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Defrost cycle performance for an air-source heat pump with a scroll and a reciprocating compressor
I'EIIU T T E R W O R T H I"~E I N E M A N N
Int. J. ReJ~'ig. Vol. 18, No. 2, pp 107 112, 1995 Copyright ~ 1995 Elsevier Science Ltd and IIR Printed in Great Britain. All rights reserved 0140-7007/95/$10.00 + 00
Defrost cycle performance for an air-source heat pump with a scroll and a reciprocating compressor V. Payne and D. L. O'Neal Energy Systems Laboratory, Department of Mechanical Engineering, Texas A & M University, College Station, TX 77843, USA Received 29 December 1992; revised 15 June 1994
A 10.6 kW nominal cooling capacity air-source heat pump was tested according to ANSI/ASHRAE Standard 116-1983 for the frost acumulation and defrost cycle. These tests required indoor conditions of 21.1 °C (70 °F) dry-bulb, 15 °C (60 °F) maximum wet-bulb, with outdoor conditions of 1.7 °C (35 °F) dry-bulb, 0.5~C (30°F) wet-bulb. The unit was tested with the original scroll compressor and a reciprocating compressor that yielded similar heating performance. Heating capacity for the scroll system peaked at 8.4 kW (2.38 tons), while the reciprocating system heating capacity peaked at 8.5 kW (2.42 tons) during the frosting period. Heating capacities for the two system configurations differed by less than 1% during the frosting period. Power demand for the scroll system peaked at 2.9 kW, and the reciprocating system power demand peaked at 3.1 kW. During the frosting period, the reciprocating system power demand averaged 7% higher than the scroll system power demand. The reciprocating system completed a defrost 5% faster than the scroll system. Scroll system defrost time was 6.8 min while reciprocating system defrost time was 6.5 min. The volume of condensate produced differed by less than 3% with 1680 ml (102.5 in 3) and 1640 ml (100 in 3) produced by the scroll and reciprocating systems, respectively. Discharge pressures during defrost were within 3% with peak values of t315 kPa (191 psia) and 1351 kPa (196 psia) for the scroll and reciprocating systems respectively. The reciprocating compressor produced higher levels of discharge superheat, peaking at 53 °C (95 °F) compared to the scroll system peak discharge superheat of 47"C (85 °F). Overall, discharge superheat for the reciprocating system averaged 18% higher than the scroll system. The reciprocating system produced defrost refrigerant flowrates that averaged 3% higher than the scroll system. Refrigerant flowrates for the scroll and reciprocating systems peaked at 3.7 kg min ~ (8.2 Ibm min-~) and 4.0 kg min-1 (8.8 Ibm min-~) respectively. (Keywordse heat pump; scroll compressor; reciprocating compressor; defrosting; testing; measurement; performance)
Performance du cycle de d6givrage d'une pompe /t chaleur off l'air est la source de chaleur, avec un compresseur/t spirale et un compresseur/t piston On a testb une pompe h chaleur o~ I'air est la source de chaleur, d'une capacitb nominale de refroidissement de 10,6 kW, conformbment h la norme A N S I / A S H R A E 116-1983 sur l'accumulation de givre et le o,cle de dbgivrage. Ces essais ont exigO des temp&atures intbrieures de 21,1 °C (70 °F) de bulbe sec, et de 15 '-C (60 °F) au maximum de bulbe humide, avec des conditions extbrieures de 1,7 °C (35 °F) de bulbe sec et 0,5 °C (30 °F) de bulbe humide. L'appareil a btb essayb avec le compresseur h spirale d'oriyine et un compresseur ~ piston, dont les performances de chauffage btaient similaires. La capacitb de chauffage du compresseur h spirale a atteint une pointe de 8,4 k W, alors que celle du compresseur h piston a atteint une pointe de 8,5 k W au cours de la pbriode de dbgivrage. Les capacitbs de chauffage de ces deux configurations d!ffOraient de moins de 1% au cours de la pbriode de 9ivrage. La consommation dWectricitb du compresseur h spirale a atteint un maximum de 2,9 kW, et celle du compresseur gt piston un maximum de 3.1 kW. Au tours de la pbriode de givrage, cette derni&e a atteint une moyenne supkrieure de 7% gt celle du compresseur h spirale, Avec le compresseur h piston, le dkgivrage a btb de 5% plus rapide qu'avec le compresseur h spirale. Avec ce dernier, la durbe de dbgivrage a bt~ de 6,8 min, alors qu'avec le compresseur gt piston, elle a btb de 6,5 min. Le volume de condensat produit diffbrait de moins de 3%, avec 1680 ml (102,5 ins) pour le compresseur ~ spirale et 1640 ml (100 in3) pour le compresseur h piston. Les pressions de refoulement au tours du dbgivrage sont restbes dans une limite de 3% avec des valeurs de pointe de 1315 kPa (191 psia ) pour h' compresseur a spirale et de 1351 kPa (196 psia) pour le compresseur gtpiston. Avec le compresseur ~ piston, des quantitbs plus blevbes de chaleur de refoulement ont btb produites: la temp&ature btait de 53 °C (95 °F), par rapport h 47 °C (85 °F) pour le compresseur ~ spirale. Globalement, la chaleur de refoulement du compresseur h piston a btb en moyenne de 18% supbrieure ~ celle du compresseur gt spirale. Avec le compresseur gt piston, les dbbits masse de frigorigbne au dbgivraoe btaient de 18% supbrieurs 71 ceux obtenus avec le compresseur h spirale. Les dbbits masse de frigorigOne ont btb de 3,7 k9 min -l (8,2 Ibm rain"1) pour le compresseur h spirale et de 4,0 k 9 min -~ (8,8 Ibm min 1) pour le compresseur gt piston.
(Mots cl6s: pompe &chaleur; compresseur fi spirale; compresseur &piston; d6givrage; essai; mesure; performance)
107
108
V. Payne and D. L. O'Nea/
The scroll compressor was first described in the patent literature in the early 1900s 1. It is increasingly being used in residential-sized heat pumps in the USA. The scroll compressor has begun to replace the reciprocating compressor in this market because of its even driving torque distribution, low pressure-drop suction process, lack of suction re-expansion effects, rotational operation, and valveless design 2. In addition, the scroll compressor has proven to be efficient and reliable while providing low noise and low vibration 3. The simple structure and compact design of the scroll compressor has also increased its popularity among heat pump manufacturers. While scroll compressors are increasingly replacing reciprocating compressors in US-manufactured heat pumps, no study has been published that directly compares the two compressors' performance during the defrost cycle of a heat pump. The purpose of this investigation was to test scroll and reciprocating compressors with comparable efficiency and capacity to assess the relative performance of the two compressors during the defrost cycle of an air-to-air heat pump. The investigation focused on the transient performance of the two different system configurations. The heat pump used in these tests was originally equipped with a scroll compressor. A reciprocating compressor was selected to replace the scroll compressor while providing the same heating capacity. The two compressors were matched for capacity and efficiency by the system manufacturer. Tests were conducted under frosting conditions, with the same indoor and outdoor heat exchangers being used by both compressors.
