An updated review of recent advances on modified technologies in transcritical CO2 refrigeration cycle

Journal Pre-proof An updated review of recent advances on modified technologies in transcritical CO2 refrigeration cycle Binbin Yu, Jingye Yang, Dandong Wang, Junye Shi, Jiangping Chen PII:

S0360-5442(19)31842-0

DOI:

https://doi.org/10.1016/j.energy.2019.116147

Reference:

EGY 116147

To appear in:

Energy

Received Date: 6 July 2019 Revised Date:

9 September 2019

Accepted Date: 16 September 2019

Please cite this article as: Yu B, Yang J, Wang D, Shi J, Chen J, An updated review of recent advances on modified technologies in transcritical CO2 refrigeration cycle, Energy (2019), doi: https:// doi.org/10.1016/j.energy.2019.116147. This is a PDF file of an article that has undergone enhancements after acceptance, such as the addition of a cover page and metadata, and formatting for readability, but it is not yet the definitive version of record. This version will undergo additional copyediting, typesetting and review before it is published in its final form, but we are providing this version to give early visibility of the article. Please note that, during the production process, errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain. © 2019 Published by Elsevier Ltd.

An updated review of recent advances on modified technologies in transcritical CO2 refrigeration cycle Binbin Yua, Jingye Yanga, Dandong Wanga, Junye Shia,b, Jiangping Chena,b* a. Institute of Refrigeration and Cryogenics, Shanghai Jiaotong University, Shanghai, China b. Shanghai High Efficiency Cooling System Research Center, Shanghai, China *corresponding author: [email protected] Tel. +(86)21 34206775 Abstract With carbon emission levels on the rise, rapid and far-reaching action is needed to counteract global warming. Among many available strategies, CO2 is nowadays more and more often proposed as a solution for heating, cooling and refrigeration purposes since the extremely low global warming potential and being natural. In order to overcome the inherently low efficiency in high-temperature conditions and high operating pressure especially in transcritical cycles, an updated review of the advances on modified technologies to solve the drawbacks of CO2 refrigeration is provided and recent progress on the energy efficiency improvement is summarized. First, the basic principles of the CO2 refrigeration cycle and important performance characteristics are discussed. Then, a detailed discussion on different modified technologies as well as their operating fundamental, technical features and performance are provided, followed by a summary of previous studies. At the end of this review, conclusion and perspectives on the future development of this field are presented. Keywords Natural refrigerant; CO2; Energy efficiency; Transcritical cycle; Global warming Nomenclature abs cP cc COP chem des DX DMS DBT FGB GWP h IHX l MAC OCR phys P T

Absorption Specific heat, kJ kg-1 K-1 Cubic centimeter Coefficient of performance Chemical Desorption Direct expansion Dedicated mechanical subcooling Dry bulb temperature Flash gas bypass Global warming potential Enthalpy, kJ kg-1 Internal heat exchanger liquid Mobile air conditioning Oil circulation rate Physical Pressure, MPa Temperature, K 1

TES vap WBT

Thermoelectric mechanical subcooling vapor Wet-bulb temperature

Greek letters

η ε

Effectiveness Compression ration

Subscripts ev gc is i opt s

Evaporation Gas cooler Isentropic Intermediate Optimum Saturation

1. Introduction The adoption of the Paris Agreement aroused the public discourse focusing on the atmospheric concentration of greenhouse gases[1]. Almost three years later, the Intergovernmental Panel on Climate Change (IPCC) has issued an alarming report on the risks of a global warming increase above 1.5 °C [2]. The report stresses the disastrous consequences of what a 2 °C increase would mean compared to a 1.5°C increase. The deceptively small difference between these two temperature increases obscures a predicted 10 cm rise in sea levels, severe Arctic sea ice decrease, and a loss of 99% of the Earth’s coral reefs. With carbon emission levels on the rise, the IPCC warns that “rapid and far-reaching” action is needed in order to keep global warming levels at 1.5°C. Among many strategies for global warming mitigation, refrigerant management is placed first on Hawken’s book [3], in which total 80 strategies are presented and analyzed. Many Hydrofluorocarbons (HFCs) used in heating, ventilation, air conditioning and refrigeration (HVAC&R) industry have high global warming potential (GWP), for example, R134a has a GWP of 1340 over a 100-year period [4]. The clear phase-out pathway of HFCs has been listed in Kigali amendment report by the 28th Meeting of the Parties to the Montreal Protocol [5]. The urgency of refrigerant substitution cannot be overstated. In this sense, natural refrigerants have aroused increasing interest in the HVAC&R field as an ultimate solution rather than seeking for new chemicals like HFCs. Among many natural refrigerants, i.e. fluids like hydrocarbons, ammonia, air, water and CO 2 . CO 2 is the only refrigerant that is non-flammable, non-toxic, 0 ODP, 1 GWP, classified as A1 [6], and can operate below 0 . No other refrigerants can meet these properties simultaneously. For this point, CO 2 can be the final solution in refrigeration systems without any environmental concern, which is a big problem for current society. Furthermore, CO 2 is inexpensive and shows higher latent heat, specific heat, density and thermal conductivity and lower viscosity, in comparison with HFCs. From a historical development perspective, in 1850, CO 2 was first introduced as a refrigerant in a British patent received by Alexander Twining [7]. The first CO 2 system was built by Lowe to produce artificial ice in the late 1860s [8]. Since then, CO 2 had been always on the market until synthetic refrigerants CFCs aggressively appeared in the 1930s with better performance and lower cost [9]. In the late 1980s, ozone-depleting refrigerants including CFCs began to be phased out. In 1990, Norway professor Gustav Lorentzen published a patent application of transcritical CO 2 system 2

intended for automobile air conditioning, which indicated the reviving of CO 2 [10]. In 1993, the experimental results of the first prototype CO 2 system were published by Lorentzen and Pettersen [11]. The prototype system was further improved by Pettersen and showed similar performance to that of R12 [12]. In 1999, the first transcritical CO 2 residential air conditioning was simulated and compared with the R22 system in a German patent [13]. Afterward, since the beginning of the 21st century, researchers have never stopped studying transcritical CO 2 cycles in various applications. To the best knowledge of the authors, several review papers on transcritical CO 2 cycle have been summarized and compiled. Kim et al. [14] presented a review for transcritical CO 2 cycle technology in various refrigeration, air-conditioning and heat pump applications based on the literature from 1994 to 2004, in which fundamental process and system design issues were elaborated. Besides, four modified cycles including internal heat exchange cycle, expansion with work recovery, two-stage cycle and flash gas bypass were introduced to improve system efficiency. Ma et al. [15] provided a comprehensive overview of transcritical CO 2 heat pump and refrigeration systems, issues including properties of supercritical CO 2 and that containing PAG lubricants, optimum pressure, novel cycles based on expanders are covered. Pradeep [16] presented a historic view of the fundamentals and application of CO 2 in low-temperature refrigeration systems especially in the food industry, he suggested that further fundamental research on unraveling the physics of boiling and condensation heat transfer of CO 2 and CO 2 -oil mixture extending to mini and micro heat exchangers is needed. Maina [17] reviewed various applications of CO 2 and situations in South Africa, the authors believe CO 2 will be the ultimate refrigerant of the future. Qi [18] reviewed the advances of CO 2 air conditioning and heat pump system in the electric vehicle. Paride et al. [19] implemented an in-depth review covering the CO 2 refrigeration plants for food retail applications from energy, environmental and economic perspective. Brian et al. [20] presented an overview of transcritical carbon dioxide heat pump systems with numerical analysis, system components, configurations and modifications of the compression and expansion process was also covered. Besides, heat transfer properties of supercritical CO 2 were reviewed by Luisa et al. [21], and various CO 2 transcritical work recovery expanders were reviewed by Simarpreet et al. [22] However, the systematic and comprehensive description of novel improvement technologies in basic CO 2 refrigeration cycles which simultaneously includes operating fundamentals, technical features and performance are limited. Many novel technologies have been innovatively developed and gotten substantial improvements in recent years, whereas these technologies have not been summarized and analyzed in the above reviews. Motivated by this point, the present study aims to present a thorough state-of-art review on an information update of numerous improvement technologies for transcritical CO 2 cycle. This review starts with a brief description of the basic principles and important performance of transcritical CO2 cycle. It is then followed by detailed discussions on different improvement technologies based on their fundamentals, features and performance. In the end, we present our brief perspectives on the prospects of this field. 2. The basic transcritical CO2 refrigeration cycle In transcritical CO2 cycle, most of the time the heat rejection takes place above the critical point in the supercritical region, no saturation condition exists and the pressure is independent of the temperature, so the “condenser” in conventional vapor compression cycle is replaced by a gas cooling device called “gas cooler”. However, the heat absorption occurs by evaporation of the refrigerant at low pressure, which is a similar case in the conventional subcritical cycle [23]. Fig. 1 shows the basic transcritical CO2 cycle, in which the lgP-h diagram was drawn based on the following assumptions: no pressure drop, the gas cooler outlet temperature of 32 , the 3

