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The impact of refrigerant charge, airflow, and expansion devices on the measured performance of an air-source heat pump--part I.

INTRODUCTION

Air-source heat pumps are used widely for residential applications in heating season. However, the performance of air-source heat pumps degrades significantly at lower ambient temperature. In winter, the refrigerant evaporating temperature tends to be lower than the freezing point, which accelerates frost formation on the outdoor coil, and defrost operation would degrade the performance further. Heat pumps working in heating mode usually have a higher pressure difference than when working in cooling mode, which results in different start-up behaviors. On the other hand, system charge level, expansion device, and indoor airflow are important factors for heat-pump design and operation.

Domanski and Didon (1984) released their mathematical model of an air-to-air heat pump equipped with a capillary tube. The model was compared to experimental data for verification. The tested air-source heat pump had a reciprocating compressor and a suction line accumulator. The authors observed that in heating mode, superheated vapor did not enter the suction line accumulator and the compressor. O'Neal et al. (1991) investigated the effect of short-tube orifice size on an air-source heat pump during the reverse-cycle defrosting. They compared five open-tube orifice sizes and reported that increasing the orifice size can result in a shorter defrosting period and, thus, higher integrated COP. Payne and O'Neal (1994) compared defrost cycle performance of an air-source heat pump with a scroll compressor to that of a reciprocating compressor for a 2.5 ton (8.79 kW) air-source heat pump using R-22 as the refrigerant. They reported that the scroll compressor showed a slightly better COP, while the reciprocating compressor led to a shorter defrosting cycle due to its higher discharge temperature. Chen and Guo (2009) studied the reverse-cycle defrosting characteristics of an 11.2 kW (3.2 ton) split-type air-source heat pump. The effect of outdoor air parameters (temperature, humidity, and frontal air velocity) on defrosting cycle performance were investigated. They reported that the defrosting time and power consumption decreased with increasing outdoor temperature and relative humidity; however, there was an optimum air velocity for defrosting performance. They explained that the defrost time and power consumption were determined by the frost density and frost mass accumulation. A larger mass of frost accumulation results in lower condensing temperature and better defrosting performance; on the other hand, denser frost requires more power and a longer time to melt.

Kim and Choi (2001) conducted an experimental study of a water-to-water heat pump using an electronic expansion valve and a capillary tube with varying system charge level, from -20% to 20% of the nominal. They concluded that the system using the capillary tube was more sensitive to the and Integration Center, Oak Ridge National Laboratories, Oak Ridge, charge variation than the system using the electronic expansion device. Corberon et al. (2011) studied the influence of evaporating source and condensing sink temperatures on the optimal refrigerant charge of a water-to-water heat pump. They reported that the optimal charge, reached at around 5-7 K (9-13[degrees]R) sub cooling degrees, was almost independent of the sink temperature. However, extra charge was needed for the optimum COP when increasing the source temperature.

In an experimental study, Kim et al. (2009) investigated the performance of a residential heat pump operating in cooling mode with single fault imposed; the heat-pump system used a thermal expansion valve (TXV) as the expansion device. They evaluated the performances of improper outdoor airflow, indoor airflow, refrigerant undercharge, and refrigerant overcharge, etc. Yoon et al. (2011) conducted a series of comparative studies for the heating mode and reported commonality between sensitive features in the heating and cooling modes. They also pointed out several different features, for example, the COP of the heating mode was more impaired by undercharge.

Shen et al. (2006, 2009) published substantial laboratory testing data for three unitary air conditioners that used R -410A refrigerant and ranged in size from 3 to 5 tons (10.5 to 17.6 kW). The experimental study covered an extensive range of charge inventory, indoor and outdoor airflow rates, expansion devices, compressor types, indoor and outdoor temperature and humidity levels, etc. However, the published data were limited to steady-state performance in cooling mode.

Manufacturers' product manuals provide heat-pump performance data, but they tend to be limited in the operation range and seldom give detailed defrost and cyclic data. In open publications, we rarely find a comprehensive study on a real heat-pump product in heating mode with varying multiple operating parameters; for the existing studies, researchers were inclined to study the effect of a single variable (charge, ambient temperature, etc.) on a single performance output (COP, defrosting degradation, etc.). Considering these factors, we launched an in-depth testing and simulation effort to investigate the impact of refrigerant charge levels, indoor airflows, outdoor operating conditions, and expansion devices on heating performance of an air-source heat pump. The work is split into two parts. The first part involves the steady-state, cyclic and defrosting tests on the equipment, and the second part involves component and system simulations, as compared to the steady test data.

EXPERIMENTAL WORK

The tested unit was a 3-ton R -22 split heat pump. The available product information is listed below.

* Nominal Charge: 7.13 lbs (3.234 kg)

* SEER: 10.00

* HSPF: 7.20

* Nominal tonnage (cooling): 3 tons (10.55 kW)

* Power supply: 1 PH AC/60 Hz/208/230 V

* Compressor type: Reciprocating piston

* Outdoor HX: Row number 1

Number of paralleled circuits 3

Tube number/each circuit 12

Frontal area 18.25 [ft.sup.2]

(1.694 [m.sup.2])

Tube Length 6.08 ft (1.853 m)

* Indoor HX:

Row number 3

Number of paralleled circuits 5

Tube number Circuit# 14/Circuit #1,

12/Circuit #2, 12/Circuit

#3, 16/Circuit #4, 18/

Circuit #5

Frontal area 2.97 [ft.sup.2] (0.276 [m.sup.2])

Tube Length 1.49 ft (0.454 m)

* Outdoor fan:

Nominal power 1/4 hp (186.4 W)

Nominal airflow rate 3100 cfm (1.463 [m.sup.3] /s)

* Indoor Blower:

Nominal power 1/3 hp (248.6 W)

Nominal airflow rate 1300 cfm (0.614 [m.sup.3] /s)

It shall be noted, for comparing the performances using different expansion devices, that the outdoor expansion device for heating mode was switched from a fixed-area expansion orifice (FEO) to a TXV. The unit was originally equipped with the FEO, and a suction line accumulator was installed to avoid liquid refrigerant entering the compressor. A four-way value was used to change the refrigerant flow direction from cooling to heating mode.

In summary, we performed 150 steady-state tests, which are summarized in Table 1, 18 cyclic tests summarized in Table 2, and 18 defrost tests summarized in Table 3.
Table 1. Steady-State Testing Matrix

Tested Parameter          FEO                    TXV

Refrigerant       70%, 85%, 100%,115%,  70%, 85%, 100%, 115%,
charge                 and 130% of the        and 130% of the
                         nominal level          nominal level

Indoor airflow        800, 1100, 1300,       800, 1100, 1300,
                        1500, 1750 cfm         1500, 1750 cfm

Indoor                  70[degrees]F         70[degrees]C
temperature

Outdoor             47[degrees]F, 35        47[degrees]F,
temperature           [degrees]F, 17         35[degrees]F,
                          [degrees]F          17[degrees]F

Note: 1 cfm = 0.0004719 [m.sup.3] /s; 70[degrees]F = 21.1
[degrees]C; 47[degrees]F = 8.3[degrees]C; 35[degrees]
F = 1.7[degrees]C; 17[degrees]F = -8.3[degrees]C;
33[degrees]F = 0.6[degrees]C

Table 2. Cyclic Testing Matrix

Tested Parameter          FEO                 TXV

Refrigerant        70%, 100%, and 130%      70%, 100%,
charge            of the nominal level  and130% of the
                                         nominal level

Indoor airflow     800, 1300, and 1750  800, 1300, and
                                   cfm        1750 cfm

Indoor                  70[degrees]F  70[degrees]F
temperature

Outdoor                 47[degrees]F   47[degrees]F
temperature

Note: 1 cfm = 0.0004719 [m.sup.3] /s; 70[degrees]F = 21.1
[degrees]C; 47[degrees]F = 8.3[degrees]C; 35[degrees]
F = 1.7[degrees]C; 17[degrees]F = -8.3[degrees]C;
33[degrees]F = 0.6[degrees]C

Table 3. Defrost Testing Matrix

Tested Parameter          FEO                   TXV

Refrigerant        70%, 100%, and 130%   70%, 100%, and 130%
charge            of the nominal level  of the nominal level

Indoor airflow     800, 1300, and 1750   800, 1300, and 1750
                                   cfm                   cfm