Test facility Testing of the system, with 10.6 kW nominal cooling capacity, was conducted in psychrometric rooms. The test apparatus consisted of the outdoor heat exchanger test section, the indoor heat exchanger test section, test heat pump, and the data acquisition system. The psychrometric rooms allowed indoor and outdoor dry-bulb and wet-bulb temperatures to be maintained within 0.3 °C (0.2 °F) during steady-state operations. The outdoor heat exchanger test section consisted of the heat pump outdoor fan/coil unit, insulated airflow duct, flow nozzles, thermocouple grids, airflow control damper, and airflow chamber booster fan. The outdoor airflow chamber was placed over the outdoor fan/coil unit to allow normal heating operation. The airflow chamber booster fan allowed the static pressure exiting the outdoor fan/coil unit to be maintained at zero to simulate free discharge conditions. Air exiting the fan/coil unit passed through turning vanes, a thermocouple grid, a dewpoint sampler, and through the flow nozzles. Pressure drop across the flow nozzles along with the dewpoint of the incoming air was used to calculate the volumetric flowrate of air through the outdoor fan/coil unit. Once the air had passed through the flow nozzles, the air temperature was measured by a thermocouple grid. The air then passed through a louvred damper and a booster fan to be recirculated in the outdoor room. The louvred damper allowed airflow to be reduced, as the outdoor heat exchanger was covered with frost during heating operation. Dewpoint temperatures were sampled before and after the outdoor fan/coil unit using chilled mirror hygrometers. The indoor heat exchanger test section consisted of the indoor fan/coil unit, airflow ductwork, and indoor airflow
Description of test points Tableau 1 Description des" points de mesure
Table !
Channel
Type
Location
0~, 5 6 7 8-11 12 13 14
Thermocouple Thermocouple Thermocouple Thermocouple Thermocouple TC-grid Dewpoint sensor TC-grid
Refrigerant line Outdoor vapour circuit Indoor coil liquid line Outdoor vapour circuit Refrigerant line Upstream indoor coil Upstream indoor coil Downstream indoor coil
15 16 17 18 19 20-24 25 26 27 28 31 32-33 34 35-37 38 39 4~41 42 43 44 45 46 47 48 49 50 51-52 53-54 55-58 59
Dewpoint s e n s o r Thermocouple Dewpoint sensor TC-grid Dewpoint s e n s o r Pressure transducer
Downstream indoor coil Upstream outdoor coil Upstream outdoor coil Downstream outdoor coil Downstream outdoor coil Refrigerantline Not used Refrigerantline Not used Refrigerantline Indoor flow chamber Not used Outdoor flow chamber Not used Refrigerant line Outdoor vapour circuit Outdoor fan Refrigerant line Total power Refrigerantline Sum 4041 Accumulator bottom Difference45~13 Chilled water Indoor nozzles Outdoor nozzles Outdoor liquid circuit Outdoor vapour circuit
Pressure transducer Pressure transducer Differential pressure Differential pressure Flowmeter Thermocouple Watt transducer Thermocouple Watt transducer Pressure transducer Pseudo channel Thermocouple Pseudo channel Thermocouple TC-grid TC-grid Thermocouple Thermocouple
chamber. Conditioned air from the indoor psychrometric room passed through a dewpoint sampler, the indoor fan/coil unit, insulated airflow ductwork, and the indoor airflow chamber. Air temperatures entering the indoor fan/coil unit were sampled by a nine-node thermocouple grid, and air temperatures exiting were measured by a 16-node thermocouple grid. Dewpoint temperatures were measured with chilled mirror hygrometers. The airflow ductwork consisted of two sets of turning vanes, which allowed the air to enter the indoor airflow chamber. The indoor airflow chamber was built to Air Movement and Control Association specifications 4. The airflow chamber allowed for accurate measurement of airflow through the indoor fan/coil unit. The outdoor heat exchanger, compressor, accumulator, and reversing valve were located in the outdoor psychrometric room, while the indoor heat exchanger and blower were located in the indoor psychrometric room. The outdoor heat exchanger was a two-row four-circuit vertical coil with wavy fins, a fin density of 7 fins c m - 1 (18 fins in-x), 9.53 mm (0.375 in) nominal diameter refrigerant tubes, and frontal dimensions of 61 cm (24 in) by 189.2 cm (75 in). Liquid- and vapour-line refrigerant circuit distribution tubes were instrumented with surface-soldered thermocouples. The lowest liquidline refrigerant distribution tube temperature was used to signal the end of a defrost operation. A three-blade axial flow fan blade with a 0.25 kW (0.33 hp) motor was used to pull air across the outdoor coil.
109
Defrost cycle performance of an air-source heat pump The indoor coil was a three-row, four-circuit A-type coil with 5.5 fins cm-1 (14 fins in-1). A centrifugal fan was included in the fan/coil unit and pulled air across the indoor heat exchanger. The indoor heat exchanger and the outdoor heat exchanger were connected by refrigerant tubing referred to as the 'liquid' and 'vapour' lines. The liquid line consisted of approximately 9.5 m (31 ft) of 9.53 mm (0.375 in) diameter tubing. The liquid line contained both the heating- and cooling-mode expansion devices. The heating-mode expansion device was a piston-type orifice with a diameter of 1.5 mm (0.059 in). The cooling/defrost-mode expansion device was a thermostatic expansion valve (TXV). The TXV superheat control bulb was attached to the vapour line close to the indoor heat exchanger. The vapour line included 9.2 m (30 ft) of 15.9 mm (0.625 in) diameter tube connecting the indoor coil to the reversing valve and 1.5 m (5 ft) of tubing connecting the reversing valve to the outdoor heat exchanger. The refrigerant circuit was instrumented with thermocouple probes, pressure transducers, and flowmeters (Table 1). Pressures throughout the refrigerant circuit were measured with 0-2.1 MPa (0-300 psia) pressure transducers calibrated with a deadweight tester. Temperature probes were placed throughout the refrigerant circuit in close proximity to the pressure transducers to allow accurate prediction of the refrigerant characteristics. Refrigerant flowrate was measured by two parallel Coriolis effect mass flowmeters in the refrigerant liquid line. Pressure drop across the mass flowmeters was limited to a maximum of 69 kPa (10 psia) at maximum refrigerant flowrate in accordance with ASHRAE Standard 116-835 . Valves were lever-actuated ball valves. These valves allowed for component testing, leak detection, and system component isolation due to possible failure. The test procedure consisted of selecting test conditions, setting the system refrigerant charge, and developing a test procedure to ensure repeatable results. The test conditions chosen were those specified for the frost build-up tests in the ASHRAE Standard 116-835. Test conditions for this procedure are given in Table 2. The refrigerant charge was set in the cooling mode for both compressor configurations, with indoor conditions of 26.7 °C (80 "F) dry-bulb, 19.4 °C (67 °F) wet-bulb, and
Table 2 Test conditions
Tableau 2 Conditionsd'essai Limits (°c)
Location
Dry-bulb ("C)
Wet-bulb UC)
Frosting
Defrosting
Indoor Outdoor
21.1 1.7
15.5 max 0.5
+0.56 +0.56
+ 1.1 ___1.1