evaporation temperature of 5 , superheat of 0 K, isentropic compression process and isenthalpic expansion. It can be seen that the high-side pressure is around 10 MPa, which is normally 5-6 times higher than conventional refrigerants, resulting in high cost and offering a threat to system reliability and safety. Theoretically, compared to a conventional vapor compression cycle, the transcritical CO2 cycle is less efficient in the same conditions. Take the R134a subcritical cycle for an example, Fig. 2 illustrates the comparison of thermodynamic cycles for R134a and CO2 in the temperature-entropy diagram, in which the exergy losses in terms of areas are obvious to be observed [14]. In Fig. 2, assuming equal evaporating temperature, equal gas cooler outlet temperature with condensing temperature, two processes are responsible for the additional thermodynamic losses of transcritical CO2 cycle, i. e. throttling loss and heat rejection loss. The larger heat rejection loss results from the much higher average CO2 temperature during the gas cooling process. The throttling loss has much to do with the refrigerant properties, with temperatures determined, larger entropy increase occurs in the throttling process due to the greater pressure difference before and after CO2 expansion device. However, the large pressure difference in transcritical CO2 cycle generates a lower compression ratio, being around three in the cooling application while the conventional cycles operate at a ratio of eight [24]. The lower compression ratio makes a compressor tend to be more efficient. It should be noted that due to better heat transfer characteristics of CO2, the approach temperature between gas cooler outlet temperature and heat sink temperature can be much lower than that between condensing temperature and heat sink temperature [25]. Even considering these benefits, the transcritical CO2 cycle shows low-efficiency performance in high ambient temperatures. Martin [26] and Yin [27] demonstrated that the COP (coefficient of performance) of a CO2 air conditioning system is lower in the 10% usage conditions at high ambient temperature (above 30 ). Brown et al., evaluated the performance merits of CO2 and R134a mobile air conditioning systems, their results from semi-theoretical models show that the COP of CO2 was lower by 21% at 32.2 and by 34% at 48.9 . The COP disparity was even greater at high speeds and ambient temperatures [28]. In addition, Compared with a conventional R134a mobile air conditioning system, the CO2 mobile air conditioning system achieved comparable cooling capacity but had COP reductions of 26% and 10% at 27 and 45 outdoor conditions, respectively [29]. a

b

Fig. 1. Basic transcritical CO2 cycle (a), and lgP–h diagram (b).

4

Fig. 2. Comparison of thermodynamic cycles for R134a and CO2 in a temperature-entropy diagram [14]. The aforementioned literature review brings the key challenges of transcritical CO2 refrigeration cycle to light, i.e. high operating pressure and low energy efficiency in high ambient temperature conditions. Therefore, researchers have been devoted to solving them through various technologies. 3. The modified transcritical CO2 cycle The comparison of the exergy losses between subcritical and transcritical cycle shows how much a transcritical CO2 cycle is penalized. Although the theoretical and experimental results indicated that the CO2 basic cycle system efficiency would be inferior, a large number of improved technologies are promising to make the actual efficiency of transcritical CO2 cycle equal or even superior to conventional subcritical cycles. Fig. 3 shows the outline of all the improvement technologies that will be elaborated in following subsections.

Fig. 3. Outline of the review for improvement technologies towards transcritical CO2 cycle. 3.1 Internal heat exchanger cycle The P-h diagram of conventional refrigeration system with classical IHX (internal heat exchanger) is shown in Fig. 4, it had been evidenced that the IHX can both improve or decrease the system performance because of the trade-off between increased capacity and discharge temperature depending on working fluids and operating conditions [30-32]. Aprea et al. [33] presented a simplified criterion to evaluate the possible advantage of adopting an IHX from a thermodynamic point of view. The criterion was stated as the following inequality based on the state 5

point in Fig. 4: cPT1 > (h1-h4) (1) where cP (kJ/kg K) is the constant pressure specific heat in the suction condition. When this inequality is verified, the use of an IHX turns out to be advantageous [34]. When it comes to CO2, the system configuration and lgP-h diagram of transcritical CO2 cycle with and without IHX is shown in Fig. 5, in which the process taking place in the low-pressure side of IHX are the same as that of the conventional refrigeration cycle. Theoretically, the criterion discussed in Ref. [33] is applicable to CO2. Thus, one can easily calculate that the use of IHX is useful for transcritical CO2 application in summer conditions with high ambient temperatures, while the internal heat exchanger does not improve the performance of the CO2 subcritical cycle [35, 44].

Fig. 4. P-h diagram of the conventional refrigeration cycle with (1-1’-2’-3-3’-4’-1) and without IHX (1-2-3-4-1) [33]. a

b

Fig. 5. A sketch of the CO2 refrigeration cycle with IHX (a), and lgP–h diagram of transcritical CO2 cycle with and without IHX (b). In fact, the benefits of utilizing IHX for transcritical CO2 cycle are quite considerable, various researches have validated this in different applications from various aspects. For mobile air conditioning, both the capacity and COP could be increased by up to 25%. Greater influence is registered at the higher temperatures of the air to the gas cooler. In detail, counter-flow IHX is better than parallel if the compressor temperature is below its limit. Larger IHX is beneficial to increase capacity and COP and reduce the optimum pressure where the maximum COP value is obtained, but the limiting size exists to prevent the 6

