- Email: [email protected]
Optimization of refrigeration system with gas-injected scroll compressor
international journal of refrigeration 32 (2009) 1544–1554
available at www.sciencedirect.com
w w w . i i fi i r . o r g
journal homepage: www.elsevier.com/locate/ijrefrig
Optimization of refrigeration system with gas-injected scroll compressor Baolong Wang*, Wenxing Shi, Linjun Han, Xianting Li Department of Building Science, Tsinghua University, Beijing 100084, PR China
article info
abstract
Article history:
Gas refrigerant injection has been proven as an effective method to improve the perfor-
Received 6 January 2009
mance of the scroll compressor and its refrigeration system under high compression ratio
Received in revised form
working conditions. Much research on the injected scroll compressor and its system has
16 June 2009
been conducted, but the universal control and design method is still lacking. A model of the
Accepted 20 June 2009
refrigeration system with a gas-injected scroll compressor is developed in this paper. With
Published online 30 June 2009
this model, the effects of gas injection on the system and component parameters are investigated. Based on the identified evaporator characteristics and thermodynamic
Keywords:
analysis, a set of general principles for the design and operation of the refrigeration or heat
Refrigeration system
pump system with a gas-injected scroll compressor is proposed. ª 2009 Elsevier Ltd and IIR. All rights reserved.
Compression system Scroll compressor Optimization Performance Intermediate injection Recommendation Design Heat pump
Optimisation d’un syste`me frigorifique a` compresseur a` spirale a` injection de gaz Mots cle´s : Syste`me frigorifique ; Syste`me a` compression ; Compresseur a` spirale ; Optimisation ; Performance ; Injection interme´diaire ; Recommandation ; Conception ; Pompe a` chaleur
1.
Background
Gas refrigerant injection or so-called economizer technology has been applied in refrigeration systems with screw or
multistage centrifugal compressors for years. Recently, it has been developed for application in scroll compressors. Compared to the reverse Carnot Cycle, the practical performance of the vapor compression refrigeration cycle is
* Corresponding author. Tel.: þ86 10 62786571; fax: þ86 10 62773461. E-mail address: [email protected] (B. Wang). 0140-7007/$ – see front matter ª 2009 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2009.06.008
international journal of refrigeration 32 (2009) 1544–1554
dis discharge e, eva evaporator glycol glycol i inner inj injection in, inlet inlet mid middle o, out outlet opt optimal sc subcooling set set point two-phase two-phase region
Nomenclature A h m Q p P T
area, m2 enthalpy, kJ/kg mass flow rate, kg/s heat transfer capacity, kW pressure, MPa power consumption, kW Temperature, K
Greek letters 4 loss factor D difference 3 iteration error
Abbreviations COP coefficient of performance EV expansion valve IHX inter heat exchanger optimal injection pressure for cooling capacity OIPCC
Subscripts c condenser cc cooling capacity com compressor
Condenser
System EV
IHX
Evaporator
Gas refrigerant injection system with flash tank Condenser Flash tank
Compressor
System EV
Injection EV
b
Compressor
Gas refrigerant injection system with IHX
Injection EV
degraded by the temperature difference in the heat exchangers, throttling loss and overheated gas compression loss. Fortunately, gas refrigerant injection seems to be a solution. Fig. 1 shows the two main configurations of gas refrigerant injection systems. By first-stage throttling (in flash tank system) or heat exchange between the main loop refrigerant and the injected loop refrigerant (in system with intermediate heat exchanger (IHX)), the throttling loss can be reduced by decreasing the specific enthalpy of the refrigerant entering the system expansion valve (EV) and the overheated gas compression loss can be cut down by decreasing the temperature in the compression pocket. Hereby, the gas refrigerant injection should enhance the performance of the vapor compression system. Actually, previous research (Winandy and Lebrun, 2002; Wang et al., 2009) has proven this:
a
1545
Evaporator
Fig. 1 – Configurations of gas refrigerant injection system (a) Gas refrigerant injection system with IHX (b) Gas refrigerant injection system with flash tank.
for the refrigeration or heat pump system under large compression ratio conditions, gas refrigerant injection cannot only increase the energy efficiency and capacity of the system but also enhance the reliability of the system by decreasing the discharge temperature of the compressor. Looking back at previous research on the gas refrigerant injection into scroll compressors, the research can be classified into two main groups. One is the thermodynamic analysis, simulation or experimental testing of the effects of gas injection on the compressor (Ayub et al., 1992; Dutta et al., 2001; Winandy and Lebrun, 2002; Yamazaki et al., 2002; Park et al., 2002; Cho et al., 2003;).The other is experimental research on the effects of gas injection on the system (Zehnder et al., 2002; Ma et al., 2003). The former kind of research can study the process of gas injection in the compressor, identify the thermodynamic principles at work, and reveal the effects of the injection parameters on the compressor performance, such as the power consumption, discharge temperature, compressor efficiency and so on. The latter kind of research can show the influence of the gas injection on the macro performance of the system, such as the capacity and COP, and demonstrate the technical potential of gas injection. However, those studies do not bring forward the design and control principles for the gas-injected system. Considering this, a few researchers attempted optimization research on the gas-injected system. Ma and Li (2007) investigated the effect of the injection pressure on an injected heat pump system with an IHX by experiments and analysis, and finally proposed an optimal injection pressure range for their system. Beeton and Pham (2003) analyzed the impact of the gas injection on an IHX system and advanced an economical selection method for the IHX. Unfortunately, these optimization studies always were focused on one specific component or parameter, such as the IHX or injection pressure. They also grossly simplified the analysis, for instance, considering the compression process of gas-injected scroll compressor as a two-stage compression process with intermediate mixing. Consequently, the available research on refrigeration systems
international journal of refrigeration 32 (2009) 1544–1554
with gas-injected scroll compressors cannot provide general principles for the optimal control and design of the injected system. Based on previously detailed research on the gas-injected scroll compressor (Wang et al., 2008, 2009), a model of the refrigeration system with a gas-injected scroll compressor is developed. With this model, the effects of gas injection on the system with IHX and component parameters are investigated. Based on the identified evaporator characteristics and thermodynamic analysis, a set of general principles for the design and operation of the refrigeration system with the gasinjected scroll compressor is proposed.
2. Model of refrigeration system with gasinjected scroll compressor Compared to the simple one-unit refrigeration system, which includes one compressor, one evaporator, one condenser, and one expansion valve, the obvious difference of the gas-injected refrigeration system is having a more complicated system configuration and a scroll compressor with multiple suction processes. Therefore, the
distributed-parameter models of the evaporator, condenser, IHX, and EV can be built up the same as they was previously. However, the lumped parameter compressor model used in the simulation of simple oneunit refrigeration system cannot be used in the simulation of the injected scroll compressor because there is continuous refrigerant inflow or outflow in the compression process and the refrigerant states in the compression pocket keep changing during the whole injection process. Therefore, a distributed-parameter model of the injected scroll compressor is needed (Wang et al., 2008), which can be used to investigate the effects of injection on the whole process of the compressor. Another important problem that must be solved prior to the simulation of the gas-injected system is handling the interfaces between the compressor, evaporator and condenser, and constructing the system simulation flowchart. The calculation of the suction and discharge parameters by the distributedparameter compressor model is related to the rotating angle and varies periodically. However, because the time scale of the compressor is far smaller than the time scale of the system, time-average interfaces must be used between the compressor model and the other components, which means the output of
Input configuration and operating parameters
Assume pe, pc and pinj
Compressor model: Tdis, m, Pcom, etc
Δ Tsc, etc
ΔTsc −Δ Tsc , set < ε sc
Modify pinj
Condenser model: hc,out,
Modify pc
IHX (inverse calculation): hihx,in, etc
hihx,in − hc ,out < εh1
Evaporator (inverse calculation): he,in, etc
he ,in − hihx,out < ε h 2
END Fig. 2 – Flowchart of modeling of gas-injected refrigeration system.
Modify pe
1546
1547
international journal of refrigeration 32 (2009) 1544–1554
compressor model should be averaged over a period for the input to the other components. For calculation time saving, the conventional sequential component method (Ding, 2007) is used. The flowchart of the simulation is shown in Fig. 2. How to converge the injection loop is the particular work of injected system simulation compared to the conventional simple refrigeration system. The convergence method used in the calculation of the injection loop is quite similar to the method used in the evaporator simulation of the simple refrigeration system. In the simple refrigeration system simulation, if the superheated degree at the outlet of the evaporator is fixed, the evaporator can be calculated from its outlet to inlet to get the inlet specific enthalpy and pressure. Because the throttling process in the EV can be considered as an adiabatic process, the inlet specific enthalpy of the evaporator should be the same as the outlet specific enthalpy of the condenser if the assumed outlet pressure is proper. Otherwise, the outlet pressure of the evaporator should be reassumed for a recalculation. In the gas injection system simulation, the refrigerant superheated degree or specific enthalpy or pressure at the injection port should be determined prior to simulation. Based on it, the injection refrigerant mass flow rate can be calculated by the injected compressor model as well as the refrigerant mass flow rate through the evaporator. So, the IHX can be similarly calculated from the outlet of injection loop (at injection port) to its inlet (at the outlet of the injection EV). Comparison between the injection loop specific enthalpy at the inlet of the
3. Effect of gas injection on system parameters Previous research (Wang et al., 2008) shows that the refrigerant injection affects the inner parameters of the compressor as well as the discharge parameters, such as the discharge temperature, specific enthalpy, and mass flow rate. The change in these discharge parameters will influence the performance of the condenser directly. At the same time, the variation of the input parameters of the condenser will affect the outlet parameter of condenser. The change of the outlet parameters of condenser as well as the use of the economizer will greatly impact the inlet states of evaporator, which will influence the performance of the evaporator. The
b 14
Heating capacity 18 16
Experimental Qe/kW
Experimental Qc/kW
a
IHX and the specific enthalpy at the outlet of the condenser can be the criteria of convergence. To validate the model, a multifunctional test plant is built (Wang et al., 2007), which can measure the performance of the injected system under different conditions. Fig. 3 shows the comparison of the test and simulation results under different injection conditions (5 gas-injected conditions and 5 no-injection conditions). It can be concluded that the complex injected system model has an average accuracy similar to the previous simple refrigeration system and is suitable for the numerical research of the gas-injected refrigeration system.
