasing the mean water temperature

from 40°F to 80°F (4.4°C to 26.7°C) would result in the unit heating capacity

increasing from 47,000 Btuh (13.8kW) to 80,900 Btu/h (23.7kW). Similarly the

COP would increase from 2.85 at a 40°F (4.4°C) mean water temperature to 3.25

at an 80°F (26.7°C) mean water temperature.

For the steady-state cooling data, a similar linear regression was made with

the data correlating well. The results indicate that decreasing mean water

temperature from 90°F (32.2°C) to 50°F (10°C) increases the capacity and COP

from 46,900 Btu/h (13.7kW) and 2.3 to 57,500 Btu/h (16.8kW) and 3.24

respectively (see figures 6 and 7)

.

The coefficient of degradation of COP, C^, due to part-load operation was

found to increase with decreasing mean water temperature in the heating mode

and increase with increasing mean water temperature in the cooling mode.

Assuming linearity and extrapolating the C^ vs. mean water temperature

data presented in figure 14 gives the following values for C^ for heating

cycle operation.

Heating Cycle Cp

Mean Water 5 cph at 2 cph at

Temperature 50% load 50% load

80°F ( 2 6 . 7

° C) o.09 0.11

40°F (4.4° C) 0.14 0.21

increase 56% 91 %

Similarly the following CD values for cooling cycle operation may also be

extrapolated from figure 14:

Cooling Cycle C

p

Mean Water 5 cph at 2 cph at

Temperature 50% load 50% load

50°F (10.0°C) 0.10 0.10

90°F ( 32.2°C) 0.18 0.16

increase 80% 60%

12
In an actual application the coefficient of degradation will have its

greatest effect on seasonal power consumption when the unit is substantially

oversized in comparison to the building heating or cooling load requirement.

This will occur in the cooling mode when the unit is sized to meet the heating

load in a climate where the summers are mild relative to the winters and the

ratio of the cooling to the heating heat loss factor is at its maximum. Conversely, a large part load heating seasonal performance degradation will occur when the unit is sized to meet the cooling load in climate where the winters

are mild relative to the summers and the ratio of the cooling to the heating

heat loss factor is at its minimum.

To aid in evaluating the importance of the change in unit capacity, COP,

and part-load performance degradation coefficient, C~, with varying water

temperatures, bin-method calculations of the seasonal performance factor were

performed for a representative city in each of the six heating regions defined

in [4].

The seasonal performance factor is the dimensionless ratio of the heating

(cooling) done by unit over a heating (cooling) season to the total electrical

energy used by the unit in the season. The heating (cooling) seasonal performance factor calculations include the effects of part-load performance through

the part-load degradation factor.

These calculations of the seasonal performance factor for this heat pump

when oversized by the maximum amount felt to be reasonable for heating

(matched to minimum design heating load) and cooling (matched to minimum

design heating load) are given in appendix A. For the minimum design heating

loads, the calculated degradations in the heating seasonal performance factors

due to part load operation shown in table A2 range from 12 percent to

13 percent for all six regions. The calculated degradation in the cooling

seasonal performance factor, due to part load operation at the minimum

cooling loads, also shown in table A2 range from eight percent to 10 percent

for all six regions.

The minimum heating seasonal efficiency for this heat pump occurs when

it is matched to its maximum design load and uses supplemental resistance

heat to meet the building load at low outdoor temperatures. In the sample

calculation of appendix A, the smallest heating seasonal performance factor

occurred with the maximum cycling rate (5 cph at 50 percent capacity)

and when the unit was matched with the maximum design heating load (280

percent of the unit heating capacity) assumed for region IV (see Table A4).

In this case the seasonal performance factor is 73 percent of the corresponding steady-state COP. The use of an undersized unit and consequently

substantial amounts of supplemental resistance heat, while resulting in

a significant seasonal efficiency loss, may be justified by the substantial

reduction in first cost and by better load matching in the cooling season.

13
9. REFERENCES

1. "Method of Testing, Rating and Estimating the Seasonal Performance

of Central Air Conditioners and Heat Pumps Operating in the Cooling

Mode" by George E. Kelly and Walter H. Parken, NBSIR 77-1271.

2. "Method of Testing, Rating and Estimating the Heating Seasonal

Performance of Heat Pumps" by Walter H. Parken, George E. Kelly

and David A. Didion, NBSIR 80-2002.

3. "Facility Design and Planning Engineering Weather Data" Departments of the Air Force, the Army, and the Navy, 1 July 1978, AFM

88-29, TM 5-785, NAVFAC P-89

4. "Proposed Rulemaking and Public Hearing Regarding Test Procedures

for Central Air Conditioners, Including Heat Pumps." Office of

Energy Conservation and Solar Applications, Department of Energy,

Federal Register, Thursday, April 19, 1979, part II, pages 23468

to 23506.

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

Pump Equipment." ANSI/ASHRAE 37-1978.

6. "Standard for Water Source Heat Pumps." ARI 320-76.

14
Figure 1. Heat Pump Installed in Test Apparatus

15
Figure 2. Heat Pump Installed in Test Apparatus

16
9.6’ TO CEILING

Figure 3. Air Flow Measurement Tunnel

17
V-i <d e s 3 O

m o Omm o CO o <fr m CO On

m o O

Cm (0-i 00 oo 00 00 00 00 00 On 00 00 00 00

-sc

•SC <r cn CO •*1 On o CM •-H o a. CM CN CM CM *“H o o On On o • • • • • • • • • •

C_) cn CO <n cn cn cn CM cn CM CM

mi 4J (D

O o o o o o o mO o •H 3 3 CM nD o n nO On nO i-n o CM 3 o nO cn nO 00 r^. no cn cn