Indoor                  70[degrees]F        70[degrees]F
temperature

Outdoor                 35[degrees]F        35[degrees]F
condition           (db)/33[degrees]F   (db)/33[degrees]F
                                  (wb)                  (wb)

Note: 1 cfm = 0.0004719 [m.sup.3] /s; 70[degrees]F = 21.1
[degrees]C; 47[degrees]F = 8.3[degrees]C; 35[degrees]
F = 1.7[degrees]C; 17[degrees]F = -8.3[degrees]C;
33[degrees]F = 0.6[degrees]C


Instrumentation

The experiments were carried out in ASHRAE Standard 116 (ASHRAE 2010) psychometric chambers consisting of two insulated rooms. The temperature and humidity are independently controlled within each room to the desired operating conditions. The indoor room is equipped with an ASHRAE Standard 116 air measurement box. The indoor air stream is connected to this box to yield a measurement of the airflow rate. The indoor airflow rate is altered by a variable-speed blower inside the box. A differential pressure transducer measures the pressure drop across the nozzle, while a pressure transducer measures the pressure at the nozzle inlets. Three 1000 [OMEGA] platinum probes (resistant thermal devices) measure the air temperature near the inlet of the nozzles. Assuming no air leakage, the accuracy of the indoor airflow rate measurement is assigned to be [+ or -] 1.0 %. A grid of eight wire type T thermocouples was installed at the inlet and exit of the indoor airflow. A grid of nine thermocouples was installed at the inlet of the outdoor airflow. A grid of six thermocouples was installed at the exit of the outdoor airflow. Air temperatures were obtained by averaging the measurements of thermal couples at each location.

couples at each location.

The refrigerant pressures and temperatures were measured within refrigerant flow at the eight locations entering and leaving the compressor, condenser, expansion device, and evaporator. The pressure transducers were calibrated by way of a dead-weight tester. For the pressure transducers, the manufacturer claimed an accuracy of [+ or -] 1%. The refrigerant temperatures were measured using sheathed probe thermocouples. These thermocouples were calibrated using an ice bath and boiling water to make sure the readings were within 0.5[degrees]F (0.9[degrees]C) of the known values. A mass flowmeter (Coriolis-type sensor) was used for measuring the refrigerant mass flow rate. The manufacturer claims an accuracy of [+ or -] 2%, while the calibration demonstrated that the accuracy is [+ or -] 0.6%. Two sight glasses were installed upstream and downstream of the mass flowmeter to identify the case of two-phase flow. A chilled-mirror sensor, having an accuracy of [+ or -] 0.2[degrees]F (0.36[degrees]C), was used to measure the inlet and outlet dew point of the evaporator airflow. Atmospheric pressure was measured using a mercury barometer with an accuracy of [+ or -] 0.03 kPa (0.0044 psi). The compressor power, outdoor fan power, and indoor blower power were measured with watt transducers with an accuracy of [+ or -] 10 W (34.12 Btu/h). The mass of refrigerant charged within the system was measured with a digital scale accurate to [+ or -] 0.005 kg (0.011 1bm). The measured data were recorded using a data acquisition system. For steady-state performance tests, the sampling interval was about ten seconds. In addition, wire thermocouples were soldered to the tube bends of the outdoor and indoor coils in order to provide local refrigerant temperature measurements and allow study of the phase distribution in the heat exchangers.

Test Procedures

Steady-State Test. For steady-state tests, the temperature and humidity in the psychrometric chambers were controlled at a desired condition for one hour prior to data being recorded. The steady-state tests simultaneously employed the air-enthalpy method and refrigerant enthalpy method. For cases where two-phase refrigerant was entering the mass flowmeter or two-phase refrigerant was leaving the evaporator, only the air-enthalpy method was applied to calculate the capacity. It shall be noted that during the steady-state test, frost and apparent water condensation was avoided by controlling the outdoor humidity.

Cyclic Test. For the first step of the cyclic test, the unit was ON, and the psychrometric chambers were operated under the required indoor and outdoor conditions for at least half an hour. After that, the data acquisition system was turned on to record the steady-state data under the required conditions for half an hour. Then the unit was shut off and the psychrometric chambers ran under the required conditions for at least twenty minutes. performance data of the steady-state test was used for calculating the cyclic degradation coefficient. The second step began right after the first step. The data acquisition system was started again to record data for the cyclic test. For a cyclic test, AHRI Standard 210/240 (AHRI 2008) specifies a test cycle with 6 minutes ON and 24 minutes OFF. The ambient conditions for both the indoor and outdoor test chambers must be steadily repeatable for no less than two complete repeatable OFF/ON cycles without a break in the cycling pattern. After that, the unit should be operated through an additional OFF/ON cycle, during which the required test data are recorded. The actual tests were conducted at the laboratory for four complete repeatable OFF/ON cycles. The last cycle was used as the test cycle. The cyclic performance tests employ the air-enthalpy method only. The indoor airflow was off/on at the same time as the compressor. During the cyclic test, all indoor wet-bulb temperatures were set below 60[degrees]F (15.6[degrees]C). All outdoor dew-point temperatures were controlled below the evaporating temperature, so that frost and water condensate on the outdoor heat exchanger were strictly avoided. The data sampling intervals were less than 4 seconds.

As regulated by AHRI standard 210/240, for heating cyclic tests, test tolerances are only required for the ON period, which can be classified as test operating tolerances and test condition tolerances. The variations during each test are compared with AHRI Standard 210/240 test tolerances in Table 4 for the unit using an FEO, and in Table 5 for the unit using a TXV. In Tables 4 and 5, the condition tolerance is the maximum permissible variation of the average value of the measurement from the standard or desired test condition; the operating tolerance is the maximum permissible variation of any measurement.
Table 4. Test Tolerances of Cyclic Tests for the Unit Using FEO

           Condition                                   Operating
           Tolerance                                   Tolerance

Cyclic     Outdoor DB  Indoor DB  Charge,   Indoor    Outdoor
Test No.    Req/Act,   Req/Act,      %     Airflow,     DB,
          [degrees]  [degrees]             scfm   [degrees]
               F           F                               F

1            47/46.98    70/69.98    70%     760.08      1.11
2            47/47.23    70/69.95    70%    1222.25      1.03
3            47/47.12    70/70.10    70%    1673.10      0.70
4            47/47.06    70/69.87   100%     739.24      0.98
5            47/47.02    70/70.17   100%    1224.32      0.99
6            47/47.33    70/69.97   100%    1628.79      0.97
7            47/47.12    70/69.91   130%     744.22      1.09
8            47/47.03    70/70.11   130%    1233.57      1.78
9            47/47.07    70/70.45   130%    1635.85      0.63
Allowable       [+ or      [+ or                 --       2.0
Tolerance       -]0.5      -]0.5

           Operating
           Tolerance

Cyclic       Indoor DB,
Test No.   [degrees]F

1            1.05
2            1.15
3            0.94
4            1.26
5            1.56
6            0.97
7            1.27
8            0.84
9            0.67
Allowable    2.0
Tolerance

Note: Tolerance of 1[degrees]F = 0.55[degrees]C, cfm =
4.72 e-4 [m.sup.3]/s, 47[degrees]F = 8.3[degrees]C,
70[degrees]F = 21.1[degrees]C.

Table 5. Test Tolerances of Cyclic Tests for the Unit Using TXV

           Condition                                   Operating
           Tolerance                                   Tolerance

Cyclic     Outdoor DB  Indoor DB  Charge,   Indoor    Outdoor
Test No.    Req/Act,   Req/Act,      %     Airflow,     DB,
          [degrees]  [degrees]               scfm   [degrees]
               F           F                               F

1            47/47.23   70/70.12      70%    726.77       0.80
2            47/46.88   70/70.06      70%   1213.53       0.39
3            47/47.06   70/70.27      70%   1639.27       0.98
4            47/47.06   70/70.02     100%    753.85       0.51
5            47/46.85   70/69.96     100%   1236.41       0.76
6            47/47.01   70/70.23     100%    1640.7       0.70
7            47/47.32   70/70.15     130%    758.44       0.25
8            47/47.23   70/70.11     130%   1229.34       0.52
9            47/47.15   70/70.03     130%   1659.96       0.68
Allowable       [+ or      [+ or                 --        2.0
Tolerance       -]0.5      -]0.5

           Operating
           Tolerance

Cyclic       Indoor DB,
Test No.   [degrees]F

1               1.00
2               1.21
3               0.64
4               1.42
5               1.80
6               1.61
7               1.16
8               1.27
9               1.42
Allowable        2.0
Tolerance

Note: Tolerance of 1[degrees]F = 0.55[degrees]C, cfm =
4.72 e-4 [m.sup.3]/s, 47[degrees]F = 8.3[degrees]C,
70[degrees]F = 21.1[degrees]C.