+8
14 12
-8 Injected
10
Non-injected
8 6
8
10
12
14
16
18
Experimental P/kW
+10
10 -10
8 Injected
6
Non-injected
4
6
8
10
12
14
Numerical Qe/kW
Numerical Qc/kW
c
12
4 6
Cooling capacity
Power consumption 5
4
+5
-5
3
Injected Non-injected
2
2
3
4
5
Numerical P/kW Fig. 3 – Gas-injected system model validation (a) Heating capacity (b) Cooling capacity (c) Power consumption.
1548
international journal of refrigeration 32 (2009) 1544–1554
Table 1 – Configuration parameters of the simulated system. (Tube-in-tube)
Length (mm)
Evaporator Condenser
6200 6200
Compressor
Base circle radius (mm)
IHX
Inside diameter of outer tube (mm)
Outside diameter of inner tube (mm)
Inside diameter of inner tube (mm)
Number of inner tube
9.52 9.52
8.72 8.72
9 10
Thickness of scroll (mm)
Height of scroll (mm)
Cycles of scroll
Refrigerant
2.71
3.70
35.7
2.78
R22
Type
plate
Rated capacity (kW)
5
37 42.6
state variation of refrigerant in the injection tube and the suction tube will affect the performance of the compressor again. Finally, the system will reach a new balance. Therefore, the optimization research on the heat pump system with refrigerant injection should firstly begin with studying the influence of injection on the components. Based on these results, the optimal design and control can be pursued for different optimization targets. For the condenser, the gas injection will increase mass flow rate in the high pressure side and decrease the inlet refrigerant temperature. These two factors have quite different influences on the condensing pressure, heat release capacity, and other parameters of the condenser. For instance, the increase of the refrigerant flow rate will increase the condensing pressure and the heat rejection capacity. However, the decrease of the refrigerant inlet temperature
0.35
b 220
0.30
180
Mass flow rate in evaporator
me/kg/h
Evaporating pressure
pe/MPa
a
will decrease the condensing pressure and the heat rejection capacity of the condenser. Therefore, direct analysis of the condenser is complicated and unadvisable. For heat pump systems with a thermal EV or electrical EV, the superheated degree at the outlet of the evaporator can be easily controlled at a constant value. If the scroll compressor has a fixed frequency, the volume flow rate in the suction tube of the compressor also can be considered as constant due to the little influence of the injection on the volumetric efficiency of the scroll compressor. Therefore, the only free variable for the evaporator is the inlet specific enthalpy. The effect of refrigerant injection on the evaporator and system can be easily considered by researching the effect of refrigerant injection on the inlet specific enthalpy of the evaporator. So, comparatively, the research on the evaporator is more effective and clear.
0.25
140
Tinlet,glycol = -5 ˚C
0.20 280
240
200
Tinlet,glycol = -5 ˚C Tinlet,glycol = 0 ˚C
Tinlet,glycol = 0 ˚C 160
100 280
120
240
he, inlet/kJ/kg
c
200
160
120
he, inlet/kJ/kg
d 0.8
Cooling capacity 14
Relative two-phase area
12
0.7
Atwo-phase/Ae
Qe/kW
10 8 6 4
Tinlet,glycol = -5 ˚C
2
Tinlet,glycol = 0 ˚C
0 280
240
200
he, inlet/kJ/kg
0.6
Tinlet,glycol = -5 ˚C
0.5
Tinlet,glycol = 0 ˚C 160
120
0.4 280
240
200
160
120
he, inlet/kJ/kg
Fig. 4 – Influence of gas refrigerant injection on the evaporator (a) Evaporating pressure (b) Mass flow rate in evaporator (c) Cooling capacity, (d) Relative two-phase area.
1549
international journal of refrigeration 32 (2009) 1544–1554
Based on the foregoing analysis, a series of numerical experiments about the gas-injected refrigeration system are conducted and the focus is put on the evaporator. Configuration parameters of the system used in simulations are shown in Table 1. Fig. 4 shows the effects of the inlet specific enthalpy decrease due to the gas refrigerant injection on the performance of the evaporator when the inlet glycol temperature and flow rate of the evaporator are kept constant. Something should be mentioned here. The aforementioned distributedparameter compressor model is used in the calculation and the variation of the volumetric efficiency is considered in the calculation. Based on the results, it can be found, over a proper range, that a decrease in the inlet specific enthalpy of the evaporator due to effect enhancement of refrigerant injection can greatly enhance the heat flow of the evaporator but leads to little decrease in the evaporating pressure or mass flow rate in evaporator. For example, when the inlet temperature of the glycol is 0 C and the inlet specific enthalpy is decreased from 258 kJ/kg to 186 kJ/kg, the cooling capacity is increased 41% (from 8.29 kW to 11.7 kW), the evaporating pressure is decreased 4.5% (from 0.336 MPa to 0.321 MPa), and the mass flow rate is decreased 3.6% (from 0.055 kg/s to 0.053 kg/s). In other words, gas injection does not decrease the evaporating temperature very much during pursuing a larger cooling/ heating capacity. Actually, it can be attributed to the increase in the wetted region area and the integrated heat transfer coefficient in the evaporator, which means the evaporator’s capacity is far from the rated capacity before gas injection and most of the inner area is unwetted (Fig. 4(d)). From this reasoning, when the system is working under a nominal condition the effect of gas injection on the capacity will be much limited compared to the system under a low evaporating temperature. This is due to the fact that the evaporator under non-injection is more fully utilized at nominal temperature, which has been validated by many experiments (Ma et al., 2003). To sum up, for a heat pump system under a low evaporating temperature or a high condensing temperature condition, which is far away from the rated condition, the decrease of the inlet specific enthalpy will largely increase the cooling capacity of the evaporator but has little influence on the evaporating temperature and refrigerant flow rate on the low pressure side. That is to say, the smaller the inlet specific enthalpy is, the larger the cooling capacity of the evaporator is, which is the basis of the following optimal analysis.
4.
Optimal control
For a certain heat pump system with gas refrigerant injection, the aim of optimal control is to adjust the controllable components in the system to pursue optimization of one or several parameters of the system. For a common refrigeration system, the optimization target can be cooling capacity, heating capacity, COP for cooling or COP for heating. Because the optimal control of the main loop is the same as the optimization of a simple one-unit refrigeration system and has been widely investigated (Sanaye and Malekmohammadi, 2004; Khan and Zubair, 2001), this paper just focuses on the optimization of the injection loop. Therefore, the content of
Fig. 5 – p-h diagram of gas injection system.
the optimal control of the injection system is focused on controlling the injection pressure and specific enthalpy.
4.1.
Cooling capacity
According to the foregoing investigation on the evaporator, for a certain refrigeration system with fixed parameters of the coolant, fixed speed of compressor, and fixed superheated degree, the maximum cooling capacity can be obtained when the refrigerant inlet specific enthalpy of the evaporator is least. For the gas injection system with an IHX, the p-h diagram and schematic of the IHX are shown in Figs. 5 and 6. The refrigerant in the main loop goes into the IHX with T1 and goes out with T2. The refrigerant in the injection loop goes into the IHX with T4 and goes out with T5. Because heat transfer is limited by the driving temperature difference, T2 cannot be below T4. In addition, the isothermal line and the isenthalpic line almost coincide in the liquid region, so T2 will never reach T6 and the specific enthalpy of the main loop at the IHX outlet must be larger than the specific enthalpy of saturated liquid under injection pressure, h6. From this reasoning, the injection pressure should be as low as possible in order to decrease the refrigerant specific enthalpy entering the evaporator. On the other hand, for a scroll compressor with certain injection ports, the injected refrigerant flow rate will decrease with the decrease of the injection pressure, which leads to the decrease of the pressure difference for injection. If the injected refrigerant mass flow rate is too small (Fig. 7(a)), the refrigerant in the main loop cannot be subcooled to the saturation temperature of the injection pressure when all of the refrigerant in the injection loop has been completely
T1
T2 IHX T4
T5 Fig. 6 – Schematic of the IHX.