outdoor conditions of 35 °C (95 °F). The unit was charged until the degree of refrigerant subcooling leaving the outdoor heat exchanger reached 5+1.1°C (9+2°F). All charging procedures followed manufacturer's recommendations. The control electronics installed by the manufacturer were removed to allow for manual initiation and termination of the defrost during heating-mode operation. Defrost was initiated at a 20% drop in peak heating capacity occurring during the frosting period. This manual procedure was intended to mimic the defrost controls supplied by the manufacturer. All tests used this criterion for the initiation of defrost. Defrost was terminated based upon temperature measurements on the lowest liquid-line circuit of the outdoor heat exchanger. Defrost was terminated when the circuit temperature reached 26.7+C (80°F). The scroll and reciprocating compressor system configurations used the same defrost scheme. Results
Frosting period The original compressor supplied by the heat pump manufacturer was a scroll, which was sized to yield a nominal cooling capacity of 10.6 kW (3 tons) during cooling. The reciprocating compressor tested with the same heat pump configuration was sized to yield the same cooling performance. Table 3 lists manufacturersupplied data describing the two compressors. The scroll and reciprocating compressors were matched as closely as possible for capacity and efficiency. Data in Table 3 were collected at a 7.2 °C (45 °F) evaporating temperature, a 54.4°C (130°F) condensing temperature, a 46.1 'C (115 OF) liquid temperature and a 35 °C (95 °F) ambient temperature. Testing of the heat pump under the experimental conditions showed that the unit equipped with the different compressors yielded approximately the same heating capacity (Table 4). Heating capacity for the two systems peaked at 8.4 kW (2.38 tons) and 8.5 kW (2.41 tons) for the scroll and reciprocating systems respectively. Heating capacity for the scroll system averaged only 0.6% lower than the reciprocating system during the frosting period. Figure 1 shows the heating performance of the differently configured systems during the frosting period. Figure 2 shows the power demand for the two systems during the frosting period. Power demand for the two systems peaked at 2.9 kW and 3.1 kW for the scroll and reciprocating systems respectively. The reciprocating system's power demand averaged 7% higher than that of the scroll system. The increased power demand of the reciprocating system caused the instantaneous COP to be lower than that of the scroll system (Figure 3). Instantaneous COP averaged 4% lower for the reciprocating system during the frosting period. The slow decline in power demand seen with the reciprocating
Table 3 Compressorcharacteristics
Tableau 3 Caracff'ristiquesdes compresseurs Type Scroll Reciprocating
Displacement (cm3rev- 1)
Capacity (kW)
Power (W)
COP
Current draw (A)
46.05 61.94
9.964 10.257
3090 3220
3.22 3.18
15.4 13.8
V. Payne and D. L. O ' N e a /
1 10
Table 4 Cycle performance comparison Tableau 4 Comparaison de la performance des cycles de d~givrage
Case
Maximum heating capacity (kW)
Integrated cyclic COP
Cycle time (rain)
Defrost time (rain)
Condensate (ml)
8.4 8.5
2.19 2.05
66.16 64.63
6.8 6.5
1680 1640
Scroll Reciprocating
10-
3
8--
>~ <
3
L32 ...... 0 (D L28 0
:24 x
X
2
6-
: 2 0 -q 16 ~
4-
0
O
SCROLL RECIP.
12 ;~
< (._)
1
8 " Indoor:
~l.lC,
[ Outdoor:
S C R O L L
15C
1.67C,
1[
wb
4
95*/* RH[
RECIP. 17
22
27
32
37
42
47
52
57
10-5-91
&
5-22-92
Indoor:
21 .IC,
15C
Outdoor:
1.67C,
95%
wb RH
0
r i i i i i i i i i i [ i T Fi ] , 1 i r ~ , i i i i q i r i i i T I P~ I~
12
< m < ~
62
0
Time Relative t o Defrost (min)
-r7~r777
12
n n I1
17
~ ~ t t ~ r r I r I ~ t777-~
22
27
32
37
F r I I P ~ n I I I r F I I ~ J n n I n n I I I~
42
47
52
57
62
0
T i m e RelaLive t o Defrost. ( r a i n )
Figure l
Frosting period heating capacities
Figure 1
CapacitOs de chauffage en pbriode de givrage
Figure 3
Frosting period instantaneous COPs
Figure 3
COP instantanbs en pbriode de yivra#e
_
L~ 0
- --
SCROLL RECIP.
10-5-91 Indoor:
&
5-22-92
21.1C,
15C
wb
o RH Outdoor: 1.67C,95'~
~
12
~
17
iPnlnlP
22
Time
27
Jpnnnll
32
Relative
r ur u l e u i n l l P J j l l ~ l n l n
37
42
to
47
52
Defrost
Figure 2
Frosting period power demands
Figure 2
Demandes d'blectricitb en pbriode de yivrage
r u ~
57
62
(min)
system was a function of the refrigerant flowrate and the outdoor fan power demand. As the outdoor heat exchanger was covered with frost, refrigerant flowrate decreased and outdoor fan power increased. The decrease in refrigerant flowrate caused the compressor power
demand to drop. The scroll system does not show a gradual decline in power demand because the reduction in power demand caused by decreasing refrigerant mass flowrate was equally matched by the increasing power demand of the outdoor fan. The reciprocating compressor consumed more energy than the scroll compressor, and therefore the outdoor fan power was a lower percentage of the total power for the reciprocating system. Another difference in the two compressors was the degree of discharge superheat produced. During the frosting period, discharge superheat peaked at 28 °C (50 °F) and 30 °C (54 °F) for the scroll and reciprocating compressors respectively. Figure 4 shows the compressor discharge and suction superheat during the frosting period. Discharge superheat gradually increased during the frosting period, owing to reduced refrigerant flowrate. As the difference in suction superheats of the two compressors was within 1%, the higher discharge superheat of the reciprocating compressor was due to its greater power demand. Figure 5 shows the refrigerant flowrates of the two compressors during the frosting period. Refrigerant flowrate for the two systems followed the same trends, peaking at 2.7 kg m i n - i (6 Ibm m i n - 1) after defrost and dropping to 1.9 kg m i n - 1 (4.2 Ibm min-1) at defrost initiation. Refrigerant flowrate for the reciprocating compressor averaged 2% higher than the scroll compressor during frost formation. Refrigerant flowrate decreased in both cases, owing to frost formation on the outdoor heat exchanger. Frost reduced the
111
Defrost cycle performance of an air-source heat pump
for the scroll and reciprocating compressors respectively. Suction pressure for the reciprocating system averaged 4% lower than the scroll system. Suction pressures fell to 102 kPa (15 psia) and 108 kPa (16 psia) for the scroll and reciprocating compressors respectively. Refrigerant flowrate during defrost peaked at 3.5 kg min- ~ (7.7 Ibm min- ~) and 3.7 kg min- ~ (8.2 Ibm min- ~) for the scroll and reciprocating systems respectively. Refrigerant flowrate for the reciprocating system averaged 3% higher than the scroll system. Figure 7 shows the refrigerant flowrates for both systems during the defrost. At defrost initiation, both systems show the refrigerant flowrate dropping near zero then steadily increasing as the reduced suction pressure causes the evaporation of more refrigerant. At 4 min the refrigerant flowrate of the reciprocating system equals that of the scroll system. After 4 min, the refrigerant flowrate of the reciprocating system increased over the scroll system, peaking at 3.7 kg min- 1 (8.2 Ibm min-a) at defrost termination. The larger refrigerant flowrate of the reciprocating system produced a greater defrost capacity than that of the scroll system. Defrosting capacity peaked at 9.0 kW
51 46 41t--
i36 "-~
-al < i e6 :< UZ
< D-q
- -
SCROLL RECIP
-
~13 L<
16
d~
8
U)
11
10-5-91 & 5-22-92 lndo~: 21.1C, 15C wb Outdoor: 1.67C, 95% RH
3-
5'?