30 CX r-^ nO nO in -om in

m

O

u • 3 CD CX o in CM cn On CM 00 nO 00 CM

CO 4-1 6 • • • • • • • • • •

CD CO 5 m CM

O On 00 CM o S 3 00 00 00 nO nO m cn m m

>H 0k CD H

m CNI no Onm 00 CM 00 in On

4-J Cm • • • • • • • • • •

CO < o a\

m on On in On o on

CO 3 r-H »"H *-H i—H

p-H *=H 4J »—1 3 4-1 CO 3

cd

O

X

U (X,

4-1 CD o CM on <fr CM CM

r—1 n CM

co 4-1

® • • • • • • • 9 •

01 CO o o On <fr On

mm On o H 3 00 |Ns n m <r cn cn

00c 3

*H

M

4-J cO

d

<D CD Cm 00 CM o CM nO *-H cn o o CM

4-J

O • • • • • • • • • •

CO On o o On On On CM o o o

01 3 00 On On nO nO nO -o nO in nO

4-» CO

4-)

O

C/0 CD 1 CJ *— C x

no OJ o

CO (-1 -C

<D CD 4-1

4-J

C4-4 a*

C/0 <4-C | J3

•H °H 4J »—H cn

O CM r-Hm On Cn cn

no J3 cr • • 4-1 *-H CM C”H CM f-H H r-H cn 1—1

s^? cr 1 1 1 1 I 1 l 1 i l 0) r—1

,£0 4J

cO •HU

H o o o «o 4-1 CM CX cfl X 3 l-l CJ o O o O o o o o o o O * X cn m »—1

cn nO f-H On o m CX

•H<D ' m 'd- CM <fr nO in nO o rt -Ono 3 0k *> 0k 0* 0k 0k 0k * 0k >

4-NH mi

m cn

H Cn CM <T On cn 0) cr c/o oQ 00 co 00 no nO <fr n in in

oo3

5^ •H

4J >

•H i-4 u 3

(-1 3 4-* •H CX W

^3-4

u V4

O

O o o O

O o o o o CD * X cn CO in On nO o o nO >

oaj ' *-H o co r^. On On in nO rH

XJTO 3 3

4-NH 4J <fr CM o CM nO •-H cn 00 CM CM >

cr c/o 00 00 00 00 nO nO <dn in

m o

•H

•SC •K * 3

<T> On On On On an 3

r>. on on On On 3 \ 1—„ —

\ —

''M.

a

CD

m

m in o o o V, \ X

4-J CNI CN CM cn co CO •—H

<r in

m

w

c0

N \ •K

Q

m

m

m in in

m nO nO nO nO

to3M

CD>oa

4-1 33

<U .c

4-)

T3

<D 'O3

t—Ho

CO

COS

V4Io

exa

cO

T3

0)M3

CO

CO

a)s

no

<D 4J

CO

•r4

<0 £** 18

was

used

to

calculate

the

coefficients

of

performance

listed

in

this

table
Table

2:

Steady-State

Heating

Test

Conditions o

•H • •u <D O

flj Vi 4J 3 • »H r—4 1—1

CO CO N" *n -a- CO CO 3 sr NT Nt <r NT NCO a . . • * • . # . a

Unit Pre

In. o o o o o O o o o o u

cO • a> aa CO ON ON

H NO ON nO ON CO o

Vi

O . . • • • . . . . .

0) C_> ft. o ON CO o o o o r-l o CO

Or o 3 VI V3 cO

CO CN CN CM CN CN CO CN CN CN

4J

cO • <u a. _C CO p- NT 1—1 m r^.

OO ON

V > Pn . . a • • • . . a

a

W o LO CN ON

m NO nO ONm in

m

a CM CN 1—1 rH rH t—1 r—1

i-4 r—4

T 1 3 vl

C/3 CO

a>

a vi O 3

•H CO cN

O o Nf o oo o CM

4J CO 00 . • • • • . • . a

a o a> -h m CN ON ON CN CN 00 CN CN 3 ^ co (Aha on ON 00 NO nO N- m m m

<u00 0) U Vl

cO 3

jC CO CN M0 o CO ON 00 r—1

*—i00 ON

O CO 00

H

O o m NT i—1 <r CN CN

co oj

h

•H Vl

W 3 cli a

CO CO CO CM CN CN CN CN CN CN

Vi

cu 3 M3m 4J o a . r-'. 00 • rH CO 00 NT p. nO

10H O. CN . • o • a

a . • 3 Cu 00 r—4

f''.

m rH nO st

m nO CO 3O Vi rH CO

ft, -o

a o o m

mm

O

Om

m

m

C CM ~o NO CD rH CN NO CN CN CO

Vi CO CJ on ON ON o O o O

O

O o

•H 4-1 <! CO rH fH rH CM CN CN CN CN CN CN 3OH a o o o m m m o o m m

fM 4H 00 00 00 oo 00 00 00

CJ rH rH iH rH r—4 iH rH T—4

rH rH

V4 3

CM CM CN CM CM CM CM CN CN CM

CH 4-1 a -h

a> c CN CO co H 33 a . • • CO CM

O Nf rp 00

ft. ON 00 r-v CM • a <% « a • Vi 4-1 o o o o o ON p. O MO CO CO

•H 3 i—4 rH 1—1 rH ON ON ON ON ON ON

<d o* 4H CV -H a c

a) 3) ON CM ON rH o co 00 30 o H a • a • • • • a a • O ft, ON o ON CT\ ON ON o ON ON o WHo o r>. ON M0 M0 MO MO

•H

C

<C hH * •K *

ON ON ON o> ON ON

Pr. p- I''. IP r>. ON ON 0> ^ rp ip

a)

m

m

m

O o o *>>.