1550
international journal of refrigeration 32 (2009) 1544–1554
outlet of IHX is still in two-phase area, which means the injected refrigerant is not utilized fully. So, there must be an optimal injection pressure (Fig. 8). Under this optimal injection pressure, the refrigerant in the main loop can be subcooled to the saturation temperature at the injection pressure and the refrigerant temperature of the injection loop at the outlet of IHX should be very close to the outlet temperature of the condenser (T1). Based on the heat transfer equation of the IHX, the optimal injection pressure for cooling capacity (OIPCC) is: h i minj pinj; opt $ h5 T1 ; pinj; opt h1 i h ¼ meva $ h1 h2 Tinj pinj; opt ; p1 :
(1)
where, pinj,opt is the optimal injection pressure; Tinj is the saturation temperature at the optimal injection pressure; minj is the mass flow rate of the injection refrigerant; and meva is the mass flow rate of the refrigerant passing through evaporator. Also note that the optimal injection specific enthalpy can be expressed as following: hinj; opt ¼ h5 T1 ; pinj; opt :
(2)
where, hinj,opt is the optimal injection specific enthalpy. To sum up, there is an optimal injection condition making the cooling capacity maximum by subcooling the refrigerant in main loop to the minimum specific enthalpy.
4.2.
Fig. 7 – Influence of the injection pressure on refrigerant specific enthalpy at the evaporator inlet (a) low injection pressure (b) high injection pressure.
vaporized and overheated to T5 (equal to T1). In contrast, if the injection pressure is too high (Fig. 7 (b)), the refrigerant in the main loop will be subcooled to the saturated temperature of injection pressure while the refrigerant of the injection loop at
Fig. 8 – Optimization of the injection system for cooling capacity.
Heating capacity
As was previously mentioned, direct analysis of the heating performance of the condenser is difficult because the decrease of the inlet temperature and increase of the mass flow rate have completely different effects on the heating capacity of the condenser. Actually, based on energy conservation, the heating capacity of the heat pump can be expressed as an equation of cooling capacity and power consumption of the compressor. Qc ¼ Qe þ 4$Pcom
(3)
where, 4 is the loss factor of the compressor power due to heat rejection from the compressor shell. Therefore, based on the former research result that the compressor power always increases with the increase of the injection pressure (Wang et al., 2008) and the preceding conclusion that the cooling capacity will reach peak when the injection pressure is the OIPcc, some conclusions about heating capacity can be made: If the injection pressure is less than the OIPCC, the heating capacity as well as the cooling capacity of the system will increase with the increase of the injection pressure; When the injection pressure is larger than the OIPCC, the heating capacity of the heat pump has the potential to increase or decrease; If the heating capacity increases with injection pressure, it is easy to see that the increase in the heating capacity comes from the increase in the power consumption; In either case, the heating capacity will begin to decrease after a certain point due to the rapid decrease in the cooling capacity. Therefore, the heating capacity will reach the maximum when the injection is at a pressure bigger than the OIPCC.
1551
international journal of refrigeration 32 (2009) 1544–1554
4.3.
COP
The cooling COP is decided by cooling capacity and the compressor power consumption (the power consumption of the fans is constant and neglected). According to the previous research, the cooling capacity and power consumption will increase with the increase of the injection pressure until the OIPCC. After that, the cooling capacity will decrease while the power consumption continues increasing. Actually, when the injection pressure is higher than the OIPCC, the injection specific enthalpy will decrease with the increase of the injection pressure. In theory, the power consumption of the compressor may be decreased due to the cooling effect of refrigerant in the compression pocket. However, experiments (Ayub et al., 1992; Dutta et al., 2001) show that liquid injection cannot decrease the power consumption of the compressor and the power consumption will keep increasing when the injection pressure is higher than the OIPCC. Therefore, the maximum cooling COP will be reached when the injection pressure is less than or equal to the OIPCC. The actual injection pressure is related to the actual performance of the system and the compressor. However, further experimental research shows that the maximum cooling COP often appears at the close region of OIPCC and sometimes even equals OIPCC. Based on the definitions of COP along with equation (3), the heating COP equals the cooling COP added the constant 4. So the heating COP will also reach maximum when the injection pressure is smaller than or equal to the OIPCC.
Table 3 – Effects of gas injection on the system (Te,inlet,glycol [ 0 8C, Te,inlet,glycol [ 40 8C) Injection pinj/ Bar
Validation and discussion
To validate the foregoing analysis and simulation, some experiments are conducted. Table 2 shows the effects of gas injection on the system performance when the glycol inlet temperatures of the evaporator and the condenser are 5 C and 30 C, respectively. Table 3 shows the effects of gas injection on the system performance when the glycol inlet temperatures of the evaporator and the condenser are 0 C and 40 C, respectively. According to the test results, the previous analysis and conclusions can be validated: (1) gas injection has little influence on the evaporating pressure; (2) the maximum cooling
Table 2 – Effects of gas injection on the system (Te,inlet,glycol [ 5 8C, Te,inlet,glycol [ 30 8C) Injection
pinj/ Bar
System performance HC/ kW
CC/ kW
Evaporator
COPC COPH hinlet/kJ/ kg
pe/ Bar
No
–
11.61
8.59
2.68
3.62
239.77
4.65
Gas
7.53 8.07 8.43 8.93 9.38
13.04 13.34 13.44 13.59 13.48
9.83 10.43 10.40 10.22 9.88
2.83 2.90 2.83 2.70 2.58
3.76 3.71 3.66 3.59 3.52
226.08 223.10 225.27 227.42 229.42
4.71 4.66 4.65 4.70 4.75
The bold data are the maximum or the minimum values, which indicate the optimal points locations.
HC/ kW
CC/ kW
Evaporator
COPC COPH hinlet/kJ/ kg
pe/ Bar
No
–
10.23
6.80
1.77
2.66
253.94
4.27
Gas
7.18 8.20 8.74 9.18 9.98
11.45 12.01 12.28 12.40 12.47
7.66 8.25 8.19 8.13 7.94
1.87 1.93 1.85 1.81 1.71
2.80 2.81 2.77 2.76 2.69
235.25 225.24 228.52 230.96 234.75
4.16 4.07 4.10 4.08 4.14
The bold data are the maximum or the minimum values, which indicate the optimal points locations.
capacity can be achieved when the refrigerant inlet specific enthalpy of the evaporator reaches the minimum; (3) the heating capacity will reach the maximum when the injection pressure is a pressure bigger than the OIPCC; (4) the maximum cooling and heating COP will be reached when the injection pressure is less than or equal to the OIPCC. To sum up, according to the experimental validation, the foregoing analysis on the optimal control is valid and accurate.
5. 4.4.
System performance
Optimal design
The optimal design of the heat pump system with gas injection includes: optimal selection of the system configuration, optimal design of the injection tube, optimal design the injection port, and optimal design of the one-way valve.
5.1.
Optimal selection of system configuration
As the aforementioned, there are two well-known configurations of the gas refrigerant injection system, the IHX system and the flash tank system. From the viewpoint of thermodynamics, the flash tank has better heat exchange between the main loop refrigerant and injection loop refrigerant. Some experiments also proved that the flash tank configuration has better performance than the one with an IHX (Ma and Zhao, 2008). However, it is a pity that the gas injected into the compressor is saturated vapor. Its temperature is far lower than the outlet refrigerant temperature of the condenser, and the cooling ability of the refrigerant in the injection loop is not fully used. In theory, a hybrid economizer combined flash tank and IHX (Fig. 9) will decrease the optimal injection pressure and increase the performance of the injection system. Preliminary numerical experiments show it could increase the cooling capacity of the system 1.5%w3.5%. Actually, the small capacity scroll refrigeration system always uses an IHX as the economizer due to its convenience for control. In this kind of gas injection system, the principle factor behind choosing the IHX is a large heat exchange efficiency, which means larger heat exchange areas and the countercurrent configuration are preferred.
1552
international journal of refrigeration 32 (2009) 1544–1554
injection pressure and increases in the cooling capacity. Therefore, the design methods to increase the refrigerant mass flow rate in the injection loop will enhance the cooling capacity of the system. Therefore, the design principle for the injection tube is clear: shorten the injection tube and enlarge the diameter of the injection tube. For shortening the injection tube, the scroll compressor with an economizer attached to the shell is a typical case (Perevozchikov and Doepker, 2003). In addition, the preliminary research finds that the heat exchange between the injection refrigerant and the environment and discharge gas in the compressor will decrease the impact of the injection and decrease the efficiency of the injection system (Fig. 10). Therefore, thermal insulation of the injection tube is quite necessary for achieving a high efficiency injection system.
From Condenser To injection port Overheated Vapor
Saturated Vapor
5.3. Saturated liquid
To system expansion valve
Fig. 9 – Hybrid Economizer.
5.2.
Optimal design of injection tube
According to the preceding analysis, the cooling capability of the injection loop, which is directly related to the mass flow rate in the injection loop, restricts decreases in the optimal
Relative cooling capacity
Relative Cooling Capacity
a 1.3 1.2
1.1 insulated non-insulated 1 0.4
0.6
0.8
1
1.2
1.4
1.6
1.4
1.6
Injection Pressure/MPa Relative COP for cooling
Relative COP for Cooling
b 1.2 1
0.8 insulated non-insulated 0.6 0.4
0.6
0.8
1
1.2
Injection Pressure/MPa Fig. 10 – Effect of thermal insulation of injection tube on the gas-injected system (a) Relative cooling capacity (b) Relative COP for cooling.