compressor discharge temperature exceeding its design limit [36, 41]. For commercial refrigeration, Joaquim et al. [38] studied the influence of IHX on a single-stage CO2 transcritical cycle by numerical simulation and experimental validation, they reported that the inclusion of an IHX significantly increases the COP. Under ambient temperature of 35 and 43 , the COP of a cycle with an IHX of 2 m length increases around 23% and 35%, respectively. COP increases more when the ambient temperature increases. In residential applications, 10% COP increment was obtained in a transcritical CO2 cycle working as a classical ‘‘split system’’ to cool air [40]. For a CO2 transcritical cycle for chiller application, Purohit et al. [45] studied the effects of IHX on system performance in an energetic and exergetic perspective. In the whole system, IHX has the least contribution towards exergy loss and maximum improvement in COP and exergetic efficiency by the IHX are 5.71% and 5.05% respectively. However, the IHX cycle also leads to a maximum rise of 24 °C in the compressor discharge temperature. Regarding the IHX effectiveness, because phase change might exist in IHX, and the specific heat changes drastically in the region near above the critical pressure, the traditional expression used to describe the heat exchange effectiveness is not suitable, a practical effectiveness expression for IHX was derived by Chen et al. [46] based on enthalpy difference, an IHX with high effectiveness is a very important factor for a CO2 transcritical system to achieve high COP. However, in another study, system COP was found to be inversely proportional to the IHX thermal effectiveness [47]. Normally, the classical position of an IHX in the above studies is all at gas cooler exit, but different positions for IHX configurations may also have different results. As shown in Fig. 6, Sanchez et al. [48] compared the energetic behavior of a transcritical CO2 refrigerating plant working with several IHX configurations: gas cooler exit (classical position), liquid receiver exit, and in both positions at the same time. Though a general improvement in COP and cooling capacity has been observed regardless of the position of the IHX, a maximum increment of 13% on COP has been obtained with two IHX at the same time.

Fig. 6. Experimental facility diagram for the investigation of the effect of IHX positions on system performance [48]. 3.2 Ejector expansion cycle 7

Since the first standard ejector cycle analyzed by Kornhauser [49] in 1990, the ejector technology has been extensively studied. In the existing literature, many reviews on ejector technology have been presented based on various working fluids [50-61], the ejector fundamentals, operating characteristics, ejector geometry effects, modeling, and basic control issues have been involved. Among them, a most comprehensive review of more than 300 papers was presented by Besagni et al. in 2016 [59], an overview of ejector refrigeration systems that developed during the past almost 30 years was classified and shown as Fig. 7, in which the transcritical CO2 ejector refrigeration system was partially discussed in the eighth module. In fact, considering the aforementioned larger throttling exergy loss of CO2 than that of other conventional refrigerants, the ejector is more promising for COP improvement in transcritical CO2 system [61, 62, 75].

Fig. 7. Overview of ejector refrigeration systems [59]. Fig. 8 (a) shows the layout of a standard ejector expansion transcritical CO2 cycle, based on this system, the first numerical study of ejector expansion transcritical CO2 refrigeration cycle was proposed in 2002 [63]. While the first detailed experimental analysis of a prototype ejector was reported by Elbel and Hrnjak in 2008 [64], in which 7% increment of system COP was obtained. And they were the first researchers to introduce a variable two-phase ejector to a transcritical R744 system by installing a needle in the motive nozzle to control the motive nozzle throat diameter. Numerous additional experimental studies by the optimized ejector and system components have then appeared in various applications: mobile air conditioning [65], the multi-ejector system in supermarket refrigeration [67], dairy plant [68], refrigerated sea water chiller [69], residential CO2 air conditioning [70], and others with a maximum COP improvement up to 28% [66, 71]. A main problem of the standard ejector cycle is the high oil circulation rate (OCR) in the evaporator because the oil in the separator can not return back to the compressor easily. Zhu et al. [72] experimentally investigated the OCR in the standard ejector transcritical CO2 cycle, and a significantly higher OCR of about 10% was observed at the evaporator inlet than that at the high-pressure side (~1%). Another ejector cycle is shown in Fig. 8 (b), in which no separator was used but with two up-wind and down-wind evaporators. It was originally proposed by Oshitani et al. [73]. In this cycle, the refrigerant flow leaving the condenser is split into two streams, in which the OCR is 8

believed to be the same and equal to that of the flow going through the condenser. Therefore, at steady state according to the mass balance, the OCR of the flow going through the two evaporators should be the same and is equal to the OCR of the rest of the system. Thus the problem of high OCR in the evaporator is not seen in this ejector cycle design [72]. Besides, Lawrence and Elbel [74] argued that this ejector cycle has more important practical advantages over the standard ejector cycle for those working fluids that offer less recovery potential, such as R134a. Based on this cycle, DENSO [75] has presented their new generation ejector for mobile air conditioning system. Fig. 9 (a) illustrates the working principle and its integration with the up-wind and down-wind evaporators. In Fig. 9 (a), an active flow ratio control (ARC) device is used to achieve optimal refrigerant flow ratio through ejector by separating the two-phase refrigerant flow into gas and liquid with centrifugal force, which is generated by the eccentric entry. Then the liquid-rich refrigerant flows into the down-wind evaporator through the fixed orifice, the remained gas-liquid two-phase refrigerant flows to the nozzle side and then enters the up-wind evaporator. Test results show that the ejector efficiency of the new generation is 3 times as high as previous ejectors. At the same time, the system with this new ejector in Fig. 9 (b) has achieved approximately 20% power saving compared with a conventional air conditioning system without an ejector. And it has also achieved 10% power saving compared with the previous ejector system [75]. a

b

Fig. 8. (a) Standard ejector expansion transcritical CO2 cycle, (b) New ejector cycle. a

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b Fig. 9. (a) new generation swirl flow ejector and its integration with up-wind and down-wind evaporator by DENSO, (b) the system with new generation ejector [75]. Besides investigating the system performance, visualization is always an efficient method to better interpret the internal complex flow phenomena and to improve the CO2 ejector geometry and model validation. Recently, with the development of visualization technology solving the high-pressure difficulty, the complicated flow phenomena, and phase change process inside the CO2 ejector have been observed in several studies through direct photography or internal physical variable measurement methods. Elbel [76] measured wall static pressure distribution of a transcritical CO2 two-phase ejector. Nakagawa et al. [77] obtained the pressure and temperature values along the diverging section of several supersonic CO2 primary converging-diverging nozzles with thermocouples and strain-gauges. The supersonic expansion two-phase flow after the nozzle in the suction chamber and the mixing chamber of a transcritical CO2 ejector were visualized by Zhu et al. [78] using a direct photography method. The results showed that the liquid fraction in the active flow in the suction chamber increases with increasing pressures of active flow and suction flow. The mixing of the active and suction flows in the mixing chamber is very fast. The expansion angle of the active flow at the nozzle exit decreases with increasing suction flow pressures. The entrainment ratio is inversely proportional to the expansion angle. Regarding the CO2 phase change phenomenon in the nozzle, Li et al. [79] investigated the transition of the phase change position in the primary converging-diverging nozzle using direct photography method, as depicted in Fig. 10. The visualization images revealed that the phase change could start after or before the throat according to the operating conditions. The phase change position moved upwards when the primary flow inlet pressure and temperature decreased simultaneously. However, from the literature, there is still no relevant theory about the supersonic multiphase flow with phase transition, resulting in the fact that research of the flow theory in the supersonic ejector is not conclusive, which further lead to the large error between CFD model calculation and experiments. What’s more, the current shock-wave theory is based on single-phase flow, the shock wave pattern inside nozzle expansion section and the initial location of choked flow for the real fluid haven’t been understood well, which may be solved by further visualization methods in the future.