—

4H CN CM CN CO CO CO rHm

m

CO

—s p» '^a

a in

m

m

m

mm MO M> sO

MO

4-1 COVlo&

CO>

(1) 00a

•rH>U

cO

4-1 CO

03>i—\

CO>co

•H

COc

cOaxw* 19
CAPACITY,

Btu/hr

Figure 4. Variation of Heating Capacity with

Entering Water Temperature for Different

Inlet-to-Outlet Water Temperature

Differentials

20
CAPACITY,

Btu/hr

Figure 5. Variation of Heating Capacity

with Mean Water Temperature

21
3.3

Figure 6. Variation of Heating COP with

Mean Water Temperature

22
Table

3:

Steady-State

Cooling

Test

Results o m 3 CM r-H CO 3 CT\ ON

Pn-i * O no

cm CM o • • u CM CM X u cu o o

•H

m

m c o o *-H ZD CM nO nO

O LO o o LO o m

CO CM CO CO CO CO CM ON on ON on ON

M0 1 I 00 CM CM O CM a> uo 00 CM • • • • • • • CM r-H CO CM CM CM CM o o LO m o LO o

ON nO CM r^. CM r^ o o co r-H co 00 CM NO nO m m m m nO

i-i 0) X fn

cO o 3: r^. Os vD i-H CM <y\

m

m

O • • • • • • • • • • 3 ex o a> m CO m r-H ON

co e ON ON o m r^- vO OO 3 3 2 E-i

r-H

S-i - 3 H 00 'mO CM CM

O 00 t—l

4-1 tM • • • • • • • • •

CO < o o ON o>

O r-HO o ON o ,3

f—1

r-H CM CN| CM -<r co < CJ • •H 3

JC rM

4-1 5 -o 3 3 O

CM r^ co i—

H

00 00 00 <r o

TO

X 3 X IX • • e • • • • • X CO

CL) 0 M0 <r in CO

m r-H ON CO *H

4-1 ON o o *-H vD On 00 o •H X!

to r-H r-H r-H *—i 1 1 X 3

as in 3M X TO

d) 3

S-i [X CM i“H oo t-H o>

O i—i > X 3 0 • • • • • • • • • O CO X m m <r m CM m r-H ON CM -H

to 00 OO ON

m

m <T NO 1—t

XS.

X

•H 0) 3Oc3X

0)

M-l I Ooox

cr

S'?

J+

co

cr

o*n X

hJ+

CO

cr

X 3

4-1 CO Q

Oo mo+0001

NOIONoI

OOo

NO

00o

no

CM

r-H CMrH !—(

vO

r—H

NO

~d"

NO

m

nO

mr-H NO

O3

cO

(U E

rC

4-1

to3

TO3

iHO3

*H3

cO

XO

4-4X

0)a3

4-1 3

0)

•r!O

O CX

O o o o o o o m o 4H

CM 3 -3 CM 00 *“H 00 00 CM CM 00 o X 3

do

u 3 CM 00 r-H CM

O 00 cn CM 3 O

4J 3 U

om

m

O CM «-H oo o m

O

era X*H

cr co

m -dm

m

m

m Cm 3

-3 £ ^ 3

>4-1 3

4-1 TO cO

X CM

w r-H •H 3 O o o o o o o o o 3 c\3 < U -3 O m CM 00 o o CM cn o 3 o 3 NO <r OO CO cn NO CO oo E

XI 3 o +T3 PQ CM *-H CO o CM NO TO X

cn-x -tr

m in in

m 3

cr co

X TO 3 3

CTs o\ ON ON ON ON ON ON ON •H 3

rH 3 3 0) -C cO ST *

23
Air

DB

Air

WB

Air

DB

Air

Wb

Air

Flow

Air

Flow,

Water

Discharge

Suction

Superheat

Superheat

Unit

Static

Condensate,

Into

Unit

Into

Unit

Out

Unit

Out

Unit

cfm

Standard

Flow,

Pressure,

Pressure,

at

Evap.,

at

Comp.,

Pressure,

lbs/hr

°F

°f

°F

°F

cfm

gpm

psig

psig

°F

°F

in.

w.u.

nO

jO

<t fA

• • fN rsj

«o

• • • • iA -CT rA

00 CM

• • CO xT

H CO CO •—

<

H CM CM CM

•>T <r

• • • O o o o © o o © o

on r>* nO sO ON CO O lA

• • o H CO o ON H CM <M CM CM CM CM CM H

nO ON CM © r>* sr tA

• • • uo iA «a H '•H -4

>o ON ON 'O H • • • • ON o no CM CM 00

vO r^» sO sO V© vO nO nO

00 o ON » sr to to to

CO sO tO 30 vO 00 CM ON vD

CM CM CM ts CM H CM o CM CM H CO m Mf

• • • • • • • • • CO

r*4 CO

p-4 'O r>» r>* CO CO CO

O U0 m © © o o m tA

ON ON ON ON ON o o ON 00 © © © © © *—

4

H © © CM CM CM CM CM CM fS CM CM

iA © «

A

© m ia © © 00 ON © © tn no NO P-* © © i—4 «—

4

© © © © ©

CM CM CM CM CM CM CM CM CM

© CO •a CM o O • • • • • • • • 00 © nO NO CO A P^

ia »a tA m tO a A A iA

A 00 o iA CO A 00

• • • • • • • • • CM CM CO CO a* 00 ON 00 •—4 sO NO NO NO A «A iA A sO

p*»» <r r—l ON CM nO CM P^

• • • • NO nO CO CM CO CO CO CO

NO sO -O nO nO nO sO NO nO

H o CM f—

4

© p-4 OO CM

• • • • • o © © © o © © ON ©

00 JO CO 00 00 00 00 00

C7N ON ON ON ON ON ON ON ON

r^. 00 oo CM CM A A

© © H r—

1

H r—

4

H \ NO NO NO NO nO NO nO vO nO

24
CAPACITY,

Btu/hr

Figure 7 . Variation of Cooling Capacity with Mean Water Temperature

25
50 60 70 80 90 100 110

MEAN WATER TEMPERATURE, °F

Figure 8. Variation of Cooling COP

with Mean Water Temperature

26
%

OF

FULL

LOAD

COP

%

OF

FULL

LOAD

CAPACITY

Test Conditions: Avg. Inlet Air Temperature = 69.9°F (21.1°C)

Avg. Inlet Water Temperature = 90.0°F (32.2°C)

Avg. Mean Water Temperature = 85.5 C F (29.7°C)

Figure 9. High Temperature Heating Cyclic Tests

27
%

OF

FULL

LOAD

COP

%

OF

FULL

LOAD

CAPACITY

Test Conditions: Avg. Inlet Air Temperature = 69. 8° F (21.0°C)