Defrost Test. As indicated in the manufacturer's literature, the defrost control is a time/temperature control. In heating mode, if the defrost thermostat installed at the header of the outdoor heat exchanger senses a temperature below 30 [degrees]F (-1.1[degrees]C), the defrost timer begins counting. After a factory-set time period of 90 minutes, the defrost cycle (cooling mode) starts automatically. After approximately 5 minutes, the defrost cycle stops automatically. The outdoor fan is off during the defrosting period, while the indoor blower stays on.

A whole defrost test was operated over 3 hours under the required test conditions and divided into two periods: the preliminary test period and the test period. The first 1.5 hours comprised the preliminary test period and included a defrost cycle at the end. The second period was the test period, which began at the termination of the defrost cycle in the preliminary test period and ended at the termination of the next automatically occurring defrost cycle. During the test procedure, the indoor airflow rate was always on, and the resistance heater was always off.

For defrost tests, there are separate tolerances during heating and during defrosting. Test tolerances during heating apply when the heat pump is in heating mode, except for the first 5 minutes after termination of the defrosting cycle. Test tolerances during defrosting apply during the defrosting cycle and during the first 5 minutes after the termination of the defrosting cycle. The variations during each test are compared with AHRI Standard 210/240 test tolerances in Tables 6 and 7 for the unit using FEO, and in Tables 8 and 9 for the unit using TXV. The detailed definitions of test tolerances during heating and during defrosting are given in AHRI Standard 210/240. Meanwhile, all indoor wet-bulb temperatures are below 60[degrees]F (15.6[degrees]C), and the data sampling intervals are less than 10 seconds.
Table 6. Test Tolerances of Defrost Tests during Heating for the Unit
Using FEO

                             Condition Tolerance

Defrost      Outdoor DB  OutdoorDew Point   Indoor DB   Charge,  Indoor
Test No.       Req/Act,       Req/Act,        Req/Act,     %    Airflow,
            [degrees]    [degrees]        [degrees]            scfm
                 F             F                 F

1              35/35.04       30.3/30.33      70/70.12     70%    763.83
2              35/34.92       30.3/30.45      70/70.03     70%   1229.70
3              35/35.01       30.3/30.43      70/70.11     70%   1653.77
4              35/35.00       30.3/30.39      70/70.17    100%    741.94
5              35/35.05       30.3/30.37      70/70.07    100%   1230.71
6              35/34.98       30.3/30.11      70/69.91    100%   1651.79
7              35/35.19       30.3/30.42      70/70.14    130%    772.15
8              35/34.95       30.3/30.42      70/70.11    130%   1240.13
9              35/35.11       30.3/30.55      70/70.02    130%   1654.80
Allowable         [+ or            [+ or         [+ or      --        --
Tolerance         -]1.0            -]0.7         -]0.5

           Operating Tolerance

Defrost     Outdoor    Outdoor   Indoor DB,
Test No.      DB,       Dew T,    [degrees]
          [degrees] [degrees]      F
               F          F

1               1.62       1.43        1.34
2               1.65       2.34        1.26
3               1.80       1.86        1.52
4               1.57       1.50        1.23
5               1.34       1.37        1.39
6               2.08       2.68        1.38
7               1.97       1.61        2.01
8               1.55       1.49        1.52
9               1.56       1.87        1.20
Allowable        2.0        1.5         2.0
Tolerance

Note: Tolerance of 1[degrees]F = 0.55[degrees]C, cfm = 4.72 e-4
[m.sup.3] /s, 35[degrees]F = 1.7[degrees]C, 30[degrees]
F = -1.11[degrees]C, 70[degrees]F = 21.1[degrees]C.

Table 7. Test Tolerances of Defrost Tests during Defrost for the
Unit Using FEO

            Operating Tolerances

Defrost    Outdoor DB,  Indoor DB,
Test No.   [degrees]  [degrees]
                F           F

1                 3.96        2.37
2                 3.62        1.20
3                 3.89        2.30
4                 5.11        1.11
5                 4.33        2.62
6                 5.37        1.26
7                 4.80        3.28
8                 6.94        1.50
9                 4.20        1.61
Allowable         10.0         4.0
Tolerance

Note: Tolerance of 1[degrees]F = 0.55[degrees]C.

Table 8. Test Tolerances of Defrost Tests during Heating for
the Unit Using TXV

           Condition Tolerance

Defrost    Outdoor DB  Outdoor Dew  Indoor DB   Charge,
Test No.     Req/Act      Point Req     Req/Act        %
          [degrees]       /Act      [degrees]
               F       [degrees]       F
                            F

1            35/35.08   30.3/30.34    70/70.14      70%
2            35/34.98   30.3/30.25    70/70.03      70%
3            35/35.01   30.3/30.42    70/70.05      70%
4            35/34.85   30.3/30.54    70/70.32     100%
5            35/35.05   30.3/30.51    70/70.02     100%
6            35/35.14   30.3/30.20    70/70.00     100%
7            35/35.04   30.3/30.37    70/70.03     130%
8            35/35.00   30.3/30.66    70/69.96     130%
9            35/35.07   30.3/30.38    70/69.99     130%
Allowable       [+ or        [+ or  [+ or -]0.       --
Tolerance       -]1.0        -]0.7           5

                       Operating Tolerance

Defrost      Indoor    Outdoor DB,   Outdoor     Indoor DB
Test No.     Airflow,  [degrees]     Dew T,    [degrees]
              scfm          F      [degrees]F      F

1            745.36         1.43          1.79       1.26
2           1222.18         1.76          2.86       1.74
3           1638.64         1.63          1.74       1.49
4            735.16         2.13          2.65       1.55
5           1224.81         2.08          1.81       1.60
6           1644.26         2.25          1.88       1.42
7            744.37         1.66          1.50       1.65
8           1226.60         0.63          1.04       0.79
9           1638.29         1.81          1.49       1.72
Allowable        --         2.0           1.5        2.0
Tolerance

Note: Tolerance of 1[degrees]F = 0.55[degrees]C, cfm =
4.72 e-4 [m.sup.3] /s, 35[degrees]F = 1.7[degrees]C,
30[degrees]F = -1.11[degrees]C, 70[degrees]F =
21.1[degrees]C.

Table 9. Test Tolerance of Defrost Tests during Defrost for
the Unit Using TXV

                   Operating Tolerances

Defrost Test No.  Outdoor DB,  Indoor DB,
                 [degrees]F [degrees]F

1                        4.64         2.02
2                        4.73         1.61
3                        2.93         1.29
4                        5.18         1.74
5                        3.42         2.29
6                        4.80         1.41
7                        4.51         1.43
8                        6.13         1.59
9                        3.54         1.87
Allowable                10.0          4.0
Tolerance

Note: Tolerance of 1[degrees]F = 0.55[degrees]C.


In Tables 6 and 8, there are several points where the outdoor dew-point variations were not controlled in the tolerance. These are due to the limitation of the test facility for maintaining the humidity in the outdoor chamber, which is a factor common to all the heat-pump equipment setups in this study. The variations appear acceptable, as most points controlled the range within 2[degrees]F (1.2[degrees]C). Defrosting testing is quasi-steady state, as it is more important to control the aver-age chamber temperatures. The defrosting data quality shall be sufficient to tell the performance difference when varying the charge levels, indoor airflow rates, and expansion devices.

Data Reduction

The air and refrigerant enthalpies were calculated using thermodynamic property functions in the Engineering Equation Solver (Klein 2004). The air enthalpy was determined from the measured dry-bulb temperature and dew point at a corresponding location. The refrigerant enthalpy was determined from the local pressure and temperature measurements. However, the refrigerant enthalpy of a two-phase mixture state could not be determined with the available measurements. In this case, only the air-side measurements and calculations were available.

Steady-State Data. The air-side total heating capacity was calculated as

[Q.sub.heating, air tot] = mair([h.sub.out, air, indoor] - [h.sub.in, air, indoor]) (1)

where hin, air, indoor and hout, air indoor were determined using the measured inlet and outlet temperatures of the indoor airflow as dry air. mair is the dry air mass flow rate.