10

Fig. 10. Schematic diagram of the visualization ejector and photography test platform [79]. 3.3 Vortex tube A vortex tube can separate the inlet gas into two simultaneous streams with different temperature, as it is shown in Fig. 11, one of the separated streams is at a higher, while another stream is at a lower temperature than that of introduced fluid. Detailed information about vortex tube including its geometry design, experiments, simulation, and so on have been reviewed by Thakare et al. [80]. It was discovered by Ranque [81] in 1933 and was further investigated by Hilsch in 1947 [82]. Hence, the vortex tube is also known as Ranque–Hilsch vortex tube. Based on the Ranque–Hilsch effect, smart guys can easily assume that it is capable to operate as an expansion device in the refrigeration cycle. In 1997, Keller [84] proposed a vapor-compression refrigeration cycle with a vortex tube, as shown in Fig. 12. Its working principle is as follows. The compressed high-temperature refrigerant is cooled in the gas cooler and flows into the intermediate cooler, in which the refrigerant is further cooled by the vapor stream from a separator. Simultaneously, the vapor stream is heated in the intermediate cooler and then enters the vortex tube, because of the Ranque–Hilsch effect, the vapor stream is separated into two streams, i.e. cold gas and hot gas. The hot fraction must be cooled in a desuperheater and then mixed with the cold gas. The mixture is then mixed with the vapor from the evaporator and the mixed three streams flow into the compressor to be compressed. Keller theoretically analyzed the cycle performance using R22, R134a, CO2 as working fluids. Results showed that the COP was increased by 5%, 10%, and 15%, respectively, compared to the conventional basic vapor compression cycle.

Fig. 11. Schematic of a vortex tube [83].

Fig. 12. Schematic of the vortex tube expansion cycle proposed by Keller [84]. Keller’s cycle seems to be a little complicated, then in 1999, Maurer proposed a transcritical refrigeration cycle with a vortex tube, as shown in Fig. 13 (a) [85]. In this system, there are no throttling valves, and more 11

importantly, two-phase flow exists in the vortex tube for this system. The cooled gas from gas cooler flows into the vortex tube and is divided into three low-pressure parts: superheated hot vapor, saturated liquid and saturated cold gas. The saturated liquid and vapor then mix and flow into the evaporator. The superheated gas is cooled in the desuperheater and mixes with the vapor from the evaporator to be compressed. Similarly, Fig. 13 (b) shows another configuration of the vortex tube transcritical CO2 refrigeration system proposed by Li et al. [86]. In this case, saturated liquid stream leaves from the cold side of the vortex, while the superheated vapor stream leaves from the hot side of the vortex. The superheated hot vapor enters an auxiliary heat exchanger to be cooled and then mixes with the saturated liquid. Assuming a 100% gas-liquid separation efficiency, this vortex tube system could provide up to a 37% increase in cycle efficiency, and 50% separation efficiency could provide about 20% increase in cycle efficiency than that of the conventional basic system [86].

a

b

Fig. 13. Vortex tube transcritical CO2 refrigeration system proposed by: (a) Maurer [85], (b) Li et al. [86]. It must be noted that from the literature, though the vortex tube transcritical CO2 refrigeration system has theoretically proved to be useful in various applications [87], no experimental results have been reported to validate such improvements and more importantly, the key is to demonstrate a functional two-phase vortex tube and its separation efficiency. 3.4 Expander Another profitable expansion work recovery technology for transcritical CO2 cycle is the expander, as the throttling loss is much higher than that of conventional working fluid due to the physical properties. The improvement potential of the system energy efficiency is largely depending on the isentropic efficiency in the expander, which is commonly lower than that in the compressor, resulting from the different states of the fluid during the compression and expansion processes, single phase during the compression and two-phase during part of the expansion, as two-phase flow is more subjected to friction losses. In this context, transcritical cycle behave better than conventional cycles, because the expansion process in transcritical cycle to a large extent concerns single phase, the fluid above the critical pressure is dense gas and only the final expansion involves two-phase fluid, moreover during the two-phase process the densities of CO2 liquid and vapor are not so much different as in conventional cycles. All that helps to keep high enough the isentropic efficiency of the CO2 expansion process, unlike what happens when trying to use expanders with the common refrigerants [90]. The expander system configuration is shown in Fig. 14, in which the expander replaces the expansion valve, the expander can operate 12

either independently or coaxially with the compressor, about 37% of compressor work can be recovered by CO2 expander and this number increases with increasing expander inlet temperature [15]. Tian et al. [88] theoretically calculate the transcritical CO2 cycle with an expander and reported 6-10% higher COP than that of the basic system. Yang et al. [89] presented both first and second law of thermodynamical analysis on the use of an expander, they concluded that the expander system prevented 50% decrease of exergy loss in the expansion process and produced 30% improvement of the overall system exergy efficiency, leading to 33% increase of COP. The core component of expander technologies have been reviewed by Ma et al. [15] and Simarpreet et al. [22] a

b

Fig. 14. Schematic of the CO2 refrigeration system with expander working independently (a), or coaxially with compressor (b). 3.5 Subcooling In the basic CO2 system, the supercritical CO2 stream from the gas cooler outlet directly flows into the expansion device to be throttled. Theoretically, the gas cooler outlet temperature for CO2 is limited by the ambient temperature, there exists an approach temperature varying with the performance of the gas cooler. After throttling, the evaporator inlet quality (vapor fraction) cannot be reduced further due to the limited gas cooler outlet temperature, and thus the specific cooling capacity is limited accordingly. To break this restriction, researchers have presented subcooling technologies at the exit of the gas cooler and proven the efficiency enhancement despite the extra consumption work in the subcooling system. The combination of subcooling and the main CO2 cycle is shown in Fig. 15. The subcooling system is used to further cool the CO2 from the gas cooler, which is the process of 3-4 in Fig. 15 (a), therefore the quality of evaporator inlet after throttling is reduced as point 5 in Fig. 15 (b) and the specific cooling capacity increases. For this cycle, there are mainly two subcooling technologies that have been extensively studied: dedicated mechanical subcooling (DMS) and thermoelectric subcooling (TES). Compared to the basic system, the overall system COPs have been validated to increase substantially for the two subcooling technologies despite the little extra work consumed in the subcooling system [91, 92]. a

b

13

Fig. 15. Schematic of the CO2 subcooling refrigeration system (a), and its lgP-h diagram (b). For the DMS method, the subcooling system in Fig. 15 (a) normally uses a subcritical vapor compression system, it rejects heat to the same heat sink as the main CO2 cycle. Llopis et al. theoretically [93] investigated the effects of subcooling degrees provided by the mechanical subcooling device with different refrigerants in a refrigeration plant for supermarket application. They found that for all the provided subcooling degrees from 2.5 -7.5 , the optimum heat rejection pressure, where the maximum COP value is obtained for the basic system, the overall cooling capacity and COP are improved by 19.7% and 13.7% for the evaporation temperature of -5 , respectively. It should be noted that for different refrigerants used in the subcooling cycle as R290, R1270, R1234yf, R161, R152a and R134a, the overall COP improvement performance are similar among them. Then, they performed an experimental study on the single-stage CO2 transcritical refrigeration plant with R1234yf DMS system, the cooling capacity and COP improvements are analyzed with maximum increments on the capacity of 55.7% and 30.3% on COP [94]. In a previous study of the authors, the performance of CO2 air conditioning system using propane mechanical subcooling cycle is theoretically analyzed in the automobile air condition, the effects of subcooling degrees and optimum sizing of the compressor’s combination are discussed, at the constant subcooling degree of 5 , the maximum COP and cooling capacity improvements are 12.97% and 17.89%, respectively. For a typical CO2 compressor with a displacement of 6cc that can meet the cooling load for a normal passenger car, a propane compressor displacement in the mechanical subcooling cycle between 6cc and 7.44 cc can meet the requirement of cooling capacity in the subcooler when the ambient temperature increases from 25 to 40 [95]. For the TES method, the subcooling system in Fig. 15 (a) normally uses a thermoelectric module based on the Peltier effect, when a DC current is put on both semiconductors, a temperature difference between them can be generated, then it can be used to subcool the CO2. This concept was first applied in the basic CO2 transcritical refrigeration plant by Schoenfield et al [96, 97]. They evaluated the effects of input current on the overall system cooling capacity and COP, a maximum increment of 3.3% and 7.9% were experimentally observed for low input current applied to the Bismuth-Telluride thermoelectric module, where the COP for the single thermoelectric module is highest. Other studies using the same method were theoretically found to have a COP enhancement from 3.3 to 37.8% [93, 98-101]. 3.6 Flash gas bypass The concept of the flash gas bypass was first put forward to overcome the challenge of two-phase flow distribution problem in evaporators, from headers to tubes [102, 103], which is an inherent problem for almost all refrigerants. For the CO2 application, many systems have employed small or micro-channel tubes in evaporators to bear the extremely high pressure, and the distribution problem becomes even more severe [14]. Moreover, in a conventional vapor compression system with direct expansion (DX), which is not refrigerant-specific, the 14