6 -4
Time ReLative to Defrost (min) Figure 4 Frosting period discharge superheats Figure 4 Surchauffes au refoulement en p&iode de oivraoe
3.0
1500) 6.0
. . •2 . 5
/
E :5.0 E
•
2~
=
ieoo
1250~ .....
175
..0
~2.0 0 ..a
o ,...2 r, 3 . 0 E-Z <
~i.5 z< c~l.O !T~ i-
10-5-91 & 5-22-92
]
~05
0.
2.0
SCROLL
RECIP.
0
Indoor: 21.1C, 15C wb Outdoor: 1.67C. 95% KH
~llllrT
la
I '1 'IK[I
17
a2
Time
'l,
27
[ II 'l I ''1
32
Relative
37
to
'[1'
'1['~'1
42
47
' , ,I],
52
Defrost
57
125 750 03
:~ o r~ r.
10
' '11
'
62
150 ~
~.~ooo !
4.0
3 5 0 ~q
0.0
SCROLL RECIP
0 ~-~-~,-~
5
Time
(min)
Figure 5 Frosting period refrigerant flowrates Figure 5 Dbbits masse de frigorigbne en p&iode de #ivraye
......
0 Relative
L °'"~':'~'~'_22_" I [ e5 ~
5 to
F0
10 15 DefrosL (rain)
Figure 6 Defrosting period discharge absolute pressures Figure 6 Pressions absolues de refoulement en p&iode de dbgivrage
air-to-refrigerant heat energy transfer, causing a reduction in the mass of refrigerant evaporated in the outdoor heat exchanger.
4U 35 "N
---
SCROLL RECIP
:7o S
b
Defrostin9 period The defrost time for the reciprocating system was 5% faster than the scroll system. This was partially due to the lower volume of condensate collected from the reciprocating system. The reciprocating system produced 1640 ml (100 in 3) of condensate, while the scroll system produced 1680 ml (103 in 3) of condensate. The smaller amount (2.4%) of condensate for the system equipped with the reciprocating compressor was probably caused by the 1.5 min shorter cycle time. Figure 6 shows the discharge pressure produced by the two compressors during the defrost. Discharge pressure for the reciprocating system averaged 3% higher than the scroll system during defrost. Discharge pressures peaked at 1315 kPa (191 psia) and 1351 kPa (196 psia)
.........
_0
~25 50
h-,
r,
~210
t4o
z< ~15
30
1.0
"~
LQ
20 K Il 10-5-91& 5-22-92
0.5 OO
z <
Indoor: 21.1C, 15C wb Outdoor: 1.67C, 95% RH
-5 Time
0 Relative
5 to
10 Defrost (rain)
il.o X 5
oo
Figure 7 Defrosting period refrigerant flowrates Figure 7 Dbbits masse de frigorio~ne en p&iode de d~givrage
112
V. Payne and D. L. O'Neal F
c-.
6
<
-2
0
-4
i4~ <~
-8
-_
-10-
4
'\~
\~\\ SCROLL -- Rb~CIP.
......................... 0
'~
~J
5
24<
10-5-91& 5-22-92
lndoor:21.1C.15Cwb Outdoor: 1.67C, 95% RH
-. . . . . . . . . . . . . 10
15
34
Time Relative to Defrost (min) Figure 8 Defrostingperiod heating capacities Figure 8 Capacit~s de chauffage en p~riode de d~givra9e (2.5 tons) and 9.1 kW (2.6 tons) for the scroll and reciprocating systems respectively. Figure 8 shows the defrost capacity for the two systems. Although the different compressors producetl comparable defrosting capacities, the reciprocating compressor produced discharge gas with a higher degree of superheat than the scroll compressor. In addition to having a higher refrigerant flowrate during defrost, the reciprocating system produced vapour with a higher degree of superheat than the scroll system. Discharge superheat for the two systems peaked at 47 °C (85 °F) and 53 °C (95 °F) respectively. Discharge superheat for the reciprocating system averaged 18% higher than that of the scroll system. Given that the reciprocating system produced higher refrigerant flowrates and higher discharge superheats than the scroll system, the reciprocating system produced higher-energy vapour to defrost the outdoor heat exchanger. Although the reciprocating system produced higher-energy refrigerant during defrost, the performance of the scroll system was the most efficient. Table 4 showed the integrated cyclic C O P for the system configured with the different compressors. For the performance measured between defrosts, the scroll system yielded an integrated C O P of 2.19 while the reciprocating system yielded an integrated C O P of 2.05. Conclusions
Frost accumulation and defrost tests of a 10.6 kW (3 ton) nominally sized air-source heat pump equipped with a scroll and then with a reciprocating compressor revealed the overall performance of both systems. The system
equipped with the scroll compressor provided a slightly higher integrated C O P of the two configurations. The 6% difference in COPs was almost negligible compared with the accuracy of the calculation (_+5%). The reciprocating compressor produced heating capacity that was within 1% of the original scroll compressor. The reciprocating compressor produced a higher discharge superheat during frosting and defrost. Given that both systems produced discharge pressures that were within 3% during defrost, the reciprocating compressor produced higher-energy discharge gas, which reduced the defrost time over that of the scroll-compressor-equipped system. These tests were meant to highlight differences in transient performance for a system equipped with two different compressor mechanisms. The data presented here are intended to provide a reference for determining differences in performance for a system with two different compressor mechanisms, all other variables being held constant. Although these tests were not a perfect comparison, owing to possible inequalities in compressor motors, the general trends should provide some insight into the transient operating characteristics of heat pumps equipped with scroll or reciprocating compressors. Both compressors were matched by the manufacturer on the basis of capacity and efficiency. Although sound measurements were not taken, the scroll compressor was noticeably quieter than the reciprocating compressor during frosting and at defrost initiation. The scroll compressor was also physically smaller than the replacement reciprocating compressor. This was due to the construction of the two compressors. The reciprocating mechanism and motor are springmounted inside the sealed casing whereas the scroll mechanism and motor are hard mounted to the casing. Acknowledgements
The authors wish to express their appreciation to the Texas Energy Research in Applications Program for funding for this project. References
Creux, L. 1905, Rotary engine, US Patent No. 801182 (1905) lnaba, T. et al. A scroll compressor with sealing means and low pressureshellside. Proc 1986 International CompressorEngineering Conf Purdue University, West Lafayette, IN (1986) Vol. III, 887-900 Ishii, N. et aL A study on dynamic behavior of a scroll compressor. Proc 1986 International Compressor Engineering Conf Purdue University, West Lafayette, IN (1986) VoL III, 901 916 Laboratory methods of testing fans for rating. ANSO/ASHRAE Standard 51-1985 (ANSI/AMCA 210-85) Methods of testing for seasonal efficiency of unitary airconditioners and heat pumps. ANSI/ASHRAE Standard 1161983
Int. J. ReJ~'ig. Vol. 18, No. 2, pp 107 112, 1995 Copyright ~ 1995 Elsevier Science Ltd and IIR Printed in Great Britain. All rights reserved 0140-7007/95/$10.00 + 00