The refrigerant side total heating capacity was calculated as

[Q.sub.heating, ref., tot]_

=[m.sub.ref] ([h.sub.in, ref, indoor] - [h.sub.out, ref, indoor]) + [power.sub.blower] (2)

Where

[h.sub.in, ref, indoor] = refrigerant enthalpy determined by the pressure and temperature entering the indoor coil

[h.sub.out,ref,indoor] = refrigerant enthalpy determined by the pressure and temperature leaving the indoor coil

[m.sub.ref] = refrigerant mass flow rate measured by the micro-motion mass flowmeter.

Coefficient of performance is given as

COP = [Q.sub.heating,air, tot]/([Power.sub.comp] + [Power.sub.fan] + [Power.sub.blower]) (3)

Cyclic Data. The following equations are used to calculate cyclic degradation coefficient:

[C.sub.D] = 1-[COP.sub.cyc] (47) / [COP.sub.ss] (47) /1- HLF (4)

HLF = [Q.sub.cyc] (47) /[Q.sub.ss] (47) * [tau] (5)

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] (6)

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] (7)

[ COP.sub.cyc ] (47) = [ Q.sub.cyc ] (47) /[ E.sub.cyc] (47) (8)

Where

[C.sub.D] = cyclic degradation coefficient

[tau] = duration of time for one complete cycle consisting of one compressor ON time and one compressor OFF time, which is 0.5 h

HLF = heating load factor

[Q.sub.ss] (47) and [COP.sub.ss] (47) = net steady-state heating capacity and COP, respectively, determined with the indoor air-enthalpy method at the same conditions as the cyclic test

47 = outdoor dry-bulb temperature at 47[degrees]F (8.3[degrees]C)

[Q.sub.cyc] (47) = net heating amount determined with the indoor air-enthalpy method during the ON period (6 min) of the test cycle

[E.sub.cyc] (47) = total energy usage during the ON period (6 min) of the test cycle

[COP.sub.cyc] (47) = COP for a cyclic test, determined using the indoor air-enthalpy method with data from the ON period (6 min) of the test cycle

Q (t) = averaged net heating capacity during one recording interval (less than 4 seconds), determined by averaging the heating capacities calculated with the measurements at the beginning and at the end of the recording interval

E(t) = averaged total power consumption during one recording interval, determined by averaging the total power consumptions measured at the beginning and at the end of the recording interval Defrost Data.The defrost performance data are calculated as

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] (9)

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] (10)

[COP.sub.DEF] (35) = [Q.sub.DEF] (35) / [E.sub.DEF] (35) (11)

where

[Q.sub.DEF] (35) = total net heating done over the test period between two defrost terminations, divided by the total length of the test period, and determined with the indoor air-enthalpy method

35 = outdoor dry-bulb temperature at 35[degrees]F (1.7[degrees]C)

[E.sub.DEF] (35) = total energy usage during the period between two defrost terminations, divided by the total length of the test period

[COP.sub.DEF] (35) = performance of coefficient during each defrost test

Q(t) = averaged net heating capacity during one recording interval (less than 10 seconds), determined by averaging the heating capacities calculated with the measurements at the beginning and at the end of the recording interval

E(t) = averaged total power consumption during one recording interval, determined by averaging the total power consumptions measured at the beginning and at the end of the recording interval

Uncertainty Analysis

The following method was used to determine the uncertainty in the calculated quantities determined from test measurements. For a given dependent variable,

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] (12)

where

[omega] A= total uncertainty associated with a dependent variable A

[ZETA]i = one of the independent variables which impacts the dependent variable A

([omega].sub.[[ZETA].sub.i]]) = uncertainty associated with an independent variable [[ZETA].sub.i]

Table 10 gives the uncertainties of the dependent variables in a steady-state test at a refrigerant charge of 100%, outdoor temperature of 35[degrees]F (1.67[degrees]C), and indoor airflow rate of 1300 cfm (0.614 [m.sup.3]/s) for the heat-pump unit using a TXV. The data for other steady-state tests have similar uncertainties. The relative uncertainties for cyclic and defrost tests should be larger than those for steady-state tests, but they are difficult to determine. As shown in Table 10, the measurements on the refrigerant side are much more accurate than the measurements on the air side.
Table 10. Uncertainties of Dependent Variables

                      Uncertainty (Absolute
                          or Relative)

Sub cooling degree      [+ or -]1.1[degrees]
                          F (0.6[degrees]C)

Superheat degree      [+ or -]1.1[degrees]F

Air-side total                  [+ or -]11.0%
heating capacity

Refrigerant side                 [+ or -]1 0%
total heating
capacity

COP (determined with            [+ or -]12 0%
air-side
measurements)


During the steady-state tests, the energy balances were in good shape according to AHRI Standard 210/240. The differences between refrigerant-side and air-side capacities were smaller than 6%.

RESULTS AND DISCUSSIONS

Steady-State Tests

Tables 11-13 present the steady-state performances at outdoor temperatures of 17[degrees]F (-8.33[degrees]C), 35[degrees]F (1.67[degrees]C), and 47[degrees]F (8.33[degrees]C). The performance data are listed as air-side total heating capacity, compressor power consumption, system COP, subcooling degree (measured at the condenser exit), and superheat degree (measured at the evaporator exit).
Table 11. Steady-State System Performance at Outdoor
Temperature of 17[degrees]F (-8.3[degrees]C)

Expansion                    FEO

                           Heating Capacity, Btu/h

Charge/    1750   1525    1300 cfm   1100    820   1750   1525
Indoor      cfm    cfm                cfm    cfm    cfm    cfm
Airflow

70%        20912  20809       19368  18168  17750  21294  20649
85%        20845  21151       19378  18056  17693  22034  21605
100%       21036  20931       19809  18756  18216  22270  21995
115%       21228  20597       19451  18063  17786  21965  21407
130%       21208  20642       19637  18329  18062  20909  20822

                         Compressor Power, kW

70%         1.96   1.97        1.99   2.01   2.06   1.95   1.96
85%            2   2.01        2.02   2.04   2.06   1.97   1.98
100%        2.02   2.04        2.05   2.06    2.1   1.97   1.99
115%        2.03   2.03        2.04   2.05   2.08   1.96   1.98
130%        2.02   2.02        2.04   2.05   2.08   2.01   2.02

                                COP

70%         2.11   2.12        2.03   1.96   1.95   2.17   2.15
85%         2.06   2.15        2.01   1.92   1.94   2.23   2.23
100%        2.07    2.1        2.03   1.97   1.96   2.25   2.26
115%         2.1   2.07           2   1.92   1.94   2.23   2.21
130%        2.11    2.1        2.02   1.95   1.97   2.09   2.12

                         Subcooling Degree, K

70%         0.12   0.09        0.10   0.09   0.02   0.52   0.53
85%         0.23   0.19        0.21   0.16   0.09   2.09   2.31
100%        0.24   0.18        0.16   0.12   0.06   6.27   6.57
115%        0.29   0.28        0.24   0.20   0.07   9.24   9.64
130%        0.17   0.20        0.16   0.10   0.04  15.82  16.11

                          Superheat Degree, K

70%         0.14   0.56        0.29   0.51   0.63   6.39   6.48
85%         0.06   0.43        0.64   0.79   0.89   5.30   5.17
100%        0.39   0.53        0.59   0.67   0.72   5.01   5.38
115%        1.14   1.23        1.25   1.29   1.31   5.42   5.19
130%        0.83   0.97        1.27   1.28   1.30   5.09   5.14

Expansion            TXV

              Heating Capacity, Btu/h

Charge/     1300    1100 cfm   820 cfm
Indoor       cfm
Airflow

70%         19437      18519    18540
85%         20093      19230    19068
100%        20414      19776    18871
115%        20098      19109    18890
130%        19249      18200    18183

                  Compressor Power, kW

70%         1.96        1.98     2.01
85%         1.99        2.01     2.03
100%        1.98           2     2.04
115%        1.99           2     2.02
130%        2.01        2.02        2

                      COP

70%         2.07        2.02     2.07
85%         2.11        2.07     2.12
100%        2.16        2.14     2.08
115%        2.11        2.07     2.11
130%        2.01        1.95     2.04

                  Superheat Degree, K

70%         0.51        0.45     0.33
85%         2.46        2.52     3.32
100%        6.83        7.03     8.13
115%        9.89       10.37    12.41
130%       16.56       17.19    19.18

                  Superheat
                  Degree, K

70%         6.59        6.38     5.81
85%         5.36        5.28     5.19
100%        5.49        5.26     5.09
115%        5.17        5.31     5.16
130%        5.28        5.18     5.02

Note: Btu/h = 0.2931 W, cfm = 4.72 e-4 [m.sup.3]
/s, K = 1.8[degrees]R.