refrigerant state at the outlet of the expansion device is in a two-phase condition, resulting in the fact that the vapor state enters the evaporator without having a significant cooling effect. One possible approach to solve the above problems simultaneously is the implementation of a Flash Gas Bypass (FGB), the idea behind this method is to bypass the vapor state flow avoiding entering the evaporator and directly sucked by the compressor. Fig. 16 shows a schematic of the FGB approach and its integration in the CO2 system. The supercritical CO2 from the gas cooler is throttled into two-phase state and enters the flash tank, where the phase separation takes place. Then the separated vapor stream flows into the bypass circuit and is sucked directly by the compressor, while the separated liquid flow with a quality equal to zero feeds the evaporator inlet header. It should be noted that the bypass valve in Fig. 16 (b) plays an extremely important role for the control of evaporator outlet conditions, in that closing the bypass valve allows more refrigerant through the evaporator, thus reducing superheat at the evaporator exit [102]. Tuo and Hrnjak investigated the FGB scenario in an R134a mobile air conditioning system and demonstrated that the COP can be significantly improved by 37%-55% over the DX system at the same cooling capacity, they concluded that this is owing to the improved refrigerant distribution and reduction of refrigerant pressure drop in the evaporator. Fig. 17 shows their obtained evaporator surface temperature by infrared imaging which is indicative of refrigerant distribution among the microchannel tubes. It is obvious that a more uniform profile is obtained in the FGB mode since only liquid refrigerant enters the inlet header, essentially eliminating the quality maldistribution [104]. In addition, the control strategy of the flash gas bypass system and dynamic behavior during start-ups and transients were clearly defined and investigated experimentally by the same research group [105]. It is unfortunate that no such control experiments have been conducted for CO2 condition. For CO2 FGB system, various studies have focused on the system performance by experimental or thermodynamical methods. Wang et al. established a parametric model of the CO2 FGB system using engineering equation solver and found a 7% improvement in COP over the basic system [106]. An experimental comparison to a conventional DX-system by Elbel and Hrnjak reveals that FGB increases the cooling capacity and COP at the same time by up to 9 and 7%, respectively [107]. a b

Fig. 16. Evaporator with flash gas bypass configuration [102] (a), and its integration in the CO2 system (b).

15

Fig. 17. Evaporator surface infrared images in FGB and DX modes [104]. 3.7 Parallel compression The layout of parallel compression is similar to the FGB system discussed above, the difference relies on the compression process, where an additional auxiliary compressor is used to compress the flash gas to gas cooler pressure rather than using the same compressor that mainly to compress the vapor from the evaporator. The main intention of this concept is to reduce throttling losses [111]. Fig. 18. shows the depicted parallel compression system, apart from using two separated compressors, the dotted line means that the parallel compression can also be realized by coupled compressor, i.e. a dual compressor with a shared single shaft, a T-shape shaft, and parallel shafts, respectively [108, 109], or using a certain number of main compressor’s cylinders for parallel compression [110]. Another way of directly compressing the flash gas to gas cooler pressure is leading the gas through a vent into the compression chamber of the main compressor [116]. In all cases, the required compression power is reduced and furthermore, the compressor provides improved efficiency at smaller compression ratio leading to the fact that refrigerating systems using parallel compression achieve at least the same level of efficiency, or are more efficient than refrigerating systems using flash gas bypass [115]. Sarkar and Agrawal compared the transcritical CO2 system with parallel compression with the other two systems, the flash gas bypass system is included. They found the parallel compression cycle not only improves the optimum cooling COP but also brings down the optimum discharge pressure. The parallel compression system is especially profitable for lower temperature applications, achieving a 47.3% COP improvement over the basic system for the chosen ranges of operating conditions, the optimal intermediate pressure is reported to vary only marginally [112]. Bell [110] carried out a theoretical and experimental study on this system considering different gas cooler outlet temperatures, suction superheat, intermediate vapor pressure and compressor swept volume ratio, it is concluded that parallel compression is more beneficial with CO2 than hydrocarbon system in terms of COP and capacity under the same conditions. For the optimization of a wide range of gas cooler outlet temperature, the compressor swept volume ratio needs to change. Da Ros [113] investigated the optimization of parallel compression, obtained by tuning discharge pressure and displacement ratio of compressors, providing values for COP and discharge pressure of both compressors in a few operating conditions. Cecchinato et al. [111] presented an important theoretical parametric analysis of the system and it is observed that the ratio between the compressor displacements is an important influencing factor on the performance. There is a relation between the intermediate pressure and the displacement ratio. Moreover, the COP is a function of both maximum and intermediate pressure, 16

it is not possible to optimize the COP over a wide range of temperatures without varying compressors’ swept volume, which is consistent to the conclusion of Bell [110]. Chesi et al. [114] studied the parallel compression both theoretically and experimentally, comparing it with the CO2 basic system. With the thermodynamic model, they found that an ideal cycle may reach COP improvements of over 30% and over 65% in terms of cooling capacity. The experimental results confirm the expected trend but the absolute value is lower, resulting from the separator efficiency and pressure losses along the lines of the system. It was found that the efficiency of the separator decreases with increasing flow rates at its entrance. Fritschi et al. [115] compared the energy efficiency of parallel compression with FGB system, it was demonstrated that in specific cases, significant increases in efficiency at least 10% can be achieved, the conditions are as follows: The gas cooler outlet temperature > 27.5 (each supercritical operating condition); The evaporation temperature < -7 ; Possibility of realizing intermediate pressures of up to 45 bar. He et al. [109] compared CO2 parallel compression with basic cycle, effectiveness of parallel compression is most pronounced in low evaporating temperature and high ambient conditions, with up to 21% increase in COP and 5.3 bar reduction in discharge pressure over the considered parametric range, while improvement in COP is generally below 10% when the system operates in subcritical conditions.