Optimal design of injection port
The injection port causes the most significant pressure drop in the injection loop, so decreasing the pressure drop of the injection port is significant to decreasing the overall pressure drop in the injection loop. However, the radial size of the injection port should be smaller than the thickness of the scroll wrap to avoid inner leakage between different pockets, so the effort to enlarge the area of the injection port should be focused on the tangential dimension. Under these constraints, hybrid injection ports following the scroll profile are developed, such as the crescent injection port. The choice of the optimal location of the injection port is more complicated due to its different influence on the cooling capacity and COP. For cooling capacity, the closer the injection port is to the starting position of compression, the lower the average pocket pressure is during injection. It also means a possible larger mass flow rate in the injection loop, a lower optimal injection pressure, a smaller inlet specific enthalpy of the evaporator, and a larger cooling capacity of the system. But for the power consumption, the closer the injection port is to the starting position of compression, the larger the added indicated work is due to injection. Fig. 11(a) shows the p-V diagram of the scroll compressor with injection. It can be found, compared to the compression process without injection, that gas injection in a scroll compressor will reduce part of the under compression loss but lead to a throttling/mixing loss in addition to the ideal compression work. In other words, if the compression work addition caused by the injection wants to be minimized, the throttling/mixing loss and ideal compression work should be minimized. Consequently, an injection pressure close to the inner discharging pressure pi,dis, which refers to the pocket pressure when the inner compression is finished, an injection port which makes the pocket pressure of starting injection close to pi,dis, and a small injection pressure drop are needed (Fig. 11 (b)). Therefore, the selection of the optimal injection port location for maximizing COP is more complicated due to its different influence on the cooling capacity and the indicated work. For the injection system with a constant injection pressure, with moving the injection port location away from the start of compression to the end of compression, the pocket pressure will continue increasing and the actual injection mass flow rate will decrease gradually. But the cooling capacity will keep
1553
international journal of refrigeration 32 (2009) 1544–1554
a
First compression pocket
Effect of injection on the p-V diagram
Orbiting scroll
Static scroll
p Reduction of under-compression loss
Injection port C
po,dis
Possible location
Compression process with injection pi,dis
of injection port
Compression process without injection Throttling/mixing loss
pinj
psuc
V
b
A more efficient injection
Injection port B Injection port A
p Reduction of under-compression loss
Fig. 12 – Injection port location.
po,dis Ideal increase of compression work pi,dis pinj
Throttling/mixing loss
psuc
V
Fig. 11 – p-V diagram of the scroll compressor with injection (a) Effect of injection on the p-V diagram (b) A more efficient injection.
constant in the prior period due to over-quality injection refrigerant and the main loop refrigerant temperature at the outlet of IHX keeps constant as the saturation temperature at the injection pressure. However, in the later period, the cooling capacity will decrease due to the lack of injection refrigerant for subcooling the main loop refrigerant. However, the indicated work will decrease in this whole process. Therefore, the maximum COP will appear when the injection port is located between the location for optimal cooling capacity and the highest possible position (at that position, the pocket pressure at the start of injection equals the injection pressure). In other words, the injection port location for optimal COP should be higher than the injection port location for optimal cooling capacity. Combined with the previous optimal control analysis of the injection pressure, it can be concluded: there will be an optimal injection pressure for the maximizing cooling capacity of the injected system with a fixed injection port location; there will be an optimal injection port location for the maximizing cooling capacity of the injected system with a fixed injection pressure; the optimal injection pressure/ injection port location for maximizing COP is either the same as the optimal injection pressure/injection port location for
maximum cooling capacity or is varied in the direction that decreases the difference between injection pressure and the pocket pressure at the start of injection. Furthermore, the injection port should be located as closely as possible to the scroll wrap of the fixed scroll. A detailed analysis can find that the injection port located in the first compression pocket (Fig. 12) will lead to an injection to the suction pocket, which will largely decrease the system performance. The only possible location of the injection port to avoid injection into the suction chamber is between injection port A and injection port B (possible location of injection port in Fig. 12). But the injection ports between injection port A and injection port B, for example the injection port C, will inject refrigerant into two symmetric compression pockets sequentially, which will lead to unbalanced injection. Therefore, the injection port positions must be close to the wrap of the static scroll, such as ports A and B shown in Fig. 12.
5.4.
Optimal design of one-way valve
A one-way valve to stop the inverse flow of refrigerant in the injection pipe can increase the injection mass flow rate and enhance the injection performance. Because the application of the one-way valve will increase the pressure drop of the injection loop, a low pressure drop configuration of the oneway valve is preferred. At the same time, a one-way valve, which can control the opening time, will make the control of the injection location and period possible, which will largely enhance the adaptability of injection system under variable conditions.
6.
Conclusion
A model of the refrigeration system with gas-injected scroll compressor is described in this paper. With this model, the effects of gas injection on the system and component parameters are investigated. Based on the identified
1554
international journal of refrigeration 32 (2009) 1544–1554
evaporator characteristics and thermodynamic analysis, a set of general principles for the design and operation of the refrigeration or heat pump system with the gas-injected scroll compressor is proposed: (1) Optimal control: (a) Cooling capacity: the maximum cooling capacity can be achieved when the injection pressure and specific enthalpy are adjusted to minimize the refrigerant specific enthalpy at the inlet of the evaporator; (b) Heating capacity: the heating capacity may be keep constant or slightly increased when the injection pressure is larger than the OIPCC. (c) COP: the maximum COP will be reached when the injection pressure is less than the OIPCC and sometimes equal to it; (2) Optimal design: (a) System Configuration: the flash tank system has better potential; in an IHX system, a heat exchanger with large heat exchange area and countercurrent configuration is preferred; (b) Injection Tube: the design principle for the injection tube is clarified: shortening the injection tube, enlarging the diameter of the injection tube, and using thermal insulation; (c) Injection Port: a large injection port with a small radial size and close to the scroll wrap is required; for maximum cooling/heating capacity, the injection port should be as close as possible to the starting position of compression; for maximum COP, the injection port should be located at a position higher than the optimal injection port for cooling capacity and sometimes precisely at the optimal injection port position; (d) One-way Valve: a low pressure drop configuration of the one-way valve is preferred; a one-way valve which can control the opening time will make the control of the injection location and period possible.
Acknowledgement Financial support from National Natural Science Foundation of China (Grant No. 50676042) is greatly appreciated.
references
Ayub, S., Bush, J.W., Haller, D.K., 1992. Liquid refrigerant injection in scroll compressors operating at high compression ration, in: Proceeding of the International Compressor Engineering Conference at Purdue, USA, pp. 561–567.
Beeton, W.L., Pham, H.M., 2003. Vapor-injection scroll compressors. ASHRAE Journal 45 (4), 22–27. Cho, H., Chung, J.T., Kim, Y., 2003. Influence of liquid refrigerant injection on the performance of an inverter-driven scroll compressor. International Journal of Refrigeration 26, 87–94. Ding, G., 2007. Recent developments in simulation techniques for vapour-compression refrigeration systems. International Journal of Refrigeration 30 (7), 1119–1133. Dutta, A.K., Yanagisawa, T., Fukuta, M., 2001. An investigation of the performance of a scroll compressor under liquid refrigerant injection. International Journal of Refrigeration 24, 577–587. Khan, Jameel-ur-Rehman, Zubair, S.M., 2001. Thermodynamic optimization of finite time vapor compression refrigeration systems. Energy Conversion and Management 42 (12), 1457–1475. Ma, G., Chai, Q., Jiang, Y., 2003. Experimental investigation of air source heat pump for cold regions. International Journal of Refrigeration 26, 12–18. Ma, G., Li, X., 2007. Exergetic optimization of a key design parameter in heat pump systems with economizer coupled with scroll compressor. Energy Conversion and Management 48, 1150–1159. Ma, G., Zhao, H., 2008. Experimental study of a heat pump system with flash-tank coupled with scroll compressor. Energy and Buildings 40, 697–701. Park, Y.C., Kim, Y., Cho, H., 2002. Thermodynamic analysis on the performance of a variable speed scroll compressor with refrigerant injection. International Journal of Refrigeration 25, 1072–1082. Perevozchikov, MM., Doepker, R., 2003. Scroll compressor with vapor injection, US patent, US 2003/0136145–A1. Sanaye, S., Malekmohammadi, H.R., 2004. Thermal and economical optimization of air conditioning units with vapor compression refrigeration system. Applied Thermal Engineering 24 (13), 1807–1825. Wang, B., Li, X., Shi, W., Yan, Q., 2007. Design of experimental bench and internal pressure measurement of scroll compressor with refrigerant injection. International Journal of Refrigeration 30, 179–186. Wang, B., Shi, W., Li, X., Yan, Q., 2008. Numerical research on the scroll compressor with refrigeration injection. Applied Thermal Engineering 28 (5–6), 440–449. Wang, B., Shi, W., Li, X., 2009. Numerical analysis on the effects of refrigerant injection on the scroll compressor. Applied Thermal Engineering 29 (1), 37–46. Winandy, E.L., Lebrun, J., 2002. Scroll compressors using gas and liquid injection: experimental analysis and modeling. International Journal of Refrigeration 25, 1143–1156. Yamazaki, H., Itoh, T., Sato, K., Kobayashi, H., 2002. High performance scroll compressor with liquid refrigerant injection. In: Proceeding of the International Compressor Engineering Conference at Purdue, USA, C22–1. Zehnder, M., Favrat, D., Hohl, H., Olivier, C., Perevozchikov, M., 2002. High performance air-water heat pump with extended application range for residential heating. In: The Seventh International Energy Agency Conference on Heating Pumping Technologies, Beijing, China, pp. 19–22.
available at www.sciencedirect.com
w w w . i i fi i r . o r g
journal homepage: www.elsevier.com/locate/ijrefrig
Optimization of refrigeration system with gas-injected scroll compressor Baolong Wang*, Wenxing Shi, Linjun Han, Xianting Li Department of Building Science, Tsinghua University, Beijing 100084, PR China
article info
abstract
Article history:
Gas refrigerant injection has been proven as an effective method to improve the perfor-
Received 6 January 2009
mance of the scroll compressor and its refrigeration system under high compression ratio
Received in revised form
working conditions. Much research on the injected scroll compressor and its system has
16 June 2009
been conducted, but the universal control and design method is still lacking. A model of the
Accepted 20 June 2009
refrigeration system with a gas-injected scroll compressor is developed in this paper. With
Published online 30 June 2009
this model, the effects of gas injection on the system and component parameters are investigated. Based on the identified evaporator characteristics and thermodynamic
Keywords:
analysis, a set of general principles for the design and operation of the refrigeration or heat
Refrigeration system
pump system with a gas-injected scroll compressor is proposed. ª 2009 Elsevier Ltd and IIR. All rights reserved.
Compression system Scroll compressor Optimization Performance Intermediate injection Recommendation Design Heat pump
Optimisation d’un syste`me frigorifique a` compresseur a` spirale a` injection de gaz Mots cle´s : Syste`me frigorifique ; Syste`me a` compression ; Compresseur a` spirale ; Optimisation ; Performance ; Injection interme´diaire ; Recommandation ; Conception ; Pompe a` chaleur
1.
Background
Gas refrigerant injection or so-called economizer technology has been applied in refrigeration systems with screw or
multistage centrifugal compressors for years. Recently, it has been developed for application in scroll compressors. Compared to the reverse Carnot Cycle, the practical performance of the vapor compression refrigeration cycle is
* Corresponding author. Tel.: þ86 10 62786571; fax: þ86 10 62773461. E-mail address: [email protected] (B. Wang). 0140-7007/$ – see front matter ª 2009 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2009.06.008
international journal of refrigeration 32 (2009) 1544–1554
dis discharge e, eva evaporator glycol glycol i inner inj injection in, inlet inlet mid middle o, out outlet opt optimal sc subcooling set set point two-phase two-phase region
Nomenclature A h m Q p P T
area, m2 enthalpy, kJ/kg mass flow rate, kg/s heat transfer capacity, kW pressure, MPa power consumption, kW Temperature, K
Greek letters 4 loss factor D difference 3 iteration error
Abbreviations COP coefficient of performance EV expansion valve IHX inter heat exchanger optimal injection pressure for cooling capacity OIPCC
Subscripts c condenser cc cooling capacity com compressor
Condenser
System EV
IHX
Evaporator
Gas refrigerant injection system with flash tank Condenser Flash tank
Compressor
System EV
Injection EV
b
Compressor
Gas refrigerant injection system with IHX
Injection EV
degraded by the temperature difference in the heat exchangers, throttling loss and overheated gas compression loss. Fortunately, gas refrigerant injection seems to be a solution. Fig. 1 shows the two main configurations of gas refrigerant injection systems. By first-stage throttling (in flash tank system) or heat exchange between the main loop refrigerant and the injected loop refrigerant (in system with intermediate heat exchanger (IHX)), the throttling loss can be reduced by decreasing the specific enthalpy of the refrigerant entering the system expansion valve (EV) and the overheated gas compression loss can be cut down by decreasing the temperature in the compression pocket. Hereby, the gas refrigerant injection should enhance the performance of the vapor compression system. Actually, previous research (Winandy and Lebrun, 2002; Wang et al., 2009) has proven this:
a
1545
Evaporator
Fig. 1 – Configurations of gas refrigerant injection system (a) Gas refrigerant injection system with IHX (b) Gas refrigerant injection system with flash tank.
for the refrigeration or heat pump system under large compression ratio conditions, gas refrigerant injection cannot only increase the energy efficiency and capacity of the system but also enhance the reliability of the system by decreasing the discharge temperature of the compressor. Looking back at previous research on the gas refrigerant injection into scroll compressors, the research can be classified into two main groups. One is the thermodynamic analysis, simulation or experimental testing of the effects of gas injection on the compressor (Ayub et al., 1992; Dutta et al., 2001; Winandy and Lebrun, 2002; Yamazaki et al., 2002; Park et al., 2002; Cho et al., 2003;).The other is experimental research on the effects of gas injection on the system (Zehnder et al., 2002; Ma et al., 2003). The former kind of research can study the process of gas injection in the compressor, identify the thermodynamic principles at work, and reveal the effects of the injection parameters on the compressor performance, such as the power consumption, discharge temperature, compressor efficiency and so on. The latter kind of research can show the influence of the gas injection on the macro performance of the system, such as the capacity and COP, and demonstrate the technical potential of gas injection. However, those studies do not bring forward the design and control principles for the gas-injected system. Considering this, a few researchers attempted optimization research on the gas-injected system. Ma and Li (2007) investigated the effect of the injection pressure on an injected heat pump system with an IHX by experiments and analysis, and finally proposed an optimal injection pressure range for their system. Beeton and Pham (2003) analyzed the impact of the gas injection on an IHX system and advanced an economical selection method for the IHX. Unfortunately, these optimization studies always were focused on one specific component or parameter, such as the IHX or injection pressure. They also grossly simplified the analysis, for instance, considering the compression process of gas-injected scroll compressor as a two-stage compression process with intermediate mixing. Consequently, the available research on refrigeration systems
international journal of refrigeration 32 (2009) 1544–1554
with gas-injected scroll compressors cannot provide general principles for the optimal control and design of the injected system. Based on previously detailed research on the gas-injected scroll compressor (Wang et al., 2008, 2009), a model of the refrigeration system with a gas-injected scroll compressor is developed. With this model, the effects of gas injection on the system with IHX and component parameters are investigated. Based on the identified evaporator characteristics and thermodynamic analysis, a set of general principles for the design and operation of the refrigeration system with the gasinjected scroll compressor is proposed.
2. Model of refrigeration system with gasinjected scroll compressor Compared to the simple one-unit refrigeration system, which includes one compressor, one evaporator, one condenser, and one expansion valve, the obvious difference of the gas-injected refrigeration system is having a more complicated system configuration and a scroll compressor with multiple suction processes. Therefore, the
distributed-parameter models of the evaporator, condenser, IHX, and EV can be built up the same as they was previously. However, the lumped parameter compressor model used in the simulation of simple oneunit refrigeration system cannot be used in the simulation of the injected scroll compressor because there is continuous refrigerant inflow or outflow in the compression process and the refrigerant states in the compression pocket keep changing during the whole injection process. Therefore, a distributed-parameter model of the injected scroll compressor is needed (Wang et al., 2008), which can be used to investigate the effects of injection on the whole process of the compressor. Another important problem that must be solved prior to the simulation of the gas-injected system is handling the interfaces between the compressor, evaporator and condenser, and constructing the system simulation flowchart. The calculation of the suction and discharge parameters by the distributedparameter compressor model is related to the rotating angle and varies periodically. However, because the time scale of the compressor is far smaller than the time scale of the system, time-average interfaces must be used between the compressor model and the other components, which means the output of
Input configuration and operating parameters
Assume pe, pc and pinj
Compressor model: Tdis, m, Pcom, etc
Δ Tsc, etc
ΔTsc −Δ Tsc , set < ε sc
Modify pinj
Condenser model: hc,out,
Modify pc
IHX (inverse calculation): hihx,in, etc
hihx,in − hc ,out < εh1
Evaporator (inverse calculation): he,in, etc
he ,in − hihx,out < ε h 2
END Fig. 2 – Flowchart of modeling of gas-injected refrigeration system.
Modify pe
1546
1547
international journal of refrigeration 32 (2009) 1544–1554
compressor model should be averaged over a period for the input to the other components. For calculation time saving, the conventional sequential component method (Ding, 2007) is used. The flowchart of the simulation is shown in Fig. 2. How to converge the injection loop is the particular work of injected system simulation compared to the conventional simple refrigeration system. The convergence method used in the calculation of the injection loop is quite similar to the method used in the evaporator simulation of the simple refrigeration system. In the simple refrigeration system simulation, if the superheated degree at the outlet of the evaporator is fixed, the evaporator can be calculated from its outlet to inlet to get the inlet specific enthalpy and pressure. Because the throttling process in the EV can be considered as an adiabatic process, the inlet specific enthalpy of the evaporator should be the same as the outlet specific enthalpy of the condenser if the assumed outlet pressure is proper. Otherwise, the outlet pressure of the evaporator should be reassumed for a recalculation. In the gas injection system simulation, the refrigerant superheated degree or specific enthalpy or pressure at the injection port should be determined prior to simulation. Based on it, the injection refrigerant mass flow rate can be calculated by the injected compressor model as well as the refrigerant mass flow rate through the evaporator. So, the IHX can be similarly calculated from the outlet of injection loop (at injection port) to its inlet (at the outlet of the injection EV). Comparison between the injection loop specific enthalpy at the inlet of the
3. Effect of gas injection on system parameters Previous research (Wang et al., 2008) shows that the refrigerant injection affects the inner parameters of the compressor as well as the discharge parameters, such as the discharge temperature, specific enthalpy, and mass flow rate. The change in these discharge parameters will influence the performance of the condenser directly. At the same time, the variation of the input parameters of the condenser will affect the outlet parameter of condenser. The change of the outlet parameters of condenser as well as the use of the economizer will greatly impact the inlet states of evaporator, which will influence the performance of the evaporator. The
b 14
Heating capacity 18 16
Experimental Qe/kW
Experimental Qc/kW
a
IHX and the specific enthalpy at the outlet of the condenser can be the criteria of convergence. To validate the model, a multifunctional test plant is built (Wang et al., 2007), which can measure the performance of the injected system under different conditions. Fig. 3 shows the comparison of the test and simulation results under different injection conditions (5 gas-injected conditions and 5 no-injection conditions). It can be concluded that the complex injected system model has an average accuracy similar to the previous simple refrigeration system and is suitable for the numerical research of the gas-injected refrigeration system.