Table 12. Steady-State System Performances at Outdoor Temperature
of 35[degrees]F (1.7[degrees]C)

Expansion                      FEO

                    Heating                                Heating
                  Capacity,                               Capacity,
                    Btu/h                                   Btu/h

Charge/     1750    1525 cfm   1300   1100    820   1750    1525 cfm
Indoor       cfm                cfm    cfm    cfm    cfm
Airflow

70%        28229       28534  27199  26410  25890  29525       28829
85%        30299       30066  28314  26737  26479  30074       29988
100%       30482       30043  28471  27044  24888  31814       30673
115%       30766       30175  28370  26966  26255  31254       30620
130%       30963       30624  28321  26764  26394  30726       29962

                  Compressor                              Compressor
                  Power, kW                               Power, kW

70%          2.2        2.24   2.27   2.31   2.47   2.26         2.3
85%          2.3        2.33   2.35   2.39    2.5   2.29        2.31
100%        2.34        2.36   2.38   2.39   2.49   2.31        2.34
115%        2.35        2.37   2.38   2.41   2.51   2.34        2.35
130%        2.37        2.39   2.41   2.42   2.51   2.42        2.43

                      COP                                       COP

70%         2.64        2.68    2.6   2.56   2.46   2.73        2.68
85%         2.75        2.75   2.63   2.53    2.5   2.75        2.78
100%        2.73        2.71   2.62   2.56   2.35   2.89        2.81
115%        2.74        2.73   2.62   2.53   2.47   2.82         2.8
130%        2.75        2.75   2.59   2.51   2.48    2.7        2.68

                  Subcooling                              Subcooling
                  Degree, K                               Degree, K

70%         0.22        0.11   0.07   0.02  -0.08   0.33        0.32
85%         2.20        1.83   0.91   0.23  -0.04   2.25        2.41
100%        5.18        4.29   3.53   2.28   0.06   7.89        7.96
115%        5.06        3.93   3.28   1.52  -0.03  12.20       12.77
130%        5.76        5.05   4.02   2.69   0.00  19.43       19.92

                  Superheat                               Superheat
                  Degree, K                               Degree, K

70%         9.14        8.61   7.98   7.23   0.13   5.25        4.78
85%         3.16       -0.03  -0.12   0.20  -0.06   3.76        3.69
100%       -0.01        0.07   0.07   0.07   0.09   4.17        3.74
115%       -0.07       -0.01  -0.18   0.03   0.06   4.03        4.02
130%       -0.68       -0.44  -0.28  -0.18  -0.08   3.68        3.76

Expansion            TXV

                    Heating
                  Capacity,
                    Btu/h

Charge/     1300    1100 cfm   820
Indoor       cfm               cfm
Airflow

70%         27204     25719   25893
85%         28053     26654   26800
100%        28331     27499   27761
115%        28559     27161   27047
130%        27744     26581   26896

                  Compressor
                  Power, kW

70%         27744     26581   26896
85%         2.33       2.36    2.45
100%        2.34       2.39    2.49
115%        2.37        2.4    2.48
130%        2.45       2.47    2.51

                      COP

70%         2.57       2.46    2.47
85%         2.65       2.55    2.57
100%        2.66        2.6    2.62
115%        2.66       2.58    2.57
130%        2.53       2.47    2.54

                  Superheat
                  Degree, K

70%         0.26       0.17    0.18
85%         2.65       2.79    3.72
100%        8.94       9.47   11.40
115%       13.28      13.97   16.44
130%       20.56      21.30   24.31

                  Superheat
                  Degree, K

70%        4.72        4.47    5.10
85%        4.19        3.87    3.64
100%       5.97        5.45    4.87
115%       4.26        4.15    3.70
130%       3.49        3.73    3.52

Note: Btu/h = 0.2931 W, cfm = 4.72 e-4 [m.sup.3]
/s, K = 1.8[degrees]R.

Table 13. Steady-State System Performance at Outdoor Temperature
of 47[degrees]F (8.3[degrees]C)

Expansion                      FEO

                    Heating                            Heating
                  Capacity,                        Capacity,
                    Btu/h                            Btu/h

Charge/     1750    1525 cfm   1300   1100    820    1750 cfm   1525
Indoor       cfm                cfm    cfm    cfm                cfm
Airflow

70%        29695       30180  28339  27751  28459       34942  34779
85%        33422       33518  31972  30966  32252       37020  35998
100%       35020       35020  34570  33060  31735       38598  37160
115%       37828       37043  34587  31817  31268       38289  37168
130%       37738       36307  33948  31432  31000       37743  36709

                  Compressor                       Compressor
                  Power, kW                        Power, kW

70%         2.25        2.29   2.34    2.4   2.61        2.44   2.48
85%         2.37        2.42   2.46   2.53   2.74         2.5   2.53
100%        2.52        2.56   2.61   2.66   2.79        2.52   2.56
115%        2.57         2.6   2.64   2.68    2.8         2.6   2.64
130%        2.61        2.62   2.64   2.66   2.78        2.72   2.77

                      COP                                COP

70%         2.73        2.79   2.66   2.62    2.6        3.06   3.07
85%         2.99           3   2.89   2.82   2.83        3.18   3.13
100%        2.98        2.99   3.01   2.88   2.74         3.3    3.2
115%        3.17        3.15   2.97   2.77    2.7        3.19   3.12
130%        3.14        3.05    2.9   2.75   2.69        3.05   2.98

                  Subcooling                       Subcooling
                  Degree, K                        Degree, K

70%         0.66        0.50   0.26   0.17  -0.02        0.30   0.14
85%         3.43        3.39   3.34   3.09   1.86        1.02   1.82
100%       10.26       10.14   7.13   9.11   6.36        7.67   8.74
115%       11.68       11.24  10.73  10.13   5.88       13.90  14.53
130%       13.41       13.13  12.29  10.93   6.98       20.98  22.57

                  Superheat                        Superheat
                  Degree, K                        Degree, K

70%        13.84       13.57  13.22  12.62  10.48        8.51   8.09
85%        12.15       11.48  11.02  10.09   6.12        4.34   4.41
100%        7.27        5.88   0.97  -0.08  -0.18        3.48   6.01
115%        0.06        0.11   0.21   0.25   0.27        3.84   3.49
130%        0.01        0.02   0.14   0.25   0.34        2.83   5.14

Expansion            TXV

                    Heating
                  Capacity,
                    Btu/h

Charge/     1300    1100 cfm    820
Indoor       cfm                cfm
Airflow

70%         31788      30087   30939
85%         33734      32292   32511
100%        34771      32986   33167
115%        35280      33726   33344
130%        33940      32437   33054

                  Compressor
                  Power, kW

70%          2.52       2.56     2.8
85%          2.57       2.63    2.82
100%         2.59       2.66    2.83
115%         2.67       2.73    2.88
130%         2.78       2.81    2.92

                      COP

70%          2.84       2.71    2.67
85%          2.95       2.85    2.79
100%         3.02       2.89    2.83
115%            3       2.88    2.81
130%         2.79       2.72    2.75

                  Superheat
                  Degree, K

70%          0.35       0.30    0.11
85%          1.78       2.06    3.25
100%         8.91       9.69   11.62
115%        14.97      15.90   19.16
130%        22.84      24.27   27.40

                  Superheat
                  Degree, K

70%          7.97       7.35    5.27
85%          4.24       4.01    3.84
100%         4.71       5.45    4.59
115%         3.72       3.51    3.10
130%         2.61       4.87    2.26

Note: Btu/h = 0.2931 W, cfm = 4.72 e-4 [m.sup.3]
/s, K = 1.8[degrees]R.


