Recent advances in vapor compression cycle technologies

international journal of refrigeration 60 (2015) 118–134

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Review

Recent advances in vapor compression cycle technologies Chasik Park a, Hoseong Lee b, Yunho Hwang b,*, Reinhard Radermacher b a b

School of Mechanical Engineering, Hoseo University, Asan 336-795, Republic of Korea Center for Environmental Energy Engineering, University of Maryland, 4164 Glenn L. Martin Hall Bldg., College Park, MD 20742, USA

A R T I C L E

I N F O

A B S T R A C T

Article history:

This paper comprehensively reviews the recent studies on advanced vapor compression cycle

Received 11 May 2015

technologies. These technologies are categorized in three groups: subcooling cycles, expan-

Received in revised form 1 August

sion loss recovery cycles, and multi-stage cycles. The subcooling cycle research is focused

2015

on a suction-line heat exchanger, thermoelectric subcooler and mechanical subcooler. The

Accepted 8 August 2015

expansion loss recovery cycles are mainly focused on utilizing an expander and ejector. The

Available online 20 August 2015

multi-stage cycle research includes a vapor or liquid refrigerant injection cycle, two-phase refrigerant injection cycle. All these advanced vapor compression cycle technology options

Key words:

are reviewed, and their effects are discussed. In recent years, the research and develop-

Vapor compression cycle

ment have been made to improve the performance of the VCC. This paper presents the

Subcooling

improved cycle options and their comprehensive review. From the review results, several

Ejector

future research needs were suggested.

Expander

© 2015 Elsevier Ltd and International Institute of Refrigeration. All rights reserved.

Injection Saturation cycle

Récents progrès dans les technologies de cycle à compression de vapeur Mots clés : Cycle à compression de vapeur ; Sous refroidissement ; Ejecteur ; Détendeur ; Injection ; Cycle à saturation

1.

Introduction

Energy saving has become an important issue due to the limited energy resources and ever increasing demands. In the US, the energy use by a space cooling, space heating, water heating,

and refrigeration represents about 76% of total energy consumption for the residential buildings (Department of Energy, 2015), in which energy systems are mainly relying on a vapor compression cycle (VCC). This VCC has inherent thermodynamic losses as compared to an ideal reverse Carnot cycle. Those are thermodynamic losses associated with single phase

* Corresponding author. Center for Environmental Energy Engineering, University of Maryland, 4164 Glenn L. Martin Hall Bldg., College Park, MD 20742, USA. Tel.: +1 301 405 5247; Fax: +1 301 405 2025. E-mail address: [email protected] (Y. Hwang). http://dx.doi.org/10.1016/j.ijrefrig.2015.08.005 0140-7007/© 2015 Elsevier Ltd and International Institute of Refrigeration. All rights reserved.

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injection cycle, two-phase refrigerant injection cycle and saturation cycle). In this paper, each cycle options are reviewed in terms of performance improvement.

Nomenclature

Symbols cp Δq h P Q s T

specific heat capacity refrigeration effect specific enthalpy pressure heat transfer capacity specific entropy temperature

2.

Subcooling cycles

Typically, the state of the refrigerant entering the expansion device in the VCC is subcooled liquid, so that liquid can be expanded in the expansion device and provides a stable refrigerant flow rate. When the degree of subcooling is increased, the typical expansion process, the isenthalpic process approaches to the isentropic process as shown in Fig. 1. Moreover, increased subcooling degree can increase the refrigeration effect ( Δq) and potentially improve the coefficient of performance (COP). Any other heat sink of appropriate temperature could be used to increase subcooling, but the following three methods are mainly applied; a suction line heat exchanger (SLHX), mechanical subcooler, and thermoelectric subcooler.

Subscripts c critical L saturated liquid vap vaporization Acronyms ASHRAE American Society of Heating, Refrigerating and Air-Conditioning Engineers COP coefficient of performance DESC double expansion subcooler DOE Department of Energy EEV electric expansion valve FTSC flash tank and subcooler IHX internal heat exchanger PR pressure ratio SLHX suction line heat exchanger TXV thermostatic expansion valve VCC vapor compression cycle

2.1.