Fig. 18. Schematic of the parallel compression system 3.8 Two-stage compression In the basic CO2 system, which is a single-stage compression cycle, minor exergy losses take place during an actual compression process, typically the isentropic efficiency of the compressor is a strong function to the compression ratio, which can be written as [121]: (2)

ηis = 1.003 − 0.121ε

this equation was obtained by fitting the experimental data of a CO2 compressor, where ε stands for compression ratio defined by the ratio of discharge and suction pressure. It is easy to expect that reducing the ε can raise the compressor’s efficiency, traditionally staged compression is an effective method to accomplish it by reducing the compression ratio for each stage. Of course, multistage compression more than two stages are more effective in reducing the exergy losses, but in actual application, multistage compression system is indeed seldom used owing to the complexity and installation cost. Moreover, if the compression ratio becomes extremely low, the reduction of isentropic efficiency will occur likewise [90]. Thus, very often for research and application, a two-stage compression system is extensively studied. Kim et al. [14] have presented a comprehensive review of two-stage transcritical CO2 systems based on the literature before 2004. Various studies indicated that apart from the heat rejection pressure, intermediate pressure and intermediate temperature are the other two critical parameters that determine the optimum COP in case of the two-stage CO2 refrigeration system, the three parameters are functionally coupled together that necessitates simultaneous optimization [117-120]. While usually the first step is 17

to determine the intercooling method to cool the vapor after the first compression [122]. However, up to now, there seems to be no perfect intercooling method with various scenarios between the two compressions. Fig. 19 presents four typical CO2 two-stage refrigeration systems with different intercooling scenarios, in which Fig. 19 (a) is the simplest system with an intercooler and uses one expansion valve and two-staged compressors. The discharge pressure of the first-stage compressor is cooled by an external fluid, usually the same for gas cooling, resulting in a significant increase in energy efficiency and reducing the compression work for both compressors, and a substantial reduction in the upper compressor’s discharge temperature. For this system, Srinivasan [123] found that equal discharge temperatures at the exit of each stage is an appropriate criterion for a substantial reduction in the maximum cycle temperatures. By defining an inter-stage pressure index as n [123]:

n=

ln Pi ln( P1s P2 h )

(3)

where Pi is the intermediate pressure, P1s is the suction pressure of the first-stage compressor and P2h is the discharge pressure of the second-stage compressor, he investigated the effects of evaporation temperatures on the optimum intermediate pressure and concluded that more compression needs to be done in the lower stage as the evaporator temperature decreases with increasing inter-stage pressure index. It validates the conclusion from Agrawal et al. [122] that

Pgc Pev , whose inter-stage pressure index is 0.5, will not yield the optimum pressure

ratio for each stage. Cecchinato et al. [124] theoretically compared the COP performance with the single-stage system at +4 , -10 and -30 evaporation temperature and 30 external cooling temperature and found about 3.1%, 10% and 20% COP enhancement for the three evaporation temperatures, respectively. It is apparent that this system is more competitive in low-temperature regions. Cavallini et al. [125] reported that up to 25% increase in the COP was found for typical air conditioning applications compared with basic CO2 system. Fig. 19 (b) is a double throttling and double compression cycle with a flash tank, the supercritical gas stream is throttled to the intermediate vapor-liquid mixture and flows into the flash tank, in which the two phases are separated, then the vapor is injected by the second compressor. The liquid phase is throttled again and enters the evaporator, the discharge vapor stream comes from the first compressor is mixed with the vapor phase from flash tank [90]. Cho et al. [126] experimentally investigated this system varying the amount of refrigerant charge, compressor frequencies and expansion valve openings and compared the results with the system in Fig. 19 (a). They found that the optimum charge of the flash gas injection system is higher than that of intercooler system because of lower compressor discharge pressure. The maximum cooling COP enhancement is measured to be 16.5% over that of the intercooler system. In addition, the discharge temperature of the second-stage compressor decreased by 5 to 7 . The COP increases with the increasing openings of the first- and second-stage expansion valves resulting from the decreased compression ratio. However, it makes some cost of the system cooling capacity as the mass flow rate through the evaporator decreased when the first-stage expansion valve opening was increased. Therefore, in this system, optimum control of both expansion valves openings is required. Aprea and Maiorino [129] carried out experimental research on this system for the optimization of heat rejection pressure varying with ambient temperatures. They presented a simplified model to predict the optimum heat rejection pressure validated by experimental results. The developed correction to predict the optimum pressure is written as [129]:

Popt =

2.7572 + 0.1304Te − 3.072 K / C 8.7946 + 0.02605Te − 105.48 K / C Tout , gc − 1 + 0.0538Te + 0.1606 K / C 1 + 0.05163Te + 0.2212 K / C

− 0.003Tout , gc + 0.174

(4)

where Popt is the optimum pressure, Te is evaporation temperature, Tgc,out is gas cooler outlet temperature, K and C 18

are two constants. a

b

c

d

Fig. 19. Schematic diagram of CO2 two-stage compression systems using different intercooling scenarios: (a) intercooler; (b) flash gas injection; (c) flash intercooling; (d) subcooler. Fig. 19 (c) shows a flash intercooling system, it is similar to the flash gas injection system in Fig. 19 (b), the difference is that the compressed vapor from the first-stage compressor enters the flash intercooler and desuperheated by evaporation of liquid CO2 with an intermediate temperature. This results in a double-effect that on the one hand, this process increases the CO2 vapor flow rate to be sucked by the second-stage compressor, on the other hand, the suction temperature for the second-stage compressor is significantly reduced compared with the flash gas injection system [2]. Zhang et al. [127] evaluated the performance of this system by thermodynamic analysis in comparison with the system in Fig. 19 (a). They found that the flash intercooling system performs a higher COP by 12.16% than that of the intercooler system. Agrawal et al. [128] presented a two-stage flash intercooling transcritical CO2 heat pump cycle and found that this method is not economical with CO2 refrigerant, unlike NH3 as the refrigerant. The system in Fig. 19 (d) replaces the flash tank with a heat exchanger named subcooler that used to cool the vapor stream before entering the expansion valve. Compared with the systems using a flash tank in Fig. 19 (b) and (c), the removal of the flash tank contributes a reduction in cost, but the temperature of the fluid entering the expansion valve before evaporator is a bit higher because there exists an approach temperature between the hot and cold fluid in the subcooler. The cycles with subcooler and the flash tank behave very similarly and the COP is 19

nearly the same in traditional subcritical applications, but for CO2 transcritical cycles, the isobaric curves in the region of dense gas or liquid do not strictly adhere to lower limit curve in the T-s diagram, so the split-cycle in general is not penalized in comparison with the open flash tank cycle [90]. Cecchinato et al. compared this system with the two systems in Fig. 19 (a) and (b) and the basic CO2 single-stage system by thermodynamic analysis. It results that the subcooler system outperforms all other options, showing 20.6% average increase in COP over the CO2 basic single-stage system at +4 evaporation temperature, 29.3% increase at -10 [124]. However, unfortunately to the best knowledge of the authors, there are no experimental results available for this CO2 two-stage system. 3.9 Evaporative precooling As mentioned in section 2, one main challenge for basic CO2 system is the intrinsic efficiency deterioration in high ambient temperature application. In other words, the CO2 gas cooler prefers inlet cooling medium with low temperature, in this sense, the evaporative cooling method to precool the air before entering the gas cooler may be a promising technology to improve the energy efficiency of the basic CO2 system. Traditionally, spraying water on the condenser is a well-known solution in refrigeration facing peak temperatures [130]. Although it shows good performance, it requires a large amount of water and leads to scaling and corrosion problems as the contact of water with fins, unless a high-cost surface treatment is applied. Considering the large difference between dry-bulb temperature (DBT) and the wet-bulb temperature (WBT) in summer for most regions in the world, evaporative cooling is nowadays more and more attractive for improving energy efficiency both by indirect and direct methods. The roots of evaporative cooling can be traced to Dr. Willis H. Carrier for his theory of adiabatic saturation and the standard psychrometric chart [131]. Fig. 20 shows a case of the process of the air precooling by a near adiabatic saturation [132].