Changes of Charge Mass under Different Ambient Temperatures in an FEO System. As indicated in Figure 1, for the fixed-orifice system, in the case of no superheat degree at the exit of the evaporator, two-phase refrigerant filled up the suction line accumulator. For an outdoor temperature of 17[degrees]F (-8.33[degrees]C), there was no superheat in all cases. For an outdoor temperature of 35[degrees]F (1.67[degrees]C), there was no superheat for refrigerant charge over 85% of nominal. Superheat degree existed for an outdoor temperature of 47[degrees]F (8.33[degrees]C) until the charge was over 115%. When liquid refrigerant begins to enter the suction line accumulator, the system performance indices (compressor power, heating capacity, COP, and subcooling degree) are less sensitive to variations in charge. As the refrigerant storage within the suction line accumulator becomes important for outdoor temperatures of 17[degrees]F (-8.33[degrees]C) and 35[degrees]F (1.67[degrees]C), the impacts of charge variation on heating capacity, compressor power, and COP under 17[degrees]F (-8.33[degrees]C) and 35[degrees]F (1.67[degrees]C) are much smaller than for those under 47[degrees]F (8.33[degrees]C).

At an outdoor temperature of 47[degrees]F and a refrigerant charge of 115%, liquid refrigerant begins to be stored within the accumulator. As a result, the compressor power was constant from a charge of 115% to a charge of 130%. However, the accumulator didn't appear to have a significant impact in this case, since the subcooling degree was still increasing and heating capacity kept dropping, as indicated in Figure 2. Under an outdoor temperature of 47[degrees]F (8.33[degrees]C), the maxi-mum COP occurred at 100% of the nominal charge. COP decreases with charge increasing from 100% to 130%, mainly due to a drop in heating capacity.

Changes of Charge Mass under Different Ambient Temperatures in a TXV System. The TXV is used to control the exit superheat degree of the outdoor heat exchanger in heating mode. The bulb of the TXV and its equalizer were installed at the outlet of the outdoor heat exchanger and before the reversing valve. In cases where the TXV performed well, the superheat degree could be controlled between 9[degrees]R (subcool ed) to 12[degrees]R (6.7 K). However, hunting could be observed in multiple cases when the averaged evaporator outlet superheat degree was generally less than 9[degrees]R (5 K). At a charge of 70%, an outdoor temperature of 47[degrees]F (8.33[degrees]C) and indoor airflow rate larger than 1100 cfm (0.519 [m.sup.3] /s), the evaporator outlet superheat degree was larger than 12[degrees]R (6.67 K), and the TXV was probably fully opened.

Figures 3-5 present COP comparisons of the heat-pump unit using a TXV and using a fixed-area expansion device, as a function of charge mass under different outdoor temperatures and an indoor airflow rate of 1300 cfm (0.614 [m.sup.3] /s). Figures 6 and 7 show the total air-side heating capacities and compressor powers under an outdoor temperature of 35[degrees]F (1.67[degrees]C) and an indoor airflow rate of 1300 cfm (0.614 [m.sup.3] /s), as a function of charge mass. Figure 8 shows the comparison in subcooling degree at the exit of the indoor heat exchanger between the TXV system and FEO system, as a function of charge mass. Figure 9 shows subcooling degree measured on tube bend surfaces along Circuit #4 in the indoor heat exchanger of the TXV system. In Figure 9, positive values indicate a superheated area of the heat exchanger and negative values indicate subcooled areas, allowing some inaccuracies due to the wire thermocouples soldered to the surface.

For the TXV system, COPs as a function of charge mass under different outdoor temperatures tend to show a similar trend. In this case, COP increases with charge mass from 70% to 100%. When charge mass ranges from 100% to 110% under different outdoor temperatures, COP appears to reach a maxi-mum value, then it drops with increasing charge mass.

As shown in Figure 3, variations in COPs of the TXV system and the FEO system both appear to be sensitive to charge mass. Under an outdoor temperature of 47[degrees]F (8.33[degrees]C), the FEO system has superheat degree for most of the charge masses (with a charge range from 70% to 115%). Hence, the FEO system performs like the TXV system. The performance of the TXV system is better than that of the FEO system except at a charge of 130%.

As indicated in Figures 4 and 5, under outdoor temperatures of 17[degrees]F (-8.33[degrees]C) and 35[degrees]F (1.67[degrees]C), the performance of the TXV system is more sensitive to variations in charge mass. In this case, the COP of the TXV system is larger than that of the FEO system in most cases. Under an outdoor temperature of 17[degrees]F (-8.33[degrees]C), the performance of the TXV system is worse than that of the FEO system, at a charge of 130%. Under an outdoor temperature of 35[degrees]F (1.67[degrees]C), the performance of the TXV system is worse than that of the FEO system, at charges of 130% and 70%. At 17[degrees]F (-8.33[degrees]C) for the TXV system, the deviation between maximum and minimum COP is 7%. At 35[degrees]F (1.67[degrees]C) for the TXV system, the deviation between maximum and minimum COP is 5%. At the lowest charge level for the TXV system, the superheat degree could be larger than the setting. In this case, the TXV is probably fully opened, and the advantage of TXV is not evident. At the highest charge level for the TXV system, refrigerant accumulates in the condenser, which makes the heat transfer in the condenser significantly degraded.

[FIGURE 2 OMITTED]

[FIGURE 3 OMITTED]

[FIGURE 4 OMITTED]

[FIGURE 5 OMITTED]

In Figure 8, the subcooling degree of the TXV system increases linearly with charge mass. However, the subcooling degree of the FEO system increases much slower with charge mass. With the charge mass over 100%, the subcooling degree of the FEO system is nearly unchanged, since the added liquid refrigerant mainly resides in the suction line accumulator. Refrigerant does not accumulate in the accumulator for the TXV system, since it is always superheated vapor entering the accumulator.

Figure 9 presents local subcooling degrees on tube bends along Circuit #4 of the indoor heat exchanger under different charges in the TXV system. Drastic changes in local subcooling degree indicate transition from vapor phase to two-phase or transition from two-phase to liquid phase. Consequently, Figure 9 can show a rough shape of the phase distribution. At a charge of 70%, there is no transition from two-phase to liquid phase, and the whole heat exchanger is a mixture of vapor and two-phase refrigerant. At a charge of 100%, the liquid phase appears to start from tube bend No. 5. At a charge of 130%, the liquid phase appears to start from tube bend No. 3, when 80% of the heat exchanger is filled with liquid. Hence, for the TXV system, most of the added liquid charge accumulates in the condenser, which causes less effective heat transfer and higher condensing pressure than for an FEO system. Higher condensing pressure would reduce compressor mass flow rate and increase compressor power, which would at last impact heating capacity and COP. As shown in Figures 6 and 7, the TXV system has a smaller heating capacity but higher compressor power at a charge of 130% under 35[degrees]F (1.67[degrees]C) than does the FEO system, which is primarily due to significantly higher condensing pressure. From a charge of 115% to a charge of 130%, the reduction in mass flow rate is 5%, which leads to a 3% reduction in heating capacity.

[FIGURE 1 OMITTED]

[FIGURE 6 OMITTED]

[FIGURE 7 OMITTED]

[FIGURE 8 OMITTED]

[FIGURE 9 OMITTED]

Cyclic Tests

Cyclic performance data for each cyclic test is listed in Table 14 for the unit using FEO and in Table 15 for the unit using TXV. The degradation coefficient of the TXV system at each condition is smaller than that of the FEO system, which indicates that the TXV system has smaller cyclic penalties. The TXV system keeps most of the system charge in the condenser when the system is off. The refrigerant confined in the condenser is in a two-phase state and is stored at a high pressure. However, unlike the TXV system, the FXO system balances the high- and low-side pressures, and much of the system charge migrates from the high side to the low side when the system is off, which causes energy loss and an increase in entropy.
Table 14. Cyclic Performance Data of the Unit Using FEO

Test  [Q.sub.ss]  [Q.sub,cyc]  HLF, --  [COP.sub.ss]  [COP.sub.cyc]
No.     (47),      (47), Btu             (47), --        (47), --
        Btu/h

1          30365      2213.26  0.1458          2.69           2.08
2          28918       2254.6  0.1559          2.74           2.20
3          31430      2667.92  0.1698          2.86           2.43
4          32506      2266.35  0.1394          2.74           2.08
5          34535       2609.9  0.1511          2.99           2.40
6          37226      2844.25  0.1528          3.21           2.57
7          31199      2417.72  0.1550          2.65           2.17
8          34996      2768.49  0.1582          2.99           2.50
9          37786      2973.71  0.1574          3.14           2.62

Test  [C.sub.D], --
No.

1         0.2638
2         0.2333
3         0.1792
4         0.2782
5         0.2328
6         0.2335
7         0.2153
8         0.1932
9         0.1962

Note: Btu = 1055 J, h = 3600 s, Btu/h = 0.2931 W.

