investigation on effects of piping on heating performance...
TRANSCRIPT
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Ziai LiDepartment of Building Science, Tsinghua University, Beijing
May 16, 2017
Investigation on effects of piping on heating performance of multi-split variable refrigerant
flow system
12th IEA Heat Pump Conference, 15th-18th May, 2017, Rotterdam
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Contents
Background
Motivation and method
Results and discussion
Conclusions and outlook
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Multi-split variable refrigerant flow (VRF) system
Types: Air-cooled heat pump type, heat recovery type, water-cooled type, gas engine heat pump type, etc.Configuration: ODU up to 70kW, IDU number up to 60, outdoor-to-indoor pipeline length up to 149m, height difference up to 46m
ü Space saving ü Fewer transportation loss ü Efficient part-load operationü Individual control of indoor unitsü Easy maintenance and management
Outdoor units
Indoor units
Refrigerant pipe
Office Hotel VillaHospital
• Widely applied in east Asian and European countries• Sales of VRF units shared larger than 40% of China
market of central air-conditioning products in 2016
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0.00.51.01.52.02.53.03.54.0
0% 20% 40% 60% 80% 100%
Syst
em C
OP
Part-loadratio
Firstfloor
Fifthfloor
Seventhfloor
FourthfloorSeventh floor: average COP =3.32 Fifth floor: average COP=3.11First floor: average COP=2.92
Issue about piping of VRF system in application
Field cooling performance of multi-split VRF systems in one office building in Beijing, 2005 (Note: applying centralized control method)
Data resource: Department of building science , Tsinghua University. Reports on public building energy saving diagnosis in the summer of 2005.
H Lpipe=L1+L2
L1
L2
Refrigerant pipe
Performance degradation caused by piping (L and H) in application
Concern on piping effect on VRF system performance
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Issue about piping of VRF system in application
Performance degradation caused by piping (L and H) in application
• Cooling mode: Lpipe ↑ à Psuc ↓ and Tshsuc ↑ à ηvcp ↓, mrcp ↓, Qe ↓ and COP ↓• Heating mode: Lpipe ↑ à Tcond ↓ à Qc↓ and COP ↓• Cooling/Heating mode: H↑à Insufficiency of ΔP of indoor EEVs or overpressure at
the entry of indoor EEVs
1
1’
23
3’
4
ΔPLP
ΔPGP
R410A, cooling mode
ΔPEEV1~N
ΔPEEV0
1
2’
23
4
5
ΔPLP
ΔPGP
R410A, heating mode
4’
ΔPEEV1~N
ΔPEEV0
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Research status of piping effect of VRF system Reference Approach Control method Condition Highlights
W. Shi, et al (2007)
Simulation, R22
Constant outlet superheat of evaporators
Rated cooling/heatingcondition (given fixed compressor speed)
The cooling and heating COP decrease 25% and 11% respectively at pipe length of 100m
K. Zheng, et al (2007)
Field test, R22 Not mentioned Field cooling/heating
condition
The cooling capacity degradation is mainly caused insufficiency of ΔP for indoor EEV caused by pressure drop along liquid pipe while the heating capacity degradation is mainly caused by the heat leakage to environment along gas pipe
Z. Guo(2008)
Experiment, R410A Not mentioned
Variable cooling conditions (given differentactive indoor units)
The effect of pipe length on performance under part-load cooling conditions is smaller than that under full load condition
X. Wang (2010)
Simulation, R410A
Constant suction pressure and continuous operation of indoor units
Rated cooling condition (given certain cooling capacity)
The cooling COP decreases as pipe length increases without bypass loop while it is almost not affected by pipe length when suction superheat is controlled by bypass loop
D. Zhou, et al (2011)
Simulation,R410A
Constant suction superheat and continuous operation of indoor units
Rated cooling condition (given certain cooling capacity)
The effect of pipe length on performance under part-load cooling conditions is less than that under full load condition
Y. Pan et al. (2012)
Simulation, R22
Constant outlet superheat of evaporators
Rated cooling condition (given fixed compressor speed)
The highest COP can be obtained when indoor units are located equally from the outdoor unit
Z. Li, et al (2016)
Simulation, R410A
Constant outlet superheat of evaporators
Rated cooling condition (given fixed compressor speed)
The cooling capacity reduces by 14% and the COP reduces by 15% as the main pipe length increases from 10 m to 190 m
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Research status of piping effect of VRF system
Brief summary
• The available simulation work focuses on pipe length effect on the cooling performance of VRF system given fixed compressor speed. There is lack of study on the piping effect on its heating performance given various heating load
• The control method of the simulated systems is either indefinite or unpopular currently
• The simulation assumed a constant subcooling degree of condenser exit instead of using refrigerant mass charge conservation under part-load conditions
• The experiment and field test study need more detailed information of control method of the tested system to analyse the piping effect under various operation conditions
This study works on the effect of pipe length and height difference on VRF system performance under part-load heating conditions with the current heating control method based on simulation
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Contents
Background
Motivation and method
Results and discussion
Conclusions and outlook
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Current heating control method of VRF systems
Compressor
Gas-liquid separator
Four-way valve
Outdoor heat exchanger
EEV0 EEV1
EEV2
EEVN
Indoor heat exchanger #1
…
Indoor heat exchanger #2
Indoor heat exchanger #N
Outdoor unit Indoor unit
Main liquid pipe
Main gas pipe
Secondary liquid pipe
Secondary gas pipe
Actuator Control Target / Set value
Indoor unit
Indoor fan User control Rated air volume
EEV1~N Automatic Target air temperature Tai =20℃& indoor HEX exit subcooling degree ΔTsc ≥ 1℃
Outdoor unit
Compressorspeed Automatic Constant discharge pressure set value
Tsat (Pdis, cp set)=54℃Outdoor fan Automatic Rated air volume
EEV0 Automatic Compressor suction superheat set value ΔTsh, suc, set
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Methodology
Model simplification
• Regardless of the effect of lubricating oil• Inlet refrigerant of gas-liquid separator is superheated and no refrigerant is
accumulated• No frosting occurred out of the outdoor heat exchanger
Establishment of steady-state heating model of VRF system
Validation of refrigerant pipe model
Component design of a VRF system
Simulation on heating performance using constant discharge pressure control method (variable pipe length and height
difference)
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Model development
[1] R. Koury, L. Machado, K. Ismail, Numerical simulation of a variable speed refrigeration system, International Journal of Refrigeration 24 (2001) 192-200.[2] Shao S, Li X, Shi W, et al. A universal simulation model of air-cooled condenser consisting of plate-fin-tube[C]//ASME 2003 Heat Transfer Summer Conference. American Society of Mechanical Engineers, 2003: 729-734.[3] S. W. Churchill, 1977. Friction-factor equation spans all fluid-flow regimes. Chemical Engineering 84: 91-92.[4] Zhou Qiangtai. Two-phase flow and heat transfer [M] . Water conservancy and hydropower press, 1990.
Variable speed compressor: Efficiency model [1]
cp th v i,cp/V vm n h=
cp cp o,cp i,cp loss( ) /h h fW m -=
o,cp s i,cpo,cp i,cp
s
|h hh h
h-
+=
v 1 i,cp o,cp, )( Pf Ph = s 2 i,cp o,cp, )( Pf Ph =
Fin-and-tube evaporator, condenser: One-dimensional distributed-parameter model [2]
EEV: Correlation for the variable area expansion device
Pressure drop along refrigerant pipeline : Two-phase flow empirical model
Single phase[3]:
Two-phase [4]:
eev i,eev i,eev o,eev( ) 2 ( )Dm C A z P Pr= -
o,eevi,eevh h= i,eev o,eev0.02005 +6.34 /DC r r=
2f
i
( )2
dP f Jdl D r
- = ×
( )1/1212 1.58 8 /f Re K -é ù= +ë û
1616 0.9
i
37530 7( ) 2.457 ln(( ) 0.27 )KRe Re D
eé ù= - +ê ú
ë û2
2L0fL0
i L
( )2fdP J
dl Df
r- = × ×
L0 8 kf c Re-= ×2 2 (2 )/2 (2 )/2 (2 )L0 1 ( 1)[ (1 ) ]q q qY bx x xf - - -= + - - +
GzL
L
( ) {1 (1 )}dP gdz
rr ar
- = × × - -
Potential pressure drop:
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Validation of refrigerant pipe model
Compressor
Gas-liquid separator
Condenser
EEV
Evaporator
Outdoor unit
Indoor unit
Liquid pipeGas pipeTP
TP
TP
TP
TP
T
P
Experimental data centre room air-conditioner for validation of refrigerant pipe model
• 5 HP variable-speed compressor• Indoor unit was located 30 m higher• 100-meter-long liquid pipe and a 100-meter-
long gas pipe• Wrapped with 18 mm insulation• Refrigerant mass flow rate is calculated based
the suction and discharge refrigerant state of the compressor by employing the compressor efficiency model
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Validation of refrigerant pipe model
Ambient temperature (oC) 45 35 25
Compressor frequency (Hz) 84 70 50 84 70 50 84 70 50
Measurement
Suction temperature (oC) 34.5 36.4 40.3 32.6 34.8 38.4 27.1 30.8 35.2
Suction pressure (MPa) 1.136 1.221 1.174 1.102 1.181 1.182 1.096 1.181 1.322
Discharge pressure (MPa) 3.612 3.478 3.205 2.953 2.818 2.567 2.631 2.547 2.418
Gas pipe pressure drop (MPa) 0.160 0.120 0.037 0.217 0.169 0.075 0.286 0.241 0.125
Liquid pipe pressure drop (MPa) 0.304 0.293 0.267 0.311 0.301 0.275 0.318 0.306 0.288
Simulation
Refrigerant mass flow rate (kg/h) 446.2 399.2 259.1 432.6 380.1 256.2 443.3 387.9 292.4
Gas pipe pressure drop (MPa) 0.155 0.121 0.043 0.191 0.148 0.074 0.227 0.173 0.092
Liquid pipe pressure drop (MPa) 0.295 0.297 0.290 0.334 0.333 0.315 0.360 0.347 0.330
Deviation of simulation
Gas pipe pressure drop -3.1% 0.8% 16.2% -12.0% -21.4% -1.3% -20.6% -19.2% -26.4%
Liquid pipe pressure drop -3.0% 1.4% 8.6% 7.4% 10.6% 14.6% 13.2% 13.4% 14.6%
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Solving procedure for VRF system heating simulation
[5] Li Z, Wang B, Li X, et al., Simulation on effects of subcooler on cooling performance of multi-split variable refrigerant flow systems with different lengths of refrigerant pipeline, Energy and Buildings, 126 (2016) 301-309.
Assuming fcp, Pi,cp, Po,MLP
Calculating mcp, Wcp, To,cp by compressor model
eh<eh' ?
End
Adjusting Pi,cp
Calculating Pi,e, hi,e, xeev0 by evaporator model and EEV model
Yes
Start
eQ<eQ' ?
Assuming mc, j, ho,SLP, j
Calculating Po,c, j, ho,c, j, xeevj by condenser model and EEV model
j=N?
emass<emass' ?
j=j+1
Adjusting Po,MLP
Inputting system configuration, operating condition, Mass_charge, ΔTsho,e,set, Po,cp,set, Qc,1~Qc,N
Calculating Pi,MLP, hi,MLP, Po,MGP, ho,MGP by liquid refrigerant pipe and gas
refrigerant pipe models, j=1
Adjusting mc, j
YesNo
No
YesNo
Yes
No
No
Calculating Pi,SLP,j, hi,SLP,j, Po,SGP,j, ho,SGP,j by liquid refrigerant pipe and gas
refrigerant pipe models
em<em' ?Adjusting fcp
Yes
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Contents
Background
Motivation and method
Results and discussion
Conclusions and outlook
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Configuration of VRF system for simulation
Compressor
Displacement, Vth 85 cm3/revRefrigerant R410A
Quantity 1Rotation speed 20~100 rps
Rated nominal cooling
performance (pipe
length=10m)
Rated rotation speed 60 rpsRefrigerant charge (circulating) 3.0 kg
Rated cooling capacity 28 kW
Rated suction & discharge pressure Tsat,suc=5.2 ℃, Tsat,dis=47.6℃
Indoor HEXRated cooling capacity 7.0 kW
Quantity 4
Outdoor HEXRated heat exchange capacity 35.4 kW
Quantity 1EEV Nominal diameter of EEV0/EEV1~N 6.0 mm/ 2.4 mm
Main pipe External diameter Do of main liquid/ gas pipe 12.7mm(liquid), 25.4mm(gas)
Addition of refrigerant charge 0.11 kg/m
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0 30 60 90 120 150 18030
40
50
60
70
80Compressor frequency
Qc=26.0kW Qc=23.6kW Qc=21.2kW Qc=18.8kW
Com
pres
sor f
requ
ency
(Hz)
Length of main gas pipe and main liquid pipe (m)
3.20
3.25
3.30
3.35
3.40
3.45
Inlet pressure of condensers Qc=26.0kW Qc=23.6kW Qc=21.2kW Qc=18.8kW
Pres
sure
(MPa
)0 30 60 90 120 150 180
3.0
3.2
3.4
3.6
3.8
4.0COP
Qc=26.0kW Qc=23.6kW Qc=21.2kW Qc=18.8kW
CO
P (W
/W)
Length of main gas pipe and main liquid pipe (m)
0.0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
4.0
Suction saturation temperature of compressor
Qc=26.0kW Qc=23.6kW Qc=21.2kW Qc=18.8kW
Tem
pera
ture
(o C)
Effect of main pipe length (Tai=20℃, Tao=7℃, RHao=50%, Tsat (Pdis, cp set)=54℃)
Results and discussion
As LMGP and LMLP increased from 5 m to 165 m, • Compressor frequency increases by 3~5 Hz to output the same heating capacity• Te reduces by about 0.5 oC• Pin of condensers reduces by 0.068 MPa (about 1oC of Tc drop) at Qc=26kW• COPcomp decreases from 3.28 to 3.11 at Qc=26 kW
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Effect of outdoor-to-indoor height difference (Tai=20℃, Tao=7℃, RHao=50%, Tsat(Pdis, cp set)=54℃)
-60 -40 -20 0 20 40 6030
40
50
60
70
80
90
100Compressor frequency
Qc=26.0kW Qc=23.6kW Qc=21.2kW Qc=18.8kW
Com
pres
sor f
requ
ency
(Hz)
Height difference between the outdoor unit and indoor units (m)
0.0
0.5
1.0
1.5
2.0
2.5
3.0
3.5Pressure difference of EEV0
Qc=26.0kW Qc=23.6kW Qc=21.2kW Qc=18.8kW
Pres
sure
diff
eren
ce(M
Pa)
-60 -40 -20 0 20 40 601.5
2.0
2.5
3.0
3.5
4.0
4.5
5.0
COP Qc=26.0kW Qc=23.6kW Qc=21.2kW Qc=18.8kW
CO
P (W
/W)
Height difference between the outdoor unit and indoor units (m)
0
5
10
15
20
25
30
35
40
Outlet subcooling degree of condensers Qc=26.0kW Qc=23.6kW Qc=21.2kW Qc=18.8kW
Subc
oolin
g de
gree
(o C)
Results and discussion
• Compressor frequency and COPcomp show little change (tiny decrease of COPcompcaused by length increase) while ΔTsc of condensers increase slightly
• Available ΔP for EEV0 decreased markedly (decreased from 2.25 MPa to 1.55 MPa at Qc=26 kW), affecting adjustability and system stability
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Contents
Background
Motivation and method
Results and discussion
Conclusions and outlook
20
Conclusions and outlook
Conclusions
• A model has been established for heating operation of multi-split VRF system applicable to analysis of piping effect and control method• According to simulation of a designed VRF system under the constant discharge pressure control method:
(1) Lengthening of horizontal pipe will reduce COP apparently in heating mode even under part-load conditions
(2) Height difference in range from -50 m to 50 m had low effect on COP but it would affect the available pressure drop of EEVs
Outlook
• Application of the developed model to study the effect of piping under more operation conditions and different control methods, aiming at a comprehensive evaluation on the application of multi-split VRF systems in buildings