Avg. Inlet Water Temperature = 50.0°F (10.0°C)

Avg. Mean Water Temperature = 44.7°F (7.1°C)

Figure 10. Low Temperature Heating Cyclic Tests

28
%

OF

FULL

LOAD

COP

%

OF

FULL

LOAD

CAPACITY

% LOAD % LOAD

Test Conditions: Avg. Inlet Dry Bulb Air Temperature = 80.0°F (26.7°C)

Avg. Inlet Wet Bulb Air Temperature = 66.8°F (19.3°C)

Avg. Inlet Water Temperature = 90.0°F (32.2°C)

Avg. Mean Water Temperature = 95.7°F (35.4°C)

Figure 11. High Temperature Cooling Cyclic Tests

29
%

OF

FULL

LOAD

COP

%

OF

FULL

LOAD

CAPACITY

% LOAD

Test Conditions: Avg. Inlet Dry Bulb Air Temperature = 80.1°F (26.7°C)

Avg. Inlet Wet Bulb Air Temperature = 65.9°F (18.8°C)

Avg. Inlet Water Temperature = 42.1°F (5.61°C)

Avg. Mean Water Temperature = 54.2°F (12. 3° C)

Figure 12. Low Temperature Cooling Cyclic Tests

30
%

OF

FULL

LOAD

COP

%

OF

FULL

LOAD

COP

0 20 40 60 80 100 0 20 40 60 80 100

% LOAD % LOAD

0 20 40 60 80 100

% LOAD

0 20 40 60 80 100

% LOAD —a

Cycling times of 4 min. on/15 off, 6 on/6 off, and 15 on/4 off —o— Cycling times of 9 min. on/38 off, 15 on/15 off, and 38 on/9 off

Figure 13. Linear Regression Fit to Data on the

Degradation of the Coefficient of Performance

at Part Load Operation

31
COEFFICIENT

OF

DEGRADATION,

C

0

0.22

Figure 14. Variation of Coefficient of Degradation

with Mean Water Temperature

32
Appendix A. SEASONAL PERFORMANCE CALCULATIONS

Six representative cities were selected for seasonal performance factor

calculations on the basis of design temperatures close to these tabulated

in reference 10.4 and for having abundant ground water. The design conditions

used in the seasonal calculations for these cities are summarized in table Al

.

The associated weather data for the calculations were taken directly from

reference 10.3.

Results of the seasonal calculations for the minimum building design load

to which the unit is likely to be matched are given in table A2. Results

when the unit is matched to the greatest likely building design load are given in table A3 and table A4. A sample calculation is shown in table A5

.

The maximum, and minimum design load requirements were calculated from the

following equations based on material presented in references 10.1 and 10.2

Maximum design cooling requirement = Q s s . c

Minimum design cooling requirement = (Regions I, II, III, IV, VI)

Q *. the lesser of ss,H1 0D,c2. or Q~

!

ss ,c

^ 65 T0D,H^

Minimum design cooling requirement = the lesser of

(Region V)

Qss,H or Q, 2.2

ss ,c

Maximum design heating requirement = the greater of

(Regions I, II, III, IV, VI)

Maximum design heating requirement = the greater of 2.2 $

(Region V)

Minimum design heating requirement = the lesser of

(Regions I, II, III, IV, VI)

Q

(65-T

s° n H

Minimum design heating requirement = the lesser of

(Region V)

where: 6 = Steady-state cooling capacity of unit, kBtu/h

ss *

c

Q „ = Steady-state heating capacity, of unit, kBtu/h

bb » U

Tod c = Temperature, outdoor design, cooling, °F

T

()d ^ = Temperature, outdoor design, heating, °F

A-l
The above equations differ from these presented in 10.2 in that Qgg

is

to be calculated at a specific application water temperature. In reference

10.2 Q ^ is taken at the 47°F rating point value, the value of Qg is

assumed to be equal to Qgs

and T

()D c is assumed to be 95°F for all regions.

The above minimum and maximum design heating and cooling requirements were

developed on the basis that:

1. The unit may be sized to meet either the design cooling or heating load.

2. The building heat gain per degree fahrenheit in the cooling season will

range from approximately equal to, to twice the building heat loss per

degree fahrenheit in the heating season.

The seasonal performance loss at the minimum design load (table A2) is totally

attributable to part-load performance degradation which is at its maximum.

Consequently, this comparison shows not only the maximum value of seasonal

performance factor degradation due to part-load performance degradation

but also the maximum error that would occur in the seasonal performance factor

calculation if the part-load performance degradation were not accounted for.

Table A3 and A4 compare bin calculations at the maximum heating and cooling

design loads respectively to the steady-state COP. Tables A3 and A4 can

be used to compare seasonal performance factors calculated for maximum and

minimum design heating and cooling loads.

The comparison to the steady-state COP in table A4 is not a valid comparison

of calculation procedures with and without part-load performance degradation.

A valid comparison would require a bin calculation of seasonal performance

degradation due to resistance heat only. The purpose of comparison to the

steady-state COP in this table is to provide a uniform basis of comparison

with the seasonal losses at the building loads used in preparing table A2.

The unit capacity and COP were calculated in each region from linear regression

fits to the data presented in figures 5, 6, 7, and 8. These are:

0ss ,c , Btuh =

C0P

s s> c %s,H> Btuh =

C0P ss,H

70,794 - 265.4 T ’ m

4.419 - 0.0236 Tm

13,010 + 848.7 T

* m

2.030 + 0.0184 Tm

where Tm

is the mean water temperature, °F, the average of the water temperature

in and out the unit. The mean water temperature was calculated from the

entering water tanperature presented in table A1 as follows:

A-2
It was assumed for the sample calculation that the water flow rate for all

applications was equal to that at the ARI cooling rating point, i.e., with

an entering water temperature of 85°F and a leaving water temperature of

95°F. This resulted in a water flow of:

Q water flow rate, lbs/h = m = ss,c = (70,794) - (265.4) (90) = 4690.8 lbs/h

C

p

AT (1) (95-85)

which is equivalent to a flow rate of (4690.8)/(60)(8.33) = 9.4 gpm.