Table 15. Cyclic Performance Data of the Unit Using TXV

Test  [Q.sub.ss]  [Q.sub,cyc]  HLF, --  [COP.sub.ss]  [COP.sub.cyc]
No.     (47),      (47), Btu             (47), --        (47), --
        Btu/h

1         31041      2494.48    0.1607        2.65           2.16
2         32633      2914.86    0.1786        2.91           2.58
3         34942      3171.67    0.1815        3.06           2.74
4         33269      2636.15    0.1585        2.82           2.25
5         34771      3087.62    0.1776        3.02           2.70
6         38598      3396.92    0.1760        3.30           2.85
7         33156      2544.71    0.1535        2.73           2.20
8         33940      3102.56    0.1828        2.79           2.69
9         37743      3417.87    0.1811        3.05           2.87

Test  [C.sub.D], --
No.

1          0.2200
2          0.1376
3          0.1266
4          0.2410
5          0.1295
6          0.1656
7          0.2288
8          0.0452
9          0.0720

Note: Btu = 1055 J, h = 3600 s, Btu/h = 0.2931 W.


The FEO and TXV systems show the same tendency for the degradation coefficient to increase with a reduction of door airflow rate. There are two factors that lead to this result. First of all, the steady-state heating capacity increases significantly with increasing indoor airflow rates, while the energy penalty associated with cycling is less sensitive to airflow. Therefore, the energy loss becomes a smaller fraction of total heating capacity with increasing indoor airflow rate. Secondly, increasing airflow leads to a smaller difference between high- and low-side pressures, which reduces the energy loss associated with cycling.

The degradation coefficients of the TXV system are significantly smaller at a charge level of 130% and indoor airflow rate larger than 1300 cfm (0.614 [m.sup.3]/s) (Tests No. 8 and 9). Tests were conducted again to validate the results, which were proved to be repeatable.

Defrost Tests

The defrost performance data for each defrost test is listed in Table 16 for the FEO unit and in Table 17 for the unit using a TXV. The defrost performance data is compared with the steady-state performance data conducted at similar indoor and outdoor conditions and indoor airflow rate. The heating capacity and COP of the relative steady-state test are presented as Q(35) and COP (35). The ratio of the defrost heating capacity to the steady-state heating capacity is presented as [Ratio.sub.Q], and the ratio of the defrost COP to the steady-state COP is presented as [Ratio.sub.COP].
Table 16. Defrost Performance Data of the Unit Using FEO

Test  Period,    Q    [Q.sub.DEF](35), Btu/h  [R.sub.Q]  COP(35),
No.      s     (35),                             %
                 Btu/h

1        5910  25992                23533.9     0.9054      2.45
2        5800  27199                22201.3     0.8163      2.60
3        5810  28229                24109.8     0.8541      2.64
4        5790  24990                23720.9     0.9494      2.34
5        5855  28471                25121.9     0.8824      2.62
6        5755  30482                26937.6     0.8837      2.73
7        5935  26496                22784.6     0.8599      2.47
8        5750  28321                25452.4     0.8987      2.59
9        5770  30963                25492.3     0.8233      2.75

Test  [COP.sub.DEF], (35)  [Ratio.sub.COP], %
No.            --

1            2.28                   0.9286
2            2.22                   0.8522
3            2.30                   0.8698
4            2.28                   0.9687
5            2.37                   0.9049
6            2.49                   0.9124
7            2.19                   0.8851
8            2.39                   0.9218
9            2.35                   0.8556

Note: Btu = 1055 J, h = 3600 s, Btu/h = 0.2931 W.

Table 17. Defrost Performance Data of the Unit Using TXV

Test  Period, s  Q(35),  [Q.sub.DEF],  [Ratio.sub.Q]  COP(35),
No.              Btu/h       Btu/h             %            --

1         5900    25995       22539.3         0.8671      2.46
2         5840    27204       23995.8         0.8821      2.57
3         5905    29525       24305.8         0.8232      2.73
4         5875    27863       23196.1         0.8325      2.61
5         5850    28331       23977.5         0.8463      2.66
6         5825    31814       25925.6         0.8149      2.89
7         5805    26998       24105.1         0.8928      2.52
8         5875    27744       23688.8         0.8538      2.53
9         5875    30726       26407.2         0.8594      2.70

Test  [COP.sub.DEF], (35)  [Ratio.sub.COP], %
No.            --

1            2.22                 0.9025
2            2.33                 0.9049
3            2.32                 0.8492
4            2.27                 0.8710
5            2.33                 0.8753
6            2.42                 0.8387
7            2.31                 0.9191
8            2.23                 0.8807
9            2.41                 0.8938

Note: Btu = 1055 J, h = 3600 s, Btu/h = 0.2931 W.


Figures 10-12 present superheat degree variations at the exit of the outdoor heat exchanger in the unit using a TXV at different operating conditions. Frost formation leads to hunting of the TXV in all situations. Higher charge inventory and indoor airflow rate leads to longer hunting times. At the end of a frost formation test, there is no superheat degree at the exit of the outdoor heat exchanger, even when using a TXV. During frost formation tests, at low charge, the performances of the two systems are not very different; at high charge, performances of the TXV system tend to be better than those of the fixed-orifice system. At high charge, there was no superheat degree in the unit using an FEO, while the unit using a TXV could still maintain a superheat degree at the beginning of the defrosting period. In this case, frost could form more quickly on the evaporator of the unit using an FEO, for it had a lower mean evaporating temperature.

[FIGURE 10 OMITTED]

[FIGURE 11 OMITTED]

[FIGURE 12 OMITTED]

CONCLUSIONS

Steady-State Tests

1. The suction line accumulator had a noticeable effect on system performance of the unit using an FEO. At low outdoor air temperature, low indoor airflow rate, and high charge inventory, two-phase refrigerant exited the outdoor heat exchanger, and then liquid refrigerant was stored in the suction line accumulator. After liquid refrigerant begins to reside in the accumulator, the system performance is insensitive to variations in charge inventory. The variations of COP were fairly small at outdoor temperatures of 17[degrees]F (-8.3[degrees]C) and 35[degrees]F (1.7[degrees]C) as compared to that at outdoor temperature of 47[degrees]F (8.3)[degrees]C), since there were superheat degrees in most charges at 47[degrees]F.

2. For the unit using a TXV, superheat always existed at the exit of the outdoor heat exchanger, so the suction line accumulator had a minor effect. The variation of charge inventory affected the system performances significantly as compared to the system using an FEO. The largest COPs occurred around 100% to 110% of nominal charge. At high charge, a significant amount of liquid refrigerant accumulated in the condenser, which led to degradations in heating capacity and COPs and increased the compressor power consumption.

3. For the unit using a TXV, the superheat degree could be controlled from 9[degrees]R (5 K) to 12[degrees]R (6.7 K), but hunting could be observed in multiple cases when the averaged evaporator outlet superheat degree was generally less than 9[degrees]R. At a charge of 70%, an outdoor temperature of 47[degrees]F (8.3[degrees]C) and indoor airflow rate larger than 1100 cfm (0.519[m.sup.3] /s), the evaporator outlet superheat degree was larger than 12[degrees]R (6.67 K), and the TXV was probably fully opened.

4. There was a 2~3 K (3.6~5.4[degrees]R) temperature change in the refrigerant flowing across the reversing valve. This indicates that there is noticeable heat transfer between the suction vapor and discharge vapor in the reversing valve.

5. The COPs of the TXV system are higher than those of the FEO system, except for the overcharged cases, when the redundant charge lodges in the accumulator for the FEO system but accumulates in the condenser coil in the TXV system.

6. The system performance indices of the FEO system are less sensitive than those of the TXV system when varying the charge, indoor airflow rate, and outdoor temperature, due to the buffering effect of the suction line accumulator in the FEO system.

7. At 100% charge level, when changing the ambient temperature from 47[degrees]F (8.3[degrees]C) to 17[degrees]F (-8.3[degrees]C), we can see approximately 40% degradation in heating capacities and 30% degradation in heating COPs, which are relatively consistent according to indoor airflow rates.

8. The heating capacity always increases with indoor airflow rate and outdoor temperature but reaches peak value at a charge level for each pair of indoor airflow rate and outdoor temperature. The heating COPs reach optimums at 47[degrees]F (8.3[degrees]C) ambient, 100% charge, and rated indoor airflow rate.