Defrost cycle performance for an air-source heat pump with a scroll and a reciprocating compressor V. Payne and D. L. O'Neal Energy Systems Laboratory, Department of Mechanical Engineering, Texas A & M University, College Station, TX 77843, USA Received 29 December 1992; revised 15 June 1994
A 10.6 kW nominal cooling capacity air-source heat pump was tested according to ANSI/ASHRAE Standard 116-1983 for the frost acumulation and defrost cycle. These tests required indoor conditions of 21.1 °C (70 °F) dry-bulb, 15 °C (60 °F) maximum wet-bulb, with outdoor conditions of 1.7 °C (35 °F) dry-bulb, 0.5~C (30°F) wet-bulb. The unit was tested with the original scroll compressor and a reciprocating compressor that yielded similar heating performance. Heating capacity for the scroll system peaked at 8.4 kW (2.38 tons), while the reciprocating system heating capacity peaked at 8.5 kW (2.42 tons) during the frosting period. Heating capacities for the two system configurations differed by less than 1% during the frosting period. Power demand for the scroll system peaked at 2.9 kW, and the reciprocating system power demand peaked at 3.1 kW. During the frosting period, the reciprocating system power demand averaged 7% higher than the scroll system power demand. The reciprocating system completed a defrost 5% faster than the scroll system. Scroll system defrost time was 6.8 min while reciprocating system defrost time was 6.5 min. The volume of condensate produced differed by less than 3% with 1680 ml (102.5 in 3) and 1640 ml (100 in 3) produced by the scroll and reciprocating systems, respectively. Discharge pressures during defrost were within 3% with peak values of t315 kPa (191 psia) and 1351 kPa (196 psia) for the scroll and reciprocating systems respectively. The reciprocating compressor produced higher levels of discharge superheat, peaking at 53 °C (95 °F) compared to the scroll system peak discharge superheat of 47"C (85 °F). Overall, discharge superheat for the reciprocating system averaged 18% higher than the scroll system. The reciprocating system produced defrost refrigerant flowrates that averaged 3% higher than the scroll system. Refrigerant flowrates for the scroll and reciprocating systems peaked at 3.7 kg min ~ (8.2 Ibm min-~) and 4.0 kg min-1 (8.8 Ibm min-~) respectively. (Keywordse heat pump; scroll compressor; reciprocating compressor; defrosting; testing; measurement; performance)
Performance du cycle de d6givrage d'une pompe /t chaleur off l'air est la source de chaleur, avec un compresseur/t spirale et un compresseur/t piston On a testb une pompe h chaleur o~ I'air est la source de chaleur, d'une capacitb nominale de refroidissement de 10,6 kW, conformbment h la norme A N S I / A S H R A E 116-1983 sur l'accumulation de givre et le o,cle de dbgivrage. Ces essais ont exigO des temp&atures intbrieures de 21,1 °C (70 °F) de bulbe sec, et de 15 '-C (60 °F) au maximum de bulbe humide, avec des conditions extbrieures de 1,7 °C (35 °F) de bulbe sec et 0,5 °C (30 °F) de bulbe humide. L'appareil a btb essayb avec le compresseur h spirale d'oriyine et un compresseur ~ piston, dont les performances de chauffage btaient similaires. La capacitb de chauffage du compresseur h spirale a atteint une pointe de 8,4 k W, alors que celle du compresseur h piston a atteint une pointe de 8,5 k W au cours de la pbriode de dbgivrage. Les capacitbs de chauffage de ces deux configurations d!ffOraient de moins de 1% au cours de la pbriode de 9ivrage. La consommation dWectricitb du compresseur h spirale a atteint un maximum de 2,9 kW, et celle du compresseur gt piston un maximum de 3.1 kW. Au tours de la pbriode de givrage, cette derni&e a atteint une moyenne supkrieure de 7% gt celle du compresseur h spirale, Avec le compresseur h piston, le dkgivrage a btb de 5% plus rapide qu'avec le compresseur h spirale. Avec ce dernier, la durbe de dbgivrage a bt~ de 6,8 min, alors qu'avec le compresseur gt piston, elle a btb de 6,5 min. Le volume de condensat produit diffbrait de moins de 3%, avec 1680 ml (102,5 ins) pour le compresseur ~ spirale et 1640 ml (100 in3) pour le compresseur h piston. Les pressions de refoulement au tours du dbgivrage sont restbes dans une limite de 3% avec des valeurs de pointe de 1315 kPa (191 psia ) pour h' compresseur a spirale et de 1351 kPa (196 psia) pour le compresseur gtpiston. Avec le compresseur ~ piston, des quantitbs plus blevbes de chaleur de refoulement ont btb produites: la temp&ature btait de 53 °C (95 °F), par rapport h 47 °C (85 °F) pour le compresseur ~ spirale. Globalement, la chaleur de refoulement du compresseur h piston a btb en moyenne de 18% supbrieure ~ celle du compresseur gt spirale. Avec le compresseur gt piston, les dbbits masse de frigorigbne au dbgivraoe btaient de 18% supbrieurs 71 ceux obtenus avec le compresseur h spirale. Les dbbits masse de frigorigOne ont btb de 3,7 k9 min -l (8,2 Ibm rain"1) pour le compresseur h spirale et de 4,0 k 9 min -~ (8,8 Ibm min 1) pour le compresseur gt piston.
(Mots cl6s: pompe &chaleur; compresseur fi spirale; compresseur &piston; d6givrage; essai; mesure; performance)
107
108
V. Payne and D. L. O'Nea/
The scroll compressor was first described in the patent literature in the early 1900s 1. It is increasingly being used in residential-sized heat pumps in the USA. The scroll compressor has begun to replace the reciprocating compressor in this market because of its even driving torque distribution, low pressure-drop suction process, lack of suction re-expansion effects, rotational operation, and valveless design 2. In addition, the scroll compressor has proven to be efficient and reliable while providing low noise and low vibration 3. The simple structure and compact design of the scroll compressor has also increased its popularity among heat pump manufacturers. While scroll compressors are increasingly replacing reciprocating compressors in US-manufactured heat pumps, no study has been published that directly compares the two compressors' performance during the defrost cycle of a heat pump. The purpose of this investigation was to test scroll and reciprocating compressors with comparable efficiency and capacity to assess the relative performance of the two compressors during the defrost cycle of an air-to-air heat pump. The investigation focused on the transient performance of the two different system configurations. The heat pump used in these tests was originally equipped with a scroll compressor. A reciprocating compressor was selected to replace the scroll compressor while providing the same heating capacity. The two compressors were matched for capacity and efficiency by the system manufacturer. Tests were conducted under frosting conditions, with the same indoor and outdoor heat exchangers being used by both compressors.