The relationship between the mean water temperature, T

m , and the inlet

water temperature, T^, assuming constant water flow at 4690.8 lbs/hr, was

then calculated for cooling by solving simultaneously the two equations:

Sss.c - 70 > 79A - 265 - A Tm>c

Q ss , c = m CPaT = (4690.8)(l)[(2)(Tm>c

-Ti

)]

giving:

Tm c = (70,794+9381.6 T

± )/(9647)

Similarly for heating operation the two equations:

$ss,H ’ 13 .° 10 + 8A8 ‘ 7 Tr,,H

Qss ,h = “ OpAT = (4690.8)(l)[2(T

i

-Tm>H )]

give:

Tm H = (9381.6 T

i-13,010)/(10,230)

Using the above equations Tm

was calculated for each city and then, using

T , the capacity and COP were calculated using the previously given linear

regression algorithms and calculated by linear interpolation of the data

given in section 6. These quantities were in turn used in conjunction with

weather data from reference 3 to calculate the seasonal performance of the

unit

.

As a sample, the calculation procedure for the city chosen as representative

of region IV, Dayton, Ohio, will be presented in detail.

The inlet water temperature for Dayton was assumed to be 53°F. This results

in a cooling mean water temperature, Tm c> of

Tm c = [(70,794)+(9381.6)(53) 3/(9647) = 58.88°F

Using this mean water temperature, the steady-state cooling capacity and COP

may be calculated as:

A-3
6 „ = (70,794)-(265.4)(58.88) = 55.167 kBtu/h 3u jC

COP = (4. 419)-(0. 0236)(58. 88) = 3.029.

j c From these two quantities the steady-state cooling power draw, E , may

be calculated: S * C

j ss ,c 3.413 COP,

ss ,c Similarly for heating:

(55.167)

(3.413)(3.029)

= 5.336 kW

Tm >H = K9381. 6)(53)-(13,010)]/(10, 230) = 47.33°F

Qgs H

= (13,010)+(848.7)(47.33) = 53.179 kBtu/h

C0P c , = (2.030) + (0.0184X47.33) =2.898

SS j n

Es s,H

(53.179)

(3.413)(2.898)

= 5.376 kW

Next, the degradation coefficients were calculated by linear interpolation

of the data presented in section 6. As an example, for a cycling rate of

2 cph at 50% load a value of C

D of 0.163 was measured at a 95.68°F mean water

temperature, and a value of C^ of 0.105 was measured at a 54.21 mean water

temperature. These data can be combined into a linear interpolation formula as

CD = 0.105 + (58.88-54.21) = 0.112

(95.68-54.2 1

)

The other required C D values for this region were similarly calculated and are summarized as follows:

Cooling

,

2 cph at 50% load 0.112

Cooling

,

5 cph at 50% load 0.121

Heating

,

2 cph at 50% load 0.132

Heating

,

5 cph at 50% load 0.194

The building loads for which seasonal calculations were to be performed were calculated using the previously given formulae and were then rounded

to the standardized values given in references 2 and 4.

A-4
The maximum design cooling requirement was equal to = 55.167 kBtu/h

rounded to 60 kBtu/h.

0ss ,c

The minimum design cooling requirement was equal to the lesser of

Qs

„ ti (T

q

-65)/(65-T ) = (53.179)(89-65) / (65-4) = 20.923 kBtu/h or

Q

* = o5il67 kBtu/h which is 20.923 kBtu/h rounded to 20 kBtu/h. SS j c The maximum design heating requirement was equal to the greater of

Qp „ „(65-Tnn u )/(Tnn „-65) = (55. 1 67 ) (65-4 ) / (2 ) (89-65 ) = 70.108 kBtu/h ir

ss ,c Q

= 5 3 .

M

kBtu^f^which is 140.216 kBtu/h rounded to 150 kBtu/h. SS j tl The minimum design heating requirement was equal to the lesser of

( 65-Tqd h

)/ 2 (T0D C" 65) = C55. 1 67 ) ( 65-4 )/ (2 ) (89-65) = 70.108 kBtu/h or’Q „

= 53.178 &Btu/h which is 53.179 kBtu/h rounded to 50 kBtu/h.

The sample bin method seasonal calculations in table A5 were performed at

the maximum design cooling requirement, 60 kBtu/h, and the maximum design

heating requirement, 150 kBtu/h.

In table A5 the representative bin outdoor temperatures, t., and fractional

bin hours, nj/N, were obtained from the data presented in reference 3. tj

is the average temperature for each of the 5°F bins of reference 3. n^/N

was obtained by dividing the number of hours in each bin by the total number

of hours above 65°F for the cooling bins and the total number of hours below

65°F for the heating bins.

The building loads as a function of temperature for cooling, BL

c (tj), and

heating BL

fj(tj), were calculated by the following equations:

(T - 65)

B^> - TCc ) ( jOD>c

-6 5 ) (DCR) ’

and

(65-T.)

BLH

(tj) - 7

J (CH )(DHR),

(65 Tqd.h'

where C

c and are experience factors which tend to improve the agreement

between calculated and measured building loads. Recommended values

(references 1 and 2) are C

c = 1.1 and = 0.77.

As an example, for cooling bin number 4,

BLC (82) = - (60) = 38.636 kBtu/h ’

A-5
and for hearing bin number 4,

V47 ) - ^

(0.77X150) = 34.082 kBtu/h'

ctor, X( t . ) , equal to M, (0/$ < . X(tj)has a maximum valueJ of 1.*

The next calculated quantity was the load factor

for cooling and to BL

H (t.)/Q

gg

for heating

For cooling bin number 4: *

X

c (82) = (38.636)/ (55. 167) = 0.700 ,

and for heating bin number 4:

X

h (47) = (34.082)/53. 179) = 0.641.

Next the part-load factor is calculated from the equation:

PLF(X) = 1 - C D (1-X(t

j

)) .