+8
14 12
-8 Injected
10
Non-injected
8 6
8
10
12
14
16
18
Experimental P/kW
+10
10 -10
8 Injected
6
Non-injected
4
6
8
10
12
14
Numerical Qe/kW
Numerical Qc/kW
c
12
4 6
Cooling capacity
Power consumption 5
4
+5
-5
3
Injected Non-injected
2
2
3
4
5
Numerical P/kW Fig. 3 – Gas-injected system model validation (a) Heating capacity (b) Cooling capacity (c) Power consumption.
1548
international journal of refrigeration 32 (2009) 1544–1554
Table 1 – Configuration parameters of the simulated system. (Tube-in-tube)
Length (mm)
Evaporator Condenser
6200 6200
Compressor
Base circle radius (mm)
IHX
Inside diameter of outer tube (mm)
Outside diameter of inner tube (mm)
Inside diameter of inner tube (mm)
Number of inner tube
9.52 9.52
8.72 8.72
9 10
Thickness of scroll (mm)
Height of scroll (mm)
Cycles of scroll
Refrigerant
2.71
3.70
35.7
2.78
R22
Type
plate
Rated capacity (kW)
5
37 42.6
state variation of refrigerant in the injection tube and the suction tube will affect the performance of the compressor again. Finally, the system will reach a new balance. Therefore, the optimization research on the heat pump system with refrigerant injection should firstly begin with studying the influence of injection on the components. Based on these results, the optimal design and control can be pursued for different optimization targets. For the condenser, the gas injection will increase mass flow rate in the high pressure side and decrease the inlet refrigerant temperature. These two factors have quite different influences on the condensing pressure, heat release capacity, and other parameters of the condenser. For instance, the increase of the refrigerant flow rate will increase the condensing pressure and the heat rejection capacity. However, the decrease of the refrigerant inlet temperature
0.35
b 220
0.30
180
Mass flow rate in evaporator
me/kg/h
Evaporating pressure
pe/MPa
a
will decrease the condensing pressure and the heat rejection capacity of the condenser. Therefore, direct analysis of the condenser is complicated and unadvisable. For heat pump systems with a thermal EV or electrical EV, the superheated degree at the outlet of the evaporator can be easily controlled at a constant value. If the scroll compressor has a fixed frequency, the volume flow rate in the suction tube of the compressor also can be considered as constant due to the little influence of the injection on the volumetric efficiency of the scroll compressor. Therefore, the only free variable for the evaporator is the inlet specific enthalpy. The effect of refrigerant injection on the evaporator and system can be easily considered by researching the effect of refrigerant injection on the inlet specific enthalpy of the evaporator. So, comparatively, the research on the evaporator is more effective and clear.
0.25
140
Tinlet,glycol = -5 ˚C
0.20 280
240
200
Tinlet,glycol = -5 ˚C Tinlet,glycol = 0 ˚C
Tinlet,glycol = 0 ˚C 160
100 280
120
240
he, inlet/kJ/kg
c
200
160
120
he, inlet/kJ/kg
d 0.8
Cooling capacity 14
Relative two-phase area
12
0.7
Atwo-phase/Ae
Qe/kW
10 8 6 4
Tinlet,glycol = -5 ˚C
2
Tinlet,glycol = 0 ˚C
0 280
240
200
he, inlet/kJ/kg
0.6
Tinlet,glycol = -5 ˚C
0.5
Tinlet,glycol = 0 ˚C 160
120
0.4 280
240
200
160
120
he, inlet/kJ/kg
Fig. 4 – Influence of gas refrigerant injection on the evaporator (a) Evaporating pressure (b) Mass flow rate in evaporator (c) Cooling capacity, (d) Relative two-phase area.
1549
international journal of refrigeration 32 (2009) 1544–1554
Based on the foregoing analysis, a series of numerical experiments about the gas-injected refrigeration system are conducted and the focus is put on the evaporator. Configuration parameters of the system used in simulations are shown in Table 1. Fig. 4 shows the effects of the inlet specific enthalpy decrease due to the gas refrigerant injection on the performance of the evaporator when the inlet glycol temperature and flow rate of the evaporator are kept constant. Something should be mentioned here. The aforementioned distributedparameter compressor model is used in the calculation and the variation of the volumetric efficiency is considered in the calculation. Based on the results, it can be found, over a proper range, that a decrease in the inlet specific enthalpy of the evaporator due to effect enhancement of refrigerant injection can greatly enhance the heat flow of the evaporator but leads to little decrease in the evaporating pressure or mass flow rate in evaporator. For example, when the inlet temperature of the glycol is 0 C and the inlet specific enthalpy is decreased from 258 kJ/kg to 186 kJ/kg, the cooling capacity is increased 41% (from 8.29 kW to 11.7 kW), the evaporating pressure is decreased 4.5% (from 0.336 MPa to 0.321 MPa), and the mass flow rate is decreased 3.6% (from 0.055 kg/s to 0.053 kg/s). In other words, gas injection does not decrease the evaporating temperature very much during pursuing a larger cooling/ heating capacity. Actually, it can be attributed to the increase in the wetted region area and the integrated heat transfer coefficient in the evaporator, which means the evaporator’s capacity is far from the rated capacity before gas injection and most of the inner area is unwetted (Fig. 4(d)). From this reasoning, when the system is working under a nominal condition the effect of gas injection on the capacity will be much limited compared to the system under a low evaporating temperature. This is due to the fact that the evaporator under non-injection is more fully utilized at nominal temperature, which has been validated by many experiments (Ma et al., 2003). To sum up, for a heat pump system under a low evaporating temperature or a high condensing temperature condition, which is far away from the rated condition, the decrease of the inlet specific enthalpy will largely increase the cooling capacity of the evaporator but has little influence on the evaporating temperature and refrigerant flow rate on the low pressure side. That is to say, the smaller the inlet specific enthalpy is, the larger the cooling capacity of the evaporator is, which is the basis of the following optimal analysis.
4.
Optimal control
For a certain heat pump system with gas refrigerant injection, the aim of optimal control is to adjust the controllable components in the system to pursue optimization of one or several parameters of the system. For a common refrigeration system, the optimization target can be cooling capacity, heating capacity, COP for cooling or COP for heating. Because the optimal control of the main loop is the same as the optimization of a simple one-unit refrigeration system and has been widely investigated (Sanaye and Malekmohammadi, 2004; Khan and Zubair, 2001), this paper just focuses on the optimization of the injection loop. Therefore, the content of
Fig. 5 – p-h diagram of gas injection system.
the optimal control of the injection system is focused on controlling the injection pressure and specific enthalpy.
4.1.
Cooling capacity
According to the foregoing investigation on the evaporator, for a certain refrigeration system with fixed parameters of the coolant, fixed speed of compressor, and fixed superheated degree, the maximum cooling capacity can be obtained when the refrigerant inlet specific enthalpy of the evaporator is least. For the gas injection system with an IHX, the p-h diagram and schematic of the IHX are shown in Figs. 5 and 6. The refrigerant in the main loop goes into the IHX with T1 and goes out with T2. The refrigerant in the injection loop goes into the IHX with T4 and goes out with T5. Because heat transfer is limited by the driving temperature difference, T2 cannot be below T4. In addition, the isothermal line and the isenthalpic line almost coincide in the liquid region, so T2 will never reach T6 and the specific enthalpy of the main loop at the IHX outlet must be larger than the specific enthalpy of saturated liquid under injection pressure, h6. From this reasoning, the injection pressure should be as low as possible in order to decrease the refrigerant specific enthalpy entering the evaporator. On the other hand, for a scroll compressor with certain injection ports, the injected refrigerant flow rate will decrease with the decrease of the injection pressure, which leads to the decrease of the pressure difference for injection. If the injected refrigerant mass flow rate is too small (Fig. 7(a)), the refrigerant in the main loop cannot be subcooled to the saturation temperature of the injection pressure when all of the refrigerant in the injection loop has been completely
T1
T2 IHX T4
T5 Fig. 6 – Schematic of the IHX.
