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Research on air-source heat pump coupled with economized vapor injection scroll compressor and ejector
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Research on air-source heat pump coupled with economized vapor injection scroll compressor and ejector Xu Shuxue, Ma Guoyuan* Beijing University of Technology, Beijing 100124, China
article info
abstract
Article history:
The performance of the heat pump can be improved further when running under low
Received 3 August 2009
temperature conditions when an ejector is used in a heat pump system coupled with
Received in revised form
economized vapor injection (EVI) scroll compressor. In this paper, the design method of the
16 March 2010
heat pump system with ejector (EVIe) is presented, and the process for designing the heat
Accepted 1 June 2010
pump with ejector has been summarized. The optimal location of the vapor injection inlets
Available online 9 June 2010
is at the place where the vapor can inject into the working chambers when they just be closed. The reasonable value for the entrainment ratio u of the ejector is between 0.1 and
Keywords:
0.2. One prototype heat pump was designed under the condition of the evaporation
Heat pump
temperature of 20 C, and an experimental setup was established to test the prototype.
Air-source
The measured results demonstrated that the heating EER of the heat pump system with
Scroll compressor
ejector could reach about 4% higher than that of the system without ejector when the
Injection
heating capacity remained nearly constant. ª 2010 Elsevier Ltd and IIR. All rights reserved.
Ejector Design Calculation
Recherches sur une pompe a` chaleur air-eau couple´e a` un compresseur a` spirale et un e´jecteur a` injection de vapeur muni d’un e´conomiseur Mots cle´s : Pompe a` chaleur ; Air-eau ; Compresseur a` spirale ; Injection ; E´jecteur ; Conception ; Calcul
1.
Introduction
The conventional air-source heat pump with single-stage compression (abbreviated SSC) cycle cannot operate efficiently and steadily for long periods in cold regions where there is a large difference between the ambient and room
temperatures. In order to solve this problem, much research has been conducted. Of the many solutions, injecting refrigerant vapor by supplementary inlets (also called economizer holes) in the scroll compressor to compose a quasi two-stage compression heat pump system, referred to as a heat pump system coupled with economized vapor injection (EVI) scroll
* Corresponding author. Tel./fax: þ86 10 67391613. E-mail address: [email protected] (M. Guoyuan). 0140-7007/$ e see front matter ª 2010 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2010.06.003
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Nomenclature a aP* dP* f fP* h k m n p pCu R T u V v
Base circle radius of scroll (m) Critical speed of the primary fluid (m s1) Throat diameter of the ejector (m) Area (m2) Section area of the nozzle (m2) Enthalpy (kJ kg1) Adiabatic index of refrigerant Mass flow rate (kg s1) Rotation speed (r min1) Pressure (kPa) mixing fluid pressure corresponding to entrainment ratio u (kPa) Gas constant for refrigerant (kJ kg1 K1) Temperature (K) Entrainment ratio Stroke volume of the compressor (m3) Specific volume (m1)
compressor (abbreviated EVI system), has been recognized as an approach with good prospects both in technology and commerce. Cho et al. (2003) and Dutta et al. (2001) researched the effects of vapor and liquid injection on the discharge temperature, power consumption, and heating performance of the scroll compressor by experiment and simulation. Wang et al. (2007) presented an integrated method for setting up an experimental bench to research the small EVI scroll compressor, including the place and size design of economizer holes, type selection and installation of the sensors. Ma et al. (2003) experimentally studied the heat pump system coupled with scroll compressor with supplementary inlets, and pointed out that the heating/cooling performance of the heat pump system could be even more improved by optimizing the design parameters. Winandy et al. (2002) developed a simplified simulation model for the scroll compressor with supplementary inlets. Zhao et al. (2006) conducted comparative research on heat pump systems with three different kinds of economizer, namely flash-tank, sub-cooler and flash-reservoir, and it was concluded that the heat pump system with the flash-tank, which has a two-stage throttling process, had better performance and reliability than the heat pump system with a sub-cooler or flash-reservoir. Nevertheless, there was some energy loss in the throttling valve in the supplementary circuit in conventional EVI heat pump system. In order to recover and utilize this energy, an ejector was used in the supplementary circuit to replace the throttling valve, and the EVI heat pump system with ejector (abbreviated EVIe system) was proposed based on the conventional EVI system. Pang et al. (2007) designed a prototype of the EVIe system and its testing setup, and the experimental results showed that the heating capacity and the heating EER increased about 10% and 4%, respectively, compared with the EVI system. And when the entrainment ratio of the ejector was 0.1, the EVIe system exhibited favorable performance under the conditions of the evaporation temperature varying from 25 C to 15 C. Ejectors that use organic working fluid have attracted much interest in recent
a hv r rP* x g 4
Dimensionless flow rate Volumetric efficiency Density (kg m3) Density of the primary fluid in the nozzle section (kg m3) Resistance factor of refrigerant flowing in the supplementary circuit Initial involute angle (rad) Involute angle (rad)
Subscripts 1, 2, 20 State point shown in Fig. 1 C Mixing fluid d Discharge H Secondary fluid P Primary fluid p* Nozzle area s Suction
years because steam jet systems suffer the disadvantage of being unable to generate refrigeration below 0 C. To use organic working fluid in the ejector not only overcomes this disadvantage of the steam jet system, but also makes use of even lower temperature waste heat energy. Chunnanond and Aphornratana (2004) provided a review on ejectors and their applications in refrigeration including the jet refrigeration cycle and many other applications of ejectors in other types of refrigeration systems. Chaiwongsa and Wongwises (2007) used the ejector as an expansion device in the vapor compression refrigeration cycle, and the effects of the nozzle diameter of the ejector on the coefficient of performance (COP) of the system were discussed. When an ejector is used in an EVI heat pump system, the running parameters of the primary fluid, secondary fluid and mixing fluid are different from that of an ejector refrigeration system or ejector vacuum equipment, and the construction design and the matching of mass flow rates must be reconsidered. This paper will present the design method of an EVIe heat pump system with a scroll compressor made by a factory in China, emphasizing on the determination of the design parameters of the ejector in the EVIe system. Also, the experimental results of the prototype developed in terms of the design method will be discussed. This research should be helpful to the development of EVIe heat pump systems and in improving their performance.
2.
EVI and EVIe heat pump systems
Fig. 1 (a) and (b) are flow charts of the EVI heat pump system and EVIe system. The EVIe system is comprised of compressor, condenser, flash-reservoir, expansion valve, evaporator, ejector and other apparatus. The compressor, condenser, flash-tank, expansion valve and evaporator are connected together to form the main circuit, and the circuit is same as conventional single-stage compression heat pump system. The compressor has supplementary inlets, the ejector
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a
3
condenser
conveniently installed, but also has a high efficiency and a wide range of operation condition.
2 6 scroll compressor
throttling valve
5
2'
1
expansion valve
4
4'
evaporator
flash-reservoir
b
3
2 condenser ejector
6 scroll compressor
check valve 1
5
2' 1
check valve 2 expansion valve
4
4'
evaporator
flash-reservoir Fig. 1 e Heat pump system coupled with vapor injection compressor, (a) EVI system (b) EVIe system.
is connected between flash-reservoir and supplementary inlets of compressor, those equipments form supplementary circuit. The flashed vapor at the top of the flash-reservoir as a primary fluid flows through the ejector, and entrained a part of the refrigerant vapor from the evaporator, which is called entrained or secondary fluid. The primary fluid and secondary fluid are mixed in ejector, and the pressure of the mixed fluid then rises to the intermediate pressure at the exit of the ejector. This part of refrigerant is sucked into compressor through supplementary inlets. Compared with the EVI system, the throttling valve in the supplementary circuit is replaced by an ejector in the EVIe heat pump system. The ejector is able to increase the pressure of the secondary fluid without mechanical energy consumption, and using it in the EVIe system can reduce the available energy loss, and the temperature of the refrigerant vapor entering into the scroll compressor through supplementary inlets. As ejector has advantages such as simple structure, low cost, no moving parts, safety for two-phase working fluid and so on, the EVIe system not only be simply composed,
3.