For cooling bin number 4 with an assumed cycling rate of 2 cph at 50% load

this results in:

PLF

c (82) = 1 - (0.112)0-0.700) = 0.966 .

For heating bin number 4 with an assumed cycling rate of 2 cph at 50%

load this results in:

PLF

H (47) = 1 - (0.132)0-0.641) = 0.953

The heat pump electrical energy input for each temperature bin is represented

by the term

n X(T.) . J J E

* N PLF(X) tss

Continuing with the exampl

rate of 2 cph at 50% load,

n . X (T.) . JN

C J I, PLF

C (X)

h sS » C

and for heating bin number

load

n X ( T . ) .

.1 H J g n

= N plf h (x) ss,h

(0.966)

(0.953)

A-6
Note that when the building load exceeds the unit capacity, the heat pump

is assumed to run continuously and X(T^ ) = PLF(X) =1. In this case the

heat pump electrical energy input is equal to:

N

As an example, for heating bin number 8:

n. . -J- Ess H = (.085X5.376) = 0.4570 kw

N

*

In this case, when the building heating load exceeds the unit heating capacity,

it is assumed that electric resistance heaters will be used to maintain the

building temperature. Their electrical power requirement is calculated as:

RH BL(T.) - Q

n

j = j ss,H

N 3.413 N

Which for heating bin number 8 is equal to:

RH _ (71.951 - 53.179X0.085) _ „ /£ -, c N 3.413

O.Ab/b

The fractional building load hours in each temperature bin are calculated as:

Hf

except when X(t.) = 1 during cooling operation, in which case the fractional

building load hours are:

Hi ^ss,c

For example, in cooling bin number 4

BL( t . ) = (0. 143X38.636) = 5.525 , N

J

in cooling bin number 7

0SS c = (0.001)(55.167) = 0.552 ,

A7
in heating bin number 4

^ BL( t . ) = (0.102)(34. 082) = 3.4764 , N

J

and in heating bin number 8

nj

BL( t . ) = (0.085)(71.951) = 6.1159 . N

J

The seasonal performance factor is equal to the summation of the fractional

building load hours divided by the summation of the total electrical energy

inputs. For example, for cooling at an assumed cycling rate of 2 cph at

50 percent load:

CSPF

4 Qss ,c> X(T j) - 1

I j=l

ni BLcOrp, X(T j ) <L

7 3.413 Z J-1

nj Xc(Tj) •

N PLF^TxT

ss » c

21.6551)

(3.413X2.2003)

2.884 »

and for heating at an assumed cycling rate of 2 cph at 50 percent load

HSPF =

15

Z >1

3.413

|

BLh(Tj)

15

z J-1

n

j xhCt j

) g +

“nplf^CxT ss,h

15

z J-1

RH

(45.0641)

(3.413X3.8171+2.3611) - 2.137
Table

Al:

Design

Conditions

and

Calculated

Unit

Performance

for

Selected

Cities

30 30 o 30 ST CM a r—4

DO DO ON NO CO

•H 04 CM

H o 30 r^. ON

4J

O • • • • • • co ;j

CD 33

CO CO co CM CM CM

303 CO DO co ON CO CO

•H ON 30 CM CM rH 30 H 04 DO sD oo O CM ON o o • • • • • • o o CM CM CM CO CO CM

>» w js CM r-» ON ON 00 rH

•H do 30 DO -3- o 3 ON oo H rH ON

CO 4J • • • • • • a. so r>» 'O’ o co nD sj- cO M o

nO NO ~o DO sl- DO

G -H * >,

30 4-1 JT CO DO DO CM o C -H '>» NO ON St VnO co DO h y 3 CM CM OO rH CM NO

rH CO 4-1 • • • • • • O O.CQ o rH CM m r* O (S jd o o

DO DO DO

m DO DO

T3 U« a <u 3 4J • O cO Oj u 3 a O <U H

a. , a

oo

M (U

C

O

H

•HO U T3 CO

CO 4-1 00

<U 3 H

33

O 00

01 Qa , a

30

H <JJ

a o h

•HO U, H O Co O U CO O 3 -H O O co

a) co

•Hoo

CD

CM

r«*

30

CO

co

on

oo

no

CM

nO

c*l do

CM

oo

ON do

ON ON

oo

do

<r

CM

ON

00

do

do

CM

CMoo o rH

4J

O o CO CO a ”3 CD •H c cO 3 G rH o <u B

>N cfl Qu 4J c O

4J rH CO

a >N c y

•H

H J2 cu CO •H cO

U o U s a 23H

ao

•H00

<uuuo

CD3

4J

cO

<Ua.a

<u 4JU

<u 4J

cO3

*3a3ou

OO

4-1 «0

CO

<u JG

>N

4-» •HO

COa

co o
Table

A2:

Comparison

of

Seasonal

Performance

Factors

to

Steady-State

Coefficients

of

Performance

for

the

Unit

Sized

at

the

Minimum

Design

Heating

and

Cooling

Loads

l-H CD CO CM -4

Oh Oh 00 00 00

CO

O oo 00 00 00 00 00

30 o • • • • • • 33

00 CO CN cn hH

43 Oh o vO cn cn ON

CO Oh 00 vO

m <r in o cn • • • • • • rH 33 CM CN CN CM CM CM

U

cu CU

S'? +j 33 o . cfl •u m * 5

CDu

•u CD o CO 4 -4 r-H cO

cO Uh Oh CM o r-H o r-H o CO Oh o ON a\ OA ON ON ON C CU 33 cn o • • • • • • o O

cu

O o •rH c o 00 CO

cu E

to «H c

vO o> -4 NO 00 o

Oh CO CO

m cn ON cu cw

Oh CO LO r^. ON NO 33u

cn • • • • • • 43 cu o CM CN CN CM CM CM a co

•HOO

CU 03 o

on ON ON

ON

ON

CMON

CMON

43

CO 14HO OO 4J c e

co cu EU

33u • • • • • • O •H

33 <+H <4H U 4H

CU <U

a CU

O

co ON CM 00 4 r-^ o o o tin •H n- 00 NO -4 o HH

Oh ON 00 NO

m r-* o OO

cn • • • • • • 3

64 33 CM CM CM CM CM CM 43 •H O c 43 m cu CO * •rl CU 43 CD o 33

cO bn CD ON l"-. CM o 00 CM •rH 3h Oh CM -H CM rH rH MH CU 33 cn O On ON ON ON ON On >4H 33