Test facility Testing of the system, with 10.6 kW nominal cooling capacity, was conducted in psychrometric rooms. The test apparatus consisted of the outdoor heat exchanger test section, the indoor heat exchanger test section, test heat pump, and the data acquisition system. The psychrometric rooms allowed indoor and outdoor dry-bulb and wet-bulb temperatures to be maintained within 0.3 °C (0.2 °F) during steady-state operations. The outdoor heat exchanger test section consisted of the heat pump outdoor fan/coil unit, insulated airflow duct, flow nozzles, thermocouple grids, airflow control damper, and airflow chamber booster fan. The outdoor airflow chamber was placed over the outdoor fan/coil unit to allow normal heating operation. The airflow chamber booster fan allowed the static pressure exiting the outdoor fan/coil unit to be maintained at zero to simulate free discharge conditions. Air exiting the fan/coil unit passed through turning vanes, a thermocouple grid, a dewpoint sampler, and through the flow nozzles. Pressure drop across the flow nozzles along with the dewpoint of the incoming air was used to calculate the volumetric flowrate of air through the outdoor fan/coil unit. Once the air had passed through the flow nozzles, the air temperature was measured by a thermocouple grid. The air then passed through a louvred damper and a booster fan to be recirculated in the outdoor room. The louvred damper allowed airflow to be reduced, as the outdoor heat exchanger was covered with frost during heating operation. Dewpoint temperatures were sampled before and after the outdoor fan/coil unit using chilled mirror hygrometers. The indoor heat exchanger test section consisted of the indoor fan/coil unit, airflow ductwork, and indoor airflow
Description of test points Tableau 1 Description des" points de mesure
Table !
Channel
Type
Location
0~, 5 6 7 8-11 12 13 14
Thermocouple Thermocouple Thermocouple Thermocouple Thermocouple TC-grid Dewpoint sensor TC-grid
Refrigerant line Outdoor vapour circuit Indoor coil liquid line Outdoor vapour circuit Refrigerant line Upstream indoor coil Upstream indoor coil Downstream indoor coil
15 16 17 18 19 20-24 25 26 27 28 31 32-33 34 35-37 38 39 4~41 42 43 44 45 46 47 48 49 50 51-52 53-54 55-58 59
Dewpoint s e n s o r Thermocouple Dewpoint sensor TC-grid Dewpoint s e n s o r Pressure transducer
Downstream indoor coil Upstream outdoor coil Upstream outdoor coil Downstream outdoor coil Downstream outdoor coil Refrigerantline Not used Refrigerantline Not used Refrigerantline Indoor flow chamber Not used Outdoor flow chamber Not used Refrigerant line Outdoor vapour circuit Outdoor fan Refrigerant line Total power Refrigerantline Sum 4041 Accumulator bottom Difference45~13 Chilled water Indoor nozzles Outdoor nozzles Outdoor liquid circuit Outdoor vapour circuit
Pressure transducer Pressure transducer Differential pressure Differential pressure Flowmeter Thermocouple Watt transducer Thermocouple Watt transducer Pressure transducer Pseudo channel Thermocouple Pseudo channel Thermocouple TC-grid TC-grid Thermocouple Thermocouple
chamber. Conditioned air from the indoor psychrometric room passed through a dewpoint sampler, the indoor fan/coil unit, insulated airflow ductwork, and the indoor airflow chamber. Air temperatures entering the indoor fan/coil unit were sampled by a nine-node thermocouple grid, and air temperatures exiting were measured by a 16-node thermocouple grid. Dewpoint temperatures were measured with chilled mirror hygrometers. The airflow ductwork consisted of two sets of turning vanes, which allowed the air to enter the indoor airflow chamber. The indoor airflow chamber was built to Air Movement and Control Association specifications 4. The airflow chamber allowed for accurate measurement of airflow through the indoor fan/coil unit. The outdoor heat exchanger, compressor, accumulator, and reversing valve were located in the outdoor psychrometric room, while the indoor heat exchanger and blower were located in the indoor psychrometric room. The outdoor heat exchanger was a two-row four-circuit vertical coil with wavy fins, a fin density of 7 fins c m - 1 (18 fins in-x), 9.53 mm (0.375 in) nominal diameter refrigerant tubes, and frontal dimensions of 61 cm (24 in) by 189.2 cm (75 in). Liquid- and vapour-line refrigerant circuit distribution tubes were instrumented with surface-soldered thermocouples. The lowest liquidline refrigerant distribution tube temperature was used to signal the end of a defrost operation. A three-blade axial flow fan blade with a 0.25 kW (0.33 hp) motor was used to pull air across the outdoor coil.
109
Defrost cycle performance of an air-source heat pump The indoor coil was a three-row, four-circuit A-type coil with 5.5 fins cm-1 (14 fins in-1). A centrifugal fan was included in the fan/coil unit and pulled air across the indoor heat exchanger. The indoor heat exchanger and the outdoor heat exchanger were connected by refrigerant tubing referred to as the 'liquid' and 'vapour' lines. The liquid line consisted of approximately 9.5 m (31 ft) of 9.53 mm (0.375 in) diameter tubing. The liquid line contained both the heating- and cooling-mode expansion devices. The heating-mode expansion device was a piston-type orifice with a diameter of 1.5 mm (0.059 in). The cooling/defrost-mode expansion device was a thermostatic expansion valve (TXV). The TXV superheat control bulb was attached to the vapour line close to the indoor heat exchanger. The vapour line included 9.2 m (30 ft) of 15.9 mm (0.625 in) diameter tube connecting the indoor coil to the reversing valve and 1.5 m (5 ft) of tubing connecting the reversing valve to the outdoor heat exchanger. The refrigerant circuit was instrumented with thermocouple probes, pressure transducers, and flowmeters (Table 1). Pressures throughout the refrigerant circuit were measured with 0-2.1 MPa (0-300 psia) pressure transducers calibrated with a deadweight tester. Temperature probes were placed throughout the refrigerant circuit in close proximity to the pressure transducers to allow accurate prediction of the refrigerant characteristics. Refrigerant flowrate was measured by two parallel Coriolis effect mass flowmeters in the refrigerant liquid line. Pressure drop across the mass flowmeters was limited to a maximum of 69 kPa (10 psia) at maximum refrigerant flowrate in accordance with ASHRAE Standard 116-835 . Valves were lever-actuated ball valves. These valves allowed for component testing, leak detection, and system component isolation due to possible failure. The test procedure consisted of selecting test conditions, setting the system refrigerant charge, and developing a test procedure to ensure repeatable results. The test conditions chosen were those specified for the frost build-up tests in the ASHRAE Standard 116-835. Test conditions for this procedure are given in Table 2. The refrigerant charge was set in the cooling mode for both compressor configurations, with indoor conditions of 26.7 °C (80 "F) dry-bulb, 19.4 °C (67 °F) wet-bulb, and
Table 2 Test conditions
Tableau 2 Conditionsd'essai Limits (°c)
Location
Dry-bulb ("C)
Wet-bulb UC)
Frosting
Defrosting
Indoor Outdoor
21.1 1.7
15.5 max 0.5
+0.56 +0.56
+ 1.1 ___1.1