1550
international journal of refrigeration 32 (2009) 1544–1554
outlet of IHX is still in two-phase area, which means the injected refrigerant is not utilized fully. So, there must be an optimal injection pressure (Fig. 8). Under this optimal injection pressure, the refrigerant in the main loop can be subcooled to the saturation temperature at the injection pressure and the refrigerant temperature of the injection loop at the outlet of IHX should be very close to the outlet temperature of the condenser (T1). Based on the heat transfer equation of the IHX, the optimal injection pressure for cooling capacity (OIPCC) is: h i minj pinj; opt $ h5 T1 ; pinj; opt h1 i h ¼ meva $ h1 h2 Tinj pinj; opt ; p1 :
(1)
where, pinj,opt is the optimal injection pressure; Tinj is the saturation temperature at the optimal injection pressure; minj is the mass flow rate of the injection refrigerant; and meva is the mass flow rate of the refrigerant passing through evaporator. Also note that the optimal injection specific enthalpy can be expressed as following: hinj; opt ¼ h5 T1 ; pinj; opt :
(2)
where, hinj,opt is the optimal injection specific enthalpy. To sum up, there is an optimal injection condition making the cooling capacity maximum by subcooling the refrigerant in main loop to the minimum specific enthalpy.
4.2.
Fig. 7 – Influence of the injection pressure on refrigerant specific enthalpy at the evaporator inlet (a) low injection pressure (b) high injection pressure.
vaporized and overheated to T5 (equal to T1). In contrast, if the injection pressure is too high (Fig. 7 (b)), the refrigerant in the main loop will be subcooled to the saturated temperature of injection pressure while the refrigerant of the injection loop at
Fig. 8 – Optimization of the injection system for cooling capacity.
Heating capacity
As was previously mentioned, direct analysis of the heating performance of the condenser is difficult because the decrease of the inlet temperature and increase of the mass flow rate have completely different effects on the heating capacity of the condenser. Actually, based on energy conservation, the heating capacity of the heat pump can be expressed as an equation of cooling capacity and power consumption of the compressor. Qc ¼ Qe þ 4$Pcom
(3)
where, 4 is the loss factor of the compressor power due to heat rejection from the compressor shell. Therefore, based on the former research result that the compressor power always increases with the increase of the injection pressure (Wang et al., 2008) and the preceding conclusion that the cooling capacity will reach peak when the injection pressure is the OIPcc, some conclusions about heating capacity can be made: If the injection pressure is less than the OIPCC, the heating capacity as well as the cooling capacity of the system will increase with the increase of the injection pressure; When the injection pressure is larger than the OIPCC, the heating capacity of the heat pump has the potential to increase or decrease; If the heating capacity increases with injection pressure, it is easy to see that the increase in the heating capacity comes from the increase in the power consumption; In either case, the heating capacity will begin to decrease after a certain point due to the rapid decrease in the cooling capacity. Therefore, the heating capacity will reach the maximum when the injection is at a pressure bigger than the OIPCC.
1551
international journal of refrigeration 32 (2009) 1544–1554
4.3.
COP
The cooling COP is decided by cooling capacity and the compressor power consumption (the power consumption of the fans is constant and neglected). According to the previous research, the cooling capacity and power consumption will increase with the increase of the injection pressure until the OIPCC. After that, the cooling capacity will decrease while the power consumption continues increasing. Actually, when the injection pressure is higher than the OIPCC, the injection specific enthalpy will decrease with the increase of the injection pressure. In theory, the power consumption of the compressor may be decreased due to the cooling effect of refrigerant in the compression pocket. However, experiments (Ayub et al., 1992; Dutta et al., 2001) show that liquid injection cannot decrease the power consumption of the compressor and the power consumption will keep increasing when the injection pressure is higher than the OIPCC. Therefore, the maximum cooling COP will be reached when the injection pressure is less than or equal to the OIPCC. The actual injection pressure is related to the actual performance of the system and the compressor. However, further experimental research shows that the maximum cooling COP often appears at the close region of OIPCC and sometimes even equals OIPCC. Based on the definitions of COP along with equation (3), the heating COP equals the cooling COP added the constant 4. So the heating COP will also reach maximum when the injection pressure is smaller than or equal to the OIPCC.
Table 3 – Effects of gas injection on the system (Te,inlet,glycol [ 0 8C, Te,inlet,glycol [ 40 8C) Injection pinj/ Bar
Validation and discussion
To validate the foregoing analysis and simulation, some experiments are conducted. Table 2 shows the effects of gas injection on the system performance when the glycol inlet temperatures of the evaporator and the condenser are 5 C and 30 C, respectively. Table 3 shows the effects of gas injection on the system performance when the glycol inlet temperatures of the evaporator and the condenser are 0 C and 40 C, respectively. According to the test results, the previous analysis and conclusions can be validated: (1) gas injection has little influence on the evaporating pressure; (2) the maximum cooling
Table 2 – Effects of gas injection on the system (Te,inlet,glycol [ 5 8C, Te,inlet,glycol [ 30 8C) Injection
pinj/ Bar
System performance HC/ kW
CC/ kW
Evaporator
COPC COPH hinlet/kJ/ kg
pe/ Bar
No
–
11.61
8.59
2.68
3.62
239.77
4.65
Gas
7.53 8.07 8.43 8.93 9.38
13.04 13.34 13.44 13.59 13.48
9.83 10.43 10.40 10.22 9.88
2.83 2.90 2.83 2.70 2.58
3.76 3.71 3.66 3.59 3.52
226.08 223.10 225.27 227.42 229.42
4.71 4.66 4.65 4.70 4.75
The bold data are the maximum or the minimum values, which indicate the optimal points locations.
HC/ kW
CC/ kW
Evaporator
COPC COPH hinlet/kJ/ kg
pe/ Bar
No
–
10.23
6.80
1.77
2.66
253.94
4.27
Gas
7.18 8.20 8.74 9.18 9.98
11.45 12.01 12.28 12.40 12.47
7.66 8.25 8.19 8.13 7.94
1.87 1.93 1.85 1.81 1.71
2.80 2.81 2.77 2.76 2.69
235.25 225.24 228.52 230.96 234.75
4.16 4.07 4.10 4.08 4.14
The bold data are the maximum or the minimum values, which indicate the optimal points locations.
capacity can be achieved when the refrigerant inlet specific enthalpy of the evaporator reaches the minimum; (3) the heating capacity will reach the maximum when the injection pressure is a pressure bigger than the OIPCC; (4) the maximum cooling and heating COP will be reached when the injection pressure is less than or equal to the OIPCC. To sum up, according to the experimental validation, the foregoing analysis on the optimal control is valid and accurate.
5. 4.4.
System performance
Optimal design
The optimal design of the heat pump system with gas injection includes: optimal selection of the system configuration, optimal design of the injection tube, optimal design the injection port, and optimal design of the one-way valve.
5.1.
Optimal selection of system configuration
As the aforementioned, there are two well-known configurations of the gas refrigerant injection system, the IHX system and the flash tank system. From the viewpoint of thermodynamics, the flash tank has better heat exchange between the main loop refrigerant and injection loop refrigerant. Some experiments also proved that the flash tank configuration has better performance than the one with an IHX (Ma and Zhao, 2008). However, it is a pity that the gas injected into the compressor is saturated vapor. Its temperature is far lower than the outlet refrigerant temperature of the condenser, and the cooling ability of the refrigerant in the injection loop is not fully used. In theory, a hybrid economizer combined flash tank and IHX (Fig. 9) will decrease the optimal injection pressure and increase the performance of the injection system. Preliminary numerical experiments show it could increase the cooling capacity of the system 1.5%w3.5%. Actually, the small capacity scroll refrigeration system always uses an IHX as the economizer due to its convenience for control. In this kind of gas injection system, the principle factor behind choosing the IHX is a large heat exchange efficiency, which means larger heat exchange areas and the countercurrent configuration are preferred.
1552
international journal of refrigeration 32 (2009) 1544–1554
injection pressure and increases in the cooling capacity. Therefore, the design methods to increase the refrigerant mass flow rate in the injection loop will enhance the cooling capacity of the system. Therefore, the design principle for the injection tube is clear: shorten the injection tube and enlarge the diameter of the injection tube. For shortening the injection tube, the scroll compressor with an economizer attached to the shell is a typical case (Perevozchikov and Doepker, 2003). In addition, the preliminary research finds that the heat exchange between the injection refrigerant and the environment and discharge gas in the compressor will decrease the impact of the injection and decrease the efficiency of the injection system (Fig. 10). Therefore, thermal insulation of the injection tube is quite necessary for achieving a high efficiency injection system.
From Condenser To injection port Overheated Vapor
Saturated Vapor
5.3. Saturated liquid
To system expansion valve
Fig. 9 – Hybrid Economizer.
5.2.
Optimal design of injection tube
According to the preceding analysis, the cooling capability of the injection loop, which is directly related to the mass flow rate in the injection loop, restricts decreases in the optimal
Relative cooling capacity
Relative Cooling Capacity
a 1.3 1.2
1.1 insulated non-insulated 1 0.4
0.6
0.8
1
1.2
1.4
1.6
1.4
1.6
Injection Pressure/MPa Relative COP for cooling
Relative COP for Cooling
b 1.2 1
0.8 insulated non-insulated 0.6 0.4
0.6
0.8
1
1.2
Injection Pressure/MPa Fig. 10 – Effect of thermal insulation of injection tube on the gas-injected system (a) Relative cooling capacity (b) Relative COP for cooling.
Optimal design of injection port
The injection port causes the most significant pressure drop in the injection loop, so decreasing the pressure drop of the injection port is significant to decreasing the overall pressure drop in the injection loop. However, the radial size of the injection port should be smaller than the thickness of the scroll wrap to avoid inner leakage between different pockets, so the effort to enlarge the area of the injection port should be focused on the tangential dimension. Under these constraints, hybrid injection ports following the scroll profile are developed, such as the crescent injection port. The choice of the optimal location of the injection port is more complicated due to its different influence on the cooling capacity and COP. For cooling capacity, the closer the injection port is to the starting position of compression, the lower the average pocket pressure is during injection. It also means a possible larger mass flow rate in the injection loop, a lower optimal injection pressure, a smaller inlet specific enthalpy of the evaporator, and a larger cooling capacity of the system. But for the power consumption, the closer the injection port is to the starting position of compression, the larger the added indicated work is due to injection. Fig. 11(a) shows the p-V diagram of the scroll compressor with injection. It can be found, compared to the compression process without injection, that gas injection in a scroll compressor will reduce part of the under compression loss but lead to a throttling/mixing loss in addition to the ideal compression work. In other words, if the compression work addition caused by the injection wants to be minimized, the throttling/mixing loss and ideal compression work should be minimized. Consequently, an injection pressure close to the inner discharging pressure pi,dis, which refers to the pocket pressure when the inner compression is finished, an injection port which makes the pocket pressure of starting injection close to pi,dis, and a small injection pressure drop are needed (Fig. 11 (b)). Therefore, the selection of the optimal injection port location for maximizing COP is more complicated due to its different influence on the cooling capacity and the indicated work. For the injection system with a constant injection pressure, with moving the injection port location away from the start of compression to the end of compression, the pocket pressure will continue increasing and the actual injection mass flow rate will decrease gradually. But the cooling capacity will keep
1553
international journal of refrigeration 32 (2009) 1544–1554
a
First compression pocket
Effect of injection on the p-V diagram
Orbiting scroll
Static scroll
p Reduction of under-compression loss
Injection port C
po,dis
Possible location
Compression process with injection pi,dis
of injection port
Compression process without injection Throttling/mixing loss
pinj
psuc
V
b
A more efficient injection
Injection port B Injection port A
p Reduction of under-compression loss
Fig. 12 – Injection port location.
po,dis Ideal increase of compression work pi,dis pinj
Throttling/mixing loss
psuc
V
Fig. 11 – p-V diagram of the scroll compressor with injection (a) Effect of injection on the p-V diagram (b) A more efficient injection.
constant in the prior period due to over-quality injection refrigerant and the main loop refrigerant temperature at the outlet of IHX keeps constant as the saturation temperature at the injection pressure. However, in the later period, the cooling capacity will decrease due to the lack of injection refrigerant for subcooling the main loop refrigerant. However, the indicated work will decrease in this whole process. Therefore, the maximum COP will appear when the injection port is located between the location for optimal cooling capacity and the highest possible position (at that position, the pocket pressure at the start of injection equals the injection pressure). In other words, the injection port location for optimal COP should be higher than the injection port location for optimal cooling capacity. Combined with the previous optimal control analysis of the injection pressure, it can be concluded: there will be an optimal injection pressure for the maximizing cooling capacity of the injected system with a fixed injection port location; there will be an optimal injection port location for the maximizing cooling capacity of the injected system with a fixed injection pressure; the optimal injection pressure/ injection port location for maximizing COP is either the same as the optimal injection pressure/injection port location for
maximum cooling capacity or is varied in the direction that decreases the difference between injection pressure and the pocket pressure at the start of injection. Furthermore, the injection port should be located as closely as possible to the scroll wrap of the fixed scroll. A detailed analysis can find that the injection port located in the first compression pocket (Fig. 12) will lead to an injection to the suction pocket, which will largely decrease the system performance. The only possible location of the injection port to avoid injection into the suction chamber is between injection port A and injection port B (possible location of injection port in Fig. 12). But the injection ports between injection port A and injection port B, for example the injection port C, will inject refrigerant into two symmetric compression pockets sequentially, which will lead to unbalanced injection. Therefore, the injection port positions must be close to the wrap of the static scroll, such as ports A and B shown in Fig. 12.
5.4.
Optimal design of one-way valve
A one-way valve to stop the inverse flow of refrigerant in the injection pipe can increase the injection mass flow rate and enhance the injection performance. Because the application of the one-way valve will increase the pressure drop of the injection loop, a low pressure drop configuration of the oneway valve is preferred. At the same time, a one-way valve, which can control the opening time, will make the control of the injection location and period possible, which will largely enhance the adaptability of injection system under variable conditions.
6.
Conclusion
A model of the refrigeration system with gas-injected scroll compressor is described in this paper. With this model, the effects of gas injection on the system and component parameters are investigated. Based on the identified
1554
international journal of refrigeration 32 (2009) 1544–1554
evaporator characteristics and thermodynamic analysis, a set of general principles for the design and operation of the refrigeration or heat pump system with the gas-injected scroll compressor is proposed: (1) Optimal control: (a) Cooling capacity: the maximum cooling capacity can be achieved when the injection pressure and specific enthalpy are adjusted to minimize the refrigerant specific enthalpy at the inlet of the evaporator; (b) Heating capacity: the heating capacity may be keep constant or slightly increased when the injection pressure is larger than the OIPCC. (c) COP: the maximum COP will be reached when the injection pressure is less than the OIPCC and sometimes equal to it; (2) Optimal design: (a) System Configuration: the flash tank system has better potential; in an IHX system, a heat exchanger with large heat exchange area and countercurrent configuration is preferred; (b) Injection Tube: the design principle for the injection tube is clarified: shortening the injection tube, enlarging the diameter of the injection tube, and using thermal insulation; (c) Injection Port: a large injection port with a small radial size and close to the scroll wrap is required; for maximum cooling/heating capacity, the injection port should be as close as possible to the starting position of compression; for maximum COP, the injection port should be located at a position higher than the optimal injection port for cooling capacity and sometimes precisely at the optimal injection port position; (d) One-way Valve: a low pressure drop configuration of the one-way valve is preferred; a one-way valve which can control the opening time will make the control of the injection location and period possible.
Acknowledgement Financial support from National Natural Science Foundation of China (Grant No. 50676042) is greatly appreciated.
references
Ayub, S., Bush, J.W., Haller, D.K., 1992. Liquid refrigerant injection in scroll compressors operating at high compression ration, in: Proceeding of the International Compressor Engineering Conference at Purdue, USA, pp. 561–567.
Beeton, W.L., Pham, H.M., 2003. Vapor-injection scroll compressors. ASHRAE Journal 45 (4), 22–27. Cho, H., Chung, J.T., Kim, Y., 2003. Influence of liquid refrigerant injection on the performance of an inverter-driven scroll compressor. International Journal of Refrigeration 26, 87–94. Ding, G., 2007. Recent developments in simulation techniques for vapour-compression refrigeration systems. International Journal of Refrigeration 30 (7), 1119–1133. Dutta, A.K., Yanagisawa, T., Fukuta, M., 2001. An investigation of the performance of a scroll compressor under liquid refrigerant injection. International Journal of Refrigeration 24, 577–587. Khan, Jameel-ur-Rehman, Zubair, S.M., 2001. Thermodynamic optimization of finite time vapor compression refrigeration systems. Energy Conversion and Management 42 (12), 1457–1475. Ma, G., Chai, Q., Jiang, Y., 2003. Experimental investigation of air source heat pump for cold regions. International Journal of Refrigeration 26, 12–18. Ma, G., Li, X., 2007. Exergetic optimization of a key design parameter in heat pump systems with economizer coupled with scroll compressor. Energy Conversion and Management 48, 1150–1159. Ma, G., Zhao, H., 2008. Experimental study of a heat pump system with flash-tank coupled with scroll compressor. Energy and Buildings 40, 697–701. Park, Y.C., Kim, Y., Cho, H., 2002. Thermodynamic analysis on the performance of a variable speed scroll compressor with refrigerant injection. International Journal of Refrigeration 25, 1072–1082. Perevozchikov, MM., Doepker, R., 2003. Scroll compressor with vapor injection, US patent, US 2003/0136145–A1. Sanaye, S., Malekmohammadi, H.R., 2004. Thermal and economical optimization of air conditioning units with vapor compression refrigeration system. Applied Thermal Engineering 24 (13), 1807–1825. Wang, B., Li, X., Shi, W., Yan, Q., 2007. Design of experimental bench and internal pressure measurement of scroll compressor with refrigerant injection. International Journal of Refrigeration 30, 179–186. Wang, B., Shi, W., Li, X., Yan, Q., 2008. Numerical research on the scroll compressor with refrigeration injection. Applied Thermal Engineering 28 (5–6), 440–449. Wang, B., Shi, W., Li, X., 2009. Numerical analysis on the effects of refrigerant injection on the scroll compressor. Applied Thermal Engineering 29 (1), 37–46. Winandy, E.L., Lebrun, J., 2002. Scroll compressors using gas and liquid injection: experimental analysis and modeling. International Journal of Refrigeration 25, 1143–1156. Yamazaki, H., Itoh, T., Sato, K., Kobayashi, H., 2002. High performance scroll compressor with liquid refrigerant injection. In: Proceeding of the International Compressor Engineering Conference at Purdue, USA, C22–1. Zehnder, M., Favrat, D., Hohl, H., Olivier, C., Perevozchikov, M., 2002. High performance air-water heat pump with extended application range for residential heating. In: The Seventh International Energy Agency Conference on Heating Pumping Technologies, Beijing, China, pp. 19–22.