The use of the SLHX has been widely applied to VCCs.The SLHX or internal heat exchanger (IHX) is often employed as a means for protecting system components.The SLHX ensures subcooled liquid refrigerant to be supplied to the expansion device inlet and superheated vapor refrigerant to the compressor inlet. Fig. 2 shows the schematic diagram and P–h diagram of the vapor compression cycle with the SLHX.The SLHX is located between the condenser outlet and expansion valve inlet and between the evaporator outlet and compressor inlet. Cold refrigerant from the suction line is used to cool down the refrigerant at condenser outlet. Reduced refrigerant temperature of the condenser outlet decreases the enthalpy of the evaporator inlet. In turn, this increases the evaporator capacity. However, the increased temperature of the compressor inlet decreases the compressor efficiency, which degrades the system performance. The increase of the suction temperature decreases the compressor volumetric efficiency which is proportional to the inverse of the suction temperature.Therefore, these two aspects should be considered in the application with the SLHX. Domanski

gas compression and isenthalpic expansion. The first loss results in high discharge refrigerant temperature, high compression work, and high condenser heat release. The second loss results in large throttling losses and low refrigeration capacity. To reduce these thermodynamic losses, many researchers investigated the improved cycle options, such as subcooling cycles (suction line heat exchanger, thermoelectric subcooler, and mechanical subcooler), expansion loss recovery cycles (expander and ejector), multi-stage cycles (a vapor or liquid refrigerant

ln P

Suction line heat exchanger

ln P

Subcooling

Isenthalpic process

Δq Increase of refrigerating effect Isentropic process

h

Fig. 1 – P–h diagram of the vapor compression cycle with liquid subcooling.

h

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3

2 Condenser

ln P Compressor 4

3

2

SLHX 4 1

6

Expansion device

5

6

1

Evaporator

h

5

Fig. 2 – Schematic diagram and P–h diagram of the vapor compression cycle with SLHX.

et al. (1994) represented that the benefit of the SLHX application depends on a combination of operating conditions and fluid properties for the refrigerants.The results showed that the COP and volumetric capacity of the VCC system with the SLHX are greatly affected by fluid properties. Klein et al. (2000) investigated the influence of the SLHX on the performance of the refrigeration cycle using a new dimensionless group for various refrigerants. The effect of the SLHX on the refrigeration capacity was correlated in terms of the temperature lift and a dimensionless grouping equal to the enthalpy of vaporization (hvap ) at the evaporator temperature divided by the product of the specific heat capacity (Cp,L ) of saturated liquid refrigerant at the evaporating temperature and the critical temperature (Tc ) . The relationship between the relative capacity index and this dimensionless quantity is shown in Fig. 3, which represents that the VCC with SLHX would be more useful for refrigerants having a relatively small value of Δhvap (Cp,L Tc ) under

Fig. 3 – Relative capacity index versus Δhvap·Cp,L −1·Tc −1 at saturated evaporating and condensing temperatures of −20 °C and 40 °C, respectively (Klein et al., 2000).

the condition of the large temperature difference between condensing and evaporating temperature. As results, the SLHX was detrimental to system performance in systems using R-22, R-32, and R-717, whereas it was useful for systems using R-507A, R-134a, R-12, R-404A, R-290, R-407C, R-600, and R-410A. Particularly, performance improvement by applying the SLHX has been extensively studied in a CO2 cycle. Lorentzen and Pettersen (1993) showed COP improvement by using the SLHX to reduce throttling losses. Several researchers (Hafner, 2000; Halozan and Rieberer, 2000; Rozhentsev and Wang, 2001) represented performance variations of a CO2 cycle with the SLHX by comparing their characteristics with those of other conventional refrigerant systems. Hwang et al. (2001) showed a 7% increase of the COP by applying the SLHX into a singlestage compression cycle, and an 18% increase of the COP by applying the SLHX into a two-stage compression cycle. Cho et al. (2007) reported that the cooling COP increased from 2.1 to 2.27, and the optimal compressor discharge pressure reduced from 9.2 to 8.7 MPa with the application of the SLHX. In automotive air conditioning systems, the influence of the SLHX using R-1234yf as working fluid was analyzed by previous studies (Jarall, 2012; Lee et al., 2011; Zhao et al., 2012; Zilio et al., 2011). R-1234yf has been proposed as a replacement for R-134a in mobile air conditioning systems due to their similar thermophysical properties. However, the cooling capacity of the R-1234yf system was smaller than that of the R-134a system when the same size compressor is used. It is because the latent heat of evaporation of R-1234yf in the evaporator is smaller than that of R-134a by 21-28% at the same saturation temperature (Cho et al., 2012). Therefore, the SLHX is applied to R-1234yf VCCs. Navarro-Esbrí et al. (2013) experimentally analyzed the effects of the SLHX on the performance of VCC system using R-1234yf. Tube-in-tube type heat exchanger was used and the effectiveness of the SLHX ranged from 17 to 25%. Experimental results showed that the cooling capacity increased by 2–9% with the SLHX application. Cho et al. (2013) showed that the cooling capacity and COP of the R-1234yf system without the SLHX decreased by up to 7% and 4.5%, respectively, as

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compared to those of R-134a system. However, those of the R-1234yf system with the SLHX decreased by 1.8% and 2.9%, implying that similar performance could be achieved by the SLHX. Pottker and Hrnjak (2015) performed an experimental study about the effect of condenser subcooling on the performance of an air conditioning system operating with R-134a and R-1234yf. The results showed the use of both the condensing subcooling and the SLHX simultaneously yielded more efficient performance improvement, especially for R-1234yf. Preissner et al. (2000) tested the performance of an R134a automotive prototype air-conditioning system with the SLHX. Tube-in-tube type heat exchanger was applied and the effectiveness of the SLHX was between 55 and 65%. At a higher condenser air temperature of 40 °C and a restrictive idling air flow rate of 1.0 m s−1, the COP and the capacity increased on the order of 5–10% with a SLHX with 60% effectiveness.

2.2.

Mechanical subcooler

Mechanical subcooling cycles utilize a small cooling system to improve the main refrigeration cycle and result in improvement of the cooling capacity. Thornton et al. (1994) developed an ideal mechanical subcooling cycle from Carnot theory and heat transfer relations. In this study, the improvement in overall COP through the use of a subcooler was found to be

ln P

approximately 10% over a range of conditions representing supermarket applications. Khan and Zubair (2000) investigated the integrated subcooler cycle which is coupled to the main cycle at exit of the condenser for rejecting the heat as shown in Fig. 4a. The author demonstrated that the COP of the system was increased by 7.5% in maximum when the subcooler saturation temperature was about arithmetic mean between the condensation and evaporation temperatures. Figure 4b shows that the dedicated mechanical subcooler cycle provides subcooling to the main cycle and both the main cycle and the subcooler cycle have their own condensers. Qureshi and Zubair (2012) studied the performance characteristics of different refrigerant combinations in the dedicated mechanical subcooler cycle. The refrigerants such as R-134a, R-410A, R-407C, R-717, and R-404A were considered as alternatives for different applications. R-134a and R-410A are often used for the applications from high temperature to medium temperature range. Therefore, these refrigerants were used in the subcooler cycle as they have relatively elevated operating temperatures as compared to the main cycle. In this study, when R-134a was used as a refrigerant in the main cycle, the subcooler cycle produced the best results in terms of COP. She et al. (2014) proposed a new subcooling cycle based on the expansion power recovery as shown Fig. 4c and Fig. 5. This proposed system applied the expander to drive the compressor in the mechanical subcooler

ln P

Enthalpy gain

Enthalpy gain

h

(a) Integrated mechanical subcooling

h

(b) Dedicated mechanical subcooling

ln P Expansion with expander

Enthalpy gain

h

(c) Mechanical subcooling based on expansion power recovery with expander Fig. 4 – P–h diagrams of various mechanical subcooling cycles.

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Fig. 5 – Schematic diagram of the mechanical subcooling based on expansion power recovery with expander (She et al., 2014).

cycle. The amount of the enthalpy gain including expander power and subcooling capacity resulted in performance improvement better than the conventional mechanical subcooling system. The proposed system achieved higher COP by 19.72% than the conventional mechanical subcooling system when R-744 was used as the refrigerant in the main refrigeration cycle and R-12 was employed for the subcooler cycle.

2.3.

Thermoelectric subcooler

The working principle of thermoelectric cooling is based on the Peltier effect. Thermoelectric modules are solid-state electronic devices that directly convert electricity to temperature difference. Thermoelectric devices have no moving parts and, therefore, are inherently reliable and require little maintenance. However, the use of thermoelectric devices and systems has been limited by their relatively low energy conversion efficiency. The COP decreases rapidly with increasing temperature lift while thermoelectric cooling is very well suited for small temperature lifts where it achieves very high COPs (Winkler et al., 2006). Therefore, it is possible to use thermoelectric technology to improve the performance of conventional VCC systems instead of utilizing a full-fledged thermoelectric coolingsystem. In this hybrid system, the thermoelectric subcooler is installed between the outlet of condenser and the inlet of expansion device. Winkler et al. (2006) reported that the use of thermoelectric subcooler in CO2 system increased the cooling capacity by 20% and the system COP by 16% through the simulation. Radermacher et al. (2007) showed that the use of the staged thermoelectric subcooler yielded more system COP improvement as compared to a single thermoelectric subcooler in theoretical study. Schoenfeld et al. (2012) experimentally investigated the performance improvement potential of the thermoelectric subcooler in a CO2 cycle. As shown in Fig. 6, the thermoelectric subcooler consists of three main components,

which are the thermoelectric modules, a cold-side heat exchanger with a microchannel, and a hot-side heat exchanger with working as a thermosyphon evaporator. This study showed that this hybrid system improved the COP of the CO2 cycle by 3.3% with a corresponding 7.9% capacity increase when the thermoelectric subcooler was effectively controlled to maximize the system COP. Sarkar (2013) simulated the energetic as well as exergetic analyses and optimizations of the CO2 refrigeration cycle with a thermoelectric subcooler. As results, the improvements of maximum cycle COP and the second law efficiency at optimum CO2 subcooling are 0.86% and 17.18%, respectively as compared to the cycle without the thermoelectric subcooler. The author concluded that use of the thermoelectric subcooler in CO2 refrigeration system not only improved the cycle cooling COP at optimum conditions, but also reduced the system highside pressure, compressor pressure ratio and compressor discharge temperature. Furthermore, another application of the thermoelectric device was attempted as a thermoelectric generator to reduce the compressor power consumption (Yilbas and Sahin, 2014). Fig. 7 shows the schematic diagram with the option of the thermoelectric generator. In this study, thermal analysis of the combined cycle, consisting of a refrigerator and thermoelectric power generator, was presented. The influence of the location of thermoelectric generator, in the combined system, on the cycle performance was investigated. The author reported that the location of the thermoelectric generator in between the condenser and the evaporator results in lower coefficient of performance of the combined system. This is because the heat rejected from the thermoelectric generator increases the heat transfer to the refrigeration system. On the other hands, the location of the thermoelectric generator in between the condenser and its ambient improves the coefficient of performance the combined cycle. The author showed the operating parameters had significant effect on the performance characteristics of the combined system.

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Fig. 6 – Schematic diagram of the vapor compression cycle with thermoelectric subcooler (Schoenfeld et al., 2012).

3.

Recovery of expansion losses

The typical VCC uses one of them such as a capillary tube, thermostatic expansion valve (TXV), and electric expansion valve (EEV) to expand a high pressure refrigerant exiting from the condenser. Theoretically, the expansion process is considered as an isenthalpic process which causes thermodynamic loss (Kornhauser, 1990). To recover this thermodynamic loss, an isentropic process is required in the expansion process. An expander or an ejector as the expansion device can be used to generate isentropic condition in the throttling process.

3.1.

Expander cycle

Fig. 8 shows the schematic diagram and P–h diagram of the expander cycle. An expander can be seen as a compressor operating in reverse. Therefore, any compressor mechanism can be used as the expander in principle. The expander enables to improve the COP of the system in two ways by increasing the cooling capacity through an isentropic process and by utilizing the recovered expansion losses for assisting the compressor, which results in a reduced compressor power consumption. Huff et al. (2002) investigated on the performance of positive displacement expanders in CO2 application. Using a

Fig. 7 – Schematic diagram of the thermoelectric generator configuration (Yilbas and Sahin, 2014).

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ln P Condenser

3

2 3

Expander

2

Compressor

4

1

1 4

Evaporator

h Fig. 8 – Schematic diagram and P–h diagram of the vapor compression cycle with expander.

work-extracting expansion process instead of the isenthalpic throttling process can improve the cycle performance significantly and may help make CO2 competitive with conventional refrigerants. First law estimations show potential improvements of the COP in the order of 40%–70% and 5%–15% for the capacity. With the expander coupled directly to the compressor shaft, the control of the high-side pressure requires special attention. Fig. 9 shows the circuitry options for high-side pressure control in expander systems. Option (a) is in a system with independent compressor and expander speeds. In options (b) and (c), the high-side pressure may be adjusted by increasing the total flow

resistance across the expansion section by introducing an additional, adjustable pressure drop. Option (d) adjusts the flow resistance across the expansion section by introducing an adjustable by-pass to the expander. In options (e) and (f), the compression is divided into two sections. The expander speed is determined by the operating conditions. The low stage compressor (e) or the high-stage compressor (f) is independent of the expander speed. Nickl et al. (2005) presented the integration of a threestage expander into a CO2 refrigeration system. The COP was improved and the discharge pressure of the main compressor was lowered. When the discharge pressure was 10 MPa and the suction pressure was 3 MPa, the improvement of the system COP compared with the throttle cycle was larger than 40%. Wang et al. (2012) conducted simulation study on a novel vanetype expander with internal two-stage expansion process for R-410A refrigeration system. The study showed that the proposed expander obtained built-in volumetric ratio up to 7.6 with the isentropic efficiency of 55.9% at 2000 rpm and theoretically improved the COP from 4.0 to 4.56, by 14.2%, under design operating condition (condensation temperature: 54.4 °C, and evaporation temperature: 7.2 °C). Subiantoro and Ooi (2013) reported an economic analysis of the expander cycle on to existing vapor compression cooling systems, particularly medium scale air conditioners. Assuming that the compressor and the expander efficiencies are 75% and 50%, respectively, the payback periods are less than 5 years for all the systems. The expander cycle has a great potential to improve the VCC performance. There are several options to extract the work by the expander. (1) The shaft of the expander can be combined with that of the compressor. (2) The generated electricity can be used for the compressor. (3) The multiple expanders can be utilized to recover the expansion loss more efficiently.The limitation on the expander cycle is the low expander efficiency. Moreover, due to the large pressure differences inside the machine one has to minimize the possibilities for internal leakage.

3.2. Fig. 9 – Circuitry options for high-side pressure control in expander systems (Huff and Radermacher, 2003).

Ejector cycle

Fig. 10 shows the schematic diagram and P–h diagram of the ejector cycle. An ejector is also known as a jet, injector or jet

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3

2

ln P

Condenser

3

2

Compressor Ejector 9 6

7

1

8

Separator 9

4

1 5

4 Expansion device

7

8

6

5 Evaporator

h

Fig. 10 – Schematic diagram and P–h diagram of the single evaporator ejector cycle.

pump in different literatures (Jiautheen and Annamalai, 2014). The ejector is mainly compose of a nozzle, mixing chamber and diffuser. Basically, the ejector is designed to mix the high pressure fluid with low pressure fluid. The refrigerant from the condenser flows through the ejector and exits the nozzle, which creates a low pressure at the nozzle outlet. The low pressure draws a refrigerant from the low temperature evaporator outlet. High pressure fluid from the condenser and low pressure fluid from the low temperature evaporator could be mixed in the mixing chamber. The pressure of mixed refrigerant recovers at the diffuser section. The ejector is an energy converter that transforms expansion losses into kinetic energy and then back to an increase pressure so that the compression work is reduced. Disawas and Wongwises (2004) investigated about the effect of heat source and heat sink temperature on the performance of refrigeration cycle using ejector as an expansion device. For R-134a, the heat source condition was varied from 6 to 18 °C, while the heat sink condition was varied from 25 to 40 °C. The authors reported that the COP improvement was decreased as the heat sink temperature was increased.The compressor pressure ratio and discharge temperature of the ejector cycle were lower than those of the typical VCC. In other research, Li and Groll (2005) also found that COP improvement of the ejector cycle was higher, especially at low inlet cold water temperature. They presented an ejector expansion transcritical CO2 cycle based on the constant pressure-mixing model. The effect of the entrainment ratio and the pressure drop in the receiving section of the ejector on the relative performance of the ejector expansion transcritical CO2 cycle was investigated using the theoretical model.They reported that the ejector expansion cycle improved the COP by more than 16% as compared to the basic VCC for typical air conditioning applications. However, if the separator does not properly separate the liquid and vapor from each other in the single phase ejector cycle in Fig. 10, the performance of the ejector cycle would be deteriorated. This is because the liquid that leaves the separator at the vapor port could be supplied to the compressor instead of the evaporator, decreases the cooling capacity and increases the compressor work, which both lower COP.

Fig. 11 shows the schematic diagram and P–h diagram of the dual evaporator ejector cycle. Lawrence and Elbel (2013, 2014) and Boumaraf et al. (2014) also named it as the alternate two-phase ejector cycle. In this cycle, the liquid at the outlet of the condenser is split into two separate streams. One stream is sent to the motive nozzle of the ejector and the other is isenthalpically throttled and sent to an evaporator. The two streams are mixed in the ejector mixing section and enter a second evaporator, where they are vaporized before returning to the compressor and condenser. This alternate ejector cycle enables to relax the constraints between the ejector entrainment ratio and the quality of the ejector outlet stream. The simulation results with R-134a and R-1234yf (Boumaraf et al., 2014) showed that the dual evaporator ejector cycle has an improvement in the COP more than 17% at Tc = 40 °C for both. When it was applied to R-1234yf, the COP was increased more than R-134a, especially at high condensing temperatures. However, R-134a represented slightly higher values of COP for both the conventional VCC and the dual evaporator ejector cycle. In the experimental results (Lawrence and Elbel, 2014), the dual evaporator ejector cycle showed maximum COP improvements of 12% with R-1234yf and 8% with R-134a as compared to a two evaporation temperature expansion valve cycle. Hafner et al. (2014) reported efficiencies and capacities for an R-744 supermarket system layout with ejectors as shown in Fig. 12. Heat from the freezing part of the system is rejected to the medium temperature part. From this part the heat is released to the different heat recovery units and external heat rejection devices like gas cooler and inter-stage cooler. The medium temperature compressors are drawing the gas from the first separator (SP-1) downstream of the ejectors. The ejectors are applied to maintain a certain pressure difference between the separator (SP-2) and the separator (SP-1).The multiejector system showed a significant COP increase as compared to the reference system (booster system with flash gas bypass and heat recovery) for both cooling and heating modes. For different climate conditions, the COP in the cooling mode increased by 5–17% and also the COP in heating mode increased between 20% and 30% as compared to the reference system.

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3

2 Condenser

ln P

Compressor

3

2

Ejector 9 6

7

High Temperature Evaporator

8

1

Expansion device

9 1 5

4 7

4

Low Temperature Evaporator

8

6

5

h

Fig. 11 – Schematic diagram and P–h diagram of the dual evaporator ejector cycle.

Lin et al. (2013) carried out the experimental investigation for the performance of an adjustable ejector used in a multi-evaporator refrigeration system as shown in Fig. 13. The adjustable ejector with a spindle to adjust the area of the nozzle throat was applied to deal with the considerable variation in primary cooling load for air conditioning system. In Fig. 13, the refrigerant flow is divided into three streams which enter evaporators 1, 2, and 3 after pressure reduction in the electronic expansion valves EEV1, EEV2, and EEV3, respectively. The refrigerant from evaporator 3 at state (8) and evaporator 2 at state (7) enter ejector 2 as primary and secondary flows, respectively. Then the two streams are mixed and leave the ejector at state (9). The mixed flow leaving ejector 2 is entrained in ejector 1 by the primary flow coming from evaporator 1. After pressure

improved, the refrigerant leaves ejector 1 at state (1).The authors described that the adjustable ejector applied in the multievaporator refrigeration system could be used to control the primary inlet pressure for better pressure recovery property. For the expander cycle, the throttling loss is recovered and transferred to the compressor, but total compressor work is not changed. Therefore, the expander efficiency is the most important factor. For the ejector cycle, the compressor work decreases with increase of the suction pressure. Moreover, heat transfer capacity increases with the use of the separator. However, the performance improvement can be limited by the ejector efficiency. Moreover, the ejector geometry is fixed for one operating condition so that the novel design should be developed for the wide range of operation.

Fig. 12 – Circuit diagram of supermarket refrigeration and heating system with multi-ejector and R-744 (Hafner et al., 2014).

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Fig. 13 – Schematic diagram and P–h diagram of the multi-evaporator refrigeration system with the adjustable ejector (Lin et al., 2013).

4.

Multi-stage cycle

As the temperature difference between evaporation and condensation increases under the condition of either very low or high ambient temperatures, the performance and reliability of the heat pump could be deteriorated. This is because the increase in irreversibility during the compression process and the reduction in the mass flow rate. Among several technologies to overcome its disadvantage, a multi-stage cycle has rapidly developed in recent years.

4.1.

Two-stage cycle

The injection techniques can be divided into liquid, vapor, and two-phase injection according to the state of the injected refrigerant. In the study of liquid injection, Dutta et al. (2001) and Cho et al. (2003) reported the influence of liquid injection on the performance of a scroll compressor, the discharge temperature of which decreased as the injection ration increased. Winandy and Lebrun (2002) investigated experimentally and theoretically for an R-22 heat pump with liquid and vapor injection. The main effect of the liquid injection was to decrease the compressor discharge temperature by about 1.2 °C for each injection ratio. The vapor injection cycle can be separated into the flash tank and subcooler vapor injection cycle. For the flash tank cycle, refrigerant vapor for injection was provided by phase separation inside the flash tank. The flash tank cycle shows unexpected flooding in the compressor at high speeds due to the difficulty of accurately controlling the amount of injection. On the other hand, injected vapor in the subcooler cycle was generated by heat exchange using the temperature difference between before and after the high-stage expansion device in the subcooler. The subcooler cycle yields a lower heating performance than the flash tank cycle even though it allows for more stable and precise cycle control through the variation of the injected amount. Heo et al. (2010) studied the heating performance of the flash tank cycle using R-410A. The COP and heating capacity of the flash tank cycle were enhanced by up to 10% and 25%, respectively, at the ambient

temperature of −15 °C as compared to those of the noninjection cycle. Wang et al. (2009) also experimented with the heating and cooling performance of the flash tank and subcooler cycles using R-410A. The flash tank and subcooler cycles showed the performance improvement over the noninjection system. In the operation condition of maximum cooling, the cooling capacity and COP were improved by 15% and 2%, respectively, at an ambient temperature of 46.1 °C. The heating COP was improved up to 23% by the flash tank cycle at an ambient temperature of −17.8 °C. Heo et al. (2011) investigated the heating performance of the heat pumps in which novel vapor injection techniques of a combined flash tank and subcooler (FTSC) cycle and a double expansion subcooler (DESC) cycle were applied as shown in Fig. 14. The DESC cycle was a modified subcooler cycle by adding an electronic expansion valve at the condenser exit to allow pressure control in the subcooler. The FTSC cycle included both a flash tank and a subcooler to increase the enthalpy difference in the evaporator due to the removal of flash gas in the flash tank and allow precise control of the injected amount. The author compared the average performance for each cycle option with that of the subcooler cycle as shown in Fig. 15. The average heating capacities of the flash tank, FTSC, and DESC cycles were higher by 14.4%, 6.0%, and 3.8%, respectively, than that of the subcooler cycle. The average COPs of all the cycle options were very similar. The author suggested the heating performance of the flash tank cycle was certainly superior to that of each of the other cycle options. However, when stable and precise cycle control is a major concern in system design, the FTSC cycle is the best selection as an alternative cycle because the flash tank cycle can experience unexpected flooding in the compressor at high speeds due to the difficulty in precise control of the injected amount. Xu et al. (2011a, 2011b, 2013) investigated on a vapor injection flash tank heat pump system, and proposed a novel cycle control strategy. Through experiment, the proportional–integral– derivative (PID) controller was able to provide accurate control of the electronic expansion valve (EEV) to reach the target superheat. It was reported that the injected vapor’s superheat can be effectively used as the control signal of the upperstage expansion valve.

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Fig. 14 – Schematic diagram of the vapor injection cycles (Heo et al., 2011).

4.2.

Saturation cycle

The saturation cycle was proposed by Lee et al. (2013) to enhance the VCC performance by reducing thermodynamic losses associated with single phase gas compression and isenthalpic expansion. In order to approach the saturation cycle, the saturation expansion and saturation compression pro-

Fig. 15 – Comparison of the average performance for each cycle option (Heo et al., 2011).

cesses were used simultaneously together with evaporation and condensing processes. Fig. 16 showed i-stage refrigerant injection VCCs, which can approach the ideal saturation vapor compression and expansion cycle. The refrigerant injection process could be repeated in many steps as long as the compressor design allows. As the stage number increases, the compression and expansion processes approach the saturation lines as shown in Fig. 17. When either the partially expanded two-phase refrigerant is separated in the phase separating tank or the partially expanded sub-stream of twophase refrigerant is heated by the main stream refrigerant from the condenser, all of separated cold vapor and a part of liquid are used for two-phase refrigerant injection to the compressor, and the low enthalpy liquid is further expanded to the following bottoming stage. The author described the saturation cycle would be able to improve not only the capacity but also the coefficient of the performance. In addition, the twophase injection technique was applied to the saturation cycle. The refrigerant state after mixing with two-phase refrigerant could be controlled by P-T sensors. Lee et al. (2015) simulated the COPs of two different working fluids under five different operating conditions as shown in Fig. 18 and Table 1. As the stage number is increased, the COPs increase. In the results, the heating COP of the four-stage cycle was higher by 42.4% for R-410A and 38.2% for the propane as compared to those of their respective single stage cycle.

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Fig. 16 – Concept of the saturation cycle (single stage cycle (a) and three multi-stage cycles (b–d)) (Lee et al., 2013).

5.

Cycle simulation

Seven different cycles are modeled using Engineering Equation Solver (EES) (Software, F-Chart, 2012). It should be noted that combination of the cycle options is not considered in this paper because its performance depends on the operating conditions and working fluids. The expansion devices are assumed to be an isenthalpic process, and the pressure drops of the heat exchangers are neglected. Moreover, the degrees of superheating and subcooling are assumed to be constant. The degree of the reheat process of the supply air is assumed 10 K in cooling mode. R-410A is used as a working fluid. The refrigerant-side is assumed as follows: • Degree of subcooling: 5 K (Whitman et al., 2004). • Degree of superheating: 5 K (Whitman et al., 2004). • Condensing temperature: 10 K higher than ambient air temperature. (Hwang, 2004). • Evaporating temperature: 10 K lower than the refrigerated air temperature. (Hwang, 2004). Fig. 17 – Schematic diagram of two-phase injected multi-stage vapor compression cycle (Lee et al., 2015).

Cycle modeling is conducted with two methods. The first approach is to assume the systems as ideal cases. This mod-

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Fig. 18 – COPs of two working fluids under various operating modes (Lee et al., 2015).

eling is to investigate the potential of the cycle options. The effectiveness of the SLHX, efficiencies of the compressor, expander, and ejector are assumed to be 1. The second approach is to assume the best values in the literature. The compressor efficiency is assumed to be 0.85. For the SLHX, the effectiveness of the heat exchanger is set to be 0.7. For the expander, the expander isentropic efficiency is defined to be 0.5 (Fukuta et al., 2006). Produced power from the expander is 100% used for the compressor. For ejector cycle, the modeling is built based on the paper (Li and Groll, 2005). Isentropic efficiency

of the nozzles is assumed to be 0.7 and that of the diffuser is assumed to be 0.7. The entrainment ratio is defined as an initial value. After calculation, it was compared with the quality of the separator, and then updated to match these two values. Regarding the refrigerant injection, the intermediate pressures are selected to result in equal pressure ratios across the compressor stages, which minimize the compressor power (Moran et al., 2010). For the saturation cycle, two-phase refrigerant is injected to compressor, and the amount is controlled to maintain the degree of superheat. All of separated cold vapor

Table 1 – Simulation conditions (ANSI/ASHRAE standard 116, 1995). Test

A – cooling mode C – cooling mode High temp – heating mode Low temp – heating mode Extreme mode

Indoor

Outdoor

DB

WB

RH

DB

WB

RH

26.7 °C 21.1 °C

19.6 °C 13.9 °C 15.6 °C

50.7% 21.4% 56.4%

32.2 °C

NA

NA

35.0 °C 27.8 °C 16.7 °C −8.3 °C −15.4 °C

NA NA 14.7 °C −9.4 °C NA

NA NA 81.1% 69.8% NA

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Table 2 – Summary of modeling assumption. Property

Value

Degree of subcooling Degree of superheating Degree of supply air reheat process Condensing temperature Evaporator temperature Compressor efficiency SLHX effectiveness Expander isentropic efficiency Ejector nozzle efficiency Ejector diffuser efficiency

5K 5K 10 K 10 K higher than ambient temperature 10 K lower than refrigerated air temperature 0.85 0.7 0.5 0.7 0.7

and a part of liquid are used for two-phase refrigerant injection to the compressor. Four-stage cycle is considered as the saturation cycle in the simulation. The modeling assumptions are summarized in Table 2. Fig. 19 shows the modeling results of the ideal efficiency case and realistic efficiency case, respectively. In this context, “the ideal efficiency case” means that all device related efficiencies (such as compressor isentropic efficiency, SLHX efficiency, expander efficiency, and ejector efficiency) are assumed to be 1; while “the realistic efficiency case” is based on the highest reported device efficiencies. In Fig. 19, the variation of the COP with cycle options and evaporation temperature are shown for these two cases. When condensing temperature

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is maintained at 45 °C, as evaporation temperature is increased, the COP increases. The use of the SLHX results in an increase of the heat capacity of the exchanger and an increase of the compressor work. Depending on two aspects, the COP of the cycle either increases or decreases. The benefit of the SLHX depends on the operating conditions and fluid properties, such as specific heat capacity, latent heat, and coefficient of thermal expansion. The effect of using the SLHX is extremely small for R-410A at high ambient temperature range. For other cycle options, as evaporation temperature is decreased, the COP improvement increases. A higher pressure ratio cycle has more potential to be improved with cycle enhancement options due to larger thermodynamic losses such as compression and throttling losses. When the cycle options are operated with ideal device efficiencies as shown in Fig. 19, the expander cycle shows the highest COP improvement. The expander cycle improves the expansion process as an isentropic process, and recovered work from the expander is used to reduce the compressor power. These two effects increase the capacity, reduce the work, and finally contribute to the COP improvement. Therefore, when the expander efficiency is assumed to be ideal, the COP improvement of the expander cycle is extraordinary. The ejector cycle also recovers the expansion loss, but its improvement is limited. It is because the ejector cycle mainly elevates the suction pressure, which increases the compressor efficiency. When the cycle options are operated with realistic device efficiencies as shown in Fig. 20, the saturation cycle shows the highest COP improvement among all cycle options.

Fig. 19 – Variation of COP with cycle options and evaporation temperature.

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Fig. 20 – Variation of COP with cycle options and evaporation temperature (real case).

6.

Conclusions

In recent years, the research and development have been made to improve the performance of the VCC. This paper presents the improved cycle options and comprehensively reviewed them. The improved cycle options include subcooling, recovery of expansion loss, and multi-stage cycle. The following conclusions are summarized from the reviewed works: • For the application of the SLHX, it can be categorized in two groups: one group has positive effect in the COP improvement (R-507A, R-134a, R-12, R-404A, R-290, R-407C, R-600, R-744, R-1234yf and R-410A) and the other group has no COP improvement (R-22, R-32, and R-717). • The use of thermoelectric device as the subcooler needs the multi-staged arrangement due to the rapid decrease of the COP with increasing temperature lift. • The performance of the ejector cycle can be increased at the low heat sink temperature. The application of the adjustable ejector enables the potential of the energy savings in multiple-evaporator refrigeration system. • In the vapor injection cycle, the flash tank cycle has the possibility of unexpected flooding in the compressor despite its superior performance. When considering the stable and precise cycle control, the FTSC cycle would be the best selection as an alternative cycle.

• The saturation cycle shows the best performance improvement among advanced cycle options. When the saturation cycle technique is applied by using four-stage cycle, the heating COPs of R-410A and the propane cycles are improved by 42.4% and 38.2%, respectively. These findings could direct the following future researches. The expander cycle has a great potential to improve the VCC performance. The limitation on the expander cycle is the low expander efficiency. Moreover, due to the large pressure differences inside the machine one has to minimize the possibilities of internal leakage.The ejector applied VCC has a large potential to improve the performance but there are some limitations such as a narrow range of operation due to the fixed geometry. When advanced vapor compression cycles are applied, the complexity of the cycles and the cost of the system increase due to the additional new components. Therefore, economic analysis on each cycle option would be recommended. Currently, the saturation cycle promises the best performance improvement, which needs more research on temperature measurement and controls. Moreover, two-phase injection compressor needs to be developed and optimized.

Acknowledgements This work was supported by the sponsors of the CEEE, University of Maryland, College Park, MD.

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