outdoor conditions of 35 °C (95 °F). The unit was charged until the degree of refrigerant subcooling leaving the outdoor heat exchanger reached 5+1.1°C (9+2°F). All charging procedures followed manufacturer's recommendations. The control electronics installed by the manufacturer were removed to allow for manual initiation and termination of the defrost during heating-mode operation. Defrost was initiated at a 20% drop in peak heating capacity occurring during the frosting period. This manual procedure was intended to mimic the defrost controls supplied by the manufacturer. All tests used this criterion for the initiation of defrost. Defrost was terminated based upon temperature measurements on the lowest liquid-line circuit of the outdoor heat exchanger. Defrost was terminated when the circuit temperature reached 26.7+C (80°F). The scroll and reciprocating compressor system configurations used the same defrost scheme. Results
Frosting period The original compressor supplied by the heat pump manufacturer was a scroll, which was sized to yield a nominal cooling capacity of 10.6 kW (3 tons) during cooling. The reciprocating compressor tested with the same heat pump configuration was sized to yield the same cooling performance. Table 3 lists manufacturersupplied data describing the two compressors. The scroll and reciprocating compressors were matched as closely as possible for capacity and efficiency. Data in Table 3 were collected at a 7.2 °C (45 °F) evaporating temperature, a 54.4°C (130°F) condensing temperature, a 46.1 'C (115 OF) liquid temperature and a 35 °C (95 °F) ambient temperature. Testing of the heat pump under the experimental conditions showed that the unit equipped with the different compressors yielded approximately the same heating capacity (Table 4). Heating capacity for the two systems peaked at 8.4 kW (2.38 tons) and 8.5 kW (2.41 tons) for the scroll and reciprocating systems respectively. Heating capacity for the scroll system averaged only 0.6% lower than the reciprocating system during the frosting period. Figure 1 shows the heating performance of the differently configured systems during the frosting period. Figure 2 shows the power demand for the two systems during the frosting period. Power demand for the two systems peaked at 2.9 kW and 3.1 kW for the scroll and reciprocating systems respectively. The reciprocating system's power demand averaged 7% higher than that of the scroll system. The increased power demand of the reciprocating system caused the instantaneous COP to be lower than that of the scroll system (Figure 3). Instantaneous COP averaged 4% lower for the reciprocating system during the frosting period. The slow decline in power demand seen with the reciprocating
Table 3 Compressorcharacteristics
Tableau 3 Caracff'ristiquesdes compresseurs Type Scroll Reciprocating
Displacement (cm3rev- 1)
Capacity (kW)
Power (W)
COP
Current draw (A)
46.05 61.94
9.964 10.257
3090 3220
3.22 3.18
15.4 13.8
V. Payne and D. L. O ' N e a /
1 10
Table 4 Cycle performance comparison Tableau 4 Comparaison de la performance des cycles de d~givrage
Case
Maximum heating capacity (kW)
Integrated cyclic COP
Cycle time (rain)
Defrost time (rain)
Condensate (ml)
8.4 8.5
2.19 2.05
66.16 64.63
6.8 6.5
1680 1640
Scroll Reciprocating
10-
3
8--
>~ <
3
L32 ...... 0 (D L28 0
:24 x
X
2
6-
: 2 0 -q 16 ~
4-
0
O
SCROLL RECIP.
12 ;~
< (._)
1
8 " Indoor:
~l.lC,
[ Outdoor:
S C R O L L
15C
1.67C,
1[
wb
4
95*/* RH[
RECIP. 17
22
27
32
37
42
47
52
57
10-5-91
&
5-22-92
Indoor:
21 .IC,
15C
Outdoor:
1.67C,
95%
wb RH
0
r i i i i i i i i i i [ i T Fi ] , 1 i r ~ , i i i i q i r i i i T I P~ I~
12
< m < ~
62
0
Time Relative t o Defrost (min)
-r7~r777
12
n n I1
17
~ ~ t t ~ r r I r I ~ t777-~
22
27
32
37
F r I I P ~ n I I I r F I I ~ J n n I n n I I I~
42
47
52
57
62
0
T i m e RelaLive t o Defrost. ( r a i n )
Figure l
Frosting period heating capacities
Figure 1
CapacitOs de chauffage en pbriode de givrage
Figure 3
Frosting period instantaneous COPs
Figure 3
COP instantanbs en pbriode de yivra#e
_
L~ 0
- --
SCROLL RECIP.
10-5-91 Indoor:
&
5-22-92
21.1C,
15C
wb
o RH Outdoor: 1.67C,95'~
~
12
~
17
iPnlnlP
22
Time
27
Jpnnnll
32
Relative
r ur u l e u i n l l P J j l l ~ l n l n
37
42
to
47
52
Defrost
Figure 2
Frosting period power demands
Figure 2
Demandes d'blectricitb en pbriode de yivrage
r u ~
57
62
(min)
system was a function of the refrigerant flowrate and the outdoor fan power demand. As the outdoor heat exchanger was covered with frost, refrigerant flowrate decreased and outdoor fan power increased. The decrease in refrigerant flowrate caused the compressor power
demand to drop. The scroll system does not show a gradual decline in power demand because the reduction in power demand caused by decreasing refrigerant mass flowrate was equally matched by the increasing power demand of the outdoor fan. The reciprocating compressor consumed more energy than the scroll compressor, and therefore the outdoor fan power was a lower percentage of the total power for the reciprocating system. Another difference in the two compressors was the degree of discharge superheat produced. During the frosting period, discharge superheat peaked at 28 °C (50 °F) and 30 °C (54 °F) for the scroll and reciprocating compressors respectively. Figure 4 shows the compressor discharge and suction superheat during the frosting period. Discharge superheat gradually increased during the frosting period, owing to reduced refrigerant flowrate. As the difference in suction superheats of the two compressors was within 1%, the higher discharge superheat of the reciprocating compressor was due to its greater power demand. Figure 5 shows the refrigerant flowrates of the two compressors during the frosting period. Refrigerant flowrate for the two systems followed the same trends, peaking at 2.7 kg m i n - i (6 Ibm m i n - 1) after defrost and dropping to 1.9 kg m i n - 1 (4.2 Ibm min-1) at defrost initiation. Refrigerant flowrate for the reciprocating compressor averaged 2% higher than the scroll compressor during frost formation. Refrigerant flowrate decreased in both cases, owing to frost formation on the outdoor heat exchanger. Frost reduced the
111
Defrost cycle performance of an air-source heat pump
for the scroll and reciprocating compressors respectively. Suction pressure for the reciprocating system averaged 4% lower than the scroll system. Suction pressures fell to 102 kPa (15 psia) and 108 kPa (16 psia) for the scroll and reciprocating compressors respectively. Refrigerant flowrate during defrost peaked at 3.5 kg min- ~ (7.7 Ibm min- ~) and 3.7 kg min- ~ (8.2 Ibm min- ~) for the scroll and reciprocating systems respectively. Refrigerant flowrate for the reciprocating system averaged 3% higher than the scroll system. Figure 7 shows the refrigerant flowrates for both systems during the defrost. At defrost initiation, both systems show the refrigerant flowrate dropping near zero then steadily increasing as the reduced suction pressure causes the evaporation of more refrigerant. At 4 min the refrigerant flowrate of the reciprocating system equals that of the scroll system. After 4 min, the refrigerant flowrate of the reciprocating system increased over the scroll system, peaking at 3.7 kg min- 1 (8.2 Ibm min-a) at defrost termination. The larger refrigerant flowrate of the reciprocating system produced a greater defrost capacity than that of the scroll system. Defrosting capacity peaked at 9.0 kW
51 46 41t--
i36 "-~
-al < i e6 :< UZ
< D-q
- -
SCROLL RECIP
-
~13 L<
16
d~
8
U)
11
10-5-91 & 5-22-92 lndo~: 21.1C, 15C wb Outdoor: 1.67C, 95% RH
3-
5'?