Design of vapor supplementary holes
If the location of the vapor injection inlets of scroll compressor varies in a certain range, there is a small effect of the positions of the injection inlets on the steady operation of the EVIe heat pump system (Cho et al., 2003; Dutta et al., 2001; Wang et al., 2007). However, to ensure that the heating performance of the system is relatively high under the condition of low temperature, the optimal location of the vapor injection inlets is at the place where the vapor can inject into the working chambers when they just be closed (Wang et al., 2007 and Ma et al., 2003), as shown in Fig. 2. The shape of the vapor injection inlets also affects the performance of EVI system. Considering the convenience to manufacture, the inlets should be round and located symmetrically at a pair of compression chambers of the scroll compressor. The inlet must be smaller than the thickness of the scroll to ensure that the two adjacent chambers are not connected by the inlets at any time. Namely, the size must fit the scroll thickness, and different scroll need different inlet size. If the profile of scroll is an involute of circle with radius of a, the coordinates of the scroll center line in coordinate xoy in Fig. 2 can be expressed as, x ¼ aðcos f þ fsin fÞ
(1)
y ¼ aðsin f fcos fÞ
(2)
Where, a is radius of base circle, m; 4 is involute angle, rad. The working chambers of scroll compressor come forth in pairs, so there must be a couple of the injection inlets for a pair of suction chambers. According to the geometric relations of
y 120
Orbiting scroll 60 Fixed scroll
0
0
x
Vapor injection inlets
Fig. 2 e Scroll compressor and the location of vapor injection.
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the scroll (Li,1998), the coordinates of the center point of one injection inlet can be derived as, 7p 1 7p xo ¼ a cos 2g q þ 7psin 2g q 2 2 2
(3)
7p 1 7p yo ¼ a sin 2g q 7pcos 2g q 2 2 2
(4)
Where g is initial involute angle, rad; q is rotation angle of the shaft, rad. The coordinates of the center of the other injection inlet can be expressed as, 9p 1 9p þ 2g q þ 9psin þ 2g q xi ¼ a cos 2 2 2
(5)
9p 1 9p þ 2g q 9pcos þ 2g q yi ¼ a sin 2 2 2
(6)
For a given scroll compressor, its scroll profile is specified and the locations of the injection inlets depend only on the shaft angle. In terms of the geometry of scroll compressor, the amount of the vapor flowing through injection inlets is remarkable when the shaft angle q varies from 0 to 120 . The locations of the pair of inlets were shown in Fig. 2 while the shaft angles are 0 , 60 and 120 , respectively. The suction chambers of scroll compressor are just closed when the shaft angle is 0 . If the injection inlets locate at the shaft angle of 0 , this means that the vapor injection begins at end of the suction process, and the locations are favorable to vapor injection. Their coordinates are: 7p 1 7p 2g þ 7psin 2g xof ¼ a cos 2 2 2
(7)
7p 1 7p yof ¼ a sin 2g 7pcos 2g 2 2 2
(8)
and 9p 1 9p xif ¼ a cos þ 2g þ 9psin þ 2g 2 2 2
(9)
9p 1 9p yif ¼ a sin þ 2g 9pcos þ 2g 2 2 2
(10)
The scroll thickness can be expresses as t ¼ 2ag. For conventional scroll compressor, 3 mm t 6 mm. For round injection inlet, its radius must be less than the thickness of scroll so that the adjacent chambers cannot be connected through the inlets at any time. Usually, the inlet radius should be satisfied t/2 rj 3t/4. For the prototype compressor for this study, the scroll thickness is 3.6 mm and the radius of the round inlet is 2.4 mm. For conventional EVI system, when the condensation temperature is 45 C and the evaporation temperature varies from 15 C to 25 C, the favorable injection pressure is between 0.6 MPa and 0.8 MPa, and the injection vapor amounts from 7%e25% of the suction vapor flow rate under this condition (Ma et al., 2003). The injection vapor amounts
only to 5% while the intermediate pressure is reduced to 0.5 MPa. Therefore, the improvement of EVI system becomes inconspicuous if the intermediate pressure is less than 0.5 MPa (Zhao et al., 2006). For example, the intermediate pressure is 0.70 MPa, and the injection vapor amount is about 12% of the vapor suction flow rate with the condensation temperature of 45 C and the evaporation temperature of 20 C. Accordingly, the heating capacity of EVI system increases approximately 11.5%.
4.
Design method of ejector
A schematic diagram of the ejector is shown in Fig. 3. When the ejector is working, the primary fluid at high pressure expands and accelerates through the primary nozzle in the ejector, and the fluid is ejected with supersonic speed to create a very low pressure region at the plane of the nozzle exit. The fluid at low pressure, called the ‘secondary fluid’, is then entrained through the port connecting the nozzle exit and mixed with the primary fluid in the mixing chamber. The fully-mixed fluid decelerates and its pressure rises in the compression chamber as its kinetic energy is converted to the potential energy of pressure. The pressure of the mixed fluid at the ejector exit, called the intermediate pressure, must match the flow rate of the vapor to be injected into the compressor. The main operating parameters of the ejector are the entrainment ratio u and the pressure values of the primary fluid, the secondary fluid and mixing fluid. u, an important parameter to express the performance of an ejector, is defined as the ratio of the mass flow rate of the entrained fluid to that of the primary fluid. For a given working fluid, the ratio u can be calculated according to the pressure values of the primary fluid, secondary fluid and mixing fluid (detailed in Huang, 1977 and Tian, 2007). The performance of the ejector can be predicted using the aerodynamic function method (Huang, 1977 and Tian, 2007). Normally, the pressure of the mixing fluid, pC, is equal to the pressure of the injection vapor in the EVI heat pump system. The compression process of the scroll compressor is divided into two stages by the vapor injection and hence consists of by three sub-processes. They are the first closed compression process before connection with the supplementary inlets, the vapor injection and mixing process and the second closed compression process after the supplementary inlets are cut off. The process is similar to a two-stage compression process but is not a real one. So, it is also referred to as a quasi two-stage compression process.
f p*
f p1
f2
f3
Mixing Fluid
Primary Fluid Secondary Fluid
Suction Section
Mixing Section
Fig. 3 e Schematic diagram of ejector.
Diffuser Section
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To sum up, the appropriate intermediate pressure must be determined in advance, and the dimensions of the main sections of the ejector can be calculated in turn. According to the steady-flow energy equation for the ejector, the following equation can be written:
0.5 0.4 -15 -20 -25
u
0.3 0.2
hC ¼
0.1 0 0.55
0.7
0.85
1
1.15
1.3
Mixing Pressure (MPa) Fig. 4 e The variation of entrainment ratio u with mixing pressure.
There is a one-way valve assembled in the supplementary circuit to prevent the reversible flow in the circuit. The pressure of the vapor in the working chamber at the end of the first compression process rises to the value of p20 , and pC p20
(11)
As mentioned above, the ratio u is an important parameter to describe the ejector performance. When the pressure values of the primary fluid and the secondary fluid are given, u is affected by the pressure of the fluid at the ejector exit, namely the mixing pressure. The variation of u with the pressure at the ejector exit is shown in Fig. 4 when the evaporation temperatures are 15 C, 20 C and 25 C, respectively. As shown in Fig. 4, the entrainment ratio u will decrease with the increase of the mixing pressure. When the mixing pressure rises to 1.0 MPa, u decreases to less than 0.1. When the mixing pressure remains constant, the entrainment ratio u increases with the increase of the pressure of the entrained vapor corresponding to the evaporation temperature. To ensure that the ejection is effective, the entrainment ratio u must keep higher value so that the ejector exhibits favorable performance. The experimental results have demonstrated that the optimal value of u is greater than 0.1 (Pang et al., 2007). In this condition, the mixing fluid pressure corresponding to entrainment ratio u, pCu, is pC pCu
(12)
Where, pC is the pressure of the mixing fluid. If the entrainment ratio u is higher than 0.2, the mixing fluid pressure becomes too low, even less than 0.5 MPa. The low mixing pressure means the vapor injection to the compression chambers becomes weak. Generally, the reasonable value of u is between 0.1 and 0.2. The mass flow rate discharged from the compressor could be increased by the vapor injection, and the heating capacity of the EVI system also increases consequently. Nevertheless, the pressure of the injection vapor, pC, has an optimal value. If the pC value is higher than the value of the intermediate pressure, the performance of the system will deteriorate (Zhao et al., 2006). The appropriate value of the intermediate pressure can be decided by experiments.
hP þ uhH 1þu
(13)
Where, hC, hP and hH are the enthalpy values of the mixing fluid, primary fluid and secondary fluid (kJ kg1). The mass flow rate of the primary fluid is determined by the throat area of the ejector and the property of the fluid, and then (14) mP ¼ fP rP aP Where, mP is the mass flow rate of the primary fluid (kg s1); fP* is the nozzle section area (m2); rP* is the vapor density of the primary fluid in the nozzle section (kg m3); and aP* is the critical speed of the primary fluid (m s1). Therefore, the dimensionless flow rate a through the supplementary circuit can be calculated by the parameters of the ejector: a¼
md ms mP ð1 þ uÞ ¼ ms nrs Vhv =60
(15)
Where, md and ms are the mass flow rates of the discharge and the suction refrigerant of the compressor (kg s1); n is the rotation speed of the compressor (r min1); rs is the density of the suction refrigerant (kg m3); V is the stroke volume of the compressor (m3); and h is the volumetric efficiency of the compressor.
evaporation temperature,''to'' , vapor injection location
ejector modle , entrainment ratio "u"
EVI system experiment
lowest injection vapor pressure
upper limit of vapor injection pressure
lower limit of vapor injection pressure
determine vapor pressure
calculate flowrate " ¦ "Á
determine mass flow rate of ejector
calculate dimension of ejector
Fig. 5 e Flow chart of design process of EVIe heat pump system.
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Cooling tower
High and low pressure switch
Pump Filter
Comp One-way valve v3
Condenser
Evaporator
v2 v1
Heat exchanger Heater
TXV
Mass flow rate meter Filter
Receiver
Hot water bulk
pressure sensor temperature sensor valve
Pump Heater
Glycol pump Cool water bulk Heater
Fig. 6 e Test system and sensor locations for the prototype.
Based on the thermodynamic model for gas charging to a rigid container, the dimensionless flow rate a can also be derived as follows: a¼
m2 m20 v20 ¼ pC p20 x m20 RkTC
(16)
Where, v20 is the refrigerant specific volume at state point 20 (m3 kg1); p20 is the pressure of the refrigerant at state point 20 (kPa); R is the gas constant for the refrigerant (kJ kg1 K1); k is the adiabatic index of the refrigerant; TC is the temperature of the mixing fluid (K); and x is the resistance factor of the refrigerant flowing through the supplementary circuit, and it is recommended to be 0.2 according to experiments (Zhao et al., 2006 and Ma et al., 2003). Replacing the specifications and technical parameters of the compressor such as stroke volume, volumetric efficiency, and rotation speed into equation (15), the throat diameter of the nozzle of the ejector, dp*, can be obtained, and other section dimensions of the ejector can be calculated consequently. This process for designing the EVIe heat pump system can be summarized as shown in the flow chart in Fig. 5.
5.
Testing system
An EVI heat pump system was designed based on the technical data of the scroll compressor manufactured by a factory
in China. The testing system of the EVI heat pump system is shown in Fig. 6. The specifications of the scroll compressor were as follows: a stroke volume of 80 cm3/rev, rating power input of 4.55 kW, rotation speed of 2800 r/min, and working fluid of R22. The operation conditions were as follows: an evaporation temperature of 20 C and condensation temperature of 45 C. The flow rate of the glycol solutions through the evaporator was measured by measuring the volumetric discharge in a graduated flask per unit time. The heating capacity is computed by the measured values by the flow rate and the temperature difference of the water flowing through the condenser, and the flow rate of the water was measured with a flow rate meter. The electric heater in the glycol-water tank was adjustable so that the solution temperature was kept constant during the measurements. The temperature of the water flowing into condenser can be adjusted by the input power of the electric heater in the water tank. And the condensation temperature varies with the water temperature. All the electric parameters of the compressor such as the input voltage, current, power and other electric parameters were recorded by a data logger. Each temperature and pressure sensor was calibrated to reduce experimental uncertainties and was connected to an Agilent data logger. The accuracy and tolerance of each sensor for measuring data was shown in Table 1. The uncertainties of the heating capacity and EER estimated by the analysis were approximately 2.6%.
Table 1 e Accuracies and tolerances of the sensors. Sensor Temperature Pressure transducer Currency acquisition unit Power acquisition unit Data logger
Accuracy
0.2 C 0.2% of 0.2% of 0.2% of 0.2% of
Tolerance
full scale full scale full scale full scale
0.2 C 0.2% of full 0.2% of full 0.2% of full 0.2% of full
scale scale scale scale
Full scale
Model
e 2.5/1.0 MPa e e e
Pt100 Huba DZFC-1 DZFC-1 HP34970A
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Table 2 e Experimental data for heat pump prototype. p1 0.24 0.24 0.24 0.24 0.24 0.25 0.24 0.24 0.24 0.24 0.24 0.24 0.24 0.24
p10
t10
p20
t20
p3
t3
p40
t40
p6
t6
Qk
P
EER
0.23 0.22 0.22 0.22 0.22 0.23 0.22 0.22 0.22 0.23 0.22 0.22 0.22 0.22
10.46 10.13 10.50 10.29 10.28 10.48 10.26 10.94 10.14 10.02 10.70 10.89 10.62 10.27
1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73
117.56 116.83 116.49 115.94 115.03 115.59 115.18 114.13 113.25 112.38 108.00 107.99 106.38 105.03
1.69 1.68 1.68 1.69 1.69 1.69 1.68 1.68 1.67 1.68 1.66 1.66 1.65 1.65
43.77 43.65 43.60 43.87 43.69 43.52 43.67 43.60 43.40 43.47 43.38 43.38 43.19 43.01
0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26
15.25 15.37 15.39 15.31 15.41 14.16 15.02 15.11 15.11 15.06 14.70 14.70 14.74 14.95
0.54 0.54 0.53 0.52 0.50 0.69 0.70 0.90 0.90 1.10 1.09 1.29 1.29
18.13 17.92 16.12 14.51 10.49 19.79 19.79 23.38 17.68 26.74 26.74 30.17 27.94
5.35 5.85 5.97 6.06 6.04 6.03 6.19 6.36 6.08 6.08 6.15 6.22 5.82 5.92
4.33 4.38 4.36 4.32 4.36 4.32 4.60 4.62 4.66 4.79 4.78 4.86 4.89 4.86
1.24 1.33 1.37 1.40 1.39 1.40 1.35 1.37 1.31 1.27 1.29 1.28 1.19 1.22
Notes: p-pressure, MPa; t-temperature, C; Qk-heating capacity, kW; P-power input, kW.
Relevant parameters were recorded only if the fluctuation of every parameter was within 1%. The experimental method and data processing accorded with Chinese national standards GB 5773-86 (the standard for performance tests for positive displacement refrigerant compressors) and GB/T 9056-1999 (the standard for performance tests for positive displacement refrigerant compressor condensing units). In order to test the performance of the heat pump under different state parameters of the injection vapor, the testing system for the conventional heat pump was rebuilt. The power of the electric heater and the mass flow rate of the glycol-water solution of the heat exchanger in the supplementary circuit can be adjusted to keep the temperature of the injection vapor steady. By switching the valves on or off, it is easy to realize the following operation modes of the testing system. For the test system in Fig. 6, it is running as SSC system when all of the valves, v1, v2 and v3, are shut off, and running as EVI system when v1 is turned on and v2 and v3 off, and running as EVIe system when v1 is turned off and v2 and v3 on.
0.3
Testing results and discussions
The main measured data and the performance of the heat pump system calculated by the data are listed in Table 2. Table 2 shows that the heat pump has higher heating capacity and heating EER when the pressure of the injection vapor is 0.5e0.7 MPa and the temperature is 18e20 C, which are also the mixing fluid state parameters of the ejector. The dimensions of the main sections of the ejector are listed in Table 3. The experimental results have been rearranged as the curves shown in Figs. 7e10.
EVI EVIe 0.2 α
6.
The variation of the dimensionless flow rate a through the supplementary circuit with the evaporation temperature is shown in Fig. 7. For both the EVI and EVIe heat pump systems, the dimensionless flow rate a decreases with the increase of the evaporation temperature, and its value varies from 0.25 to 0.03 when the evaporation temperature goes up from 25 C to 15 C. This is because the pressure difference between the injection vapor and the refrigerant in the working chambers of the scroll compressor will be bigger when the evaporation temperature becomes lower. Under the same pressure of the injection vapor, the dimensionless flow rate a of the EVIe system is greater than that of the EVI system. For example, a is 0.24 for EVIe system but 0.17 for EVI system when the evaporation temperature is 23 C. This is because the ejector entrained the vapor at the exit of the evaporator with lower enthalpy and specific volume, and it gives the injection vapor a lower superheating degree and specific volume. Under the design condition of an evaporation temperature of 20 C, the ejector can achieve better performance, and it also reaches the maximum difference of the flow rate a between the EVIe and EVI system, and the difference value is about 0.08. The variations of the heating capacity with the evaporation temperature are shown in Fig. 8. The heating capacity of the
0.1
0 -25
Table 3 e Area of main section of ejector (mm2). f* 0.196
fp1
f2
f3
0.318
1
0.494
-23
-20
-17
-15
Evaporation Temperature
Fig. 7 e The variation of dimensionless flow rate a of the injection vapor with evaporation temperature.
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3
8
Heatin g EER
Heating Capacity (kW)
12
1.5 EVI EVIe SSC
EVI EVIe SSC
4
0 -25
0 -25
-23
-20
-17
-15
Evaporation Temperature
Fig. 8 e The variation of heating capacity with evaporation temperature.
EVIe system is slightly greater than that of the EVI system, and the difference of the heating capacity increases with the decrease of the evaporation temperature. No matter EVI or EVIe system, the heating capacity reduces from 9.6 kW to 9.0 kW with variation of evaporation temperature from 15 C to 17 C, and the capacity difference between the two systems is neglectable. However, the capacity difference became remarkable as the evaporation temperature decreases further, and is about 1.1 kW while the evaporation temperature goes down to 20 C. This is because the injection vapor in the EVIe system has lower enthalpy, and it results in lower enthalpy of the discharge vapor from the compressor, even though the flow rate a of the EVIe system is greater than that of the EVI system. The heating capacity is equal to the product of the enthalpy difference between the inlet and exit of the condenser and the mass flow rate of the discharge vapor, and the heating capacity increases only when the product becomes greater. When the evaporation temperature becomes low, the mass flow rate of the injection vapor becomes dominant compared
Po wer In p u t (kW)
5
4 EVI EVIe SSC
3 -25
-23
-20
-17
Evaporation Temperature
Fig. 9 e The variation of power input with evaporation temperature.
-15
-23
-20
-17
-15
Evaporation Temperature
Fig. 10 e The variation of heating EER with evaporation temperature.
with the discharge enthalpy, so their product increases with the mass flow rate. The variations of the power input of the compressor with the evaporation temperature are shown in Fig. 9. The power input of the EVI heat pump system decreases very slowly with the increase of the evaporation temperature and is 1.5e3.4% higher than that of the conventional heat pump. This is because the mass flow rate of the injection vapor rises with the increase of the evaporation temperature in both the EVIe and EVI system. The power input of the EVIe system is 4% lower than that of the EVI system at the same evaporation temperature. This demonstrates that the ejector can effectively recover a part of the pressure energy of the refrigerant in the supplementary circuit, and then the compression process of the system can be further improved. When the evaporation temperature is 20 C, the ejector performs better under the design conditions, and so the EVIe heat pump system exhibits favorable performance, as the compression efficiency is further increased by the ejector. The variations of heating EER with the evaporation temperature are shown in Fig. 10. Using an ejector in the EVIe heat pump system can improve the working conditions of the scroll compressor, so the compressor has a relatively high efficiency compared with the EVI system when it is operated at low temperature. For example, the heating EER is 2.15 for EVIe system and 1.64 for EVI system when the evaporation temperature is 20 C. The heating EER of the EVIe heat pump system rises about 3e5% compared with the EVI system while the heating capacity remains nearly constant. The improvement of the EER becomes more remarkable as the evaporation temperature falls. This demonstrates that the improvement by using the ejector in the heat pump system is more remarkable under the conditions of low evaporation temperature. On the other hand, the improvement of the heating capacity by using the ejector becomes relatively small with the increase of the evaporation temperature. When the evaporation temperature goes up to 15 C, the heating capacity of the EVIe system is very close that of the EVI system. This is because the dimensionless flow rate a through
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the supplementary circuit decreases from 20 to 30%e10%, as shown in Fig. 7, when the evaporation temperature is increased from 25 C to 15 C. In both the EVIe system and the EVI system, the improvement of the heating capacity is weakened when the flow rate a becomes slow. Thus, there is no remarkable improvement from using the ejector to replace the throttling valve in the supplementary circuit if the evaporation temperature is more than 15 C.
7.
Conclusions
The design method of the heat pump system coupled with the EVI scroll compressor and ejector was presented, and the prototype was tested with evaporation temperature from 25 C to 15 C. The following conclusions could be made. (1) The optimal location of the vapor injection inlets is at the place where the vapor can inject into the working chambers when they just be closed, and the size must fit the scroll thickness to ensure that the two adjacent chambers are not connected by the inlets at any time. For a scroll with the circle involute profile, its thickness, t, is usually between 3 mm and 6 mm, and the radius of round injection inlets should be satisfied t/2 rj 3t/4. (2) To ensure that the ejection is effective, the entrainment ratio u must keep higher value so that the ejector exhibits favorable performance. Generally, the reasonable value of u is between 0.1 and 0.2. (3) The experimental results of the prototype demonstrate that the heating EER of the EVI heat pump system with ejector is about 4% higher than that of the system without an ejector while the heating capacity is kept nearly constant.
Acknowledgments We are very grateful to the support of National Natural Science Foundation No. 50776001, China, and express our heartfelt thanks to Mr Wu Xuezhi, Pang Zongzhan, Peng Long and Miss Yu Lihua for their effective help in the development and experiment of the prototype.
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references
Chaiwongsa, P., Wongwises, S., 2007. Effect of throat diameters of the ejector on the performance of the refrigeration cycle using a two-phase ejector as an expansion device. International Journal of Refrigeration 24, 1e8. Chinese national standards GB 5773-86, the standard for performance tests for positive displacement refrigerant compressors. Chinese national standards GB/T 9056-1999, the standard for performance tests for positive displacement refrigerant compressor condensing units. Cho, H., Chung, J.T., Kim, Y., 2003. Influence of liquid refrigerant injection on the performance of an inverterdriven scroll compressor. International Journal of Refrigeration 26, 87e94. Chunnanond, K., Aphornratana, S., 2004. Ejectors: applications in refrigeration technology. Renewable and Sustainable Energy Reviews 8, 129e155. 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, 577e587. Huang, Q.Y., 1977. Ejector (in Chinese). Science Press, Beijing, pp. 1e132. Li, L.S., 1998. Scroll Compressor (in Chinese). China Machine Press, Beijing, pp. 14e32. Ma, G.Y., Chai, Q.H., Jiang, Y., 2003. Experimental investigation of air-source heat pump for cold regions. International Journal of Refrigeration 26 (1), 12e18. Pang, Z.Z., Ma, G.Y., Yu, L.H., 2007. Experimental Study on Quasi Two-stage Compression Heat Pump System Coupled with Ejector. 22nd International Congress of Refrigeration, Beijing. Tian, Q., 2007. Solar-powered Ejector Refrigeration (in Chinese). Science Press, Beijing, pp. 112e146. Wang, B.L., Li, X.T., Shi, W.X., Yan, Q.S., 2007. Design of experiment bench and internal pressure measurement of scroll compressor with refrigerant injection. International Journal of Refrigeration 30, 179e186. Winandy, E., Saavedra, O.C., Lebrun, J., 2002. Experimental analysis and simplified modeling of a hermetic scroll refrigeration compressor. Applied Thermal Engineering 22, 107e120. Zhao, H.X., Liu, S.G., Liu, Z.B., Ma, G.Y., 2006. Experiment study on heat pump system with scroll compressor flash tank. Acta Energiae Solaris Sinica 27 (5), 377e381.
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w w w . i i fi i r . o r g
journal homepage: www.elsevier.com/locate/ijrefrig
Research on air-source heat pump coupled with economized vapor injection scroll compressor and ejector Xu Shuxue, Ma Guoyuan* Beijing University of Technology, Beijing 100124, China
article info
abstract
Article history:
The performance of the heat pump can be improved further when running under low
Received 3 August 2009
temperature conditions when an ejector is used in a heat pump system coupled with
Received in revised form
economized vapor injection (EVI) scroll compressor. In this paper, the design method of the
16 March 2010
heat pump system with ejector (EVIe) is presented, and the process for designing the heat
Accepted 1 June 2010
pump with ejector has been summarized. The optimal location of the vapor injection inlets
Available online 9 June 2010
is at the place where the vapor can inject into the working chambers when they just be closed. The reasonable value for the entrainment ratio u of the ejector is between 0.1 and
Keywords:
0.2. One prototype heat pump was designed under the condition of the evaporation
Heat pump
temperature of 20 C, and an experimental setup was established to test the prototype.
Air-source
The measured results demonstrated that the heating EER of the heat pump system with
Scroll compressor
ejector could reach about 4% higher than that of the system without ejector when the
Injection
heating capacity remained nearly constant. ª 2010 Elsevier Ltd and IIR. All rights reserved.
Ejector Design Calculation
Recherches sur une pompe a` chaleur air-eau couple´e a` un compresseur a` spirale et un e´jecteur a` injection de vapeur muni d’un e´conomiseur Mots cle´s : Pompe a` chaleur ; Air-eau ; Compresseur a` spirale ; Injection ; E´jecteur ; Conception ; Calcul
1.
Introduction
The conventional air-source heat pump with single-stage compression (abbreviated SSC) cycle cannot operate efficiently and steadily for long periods in cold regions where there is a large difference between the ambient and room
temperatures. In order to solve this problem, much research has been conducted. Of the many solutions, injecting refrigerant vapor by supplementary inlets (also called economizer holes) in the scroll compressor to compose a quasi two-stage compression heat pump system, referred to as a heat pump system coupled with economized vapor injection (EVI) scroll
* Corresponding author. Tel./fax: þ86 10 67391613. E-mail address: [email protected] (M. Guoyuan). 0140-7007/$ e see front matter ª 2010 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2010.06.003
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Nomenclature a aP* dP* f fP* h k m n p pCu R T u V v
Base circle radius of scroll (m) Critical speed of the primary fluid (m s1) Throat diameter of the ejector (m) Area (m2) Section area of the nozzle (m2) Enthalpy (kJ kg1) Adiabatic index of refrigerant Mass flow rate (kg s1) Rotation speed (r min1) Pressure (kPa) mixing fluid pressure corresponding to entrainment ratio u (kPa) Gas constant for refrigerant (kJ kg1 K1) Temperature (K) Entrainment ratio Stroke volume of the compressor (m3) Specific volume (m1)
compressor (abbreviated EVI system), has been recognized as an approach with good prospects both in technology and commerce. Cho et al. (2003) and Dutta et al. (2001) researched the effects of vapor and liquid injection on the discharge temperature, power consumption, and heating performance of the scroll compressor by experiment and simulation. Wang et al. (2007) presented an integrated method for setting up an experimental bench to research the small EVI scroll compressor, including the place and size design of economizer holes, type selection and installation of the sensors. Ma et al. (2003) experimentally studied the heat pump system coupled with scroll compressor with supplementary inlets, and pointed out that the heating/cooling performance of the heat pump system could be even more improved by optimizing the design parameters. Winandy et al. (2002) developed a simplified simulation model for the scroll compressor with supplementary inlets. Zhao et al. (2006) conducted comparative research on heat pump systems with three different kinds of economizer, namely flash-tank, sub-cooler and flash-reservoir, and it was concluded that the heat pump system with the flash-tank, which has a two-stage throttling process, had better performance and reliability than the heat pump system with a sub-cooler or flash-reservoir. Nevertheless, there was some energy loss in the throttling valve in the supplementary circuit in conventional EVI heat pump system. In order to recover and utilize this energy, an ejector was used in the supplementary circuit to replace the throttling valve, and the EVI heat pump system with ejector (abbreviated EVIe system) was proposed based on the conventional EVI system. Pang et al. (2007) designed a prototype of the EVIe system and its testing setup, and the experimental results showed that the heating capacity and the heating EER increased about 10% and 4%, respectively, compared with the EVI system. And when the entrainment ratio of the ejector was 0.1, the EVIe system exhibited favorable performance under the conditions of the evaporation temperature varying from 25 C to 15 C. Ejectors that use organic working fluid have attracted much interest in recent
a hv r rP* x g 4
Dimensionless flow rate Volumetric efficiency Density (kg m3) Density of the primary fluid in the nozzle section (kg m3) Resistance factor of refrigerant flowing in the supplementary circuit Initial involute angle (rad) Involute angle (rad)
Subscripts 1, 2, 20 State point shown in Fig. 1 C Mixing fluid d Discharge H Secondary fluid P Primary fluid p* Nozzle area s Suction
years because steam jet systems suffer the disadvantage of being unable to generate refrigeration below 0 C. To use organic working fluid in the ejector not only overcomes this disadvantage of the steam jet system, but also makes use of even lower temperature waste heat energy. Chunnanond and Aphornratana (2004) provided a review on ejectors and their applications in refrigeration including the jet refrigeration cycle and many other applications of ejectors in other types of refrigeration systems. Chaiwongsa and Wongwises (2007) used the ejector as an expansion device in the vapor compression refrigeration cycle, and the effects of the nozzle diameter of the ejector on the coefficient of performance (COP) of the system were discussed. When an ejector is used in an EVI heat pump system, the running parameters of the primary fluid, secondary fluid and mixing fluid are different from that of an ejector refrigeration system or ejector vacuum equipment, and the construction design and the matching of mass flow rates must be reconsidered. This paper will present the design method of an EVIe heat pump system with a scroll compressor made by a factory in China, emphasizing on the determination of the design parameters of the ejector in the EVIe system. Also, the experimental results of the prototype developed in terms of the design method will be discussed. This research should be helpful to the development of EVIe heat pump systems and in improving their performance.
2.
EVI and EVIe heat pump systems
Fig. 1 (a) and (b) are flow charts of the EVI heat pump system and EVIe system. The EVIe system is comprised of compressor, condenser, flash-reservoir, expansion valve, evaporator, ejector and other apparatus. The compressor, condenser, flash-tank, expansion valve and evaporator are connected together to form the main circuit, and the circuit is same as conventional single-stage compression heat pump system. The compressor has supplementary inlets, the ejector
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a
3
condenser
conveniently installed, but also has a high efficiency and a wide range of operation condition.
2 6 scroll compressor
throttling valve
5
2'
1
expansion valve
4
4'
evaporator
flash-reservoir
b
3
2 condenser ejector
6 scroll compressor
check valve 1
5
2' 1
check valve 2 expansion valve
4
4'
evaporator
flash-reservoir Fig. 1 e Heat pump system coupled with vapor injection compressor, (a) EVI system (b) EVIe system.
is connected between flash-reservoir and supplementary inlets of compressor, those equipments form supplementary circuit. The flashed vapor at the top of the flash-reservoir as a primary fluid flows through the ejector, and entrained a part of the refrigerant vapor from the evaporator, which is called entrained or secondary fluid. The primary fluid and secondary fluid are mixed in ejector, and the pressure of the mixed fluid then rises to the intermediate pressure at the exit of the ejector. This part of refrigerant is sucked into compressor through supplementary inlets. Compared with the EVI system, the throttling valve in the supplementary circuit is replaced by an ejector in the EVIe heat pump system. The ejector is able to increase the pressure of the secondary fluid without mechanical energy consumption, and using it in the EVIe system can reduce the available energy loss, and the temperature of the refrigerant vapor entering into the scroll compressor through supplementary inlets. As ejector has advantages such as simple structure, low cost, no moving parts, safety for two-phase working fluid and so on, the EVIe system not only be simply composed,
3.
Design of vapor supplementary holes
If the location of the vapor injection inlets of scroll compressor varies in a certain range, there is a small effect of the positions of the injection inlets on the steady operation of the EVIe heat pump system (Cho et al., 2003; Dutta et al., 2001; Wang et al., 2007). However, to ensure that the heating performance of the system is relatively high under the condition of low temperature, the optimal location of the vapor injection inlets is at the place where the vapor can inject into the working chambers when they just be closed (Wang et al., 2007 and Ma et al., 2003), as shown in Fig. 2. The shape of the vapor injection inlets also affects the performance of EVI system. Considering the convenience to manufacture, the inlets should be round and located symmetrically at a pair of compression chambers of the scroll compressor. The inlet must be smaller than the thickness of the scroll to ensure that the two adjacent chambers are not connected by the inlets at any time. Namely, the size must fit the scroll thickness, and different scroll need different inlet size. If the profile of scroll is an involute of circle with radius of a, the coordinates of the scroll center line in coordinate xoy in Fig. 2 can be expressed as, x ¼ aðcos f þ fsin fÞ
(1)
y ¼ aðsin f fcos fÞ
(2)
Where, a is radius of base circle, m; 4 is involute angle, rad. The working chambers of scroll compressor come forth in pairs, so there must be a couple of the injection inlets for a pair of suction chambers. According to the geometric relations of
y 120
Orbiting scroll 60 Fixed scroll
0
0
x
Vapor injection inlets
Fig. 2 e Scroll compressor and the location of vapor injection.
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the scroll (Li,1998), the coordinates of the center point of one injection inlet can be derived as, 7p 1 7p xo ¼ a cos 2g q þ 7psin 2g q 2 2 2
(3)
7p 1 7p yo ¼ a sin 2g q 7pcos 2g q 2 2 2
(4)
Where g is initial involute angle, rad; q is rotation angle of the shaft, rad. The coordinates of the center of the other injection inlet can be expressed as, 9p 1 9p þ 2g q þ 9psin þ 2g q xi ¼ a cos 2 2 2
(5)
9p 1 9p þ 2g q 9pcos þ 2g q yi ¼ a sin 2 2 2
(6)
For a given scroll compressor, its scroll profile is specified and the locations of the injection inlets depend only on the shaft angle. In terms of the geometry of scroll compressor, the amount of the vapor flowing through injection inlets is remarkable when the shaft angle q varies from 0 to 120 . The locations of the pair of inlets were shown in Fig. 2 while the shaft angles are 0 , 60 and 120 , respectively. The suction chambers of scroll compressor are just closed when the shaft angle is 0 . If the injection inlets locate at the shaft angle of 0 , this means that the vapor injection begins at end of the suction process, and the locations are favorable to vapor injection. Their coordinates are: 7p 1 7p 2g þ 7psin 2g xof ¼ a cos 2 2 2
(7)
7p 1 7p yof ¼ a sin 2g 7pcos 2g 2 2 2
(8)
and 9p 1 9p xif ¼ a cos þ 2g þ 9psin þ 2g 2 2 2
(9)
9p 1 9p yif ¼ a sin þ 2g 9pcos þ 2g 2 2 2
(10)
The scroll thickness can be expresses as t ¼ 2ag. For conventional scroll compressor, 3 mm t 6 mm. For round injection inlet, its radius must be less than the thickness of scroll so that the adjacent chambers cannot be connected through the inlets at any time. Usually, the inlet radius should be satisfied t/2 rj 3t/4. For the prototype compressor for this study, the scroll thickness is 3.6 mm and the radius of the round inlet is 2.4 mm. For conventional EVI system, when the condensation temperature is 45 C and the evaporation temperature varies from 15 C to 25 C, the favorable injection pressure is between 0.6 MPa and 0.8 MPa, and the injection vapor amounts from 7%e25% of the suction vapor flow rate under this condition (Ma et al., 2003). The injection vapor amounts
only to 5% while the intermediate pressure is reduced to 0.5 MPa. Therefore, the improvement of EVI system becomes inconspicuous if the intermediate pressure is less than 0.5 MPa (Zhao et al., 2006). For example, the intermediate pressure is 0.70 MPa, and the injection vapor amount is about 12% of the vapor suction flow rate with the condensation temperature of 45 C and the evaporation temperature of 20 C. Accordingly, the heating capacity of EVI system increases approximately 11.5%.
4.
Design method of ejector
A schematic diagram of the ejector is shown in Fig. 3. When the ejector is working, the primary fluid at high pressure expands and accelerates through the primary nozzle in the ejector, and the fluid is ejected with supersonic speed to create a very low pressure region at the plane of the nozzle exit. The fluid at low pressure, called the ‘secondary fluid’, is then entrained through the port connecting the nozzle exit and mixed with the primary fluid in the mixing chamber. The fully-mixed fluid decelerates and its pressure rises in the compression chamber as its kinetic energy is converted to the potential energy of pressure. The pressure of the mixed fluid at the ejector exit, called the intermediate pressure, must match the flow rate of the vapor to be injected into the compressor. The main operating parameters of the ejector are the entrainment ratio u and the pressure values of the primary fluid, the secondary fluid and mixing fluid. u, an important parameter to express the performance of an ejector, is defined as the ratio of the mass flow rate of the entrained fluid to that of the primary fluid. For a given working fluid, the ratio u can be calculated according to the pressure values of the primary fluid, secondary fluid and mixing fluid (detailed in Huang, 1977 and Tian, 2007). The performance of the ejector can be predicted using the aerodynamic function method (Huang, 1977 and Tian, 2007). Normally, the pressure of the mixing fluid, pC, is equal to the pressure of the injection vapor in the EVI heat pump system. The compression process of the scroll compressor is divided into two stages by the vapor injection and hence consists of by three sub-processes. They are the first closed compression process before connection with the supplementary inlets, the vapor injection and mixing process and the second closed compression process after the supplementary inlets are cut off. The process is similar to a two-stage compression process but is not a real one. So, it is also referred to as a quasi two-stage compression process.
f p*
f p1
f2
f3
Mixing Fluid
Primary Fluid Secondary Fluid
Suction Section
Mixing Section
Fig. 3 e Schematic diagram of ejector.
Diffuser Section
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To sum up, the appropriate intermediate pressure must be determined in advance, and the dimensions of the main sections of the ejector can be calculated in turn. According to the steady-flow energy equation for the ejector, the following equation can be written:
0.5 0.4 -15 -20 -25
u
0.3 0.2
hC ¼
0.1 0 0.55
0.7
0.85
1
1.15
1.3
Mixing Pressure (MPa) Fig. 4 e The variation of entrainment ratio u with mixing pressure.
There is a one-way valve assembled in the supplementary circuit to prevent the reversible flow in the circuit. The pressure of the vapor in the working chamber at the end of the first compression process rises to the value of p20 , and pC p20
(11)
As mentioned above, the ratio u is an important parameter to describe the ejector performance. When the pressure values of the primary fluid and the secondary fluid are given, u is affected by the pressure of the fluid at the ejector exit, namely the mixing pressure. The variation of u with the pressure at the ejector exit is shown in Fig. 4 when the evaporation temperatures are 15 C, 20 C and 25 C, respectively. As shown in Fig. 4, the entrainment ratio u will decrease with the increase of the mixing pressure. When the mixing pressure rises to 1.0 MPa, u decreases to less than 0.1. When the mixing pressure remains constant, the entrainment ratio u increases with the increase of the pressure of the entrained vapor corresponding to the evaporation temperature. To ensure that the ejection is effective, the entrainment ratio u must keep higher value so that the ejector exhibits favorable performance. The experimental results have demonstrated that the optimal value of u is greater than 0.1 (Pang et al., 2007). In this condition, the mixing fluid pressure corresponding to entrainment ratio u, pCu, is pC pCu
(12)
Where, pC is the pressure of the mixing fluid. If the entrainment ratio u is higher than 0.2, the mixing fluid pressure becomes too low, even less than 0.5 MPa. The low mixing pressure means the vapor injection to the compression chambers becomes weak. Generally, the reasonable value of u is between 0.1 and 0.2. The mass flow rate discharged from the compressor could be increased by the vapor injection, and the heating capacity of the EVI system also increases consequently. Nevertheless, the pressure of the injection vapor, pC, has an optimal value. If the pC value is higher than the value of the intermediate pressure, the performance of the system will deteriorate (Zhao et al., 2006). The appropriate value of the intermediate pressure can be decided by experiments.
hP þ uhH 1þu
(13)
Where, hC, hP and hH are the enthalpy values of the mixing fluid, primary fluid and secondary fluid (kJ kg1). The mass flow rate of the primary fluid is determined by the throat area of the ejector and the property of the fluid, and then (14) mP ¼ fP rP aP Where, mP is the mass flow rate of the primary fluid (kg s1); fP* is the nozzle section area (m2); rP* is the vapor density of the primary fluid in the nozzle section (kg m3); and aP* is the critical speed of the primary fluid (m s1). Therefore, the dimensionless flow rate a through the supplementary circuit can be calculated by the parameters of the ejector: a¼
md ms mP ð1 þ uÞ ¼ ms nrs Vhv =60
(15)
Where, md and ms are the mass flow rates of the discharge and the suction refrigerant of the compressor (kg s1); n is the rotation speed of the compressor (r min1); rs is the density of the suction refrigerant (kg m3); V is the stroke volume of the compressor (m3); and h is the volumetric efficiency of the compressor.
evaporation temperature,''to'' , vapor injection location
ejector modle , entrainment ratio "u"
EVI system experiment
lowest injection vapor pressure
upper limit of vapor injection pressure
lower limit of vapor injection pressure
determine vapor pressure
calculate flowrate " ¦ "Á
determine mass flow rate of ejector
calculate dimension of ejector
Fig. 5 e Flow chart of design process of EVIe heat pump system.
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Cooling tower
High and low pressure switch
Pump Filter
Comp One-way valve v3
Condenser
Evaporator
v2 v1
Heat exchanger Heater
TXV
Mass flow rate meter Filter
Receiver
Hot water bulk
pressure sensor temperature sensor valve
Pump Heater
Glycol pump Cool water bulk Heater
Fig. 6 e Test system and sensor locations for the prototype.
Based on the thermodynamic model for gas charging to a rigid container, the dimensionless flow rate a can also be derived as follows: a¼
m2 m20 v20 ¼ pC p20 x m20 RkTC
(16)
Where, v20 is the refrigerant specific volume at state point 20 (m3 kg1); p20 is the pressure of the refrigerant at state point 20 (kPa); R is the gas constant for the refrigerant (kJ kg1 K1); k is the adiabatic index of the refrigerant; TC is the temperature of the mixing fluid (K); and x is the resistance factor of the refrigerant flowing through the supplementary circuit, and it is recommended to be 0.2 according to experiments (Zhao et al., 2006 and Ma et al., 2003). Replacing the specifications and technical parameters of the compressor such as stroke volume, volumetric efficiency, and rotation speed into equation (15), the throat diameter of the nozzle of the ejector, dp*, can be obtained, and other section dimensions of the ejector can be calculated consequently. This process for designing the EVIe heat pump system can be summarized as shown in the flow chart in Fig. 5.
5.
Testing system
An EVI heat pump system was designed based on the technical data of the scroll compressor manufactured by a factory
in China. The testing system of the EVI heat pump system is shown in Fig. 6. The specifications of the scroll compressor were as follows: a stroke volume of 80 cm3/rev, rating power input of 4.55 kW, rotation speed of 2800 r/min, and working fluid of R22. The operation conditions were as follows: an evaporation temperature of 20 C and condensation temperature of 45 C. The flow rate of the glycol solutions through the evaporator was measured by measuring the volumetric discharge in a graduated flask per unit time. The heating capacity is computed by the measured values by the flow rate and the temperature difference of the water flowing through the condenser, and the flow rate of the water was measured with a flow rate meter. The electric heater in the glycol-water tank was adjustable so that the solution temperature was kept constant during the measurements. The temperature of the water flowing into condenser can be adjusted by the input power of the electric heater in the water tank. And the condensation temperature varies with the water temperature. All the electric parameters of the compressor such as the input voltage, current, power and other electric parameters were recorded by a data logger. Each temperature and pressure sensor was calibrated to reduce experimental uncertainties and was connected to an Agilent data logger. The accuracy and tolerance of each sensor for measuring data was shown in Table 1. The uncertainties of the heating capacity and EER estimated by the analysis were approximately 2.6%.
Table 1 e Accuracies and tolerances of the sensors. Sensor Temperature Pressure transducer Currency acquisition unit Power acquisition unit Data logger
Accuracy
0.2 C 0.2% of 0.2% of 0.2% of 0.2% of
Tolerance
full scale full scale full scale full scale
0.2 C 0.2% of full 0.2% of full 0.2% of full 0.2% of full
scale scale scale scale
Full scale
Model
e 2.5/1.0 MPa e e e
Pt100 Huba DZFC-1 DZFC-1 HP34970A
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Table 2 e Experimental data for heat pump prototype. p1 0.24 0.24 0.24 0.24 0.24 0.25 0.24 0.24 0.24 0.24 0.24 0.24 0.24 0.24
p10
t10
p20
t20
p3
t3
p40
t40
p6
t6
Qk
P
EER
0.23 0.22 0.22 0.22 0.22 0.23 0.22 0.22 0.22 0.23 0.22 0.22 0.22 0.22
10.46 10.13 10.50 10.29 10.28 10.48 10.26 10.94 10.14 10.02 10.70 10.89 10.62 10.27
1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73 1.73
117.56 116.83 116.49 115.94 115.03 115.59 115.18 114.13 113.25 112.38 108.00 107.99 106.38 105.03
1.69 1.68 1.68 1.69 1.69 1.69 1.68 1.68 1.67 1.68 1.66 1.66 1.65 1.65
43.77 43.65 43.60 43.87 43.69 43.52 43.67 43.60 43.40 43.47 43.38 43.38 43.19 43.01
0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26 0.26
15.25 15.37 15.39 15.31 15.41 14.16 15.02 15.11 15.11 15.06 14.70 14.70 14.74 14.95
0.54 0.54 0.53 0.52 0.50 0.69 0.70 0.90 0.90 1.10 1.09 1.29 1.29
18.13 17.92 16.12 14.51 10.49 19.79 19.79 23.38 17.68 26.74 26.74 30.17 27.94
5.35 5.85 5.97 6.06 6.04 6.03 6.19 6.36 6.08 6.08 6.15 6.22 5.82 5.92
4.33 4.38 4.36 4.32 4.36 4.32 4.60 4.62 4.66 4.79 4.78 4.86 4.89 4.86
1.24 1.33 1.37 1.40 1.39 1.40 1.35 1.37 1.31 1.27 1.29 1.28 1.19 1.22
Notes: p-pressure, MPa; t-temperature, C; Qk-heating capacity, kW; P-power input, kW.
Relevant parameters were recorded only if the fluctuation of every parameter was within 1%. The experimental method and data processing accorded with Chinese national standards GB 5773-86 (the standard for performance tests for positive displacement refrigerant compressors) and GB/T 9056-1999 (the standard for performance tests for positive displacement refrigerant compressor condensing units). In order to test the performance of the heat pump under different state parameters of the injection vapor, the testing system for the conventional heat pump was rebuilt. The power of the electric heater and the mass flow rate of the glycol-water solution of the heat exchanger in the supplementary circuit can be adjusted to keep the temperature of the injection vapor steady. By switching the valves on or off, it is easy to realize the following operation modes of the testing system. For the test system in Fig. 6, it is running as SSC system when all of the valves, v1, v2 and v3, are shut off, and running as EVI system when v1 is turned on and v2 and v3 off, and running as EVIe system when v1 is turned off and v2 and v3 on.
0.3
Testing results and discussions
The main measured data and the performance of the heat pump system calculated by the data are listed in Table 2. Table 2 shows that the heat pump has higher heating capacity and heating EER when the pressure of the injection vapor is 0.5e0.7 MPa and the temperature is 18e20 C, which are also the mixing fluid state parameters of the ejector. The dimensions of the main sections of the ejector are listed in Table 3. The experimental results have been rearranged as the curves shown in Figs. 7e10.
EVI EVIe 0.2 α
6.
The variation of the dimensionless flow rate a through the supplementary circuit with the evaporation temperature is shown in Fig. 7. For both the EVI and EVIe heat pump systems, the dimensionless flow rate a decreases with the increase of the evaporation temperature, and its value varies from 0.25 to 0.03 when the evaporation temperature goes up from 25 C to 15 C. This is because the pressure difference between the injection vapor and the refrigerant in the working chambers of the scroll compressor will be bigger when the evaporation temperature becomes lower. Under the same pressure of the injection vapor, the dimensionless flow rate a of the EVIe system is greater than that of the EVI system. For example, a is 0.24 for EVIe system but 0.17 for EVI system when the evaporation temperature is 23 C. This is because the ejector entrained the vapor at the exit of the evaporator with lower enthalpy and specific volume, and it gives the injection vapor a lower superheating degree and specific volume. Under the design condition of an evaporation temperature of 20 C, the ejector can achieve better performance, and it also reaches the maximum difference of the flow rate a between the EVIe and EVI system, and the difference value is about 0.08. The variations of the heating capacity with the evaporation temperature are shown in Fig. 8. The heating capacity of the
0.1
0 -25
Table 3 e Area of main section of ejector (mm2). f* 0.196
fp1
f2
f3
0.318
1
0.494
-23
-20
-17
-15
Evaporation Temperature
Fig. 7 e The variation of dimensionless flow rate a of the injection vapor with evaporation temperature.
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i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 4 ( 2 0 1 1 ) 1 5 8 7 e1 5 9 5
3
8
Heatin g EER
Heating Capacity (kW)
12
1.5 EVI EVIe SSC
EVI EVIe SSC
4
0 -25
0 -25
-23
-20
-17
-15
Evaporation Temperature
Fig. 8 e The variation of heating capacity with evaporation temperature.
EVIe system is slightly greater than that of the EVI system, and the difference of the heating capacity increases with the decrease of the evaporation temperature. No matter EVI or EVIe system, the heating capacity reduces from 9.6 kW to 9.0 kW with variation of evaporation temperature from 15 C to 17 C, and the capacity difference between the two systems is neglectable. However, the capacity difference became remarkable as the evaporation temperature decreases further, and is about 1.1 kW while the evaporation temperature goes down to 20 C. This is because the injection vapor in the EVIe system has lower enthalpy, and it results in lower enthalpy of the discharge vapor from the compressor, even though the flow rate a of the EVIe system is greater than that of the EVI system. The heating capacity is equal to the product of the enthalpy difference between the inlet and exit of the condenser and the mass flow rate of the discharge vapor, and the heating capacity increases only when the product becomes greater. When the evaporation temperature becomes low, the mass flow rate of the injection vapor becomes dominant compared
Po wer In p u t (kW)
5
4 EVI EVIe SSC
3 -25
-23
-20
-17
Evaporation Temperature
Fig. 9 e The variation of power input with evaporation temperature.
-15
-23
-20
-17
-15
Evaporation Temperature
Fig. 10 e The variation of heating EER with evaporation temperature.
with the discharge enthalpy, so their product increases with the mass flow rate. The variations of the power input of the compressor with the evaporation temperature are shown in Fig. 9. The power input of the EVI heat pump system decreases very slowly with the increase of the evaporation temperature and is 1.5e3.4% higher than that of the conventional heat pump. This is because the mass flow rate of the injection vapor rises with the increase of the evaporation temperature in both the EVIe and EVI system. The power input of the EVIe system is 4% lower than that of the EVI system at the same evaporation temperature. This demonstrates that the ejector can effectively recover a part of the pressure energy of the refrigerant in the supplementary circuit, and then the compression process of the system can be further improved. When the evaporation temperature is 20 C, the ejector performs better under the design conditions, and so the EVIe heat pump system exhibits favorable performance, as the compression efficiency is further increased by the ejector. The variations of heating EER with the evaporation temperature are shown in Fig. 10. Using an ejector in the EVIe heat pump system can improve the working conditions of the scroll compressor, so the compressor has a relatively high efficiency compared with the EVI system when it is operated at low temperature. For example, the heating EER is 2.15 for EVIe system and 1.64 for EVI system when the evaporation temperature is 20 C. The heating EER of the EVIe heat pump system rises about 3e5% compared with the EVI system while the heating capacity remains nearly constant. The improvement of the EER becomes more remarkable as the evaporation temperature falls. This demonstrates that the improvement by using the ejector in the heat pump system is more remarkable under the conditions of low evaporation temperature. On the other hand, the improvement of the heating capacity by using the ejector becomes relatively small with the increase of the evaporation temperature. When the evaporation temperature goes up to 15 C, the heating capacity of the EVIe system is very close that of the EVI system. This is because the dimensionless flow rate a through
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 4 ( 2 0 1 1 ) 1 5 8 7 e1 5 9 5
the supplementary circuit decreases from 20 to 30%e10%, as shown in Fig. 7, when the evaporation temperature is increased from 25 C to 15 C. In both the EVIe system and the EVI system, the improvement of the heating capacity is weakened when the flow rate a becomes slow. Thus, there is no remarkable improvement from using the ejector to replace the throttling valve in the supplementary circuit if the evaporation temperature is more than 15 C.
7.
Conclusions
The design method of the heat pump system coupled with the EVI scroll compressor and ejector was presented, and the prototype was tested with evaporation temperature from 25 C to 15 C. The following conclusions could be made. (1) The optimal location of the vapor injection inlets is at the place where the vapor can inject into the working chambers when they just be closed, and the size must fit the scroll thickness to ensure that the two adjacent chambers are not connected by the inlets at any time. For a scroll with the circle involute profile, its thickness, t, is usually between 3 mm and 6 mm, and the radius of round injection inlets should be satisfied t/2 rj 3t/4. (2) To ensure that the ejection is effective, the entrainment ratio u must keep higher value so that the ejector exhibits favorable performance. Generally, the reasonable value of u is between 0.1 and 0.2. (3) The experimental results of the prototype demonstrate that the heating EER of the EVI heat pump system with ejector is about 4% higher than that of the system without an ejector while the heating capacity is kept nearly constant.
Acknowledgments We are very grateful to the support of National Natural Science Foundation No. 50776001, China, and express our heartfelt thanks to Mr Wu Xuezhi, Pang Zongzhan, Peng Long and Miss Yu Lihua for their effective help in the development and experiment of the prototype.
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references
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