CU

O CJ> • • • • • t a» 43

a

u oo <4H CM

00

O

O rH cn NO ON rH c cu Uh i—I NO o in •4 CM •H 3

CU -4 •4 NO C'H ON r~H rH rH

cn • • • • • • o CO CJ CM CM CM CM CM CM o > o

CU

cu 43 33 CO E 00 33 43 43 3 c w* \ CD E •rH 33 3 14H | •H 43 CO 4-1 in in O O O O O >N c CO O 0Q CM cn 4 m in m 43

•rH <U rJ 33 CU CO 33 3 CU rH 43

CO CO >

cu . g 00 33 cu 33 CU 3 e A 43 43 (H E •H 33 3 o o o o o o CO 3

•rH iH CO 43

m 4 4 CM CM CM 43 CO 43 c O O ca CO •H CO

•rH

O rJ 33 1 U

32

U >N cocu 43 CO Q.

CO Oh E

CD CU o ® c •rH 43 U 43 o rH CD 33

43 o IH o CD CD Oh CU (U 43 (U •H (3 CO cO 33 • 43 C rH 33 O CU E 43 CU co >> CO Vh a 43 3 o >H 5

43 rH CO

g >' c o CD 3

rH bl 33 cu CO •rl CO •H 43 rH o o O 32 a s H CO CO

CO Vh

C

cn a> o

CLi arl OEM u ® <u O 4J J-l
Table

A3.

Comparison

of

Cooling

Seasonal

Performance

Factors

to

Steady-State

Coefficients

of

Performance

for

the

Unit

Sized

at

the

Maximum

Design

Cooling

Loads.

TJ

COO6s? OmX

a.om

CL. cn

cj

Je

CL, coo 0) Xu o sr vO 00 —H

CO 14-1 CM CM CM <r in •o o

on ON ON ON ON ON

©

CJc

CO

ea • • • • • •

vO CM CM cn E

00 00 •—4

m —4 o

CO sf NO 00 o oo 4-1 • • • • • • l-l CM CM CM CM cn CM ©

CL

«4-l o

X)

CO * o CO

i—4 CO u. a.

if-S CL,

O ON cn cn CM '3- 00

O CO CJ CM cn cn in

m SIm CJ CJ ON On On ON ON ON • • • • • •

4-1 CO X

CL Cl4 o srm MIm rs o CL, o cn GO NO CM

CO srm VO oo o oo

CM CJ • • • • • •

CNI CM CM CM cn CM

4Jc©

<4-t <uoo2

•H

I—I ooo©Xu

4-1o § C 00

00C • JE

•H CO r-4 CO 3

><

©

O

O 4-1 CO Q O t-J CO

5S CJ

Om om om ovO onO om co

00©

OS >w

(U3

r—4

CO>©

©

4-1

M

CO 3

4-1 4-1 CO CO I U

>n 3

T3 CL

cO

g

0)©

4-1 4J

CO L4 © ©

-C 4-1 4-1 (05

co •H r-4<0 COG

CO

O

CL, -H 8

00©

O L4 K >
Table a

4.

Comparison

of

heating

Seasonal

Performance

Factors

to

Steady-state

Coefficients

of

Performance

for

the

Unit

Sized

at

the

Maximum

Uesign

heating

Loau

X

CO

co

-14 04 "d* rom 00 sO CM

^4 o ,-4 r—1

r-4 40

O

.0 CJ On r'-. 30

-d 30 • • • • • •

c0o Ul 'H 34 4j ! CO O r*H O CO H •—4 3

si 04 <r r>* a> rg >o UO

04 ON jO rH

O CO -1

4-1 Si • • • • • • cO

-o t4 <N| CM ev| eg CM

a0

si •1-1 a, 30 o 01 lO <uo <u •u c sz C to r—1 O m O <M u

a) 4-1 4-1 • • • • • • o CO CO O O T 1 X) & X> 4-1 Cl •H CU CO co CM 03 U CO 03

a> 0) si Oa§0 * <4-1 COX 30 f-g ON cu J4 04 co <T) co 3 a.

04 o ON ON 3\ 30 CO CJ • • • • • • 4-1 03 0

-o 4-1 CO c o <U rH •H

sO -O’

O CM

O

04 r—4

cO CO •H o 04 O CON 30 r-^ O 00 '-H lO CO • • • • • • NW co CM CM Cs| csl CM cu

4-1 0

i0 0

30

Q, c

CJ eU

•H

u

U

CNIU c CO c CO 4-4 rH O tn X> M" <U

at 4J to • • • • • • s:

CJ co CU

O

O r—1

eg nO

Cl

H o3 CO CM

V

cu CO s:

04 cu 4-1 3$

4-1o a c so jc 3 IOC a -H -O D

•H

w -u nj -ui ^ tl n) O 3

to

a a) _q 3*5 s 03

<u3

iH

c0>

CU u

c0

4-» XI>> 'O

CO

<u

1) si

4-1

UJ •H

CO

CO Cl CO <u a, a. o a o <u 03 -o *
•r“) CM o o o m pH cn

—1 St CM ON CM ON

X ONX O'

XX

O

X CM pH

LO cn o m in ON ism pH

X O'

X

X cn o mX pH IS rs

m

X

X s-** x ON o CM m m CM X m Is pH <fr pH pH ON r^. O pH cn X X

X o ON 00 in X cn CM X X m <r X o cn pH cn CM CM co X cn pH o CD • • • • • • • • • CO m m st pH CM m st X IS X cn CM pH X CM *<r "I c O m 00 r-s pH pH X O' pH ON IS X st O' m X pH pH X <f ON O ON cn X <N X

32 1 2: cn st st CO cn pH pH o O

X

0d| • CM • U

•o X-N

«0 /-s

X o •«n 'w' -J 4J

x o m

X st o m CM

X ON 00 IS o O cn

m cn inX pH st cn

-Jm

X pH <r pH •H O' O' 00

m 00 cn pH ON rs o CO

X stX

X IS c xx O-i o ON m CO X cn o 00 ON oo 00 pH CM m O' ON pH X CO pH o X

oo V-/ o o CO in in st pH CM o H CM cnX

X st CM pH pH o o o o X •H •n z • • m c CM X

0) Qg

X

D /-N v-x ON st O' St ON rs pH

X o m ON e X til pH pH CO X cn H 00 CM X ON cn X ON

H pH pH pH pH pH pH pH pH

•H 13 O' ON ON ON 00 00 00 00 ON O' ON

acO* • UJ >HH X c #rn o *\4J U<

x o o CO

H rs cn rs o m ON pH CM o o cn

m

X

m

X pH pH

•H

Xm

X pH cn 00 00 rs o CM

X 00 co 00 pH O' IS o cn

X

X

X rs

00 a- o ON rs

m r*>»

m cn o oo oo rs

X fs pH CM

m ON ON pH

X

X pH o pH

cu o o cn

m

m St CM o pH CM cn St

X

X CM pH

H o o o o X

a: •1-1 25 • • c CM X eo

•H

4J

X O'

X cn o CM

X ON cn

X

m 'w' pH pH

x

X <3- CM ON CO o CMm rs pH pH pH pH pH pH pH pH pH pH

pHD

CO

a

O' ON ON ON 00 CO ON ON ON ON o a, pH

co o —N

T3 •r-)

X o stX CM is in cn

H ON rs

O ur pH

—

H o o ON 00 CO o 00

X St pH O' *H pH pH •H •H pH pH pH pH

r* ON rs st CM o CM St

X 00 ON

4J

X

a» sc

•H

0Q is st o X cn ON

m o 00m CM ON

X st pH oom CM o rs st pH -S CM X o cn is o St 00 St H 00 St H 00 in X m CM X X CM

0) •rn cn o X CM ON

mX pHX o m o st ON

X cn

X CM CM . J • • • • • • • • • • • • • • • • • • • • • • a V-/ CM pH o CO m m m st cn cn CM pH pH o o ON ON X X 6 J X m co CM pH pH CM cn st m X rs X ON o o pH CM X 59 CQ pH pH pH pH pH

to• m »H

X

X CO is St rH o St CM

X st rs

m stX CM CM r"sX pH -o -a < •r-) lz o *H st O rs CO CM H O o o pH ^H X m cn CM pH O o o 03 CO c 1 o o o *—H CM CM CM pH i-H pH pH pH pH pH o o o O O o o o o o

a> pH pH pHXl

S'? ss

to •r~l CM CM rs CM CM rs CM is CM rs CM CM CM CM

X

X o o H u O' ON CO 00 X X m in <r cn cn CM CM pH pH 1 t X m

4-? 4-J CO CO

•rn

X J= a c.

0k CJ

u • *H pH o co CO CM X z r^. X in <r cn CM pH X pH CM cn St m X oo O' o pH CM cn <r X 4H O pH pH pH pH pH pH o c H H /"-v /—N

CO

X

0Q 's/ s-/
NBS-114A (rev. »-78)

U.S. DEPT. OF COMM.

BIBLIOGRAPHIC DATA

SHEET

1. PUBLICATION OR REPORT NO. NBSIR 81-2287

4 , Wim-. H '"S'?

§1 wMi

P 2*

4. TITLE AND SUBTITLE

LABORATORY TESTS OF A RESIDENTIAL UNITARY WATER-SOURCE

HEAT PUMP

5. Publication Date November 1982

7. AUTHOR(S)

William J. Mulroy and George E. Kelly

8. Performing Organ. Report No.

9. PERFORMING ORGANIZATION NAME AND ADDRESS

NATIONAL BUREAU OF STANDARDS

DEPARTMENT OF COMMERCE

WASHINGTON, DC 20234

11. Contract/Grant No.

12. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street, city. State, ZIP)

Department of Energy

20 Massachusetts Ave 0 » NW

Washington, D.C. 20585

13. Type of Report & Period Covered

15. SUPPLEMENTARY NOTES

(_”] Document describes a computer program; SF-185, FIPS Software Summary, is attached.

16. ABSTRACT (A 200-word or leas factual summary of moat significant information. If document includes a significant bibliography or

literature survey, mention it here.)

The performance of a residential heat pump was measured in the laboratory over

a broad range of source water temperatures (40°F to 90°F). Tests were performed

in both heating and cooling operational modes and for both steady-state and

cyclic operation.

For both heating and cooling operations, the unit capacity and COP were found

to be linear functions of the average of the unit source and outlet water

temperatures. In heating, the unit capacity, COP, and part load performance

increased with increasing water temperature 0 In cooling, the unit capacity,

COP, and part load performance decreased with increasing water temperature

The measured degradation coefficients ranged from 0.09 to 0 o 21 for heating

and from 0.10 to 0.18 for cooling 0 An appendix is included in which the effect

of the degradation coefficient and of supplemental resistance heat on the

unit heating and cooling seasonal performance factors is calculated.

17. KEY WORDS (six to twelve entries; alphabetical order; capi talize only the first letter of the first key word unless a proper name; separated by semicolons)

Central heating equipment; cooling; heat pumps; heating; heating seasonal performance;

heating seasonal performance factor; test method; water source heat pumps

18. AVAILABILITY HXt Unlimited

For Official Distribution. Do Not Release to NTIS

[3] Order From Sup. of Doc., U.S. Government Printing Office, Washington, DC

' 20402, SD Stock No. SN003-003-

IXl Order From National Technical Information Service (NTIS), Springfield, VA. 22161

19. SECURITY CLASS

(THIS REPORT)

UNCLASSIFIED

20. SECURITY CLASS

(THIS PAGE)

UNCLASSIFIED

21. NO. OF

PRINTED PAGES

51

22. Price

$9.00

US COMM- DC