9. Compressor power consumptions decrease with the charge level and outdoor temperature but increase with reduction of indoor airflow rate.

Cyclic Tests

1. The degradation coefficient of the TXV system at each condition is smaller than that of the FEO system, which indicates that the TXV system has smaller cyclic penalties. This is because the TXV keeps most of the system charge at high pressure in the condenser when the system is off, whereas the FEO system allows pressure equalization and refrigerant migration during the off period.

2. The FEO and TXV systems showed the same tendency for an increased degradation of coefficient with a reduction of indoor airflow rate. The steady-state heating capacity increases significantly with increasing indoor airflow rates, while the energy penalty associated with cycling is less sensitive to airflow. Therefore, the energy loss becomes a smaller fraction of total heating capacity with increasing indoor airflow rate. Also, increasing airflow leads to a smaller difference between high- and low-side pressures, which reduces the energy loss associated with cycling.

3. With changing indoor airflow rate and charge level, the degradation coefficients of the TXV system range from 0.05 to 0.24, and those of the FEO system range from 0.18 to 0.26.

4. The degradation coefficient doesn't appear to have a linear relationship with charge level in either the FEO or TXV systems. The charge variation might have two contradictory effects on the cyclic performance. Greater charge might lead to more time for migration; on the other hand, it might lead to more energy (condensing pressure) built up in the TXV system, and a large amount of liquid refrigerant stored in the suction line accumulator, which boost the refrigerant flow rate at the starting period. We observed significantly lower degradation coefficients for the TXV system at 130% charge and high indoor flow rates.

Defrost Tests

1. For the unit using a TXV, frost formation led to hunting of the TXV in all situations. High charge inventory and high indoor airflow rate led to longer hunting times. At the end of a frost formation test, there was no superheat degree at the exit of the outdoor heat exchanger, even when using a TXV.

2. During frost formation tests, at low charge, the performances of the two systems were not much different; at high charge, performances of the TXV system were better than those of the FEO system. At high charge, there was no superheat degree in the unit using an FEO, while the unit using a TXV could still maintain a superheat degree at the beginning of the defrosting period. In this case, frost could form more quickly on the evaporator of the unit using an FEO, because it had a lower mean evaporating temperature.

3. The COPs and heating capacities of defrost tests appear 10% to 20% lower than those of the steady-state tests.

4. The defrosting penalties ([Ratio.sub.Q] and [Ratio.sub.cop]) don't have a consistent relationship with the charge level in either the FEO or TXV systems. Frost formation lowers the evaporating pressure, which leads to more charge in the accumulator or the condenser coil. The optimum charge at the defrosting condition would differ from that at the steady-state condition, so we can't see a linear relationship for the penalty ratios changing with the charge level.

5. The frost formation has less detrimental effect on the system at lower indoor airflow rate, due to its higher evaporating pressure.

RECOMMENDATIONS

We recommend utilizing the given data in the following ways:

1. Fault diagnosis: The data could tell engineers what to expect if the heat-pump system is undercharged (leaking), overcharged (technician's fault), or if the air coil is clogged (indoor airflow reduced). Some correction factors can be estimated to scale down nominal performance to the faulty conditions, which can be extended to the heat pumps having other tonnages.

2. Energy simulation: The data scale is sufficient to fit equipment curves for building energy simulation soft-ware. In particular, it can provide performance curves of faulty equipment (undercharged, clogged, etc.) on which we can base annual energy simulation to see how much energy is wasted.

3. Design variation: We make a thorough comparison of the units using TXV and FEO systems. The engineer can see the advantages or disadvantages, in terms of steady-state and dynamic performances, of switching the expansion devices. Some manufacturers don't provide product performance down to 17[degrees]F (-8.3[degrees]C) ambient. Our data can be used to estimate the amount of performance degradation from 47[degrees]F (8.3[degrees]C) rated condition to 17[degrees]F (- 8.3[degrees]C) cold ambient.

NOMENCLATURE

COP = coefficient of performance, --

[C.sub.D] = cyclic degradation coefficient, --

E = energy amount, J (Btu)

HLF = heating load factor, --

HSPF = heating seasonal performance factor,

W/W (Btu/h/W)

h = enthalpy, J/kg (Btu/lbm)

m = mass flow rate, kg/s (lbm/h)

Power = power consumption, W (Btu/h)

Q = heat transfer amount, J (Btu

Q = heat transfer rate, W (Btu)

[Q.sub.heating, air, tot] = steady-state heating capacity measured at indoor air side, W (Btu/h)

[Q.sub.heating, air, tot] = steady-state heating capacity measured at refrigerant side, W (Btu/h)

SEER = seasonal energy efficiency ratio, W/W (Btu/h/W)

scfm = cubic feet per minute flow rate at standard air density (1.23 kg/[m.sup.3] - 0.0768 lbm/[ft.sup.3] ), [m.sup.3] /s (cfm)

Subscripts

air = air side

blower = indoor blower

comp = compressor

cyc = cyclic performance

DEF = defrosting performance

fan = outdoor fan

heating = heating mode

in = inlet

indoor = indoor side

out = outlet

outdoor = outdoor side

ref = refrigerant side

ss = steady-state performance

tot = total capacity

REFERENCES

ASHRAE. 2010. ANSI/ASHRAE Standard 116, Methods of Testing for Rating Seasonal Efficiency of Unitary Air Conditioners and Heat Pumps. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

AHRI. 2008. ANSI/AHRI Standard 210/240, Performance Rating of Unitary Air-Conditioning and Air-Source Heat Pump Equipment. Arlington, VA: Air-Conditioning, Heating, and Refrigeration Institute.

Chen, Y.-G., and X.-M. Guo. 2009. Dynamic defrosting characteristics of air source heat pump and effects of outdoor air parameters on defrost cycle performance. Applied Thermal Engineering 29(13):2701-07.

Choi, J., and Y. Kim. 2004. Influence of the expansion device on the performance of a heat pump using R407C under a range of charging conditions. International Journal of Refrigeration 27(4):378-84.

Corberon, J.M., I. Martinez-Galvon, S. Martinez-Ballester, J. Gonzolvez-Macio, and R. Royo-Pastor. 2011. Influence of the source and sink temperatures on the optimal refrigerant charge of a water-to-water heat pump. International Journal of Refrigeration. Available online January 20, 2011.

Domanski, P., and D. Didion. 1984. Mathematical model of an air-to-air heat pump equipped with a capillary tube. International Journal of Refrigeration 7(4):249-55.

Kim, M., W.V. Payne, P.A. Domanski, S.H. Yoon, and C.J.L. Hermes. 2009. Performance of a residential heat pump operating in the cooling mode with single faults imposed. Applied Thermal Engineering 29(4)770-78.

O'Neal, D.L., K. Peterson, and N.K. Anand. 1991. Effect of short-tube orifice size on the performance of an air source heat pump during the reverse-cycle defrost. International Journal of Refrigeration 14(1):52-7.

Payne, W.V., and D.L. O'Neal. 1994. Defrost cycle performance for an air-source heat pump with a scroll and a reciprocating compressor. International Journal of Refrigeration 18(2):107-12.

Shen, B., E.A. Groll, and J.E. Braun. 2006. Improvement and validation of unitary air conditioner and heat pump simulation models at off-design conditions. PhD thesis, Herrick Labs, Purdue University.

Shen, B., E.A. Groll, and J.E. Braun. 2009. Improved methodologies for simulating unitary air conditioners at off-design conditions. International Journal of Refrigeration 32(7):1837-49.

Klein, S.A. 2004. User Manual of Engineering Equation Solver (EES). Middleton, WI: F-Chart Software.

Yoon, S.H., W.V. Payne, and P.A. Domanski. 2011. Residential heat pump heating performance with single faults imposed. Applied Thermal Engineering 5:76571.

Bo Shen, PhD Member ASHRAE

James E. Braun, PhD Fellow ASHRAE

Eckhard A. Groll, PhD Fellow ASHRAE

Bo Shen is a research scientist in the Building Technologies Research and Integration Center, Oak Ridge National Laboratories, Oak Ridge, TN. James E. Braun and Eckhard A. Groll are professors in the Mechanical Engineering School, Purdue University, West Lafayette, IN.
COPYRIGHT 2011 American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Inc.
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Author:Shen, Bo; Braun, James E.; Groll, Eckhard A.
Publication:ASHRAE Transactions
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Geographic Code:1USA
Date:Jul 1, 2011
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