6 -4
Time ReLative to Defrost (min) Figure 4 Frosting period discharge superheats Figure 4 Surchauffes au refoulement en p&iode de oivraoe
3.0
1500) 6.0
. . •2 . 5
/
E :5.0 E
•
2~
=
ieoo
1250~ .....
175
..0
~2.0 0 ..a
o ,...2 r, 3 . 0 E-Z <
~i.5 z< c~l.O !T~ i-
10-5-91 & 5-22-92
]
~05
0.
2.0
SCROLL
RECIP.
0
Indoor: 21.1C, 15C wb Outdoor: 1.67C. 95% KH
~llllrT
la
I '1 'IK[I
17
a2
Time
'l,
27
[ II 'l I ''1
32
Relative
37
to
'[1'
'1['~'1
42
47
' , ,I],
52
Defrost
57
125 750 03
:~ o r~ r.
10
' '11
'
62
150 ~
~.~ooo !
4.0
3 5 0 ~q
0.0
SCROLL RECIP
0 ~-~-~,-~
5
Time
(min)
Figure 5 Frosting period refrigerant flowrates Figure 5 Dbbits masse de frigorigbne en p&iode de #ivraye
......
0 Relative
L °'"~':'~'~'_22_" I [ e5 ~
5 to
F0
10 15 DefrosL (rain)
Figure 6 Defrosting period discharge absolute pressures Figure 6 Pressions absolues de refoulement en p&iode de dbgivrage
air-to-refrigerant heat energy transfer, causing a reduction in the mass of refrigerant evaporated in the outdoor heat exchanger.
4U 35 "N
---
SCROLL RECIP
:7o S
b
Defrostin9 period The defrost time for the reciprocating system was 5% faster than the scroll system. This was partially due to the lower volume of condensate collected from the reciprocating system. The reciprocating system produced 1640 ml (100 in 3) of condensate, while the scroll system produced 1680 ml (103 in 3) of condensate. The smaller amount (2.4%) of condensate for the system equipped with the reciprocating compressor was probably caused by the 1.5 min shorter cycle time. Figure 6 shows the discharge pressure produced by the two compressors during the defrost. Discharge pressure for the reciprocating system averaged 3% higher than the scroll system during defrost. Discharge pressures peaked at 1315 kPa (191 psia) and 1351 kPa (196 psia)
.........
_0
~25 50
h-,
r,
~210
t4o
z< ~15
30
1.0
"~
LQ
20 K Il 10-5-91& 5-22-92
0.5 OO
z <
Indoor: 21.1C, 15C wb Outdoor: 1.67C, 95% RH
-5 Time
0 Relative
5 to
10 Defrost (rain)
il.o X 5
oo
Figure 7 Defrosting period refrigerant flowrates Figure 7 Dbbits masse de frigorio~ne en p&iode de d~givrage
112
V. Payne and D. L. O'Neal F
c-.
6
<
-2
0
-4
i4~ <~
-8
-_
-10-
4
'\~
\~\\ SCROLL -- Rb~CIP.
......................... 0
'~
~J
5
24<
10-5-91& 5-22-92
lndoor:21.1C.15Cwb Outdoor: 1.67C, 95% RH
-. . . . . . . . . . . . . 10
15
34
Time Relative to Defrost (min) Figure 8 Defrostingperiod heating capacities Figure 8 Capacit~s de chauffage en p~riode de d~givra9e (2.5 tons) and 9.1 kW (2.6 tons) for the scroll and reciprocating systems respectively. Figure 8 shows the defrost capacity for the two systems. Although the different compressors producetl comparable defrosting capacities, the reciprocating compressor produced discharge gas with a higher degree of superheat than the scroll compressor. In addition to having a higher refrigerant flowrate during defrost, the reciprocating system produced vapour with a higher degree of superheat than the scroll system. Discharge superheat for the two systems peaked at 47 °C (85 °F) and 53 °C (95 °F) respectively. Discharge superheat for the reciprocating system averaged 18% higher than that of the scroll system. Given that the reciprocating system produced higher refrigerant flowrates and higher discharge superheats than the scroll system, the reciprocating system produced higher-energy vapour to defrost the outdoor heat exchanger. Although the reciprocating system produced higher-energy refrigerant during defrost, the performance of the scroll system was the most efficient. Table 4 showed the integrated cyclic C O P for the system configured with the different compressors. For the performance measured between defrosts, the scroll system yielded an integrated C O P of 2.19 while the reciprocating system yielded an integrated C O P of 2.05. Conclusions
Frost accumulation and defrost tests of a 10.6 kW (3 ton) nominally sized air-source heat pump equipped with a scroll and then with a reciprocating compressor revealed the overall performance of both systems. The system
equipped with the scroll compressor provided a slightly higher integrated C O P of the two configurations. The 6% difference in COPs was almost negligible compared with the accuracy of the calculation (_+5%). The reciprocating compressor produced heating capacity that was within 1% of the original scroll compressor. The reciprocating compressor produced a higher discharge superheat during frosting and defrost. Given that both systems produced discharge pressures that were within 3% during defrost, the reciprocating compressor produced higher-energy discharge gas, which reduced the defrost time over that of the scroll-compressor-equipped system. These tests were meant to highlight differences in transient performance for a system equipped with two different compressor mechanisms. The data presented here are intended to provide a reference for determining differences in performance for a system with two different compressor mechanisms, all other variables being held constant. Although these tests were not a perfect comparison, owing to possible inequalities in compressor motors, the general trends should provide some insight into the transient operating characteristics of heat pumps equipped with scroll or reciprocating compressors. Both compressors were matched by the manufacturer on the basis of capacity and efficiency. Although sound measurements were not taken, the scroll compressor was noticeably quieter than the reciprocating compressor during frosting and at defrost initiation. The scroll compressor was also physically smaller than the replacement reciprocating compressor. This was due to the construction of the two compressors. The reciprocating mechanism and motor are springmounted inside the sealed casing whereas the scroll mechanism and motor are hard mounted to the casing. Acknowledgements
The authors wish to express their appreciation to the Texas Energy Research in Applications Program for funding for this project. References
Creux, L. 1905, Rotary engine, US Patent No. 801182 (1905) lnaba, T. et al. A scroll compressor with sealing means and low pressureshellside. Proc 1986 International CompressorEngineering Conf Purdue University, West Lafayette, IN (1986) Vol. III, 887-900 Ishii, N. et aL A study on dynamic behavior of a scroll compressor. Proc 1986 International Compressor Engineering Conf Purdue University, West Lafayette, IN (1986) VoL III, 901 916 Laboratory methods of testing fans for rating. ANSO/ASHRAE Standard 51-1985 (ANSI/AMCA 210-85) Methods of testing for seasonal efficiency of unitary airconditioners and heat pumps. ANSI/ASHRAE Standard 1161983