Fig. 20. Process of the air precooling by a near adiabatic saturation [132]. The principle underlying is the easy conversion of sensible to latent heat in that nonsaturated air is cooled by exposure to the wetted pads. This process tends to progress until the air is saturated and the air temperature falls. The saturation efficiency of an evaporative cooler ηs, as expressed in Eq. (4), represents the ratio of the actual decrease in air temperature to the maximum possible one, i.e. the temperature difference between the dry bulb and wet bulb air [133]. More details on the evaporative cooler can be seen in references [134-138], hereafter we focus on its application in the basic CO2 system.

ηs =

tdry − tair ,out

(5)

tdry − t wet

One trade-off needs to consider is the balance between water consumption and COP improvement for the practical application in the CO2 system. Fig. 21 shows two possible arrangements, where 100% (a) or only 33% (b) 20

of the gas cooler inlet air is precooled. Girotto et al. [139, 140] analyzed the two solutions and found that the COP improvement is 17% for the 30% precooling solution and 27% for the 100% precooling solution. The water consumption is reduced by 70% by replacing the 100% precooling solution to the 30% precooling one. Table 1 summarises their results. Thus, the best option must be chosen with reference to the actual climate of the area where the gas cooler will be installed and water availability is to be considered [90]. a b

Fig. 21. Application of evaporative cooler in the basic CO2 system with 100% air precooled (a), and 30% air precooled (b). Table. 1 COP and water consumption of two different pre-cooling solutions, for 40°C and 50% relative humidity external conditions [140]. Solutions

Air temperature after evaporative cooling [°C]

CO2 temperature out of the gas cooler [°C]

Optimal high pressure [bar]

COP [-]

COP (with fan power input) [-]

Water consumption [(m3 /h)/kW]

No precooling 100% precooling 30% precooling

40.0 32.2 32.2

41.5 34.7 35.6

116 100 114

1.38 1.75 1.6

1.38 1.71 1.59

0.000926 0.000291

3.10 CO2-based mixture One of the challenges of the above-mentioned technologies to improve the basic system is the increase in system complexity, they all require additional equipment to achieve the target, which definitely necessitates additional costs and increased installation space, thus decreases the economical efficiency. In such case, some CO2-based mixtures including refrigerant blends and CO2-absorption cofluids, which are all based on CO2, have become a promising drop-in technology without any change to the basic system configuration. 3.10.1 CO2-based refrigerant blends Since the two inevitable problems caused by using CO2 as a refrigerant in basic system, as safety issues arising from high pressure and operation efficiency, one can easily expect that adding another refrigerant to CO2 could improve system efficiency and reduce operating pressure, of course, the adding refrigerant must meet the requirements of low operating pressure, high efficiency and low GWP. According to the temperature glide occurs during the evaporation or condensation process, the refrigerant blends can be classified into zeotropes, 21

near-azeotropes and azeotropes based on the difference of normal boiling temperature for each refrigerant. Taking the mixture of CO2/R41 for an example, in Fig. 22, the variations in the temperature and composition of CO2/R41 mixtures at fixed pressures constitute the bubble line and dew line, the two lines divide the diagram into three regions: the subcooled liquid, superheated vapor, and two-phase regions. The dew point temperature of the mixed refrigerant with a certain R41 mass fraction is slightly higher than the bubble point temperature. A temperature glide occurs simply because of this factor during the evaporation or condensation process. For fixed pressures of 3, 4, and 5 MPa, the maximum temperature glides are 1.75 °C, 1.55 °C, and 1.25 °C, respectively, which are slightly higher than azeotropic refrigerants without temperature glide, but substantially smaller than those of zeotropic mixtures [141].

Fig. 22. Temperature-composition diagram of CO2/R41 blends at fixed pressure [142]. In the earlier stages, the zeotropic mixtures were originally studied in the autocasacde refrigeration system showing the merit of low operating pressure and a small amount of charge [145]. Then they were studied in the conventional vapor compression system, among which propane seems to be the most popular additive. Propane has a better refrigerating effect and a much lower pressure than CO2, moreover, its flammability can be restrained by CO2. In previous research [141], an automobile air-conditioning system using CO2-propane mixture as a refrigerant was designed and experimentally investigated, a maximum COP improvement of 22% was achieved and the discharge pressure was substantially reduced. Kim et al. [143] experimentally evaluated the performance potentials of CO2-propane mixtures in a water-cooled system. The results demonstrated that the COP of CO2-propane mixture with a mass fraction of 60%/40% was 12.5% higher than that of CO2. Ju [144] found that CO2-propane with a mass fraction of 12%/88% was the most suitable natural substitute for R22 in a heat pump water heater, the optimum heating COP and capacity of which were 11.00% and 17.50% higher than those of the R22, respectively. However, it is also reported that zeotropic mixtures exhibit the problem of composition shifts and are sensitive to the leaking or recharging process from the application point [146,147]. Therefore, near-azeotropic refrigerants with little temperature glide provide another option, the properties of R41 are similar to those of CO2, especially the almost same normal boiling point. As illustrated in Fig. 22, the temperature glide of this kind of mixture among the phase change process is extremely small, making it more applicable than zeotropic mixtures. It has been demonstrated that azeotropic CO2/R41 was the most suitable candidate owing to its high COP and low high-side pressure simultaneously evaluated with other 10 types of low-GWP CO2-based mixtures in a heat pump water heater [148]. 22

Wang et al. [149] theoretically evaluated a CO2/R41 mixture refrigerant compared to pure CO2 and found that the CO2/R41 mixture could effectively reduce the irreversible loss of each component, increase the exergy efficiency of pure CO2 systems, reduce the optimal high pressure by 28.62% maximally, and achieve a maximum COP improvement of 20.52%. It is deserving to study further on the heat transfer properties in the heat exchangers and compressor optimization for the system using CO2-based mixtures, as the drop-in system has some improvement potential regarding the heat transfer and compressor performance [142]. 3.10.2 CO2-absorption cofluids In the basic transcritical CO2 system, the transcritical operation principle acts a much higher operating pressure than conventional refrigerants, leading to much higher cost and poor reliability of the system. Prior work has tried to eliminate this problem and simultaneously improve system efficiency by replacing the evaporation and gas cooling with absorption into and desorption out of an absorbing fluid. Thereby the CO2-absorption cofluid circulates in the CO2 absorption-compression refrigeration system shown in Fig. 23, this cycle can be regarded as an extreme case of a zeotropic refrigerant cycle for the reason of totally non-volatile absorbing fluid rather than a refrigerant with a different normal boiling temperature to CO2. In Fig. 23, the evaporator and gas cooler are replaced by the absorber and desorber, the two-phase flow of the CO2-absorbing cofluid is compressed in the compressor known as a “wet compression”. In the absorber, the mixture is cooled by ambient and vapor CO2 is absorbed into the cofluid. Then the mixture expands during the throttling process and CO2 begins to desorb. In the desorber, the expanded low-temperature mixture is heated by the heat source and more CO2 is extracted into the vapor phase. Finally, the two-phase CO2-absorbing cofluid flows into the compressor to start another new cycle. Due to the absorbing fluid absorbs CO2 in the absorber, the discharge pressure is significantly decreased to no more than 35 bar so that the existing components for conventional-refrigerant based vapor compression systems can be adopted directly [150].

Fig. 23. Schematic diagram of the CO2 absorption-compression refrigeration system. In the earlier stage, from the 1990s to 2000s, the adding cofluids in this concept were mainly acetone [151] and a range of organic liquids [152-156, 166, 167] based on physical absorption mechanism, which can be written as [157]: abs (6)  → CO2(phys) CO2(vap) ← des

However, several concerns of these physical absorbents have prevented their extensive study and application in the past decade. First, the COP and cooling capacity are substantially lower than that of conventional refrigerants 23

like R134a [150, 152]. This defect is attributed to the limitation of CO2 absorption capacities of the physical absorbents, because the system COP is a strong function of the molar fraction of the partial molar enthalpy of CO2 absorption [158], for common organic liquids, this absorption enthalpy is around -12 kJ mol-1 [158]. Second, for organic solvents, the dilation is commonly observed during gas dissolution, resulting in significant volume expansion, Aki et al. found that when 40% molar fraction of CO2 was dissolved, the volumes increased by 57% and 33% for acetonitrile and ethylacetate, respectively [159]. Other fetal defects are the corrosion, thermally instability and volatility, which can corrode alloy steel pipes and other components, leat to the loss of absorbents and can be lost into the CO2vapor. Owing to the drawbacks of conventional absorbing fluids, recently ionic liquids (ILs) are presented to be promising solutions. Compared with conventional fluids, ILs are mostly thermally and chemically stable, they are non-volatile with negligible vapor pressure, it can also serve as a lubricant substitute owing to good lubrication properties [164]. Furthermore, the most important advantage of ILs is that their physicochemical properties can be adjusted according to their target application by tuning the structure of the ions or adding functional groups [160,161]. When it comes to CO2 absorption, Blanchard et al. reported that supercritical CO2 is highly soluble in ILs, whereas ILs do not dissolve in supercritical CO2 [165]. ILs can be classified to conventional Ionic Liquids that follow a physical absorption mechanism and functionalized Ionic Liquids that allow chemical binding of CO2 to the functional groups. A major concern of conventional ILs is that their CO2 absorption capacity is not competitive to functional ILs, contributing to low COP, though it is higher than that of organic solvents [162]. For functionalized ILs, the CO2 bind favorably to the functional group, such as carboxylate functionality and amine functionality, the chemical absorption mechanism can be written as [157]: abs (7)  → ACO2(chem) CO2(vap)+A(l) ← des

This concept is not new, it has been widely used for CO2 capture employing aqueous amines or ammonia solutions, in which intensive energy is required to break the chemical bonds between the absorbents and the absorbed CO2 in the regeneration step (∆H=−80 to −64 kJ mol-1 (bicarbonate formation) or -101 kJ mol-1 (carbonate formation)), which represents a high operation expense due to the energy cost related to the desorption step [160]. But for ILs, typically absorption enthalpies are on the order of −50 kJ mol-1 or less [163]. This suggests that, compared with standard amine solutions, only half of the energy is required to remove the same amount of CO2 from ILs. Based on the chemical CO2 absorption by using ILs, Mozurkewich et al. investigated the CO2-ILs cofluids in a refrigeration cycle and they found that the chemically absorbing cofluids with strong affinity for CO2 performs higher specific cooling capacity and higher COP than physical cofluids, compared with transcritical CO2 system, the studied system also has a higher COP and much lower operation pressure [157]. 4. Summary of previous studies According to the above main findings associated with every single technology, the energy efficiency improvements of the modified technologies that have been reviewed above over the basic CO2 transcritical cycle are summarized in Fig. 24. Also, in Table. 2, the SWOT (Strength, Weakness, Opportunity and Threat) analysis for each technology are shown.

24

Fig. 24. Summary of the energy efficiency improvements for the modified technologies. Table.2 SWOT (Strength, Weakness, Opportunity and Threat) analysis SWOT technology

Strength

Weakness

IHX

Simple apparatus, good compatibility with other technologies Compact, high efficiency

Limited COP Necessary for all CO2 improvement refrigeration application

Leading to compressor temperature

Expansive

Leading to the circulation problem

Vortex tube

Simple structure

Expander

High expansion loss recovery, good compatibility with other technologies Simultaneously improving the cooling capacity and COP Without the problem of two-phase flow distribution in evaporators

No practical application A large machine (like another compressor)

Ejector

Subcooling

Flash bypass

Parallel compression

gas

High performance

Opportunity

Threat

Commercial refrigeration, mobile air conditioning Potential industry refrigeration Applicable in relatively large systems

higher discharge

oil

Requiring high separation efficiency High-speed operation machine, needing good maintenance

Another set of system needed

Commercial refrigeration, industry refrigeration

Too high subcooling degree adverse instead

N/A

Potential for applications

Control strategy and understanding the dynamic behavior of the bypass flow are essential for system performance It is not possible to optimize the COP over a

COP Another compressor

Commercial refrigeration 25

all

needed

Two-stage compression Evaporative precooling CO2-based mixture

Reducing compression for each stage Easy

the ratio

No extra equipment, drop-in

Relatively complicated

Applicable in relatively large systems Potential for all applications

Water consumption Some flammable additive

Potential for applications

all

wide range of temperatures without varying compressors’ swept volume Requiring appropriate inter-cooling method the balance between water consumption and COP improvement Requiring appropriate additive to guarantee performance

5. Conclusion and Perspectives This review summarizes recent advances in modified technologies for the transcritical CO2 refrigeration cycle during the last two decades, these technologies are found to be promising for the optimization of energy efficiency and operating pressure. This study is expected to help in identifying the appropriate solution for CO2 transcritical refrigeration considering the different applications. The main conclusions and perspectives for the future are as follow:
At this moment, studies on IHX is the most comprehensive, there is no doubt to use it in transcritical CO2 refrigeration;
Various drawbacks for ejector need to be overcome, such as high cost and considerable complexity, the complex internal flow conditions need further visualization based on advanced methods as well as CFD model. The shock wave pattern inside the nozzle expansion section and the initial location of choked flow for the real fluid should be understood well and conclusive in the future.
Primary experimental validation should be conducted on vortex tube, thermoelectric subcooling and CO2-absorption liquids to figure out their effect on CO2 cooling performance and energy efficiency;
The main challenge for the expander, subcooling, flash gas bypass and two-stage compression is their system complexity, the final choice therefore results from the best trade-off between increased installation costs and decreased operating costs, taking into account also the reliability and the safety characteristics of the system;
A detailed techno-economical analysis is required to estimate the life-cycle cost among different technologies;
Although there has been significant progress in understanding the CO2 dissolution behaviors and in improving the CO2 solubility, other important properties (e.g.viscosity, absorption and desorption rate, heat capacity) should be given more consideration to search for better cofluids;
More realistic studies are needed for the mass and heat transfer and the behavior of hardware components, as well as consideration of a broader range of environmental conditions for each optimized system. It should be noted that this review only covers the modified systems using a single technology mentioned above, the results of improvement potential by two or more technologies integrated simultaneously in one system are not included, though it is very often seen in the practical application level. Researchers can freely study the potential combinations with the technologies mentioned above and certificate their applicability in the future. Acknowledgment This project was supported by the National Natural Science Foundation of China (NO.51776119).

26

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Highlights  The principles of the CO2 refrigeration cycle and performance are discussed  A detailed discussion on different modified technologies is provided  Recent progress on the energy efficiency improvement is summarized  Perspectives on the future development, research opportunities are presented

Declaration of interests ☒ The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. ☐The authors declare the following financial interests/personal relationships which may be considered as potential competing interests: