R600a

Accepted Manuscript Title: Performance evaluation of a vapor injection refrigeration system using mixture refrigerant r290/r600a Author: José Vicente Hallak d'Angelo, Vikrant Aute, Reinhard Radermacher PII: DOI: Reference:

S0140-7007(16)00025-6 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.01.019 JIJR 3248

To appear in:

International Journal of Refrigeration

Received date: Revised date: Accepted date:

4-12-2015 5-1-2016 21-1-2016

Please cite this article as: José Vicente Hallak d'Angelo, Vikrant Aute, Reinhard Radermacher, Performance evaluation of a vapor injection refrigeration system using mixture refrigerant r290/r600a, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.01.019. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. 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.

Performance Evaluation of a Vapor Injection Refrigeration System Using Mixture Refrigerant R290/R600a

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Performance Evaluation of a Vapor Injection Refrigeration System Using Mixture Refrigerant R290/R600a José Vicente Hallak d’Angeloa, Vikrant Auteb, Reinhard Radermacherb a

Department of Chemical Systems Engineering, School of Chemical Engineering, University of Campinas – Av. Albert Einstein, 500, Bloco A, Campinas – SP, CEP: 13083-852, Brazil. b

Center for Environmental Energy Engineering, University of Maryland, College Park, 4164 Glenn L. Martin Hall Bldg., MD 20742, USA.

Highlights:  A vapor injection refrigeration cycle using mixture R290/R600a was studied.  Two different cases related to the choice of reference temperatures were evaluated.  A parametric analysis of the main operating variables of the cycle was performed.  Best performance was achieved between 40-50 wt% R290 and expansion ratio of 50%.

ABSTRACT This work presents a performance evaluation of a vapor injection refrigeration system using a mixture refrigerant R290/R600a, through steady-state simulations used to accomplish a parametric analysis considering the influence of the refrigerant composition over the following parameters: COP; compressor power; refrigerant mass flow rate; refrigerant temperature glide; mass flow ratio between vapor and feed streams in the flash tank; liquid and vapor composition of flash tank outlet streams and compression ratio. Two cases, denominated A and B, considering different fixed temperatures at the refrigeration system were studied and their performances were compared with the one of a basic vapor compression cycle. A maximum COP was obtained for a mixture containing 40 wt% of R290. COP of vapor injection refrigeration cycle is 16-32% greater than the one of a vapor compression cycle, depending on the composition of the mixture refrigerant and pressure drop at the cycle upper-stage expansion valve.

Keywords: simulation; mixture refrigerant; vapor injection; R290; R600a; hydrocarbon.

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Corresponding author: José Vicente Hallak d’Angelo, [email protected]

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1. Introduction Cold utilities are essential in different stages of many industrial chemical processes and they are provided by large scale refrigeration systems which are great energy consumers, presenting a significant impact on the final cost of the product, so it is very important that these refrigeration systems operate in an optimized way in order to reduce production costs. Recent studies in the area of refrigeration cycles, with regard to the search for more efficient systems, have concentrated their efforts mainly in two specific areas: the development of new technologies, using alternative cycles to obtain a better performance and the use of new refrigerants, pure or mixtures, aiming both energetic and environmental aspects. Different alternative cycles have been studied and there are many promising options to increase the thermodynamic efficiency when compared to the traditional vapor compression cycle. Currently, some of the most explored alternative refrigeration cycles are: ejector refrigeration, cascade and refrigerant injection systems. The search for alternative refrigerants is based on two main prerogatives: provide a better thermodynamic performance of the refrigeration cycle and reduce the environmental impact. ODP (ozone depleting potential) and GWP (global warming potential) are the most commonly used environmental metrics and represent a direct impact of the refrigerant. TEWI (total equivalent warming impact) and LCCP (life cycle climate performance) are additional metrics that help to evaluate the impact of any system or process which requires energy input and indirectly affects the environment (Makhnatcha and Khodabandeha, 2014). The use of mixed refrigerants (with two or more components) provides an opportunity to adjust some desirable properties to obtain a final fluid for some specific application with acceptable environmental metrics. The necessity of using new refrigerants arose mainly due to international environmental treaties, like Montreal Protocol (1987) and Kyoto Protocol (1997), which forced the phasing out of the production of numerous substances that are responsible for ozone depletion and for the greenhouse effect, causing a global warming. This included refrigerant CFCs and HCFCs, both of which have high ODP and GWP. Then, in substitution of these refrigerants, some hydrofluorocarbon (HFC), natural refrigerants (like hydrocarbons, HC) and most recently hydrofluoroolefin (HFO) have been tested because of their shorter atmospheric lifetimes when compared to CFCs, being largely destroyed in the lower atmosphere by reaction with OH radicals (Wallington et al., 2010). This work aims at a performance evaluation of vapor injection refrigeration system with a flash tank using mixture refrigerant R290/R600a developing a parametric analysis of the main process operating variables of this system, as a function of the mass composition of the mixture refrigerant and the expansion ratio at the upper-stage valve. This parametric analysis is based on results obtained from steady state computer simulations of the system. The pair R290/R600a was chosen for being a non-azeotropic refrigerant mixture and because of their low-GWP as discussed further. 4 Page 4 of 34

2. Literature review A detailed review of all types of alternative refrigeration cycles and alternative refrigerants is not the scope of this work, but as a manner to justify the objectives presented in the previous item, a literature review of recent works in these fields is presented here. 2.1 Alternative refrigeration cycles Considering ejectors systems, Chunnanond and Aphornratana (2004) presented a literature review on ejectors and their applications in refrigeration technology; Sumeru et al. (2012) performed a review on two-phase ejector as an expansion device in vapor compression cycle; Chen et al. (2013) presented a study about some recent developments in ejector refrigeration technologies and Chen et al. (2015) wrote a review on versatile ejector applications in refrigeration systems. Many recent works present experimental and/or simulations studies involving ejectors refrigeration systems, also considering exergetic and energetic analyses and process optimization. Referring to cascade refrigeration systems, Tan et al. (2015) have combined an auto-cascade with an ejector using mixed refrigerant, developing a novel cycle. Cimsit and Ozturk (2012) presented an analysis of compression-absorption cascade refrigeration cycles and later (Cimsit et al., 2015) they presented a detailed exergy-based thermoeconomic analysis and optimization of a LiBr/H2OR134a compression-absorption cascade refrigeration. Many recent works have studied a combined vapor compression-absorption cascaded refrigeration system, applied to some different processes, like: production of liquefied natural gas (LNG); water chilling; pre-cooling of LNG production and cycles with very low evaporating temperatures, around -170 oC, which can be used for many different applications. The technique of refrigerant injection consists basically in injecting the refrigerant from the condenser outlet to the suction line of the compressor or to its sealed pocket. Xu et al. (2011) presented a review of refrigerant injection for heat pumping and air conditioning systems, discussing the challenges of this technique, which can be classified into two types: liquid or vapor refrigerant injection. Vapor injection is used to improve the cooling/heating capacity keeping the same stroke volume of the compressor (Xu et al., 2011). Considering a system with vapor refrigerant injection, there are typically two different configurations used: vapor injection with a flash tank (FTVI) and vapor injection with an internal heat exchanger (IHX). Wang et al. (2009a) have studied both configurations using R410a and concluded that they have shown a comparable performance improvement, but Mathison et al. (2011) stated that this will happen only when the heat exchanger has an effectiveness of 100%. So vapor injection refrigeration systems with a flash tank offer a greater potential for performance improvements, but the IHX configuration is easier to control. Ma and Zhao (2008) have proved experimentally that the FTVI has a better performance with a COP 4.3% higher than the one for IHX configuration (the authors didn’t present an analysis about the uncertainties of this value). Lee et al. (2013) have presented a study about the potential benefits of saturation cycle with twophase refrigerant injection, performing simulations for single and multi-stages systems, using 12 pure refrigerants and 2 refrigerant mixtures, concluding that there is a significant improvement in 5 Page 5 of 34

the cooling and heating capacities and a reduction in the compressor work, when comparing with the basic vapor compression cycle. Many other authors have studied vapor injection refrigeration systems with a flash tank, using experimental and/or theoretical approaches, for both cooling and heating applications. Table 1 presents a summary of some studies considering this kind of system only for cooling purposes. Many other works using this system for heat pump applications can be found in the literature as well as a combination of all alternative cycles mentioned previously. All studies in Table 1 were developed for steady state conditions. 2.2. Alternative refrigerants Calm (2008) presented a review of the next generation of refrigerants evaluating the progression of refrigerants, shown by Fig. 1. Sarbu (2014) presented a review on substitution strategy of nonecological refrigerants used in vapor compression systems for both air-conditioning and heat pump; Hammad and Alsaad (1999) presented an application of hydrocarbon mixtures for domestic refrigerators, which was also discussed by Granryd (2001) in an overview of hydrocarbons as refrigerants. Considering the main refrigerants used in industrial, commercial and domestic applications, the sequence of substitutions may be summarized as: R12 was replaced by R22 or R134a, then some studies to replace these refrigerants using for example R410a (which is a HFC zeotropic mixture of R32 and R125) have been performed, and more recently the use of some natural refrigerants (R717-NH3, R744-CO2, R290-C3H8, R600-C4H10 and R600a-iC4H10) and hydrofluoroolefins – HFO’s (e.g. R1234yf), because of their low GWP and zero ODP. In the literature there are many papers discussing the replacement of refrigerants, performing both experimental and theoretical studies. A full review of these works is out of the scope of this paper, but considering the recent literature, there are many works dealing with the mixture refrigerant R290/R600a, which is a natural non-azeotropic refrigerant mixture (NARM). Kruse (1981) discussed the advantages of using NARM’s in heat pumps in comparison with conventional refrigerants, pointing out their advantages as: energetic improvements by using sliding temperatures during condensing and evaporation processes in counter-flow heat exchangers; extension of the application limits offering great adaptability regard to capacity, pressure and temperature levels; continuous capacity control by changing the mixture composition and application in a cascade cycle for the large temperature difference between heat source and heat sink (Miyara et al., 1992). The mixture R290/R600a can be used to replace R134a and has zero ODP and a very low GWP for a wide range of composition (Calm, 2008). Seeking for new refrigerants or mixed refrigerants, after the establishment of the environmental treaties mentioned previously, hydrocarbons have gained some attention and the pair R290/R600a has been studied intensively by many researchers. Referring specifically to works that have studied this mixture refrigerant, a summary of some recent (considering the last 10 years) papers found in the open literature is presented in Table 2. 6 Page 6 of 34

The only work found by the authors in the open literature, considering the combination of a vapor injection refrigeration system with a flash tank and the use of mixture refrigerant R290/R600a was the one of Swinney et al. (1998), showing that there is a lack in the literature about studies involving this system. HC blends, like the mixture refrigerant R290/R600a, also offers some disadvantages because of their problem of flammability and limitation in the charge quantity due to safety (fire hazard) regulations (Sekhar and Lal, 2005). 3. Methodology 3.1 Cycle description The schematic diagram of the cycle is presented in Fig. 2 and a short description is as follows (numbers in parentheses refer to process streams): mixed refrigerant as saturated vapor (#1) is injected in the first stage of the compressor at low pressure and is compressed isentropically until an intermediate pressure, producing a stream of superheated vapor (#2) that is at the same pressure of the vapor stream coming from the flash tank (#9). Stream #9 is injected in the compressor and mixed with stream #2 in an inner chamber, generating stream #3, which is compressed in a second stage of the compressor, increasing its pressure even more, generating the compressor discharge stream (#4) that flows through the condenser. A saturated liquid stream (#5) leaves the condenser and feeds the upper-stage expansion valve. At the outlet of this valve, a two-phase stream (#6) is formed and then the phases are completely separated in the flash tank, generating a vapor stream (#9) and a liquid stream (#7) which are in equilibrium at the same pressure and temperature. The liquid phase (#7) goes through the lower-stage expansion valve, generating another two-phase stream (#8) that circulates through the evaporator, leaving as saturated vapor (#1), closing the cycle. Due to the use of the flash tank, the liquid that enters the evaporator will have a lower enthalpy when compared with a single-stage cycle, causing a higher enthalpy variation in the evaporator, reducing refrigerant mass flow rate, when considering a fixed cooling capacity. A qualitative diagram considering only mixture refrigerant pressure and enthalpy (P-h) values is also presented in Figure 2 to illustrate physical states assumed by this mixture refrigerant during the cycle. A basic vapor compression cycle (VCC) with four components (condenser, expansion valve, evaporator and compressor) was also simulated in order to compare the results obtained with the ones from the vapor injection cycle. Considering Fig. 2, the vapor compression cycle presents only streams 1, 4, 5 and 8. 3.2 Refrigerant temperature glide McLinden and Radermacher (1987) have presented some methods for comparing the performance of pure and mixed refrigerants in vapor compression cycles. As they have mentioned in their work, for mixture refrigerants, there is an additional degree of freedom, the composition of the system, which causes a change of saturation temperature during constant pressure evaporation and condensation processes. The refrigerant temperature glide (RTG) is defined as the total temperature difference between the onset of the phase change and its completion and a large RTG can improve the COP of mixtures when compared to pure fluids. Högberg et al. (1993) also presented some calculation methods for comparing the performance of pure and mixed 7 Page 7 of 34

working fluids, considering the possibility of using the external conditions, i.e. the application considered. To analyze the performance of refrigeration cycles using mixture refrigerants, considering only two fixed characteristic refrigerant temperatures, there are different possible choices, but the main ones are (McLinden and Radermacher, 1987): 1 – evaporator inlet and condenser saturated vapor temperatures (Case A), 2 – evaporator and condenser outlet temperatures (Case B). Depending on the choice made, a different behavior of the COP as a function of composition can be observed. For Case A it shows a curve with a maximum COP and for case B, one with a minimum. For example, for the mixture R290/R600a, He et al. (2014) presented a curve of COP for a refrigeration cycle considering Case A and Swinney et al. (1998) one for Case B. Therefore, it is of great importance to know how such choices influence the analysis of cycle’s performance. It seems that for Case B there would be no sense in discussing the use of mixed refrigerants since any composition between the pure fluids leads to a COP that is lower than the ones obtained with the pure components, but it is important to analyze some other output variables as well and not only the COP. So, for the sake of comparison, in this work both cases are analyzed. The values of the two temperatures chosen, for both cases A and B, were defined based on the ASHRAE standard single point condition used for rating refrigerator/freezer compressors which are: 249.85 K (-23.3 oC) for evaporating temperature and 327.55 K (54.4 oC) for condensing temperature, considering the existence of a great number of studies that have used these temperatures and because of their common use in standard tests performed by compressors manufacturers, especially the ones which are suitable for industrial applications. Table 3 shows the application of these choices of reference temperatures for each one of the cases mentioned and how the refrigerant temperature glide is calculated. Cases A and B were also applied to a basic vapor compression cycle used to compare COP values with the ones from FTVI cycles. 3.3 Modeling A simulation representing the cycle described in item 3.1 was developed using process simulator Aspen Hysys® version 7.3 (Aspentech, 2015). The thermodynamic package chosen was the PengRobinson equation of state (PR-EOS), which is specially indicated for hydrocarbons (Peng and Robinson, 1976). The binary interaction parameter (kij) was estimated internally by the process simulator and it is equal to 0.01041 and the mixing rules used are the standard ones for PR-EOS, i.e.,

n

a m ix 

n

 i 1

j 1

xi x j  ai a j

  1  k  and 0 .5

ij

n

b m ix 



x i bi

, where i and j are the pure components of the

i 1

mixture constituted of n components, xi and xj are the mole fraction of components i, j and ai ,bi are the PR-EOS parameters for pure components.

This software was chosen because of its reliability and potential to evaluate the performance of complex industrial processes, allowing future optimization of these processes using some software 8 Page 8 of 34

built-in tools, so the simulation of the refrigeration cycle studied in this work can be coupled with an industrial process simulation to produce cold utilities for the process, for example: cold water for distillation columns condensers, cold thermal fluids for jacked chemical reactors and cold fluids for other heat exchangers in the process. The following assumptions were adopted for the model and applied to both refrigeration cycles (FTVI and VCC):  the system is operating at steady state, no fluctuations of any operating variable is considered;  pressure drop is neglected in heat exchangers and also in the flash tank;  all equipment are adiabatic, including the flash tank, with no heat losses to the environment;  condenser outlet stream is saturated liquid (for all the composition range of the refrigerant);  evaporator outlet stream is saturated vapor (for all the composition range of the refrigerant);  streams (vapor and liquid) leaving the flash tank are in equilibrium at the same pressure and temperature;  a thermal load of 1 kW was assumed to be the cooling capacity of the evaporator, in order to allow easier scale changes, depending on the application;  compressor (both stages) are considered isentropic;  valves are isenthalpic. To perform a parametric analysis of the refrigeration cycles studied through the simulation, two basic variables were chosen as inputs:  expansion ratio (ER) in the upper-stage valve, which generates a two-phase stream (#6) that is separated in the flash tank. This expansion ratio is defined by Eq. (1) and three different levels of expansion were studied: 30%, 50% and 70%; ER %



 P5  P6      100 P5  

(1)

 mixture refrigerant composition, in wt%, in the condenser outlet stream (#5). The entire range of composition, from 0 to 100% of R290 was evaluated in intervals of 10%. With these two variables as inputs, a fixed cooling capacity (Qevap) of 1 kW, fixed reference temperatures (item 3.2) and the previous assumptions, a parametric analysis of the refrigeration system was performed using the model developed, evaluating the influence of the inputs over the following output variables: COP, required compressor power (both stages); mass flow ratio between vapor and feed streams in the flash tank (V/F); mass composition of outlet streams from flash tank (xi and yi) ; refrigerant mass flow rate at stream #5 (m5); evaporator and condenser refrigerant temperature glide (RTG) and compression ratio (CR) in the compressor. These variables were calculated by the following equations: Q evap

COP  W

S1

W

(2) S2

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Q e v a p ( k W )  m 1  h1  h 8

W W

S1

S2



(3)

 kW 

 m 1  h 2  h1 

(4)

 kW 

 m 3  h 4  h3 

(5) (6)

m3  m2  m9 W

to ta l

W

V F



S1

W

(7)

S2

m9

(8)

m6

m vapor

q u a lity ( m a s s f r a c tio n ) 

(9)

m to ta l xi 

yi 

m ass of com ponent i

(10)

to ta l m a s s o f liq u id s tr e a m m ass of com ponent i

(11)

to ta l m a s s o f v a p o r s tr e a m R T G  T s a t _ v a p  T s a t _ liq

CR 

Pd is c h a rg e (# 4 )

(12) (13)

Ps u c tio n (# 1 )

3.4 Model validation Before validating the model/simulation using literature data, some thermodynamic properties of the simulator were checked comparing them with values obtained from Refprop © version 9.1 (Lemmon et al., 2013), which is a very traditional and reliable data bank for refrigerant properties. Table 4 presents the values of molar mass (MM), normal boiling point temperature (Tb / K), critical temperature (Tc / K), critical pressure (Pc / MPa) and acentric factor () for both pure R290 and R600a. Saturation pressure, which is an essential thermodynamic property, was also checked comparing once again values from Refprop© 9.1 and Aspen Hysys® v. 7.3. Deviations between these values were calculated by Eq. (14) and in the range from 243.15 K to 333.15 K, which was the one used in the parametric analysis, the average deviation was 0.41 % for R290 and 0.85% for R600a.

D e v ia tio n  %





 abs  REF  AH  REF 



  100 

(14)

REF being the value obtained from Refprop and AH from Aspen Hysys. Figure 3 shows a plot of saturation pressure against temperature, considering pure fluids and also some mixtures with different compositions (90/10 and 50/50 wt% R290/R600a). 10 Page 10 of 34

The thermodynamic package reliability to estimate phase equilibrium composition is a crucial point since significant errors in estimating the compositions of the flash tank outlet streams may compromise the results of the simulations. A set of experimental data taken from literature (Lim et al., 2004) was used and compared with estimated vapor-liquid equilibrium data from the simulator. The experimental uncertainty of the mole fraction by Lim et al. (2004) was  0.002. When comparing the absolute deviation between experimental and calculated data using the simulator, about 35% of calculated equilibrium data for the liquid phase and 25% for the vapor phase were inside the experimental error declared. The average absolute deviation for the mole fraction of R290 between experimental and simulated data was 0.006 mole for both phases, and the average relative deviation was 1.5% for liquid phase and 1.2% for the vapor phase. These errors were considered satisfactory for the purpose of a parametric analysis. A paper with exactly the same system studied in this work, a vapor injection refrigeration cycle with flash tank using mixture refrigerant R290/R600a, was not found by the authors in the open literature. Because of this, a model validation procedure was performed comparing the current model with some similar cycles taken from literature: Swinney et al. (1998) performed a simulation of a refrigeration cycle using the mixture R290/R600a and a pool-boiling system (which for practical purposes can be considered quite similar to an FTVI system) and He et al. (2014) have studied the use of mixture refrigerant R290/R600a in a large capacity chest freezer. Swinney et al. (1998) considers the following operating conditions: compressor polytropic efficiency is 75%; outlet stream from condenser is fixed as saturated liquid at 313.15 K (40 oC); pressure drop in heat exchangers and mixture enthalpy are neglected; outlet stream from evaporator is fixed as saturated vapor at 248.15 K (-25 oC) and a hypothetical thermal load of 2000 kW in the evaporator was adopted. Fig. 4 presents the values of COP, defined as the ratio between heat load in the evaporator (2000 kW) and power input in the compressor, for both literature and simulated data.

The differences observed in Fig. 4 are basically due to two reasons: the reference literature has used another thermodynamic fluid package (SRK equation of state) and the values were presented only in the form of graphics, so there is probably a reading error. Nevertheless, the comparison allows checking that the same COP behavior was observed and that the values are very close. In Swinney et al. (1998) temperatures at condenser and evaporator outlet streams were fixed, leading to a minimum COP behavior. A comparison between simulated data from He et al. (2014) and the ones obtained using Aspen Hysys® is presented in Table 5. In this validation the inputs are: compressor power (W); saturated liquid state at condenser outlet and at 327.55 K (54.4 oC); saturated vapor state at evaporator outlet and at 249.85 K (-23.3 oC). The outputs are: condensing pressure (Pcond), evaporating pressure (Pevap) and COP. Deviations are calculated using Eq. (14) substituting REF by values from He et al. (2014).

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From Table 5 it is possible to see that for R290 the deviations are very small, in good agreement with what was expected since the Peng-Robinson equation of state is adequate for estimating hydrocarbons properties. For the mixture R290/R600a greater deviations were observed, showing the importance of using accurate mixing rules and a binary interaction parameter. Nevertheless, since the enthalpies are not significantly affected by pressure, the deviation of COP, calculated based on enthalpies was very small, showing that the model built using Aspen Hysys® simulator is adequate to perform a thermodynamic analysis of the system of mixture refrigerants. It is important to say that He et al. (2014) did not mention in their work the fluid package used to perform the simulations, so a further discussion about the reasons of the deviations observed is not possible, but the deviations are probably due to the use of a different binary interaction parameter. After this careful validation procedure, the process simulator and the model built for a vapor injection refrigeration cycle with a flash tank were considered reliable to perform a parametric analysis of this system using a mixture refrigerant of R290/R600a. 4. Results and discussion In all following figures, the parametric analyses of the variables are performed as a function of R290 mass fraction in stream #5 (in wt%), which is the condenser outlet stream fixed as saturated liquid and also as a function of three different expansion ratio - ER (in %) in the upper-stage valve for both Cases A and B (item 3.2), when applicable. Fig. 5 and Fig. 6 present COP behavior for FTVI and VCC for both cases A and B. As expected, based on previous works from literature (Swinney et al., 1998 and He et al., 2014), Case A shows a maximum point for the COP, for mixtures around 40 wt% of R290, while Case B shows a minimum point, around 50 wt% of R290, for all expansion ratios used. From 0 to 30 wt% of R290 an ER of 70% seems to be the slightly more adequate, but from this point till 100 wt% of R290, an ER of 50% is better. For Case B there is a wider range of composition (from 30 to 70 wt%) where the COP for ERs of 70% and 50% are very close. Fig. 5 and Fig. 6 shows that the COP for a VCC is always lower than the one of a FTVI cycle, independently of the choice of temperatures (Cases A or B). Apparently it seems that there are no advantages in considering the choice of temperatures of Case B to evaluate the performance of mixed refrigerants, since they always lead to lower values of COP than the pure refrigerants. The presence of the more volatile component (R290) tends to raise the compressor discharge pressure and the presence of the less volatile component (R600a) tends to reduce the compressor suction pressure, increasing the overall compression ratio and hence reducing COP. This is why a parametric analysis to extend the comparison between Cases A and B to other process variables beyond the COP is important, trying to obtain an overview of some other pros and cons that may exist between them.

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For example, when considering the COP gain given by Eq. (15), between the FTVI and VCC cycles, as shown in Fig. 7, it is possible to see that the FTVI present a COP from 16 to 32% greater than the one of the VCC, depending on the composition of mixture refrigerant and ER used, for both Cases A and B.

C O Pg a in  %



 C O PF T V I      100  C O PV C 

(15)

It is possible to see that Case B always presents a greater gain of COP when compared to Case A, at the same R290 wt% and ER. For cases of retrofit from a VCC operating with a pure substance to a FTVI with mixture refrigerant, if the control system was project to control the temperatures used in Case B, this can be a better option than implementing a new control system for Case A. Fig. 8 and Fig. 9 present the refrigerant mass flow rate (in kg.s-1) in stream #5 for both cases of FTVI and also for the VCC. The mass flow rate is adjusted by the simulator to accomplish the fixed cooling capacity (1 kW), also considering the other inputs of the simulation. It is possible to observe that for Case A the refrigerant mass flow rate passes through a minimum in the same range that the COP assumes maximum value, showing an additional advantage of fixing the temperatures defined for Case A since the reduction of mass flow rate of the mixture refrigerant contributes to reduce the size of the equipment involved in the cycle, reducing fixed costs. Refrigerant mass flow rate for VCC is always higher than the ones of FTVI (both cases), at the same R290 wt% and ER, showing that FTVI presents a better performance when compared to VCC.

For Case B shown in Fig. 9 it is possible to see that the mass flow rate is not so sensible to the composition of the mixed refrigerant in stream #5. The effect of the ER is qualitatively the same as observed for Case A. This reduced influence of composition over the mass flow rate can be explained by the choice of temperatures made. Fixing the temperature of saturated liquid at the condenser outlet and the saturated vapor at the evaporator outlet, the variation of enthalpy between the inlet and outlet of the evaporator is less sensible to the composition. Fig. 10 shows how temperature varies in stream #8 for Case B (for Case A, T8 is fixed as -23.3 oC). It can be seen that a minimum temperature is achieved, between 50 and 60 wt% of R290 and that this value increases with decreasing ER in the upper-stage valve. As T1, temperature at the evaporator outlet stream, is fixed as -23.3 oC, the lower the temperature in stream #8, the better the heat transfer process and the performance of the cycle. So, for Case B, it would be desirable a higher ER, considering the evaporator performance. Fig. 11 shows how temperature varies in stream #5 for Case A (for Case B, T5 is fixed as 54.4 oC). Since this stream is located before the upper-stage valve, it was plotted only as a function of mixed refrigerant composition. The minimum temperature of stream #5 is achieved for a R290 13 Page 13 of 34

mass fraction of 40 wt%, which is the same composition of maximum COP. Lower temperatures in stream #5 lead to lower temperatures in stream #8, contributing to increase the COP, but on the other hand, for an effective heat transfer in the condenser, colder utilities will be necessary, which can increase operating costs. Stream #5 is the same in both FTVI and VCC.

The importance of the influence of composition over the temperatures of streams #5 and #8 is better visualized when the refrigerant temperature glide (RTG) is analyzed, since the greater the RTG, the better the heat transfer process in evaporator and condenser, increasing the performance of the refrigeration cycle. Fig. 12 presents the RTG (in K) for both FTVI cases. For the condenser, there is no influence of the ER in the upper-stage valve since the pressure is constant between the condenser inlet and outlet streams. Fig. 12 shows that the RTG for Case A is greater than the one of Case B for the range between 20 to 70 wt% of R290, with a maximum at 40 wt%, which is the same point of maximum COP for Case A. However the difference observed for the RTG between the cases is not so significant, since the greatest difference observed is less than 0.5 K. Fig. 13 presents the behavior of RTG in the evaporator for FTVI Case A (Case B presents a very similar behavior and it is not shown here) and for a VCC. The VCC presents higher values of RTG from 0 to 40 wt% of R290 and the FTVI with an ER of 70% presents the lowest values in this range. Above 40 wt% of R290 occurs an inversion in this behavior and a FTVI with ER of 70% assumes the greatest values. When the concentration of R290 increases, it is expected that the expansion in the lower-stage valve generates a two-phase mixture with a vapor phase richer in R290, since this is the less volatile component. The lower the pressure, the more R290 in the vapor phase for a fixed temperature, like in stream #8 for Case A. This will correspond to a lower saturate liquid temperature too, tough increasing the RTG. Fig. 14 presents the compressor power input (in kW) for the FTVI (Case A) necessary in both stages of the compressor. S1 stands for stage 1 and S2 for stage 2 and the % values refer to the ER in the upper-stage valve. It is possible to see that the total compressor power goes through a minimum at around 40 to 50 wt% of R290 for all ER studied. This minimum power contributes to increase COP. In the second stage of the compressor, the power consumption increases with the concentration of R290 because this is the most volatile component, but this increasing behavior is very slight for the range from 0 to 40 wt% of R290 and from this point it becomes more pronounced for an ER of 70%. For an ER of 30%, total compressor power presents the greatest value because mass flow rate in stream #7 increases (see Fig. 8 and 9), requiring more power in the first stage, but the total compressor power inputs for ER’s of 70% and 50% are very close because of the distribution of the power inputs in each stage. In these cases it is interesting to observe that the power distribution in both stages is inverted depending on the refrigerant composition. For low concentrations of R290, the distribution (proportion between W S1 and WS2) 14 Page 14 of 34

is better when an ER of 70% is used and for high concentrations of R290, it is better for an ER of 50%. This can be useful when defining how a practical compressor in two stages should be projected. Fig. 15 presents the compressor power input for the FTVI (Case B). Once again the cycle with an ER of 30% presents the greatest total power consumption in the compressor and for ER’s 70% and 50% the total power is almost the same. But the individual compressor power of each stage is quite different from Case A. For Case B, there is a better distribution of power consumption for an ER of 70% from 0 to 40 wt% of R290 and for an ER of 50% this distribution of power consumption seems to be better for almost the entire range of R290 composition. It is also remarkable that the compressor power in the second stage increases linearly for all ER’s, meaning that the choice of temperatures once again influences the power consumption in the compressor, though affecting the COP. The compression ratio given by Eq.(14) was also analyzed and its behavior is shown in Fig. 16 for FTVI (both cases) and VCC. For FTVI Case B, VCC presents the lowest compression ratio for the entire range of refrigerant composition, because there is only one expansion stage in this cycle while for FTVI there are two stages and as expected, the greater the expansion ratio, the greater the compression ratio as well. For Case A the compression ratios are almost the same regardless the expansion ratio and the cycle, varying only with mixture refrigerant composition, which may represent another advantage of choosing the reference temperatures of Case A. This is due the choice of temperatures that results in equal pressures at the same points of the cycle. When the refrigerant liquid stream passes through the upper-stage valve, a two-phase stream is produced in the expansion process and the composition of these phases is quite different since this is a non-azeotropic refrigerant mixture. The expansion ratio in this valve also influences the proportion between the vapor and liquid phases affecting the performance of both the evaporator and the compressor and hence the performance of the cycle. Fig. 17 shows the distribution of mass composition for the vapor phase (y, stream #9) and liquid phase (x, stream #7) as a function of mixture refrigerant composition at stream #5 and also as a function of the expansion ratio in this valve, considering temperatures for Case A. From Fig. 17 it is possible to observe that the composition of vapor phase was not significantly affected by the expansion ratio in the valve, while the liquid phase composition suffers some influence (the behavior and values are quite the same for Case B and this is why only the results for Case A are presented). The lower the pressure in stream #6 (higher values of the expansion ratio), the smaller the concentration of R290 in liquid phase since it is the most volatile component and tends to go the vapor phase at low pressures more than R600a does. But vapor phase composition is less sensible to the expansion ratio, so it is necessary to produce more vapor phase from the expansion. This can be seen in Fig. 18 that shows the mass flow ratio (V/F) between the flash tank vapor outlet stream (#9) and the feed stream (#6) for both cases studied. For Case A the addition of R290 to a pure R600a decreases the mass flow rate of the vapor phase from the flash tank in the range from 0 to 40 wt%, when compared to pure R600a. Only for 15 Page 15 of 34

mixture refrigerant with more than 50 wt% of R290 the mass flow rate of vapor phase is bigger than for pure R600a for all expansion ratios evaluated. Since for case A the temperature fixed is the one for saturated vapor at the condenser pressure, the presence of R290 tends to decrease the saturation pressure at the condenser and when expansion occurs, less vapor phase is produced. But after a certain concentration, there will be a significant amount of R290 in the system and since it is less volatile than R600a, an increase in vapor phase mass flow is observed. For all the range of R290 composition in stream #5, the lower the pressure in stream #6, more vapor phase is generated. For Case B any addition of R290 to pure R600a will lead to an increase in the mass flow ratio. This is because in this case, the fixed temperature is the outlet stream of the condenser and the state is fixed as saturated liquid. So, the pressure is high enough to ensure this state and the addition of a less volatile component will always increase vapor phase mass flow rate. To check this, a plot of pressure at stream #6 is presented in Fig. 19. Fig. 19 shows that pressure in both streams (#5 and #6) increases with R290 wt% in stream #5, but their behavior is different depending on the case studied. It is also possible to observe that stream #6 pressures for both cases are closer at higher ER, i.e. the more vapor is generated during the expansion process in the upper-stage valve, the less the choice of temperatures influences the pressure of the system, which becomes more a function of the mixture refrigerant composition. Investigating the behavior of temperature at stream #6, shown in Fig. 20, it is possible to observe that for Case B this temperature does not vary significantly at high pressures (low ER) of stream #6, because a lower quantity of vapor is generated (as shown in Fig. 18). But as long as the pressure drop in the valve increases (higher ER), temperature of stream #6 becomes lower and T6 for Case A is more sensitive to R290 wt% in stream #5 than it is for Case B, passing through a minimum. Finally, vapor fraction at stream #8, after the lower-stage expansion valve, was analyzed and the results are shown in Fig. 21. The greater the proportion of vapor phase, the worse for an effective heat exchange in the evaporator since less latent heat will be exchanged, demanding a greater refrigerant mass flow rate for the same cooling capacity. As can be seen in Fig. 21, Case A produces the lowest vapor fraction for all expansion rates studied. There is a minimum range of vapor fraction that coincides with the range of maximum COP, i.e. from 40 to 50 wt% of R290. The basic vapor compression cycle was also checked for both cases and they present the greatest vapor fraction of all, showing that this cycle is always less efficient than the vapor injection cycle with flash tank. Conclusions In this paper a parametric analysis of the main variables of a refrigeration system, based on simulations of a flash tank with vapor injection system using a mixture refrigerant R290/R600a was presented and the thermodynamic performance of this system was evaluated. Two different cases, denominated A and B, considering the choice of temperatures at two different points of the 16 Page 16 of 34

system were studied, together with a basic vapor compression cycle, for the sake of comparison. The parametric analysis was developed considering the influence of the mixture refrigerant composition at the condenser outlet stream and also the expansion ratio in the upper-stage valve of the system. In practical terms, for both cases A and B of FTVI systems an efficient control of the desirable temperatures is difficult to implement since the choices of temperatures are related to very specific physical states. It seems that Case B is a little easy to control since it is related to the temperatures at the outlet streams of the condenser and evaporator, while Case A is related to a temperature that occurs inside the condenser and at the outlet of the lower-stage valve. For cases of retrofit of a vapor compression cycle projected with a control system acting on the condenser and evaporator outlet temperatures, the option by a FTVI Case B seems to be easier to implement. It is also important to consider the application of the refrigeration cycle, taking into consideration the required inlet and outlet temperatures of the heat transfer fluid, which may define what choice of fixed temperatures would be better for the heat exchange process. Considering the operating variables analyzed in this work and the thermodynamic performance of the cycles, the main conclusions of this work are:  when using a mixture refrigerant of R290 and R600a, the ideal composition of the refrigerant is between 30 to 50 wt% of R290, with a maximum at 40 wt% for Case A. The 50 wt% mixture seems to be a good option since there is already a commercial mixture refrigerant available at this composition;  when comparing with a basic vapor compression cycle, both cases of FTVI cycles have shown always a greater COP;  considering the expansion ratio in the upper-stage valve, it was verified that in the range of COP maximum the expansion ratio of 50% presents a slightly better COP than the one of 70%. So if it is possible to operate with higher pressures at stream #6, reducing the power consumption in the compressor, this would be a better option;  the mass flow rate of mixture refrigerant is also minimum when COP is maximum for Case A and it is less sensible for Case B;  for Case B, the greater the expansion in the upper-stage valve, the lower the temperature in evaporator inlet (stream #8), which flavors the heat exchange between the heat transfer fluid;  considering the refrigerant temperature glide in both evaporator and condenser, no significant differences were found when comparing Cases A and B; on the region of maximum COP the expansion ratio of 50% gives also good results for the refrigerant temperature glide;  for Case A, the proportional distribution of compressor power between the stages seems to be the best since, presenting an equal variation during the entire range of mixed refrigerant composition, which may be an advantage to operate the compressor and the injection system. Both cases A and B for the temperature choice, described at item 3.2 were studied for the sake of comparison of cycle performance. No works presenting a comparison between the results of these two choices and considering the analysis of a FTVI using a mixture refrigerant of R290/R600a were 17 Page 17 of 34

found by the authors in the open literature. The simulations performed in this work and the following parametric analysis of the main operating variables may be useful to evaluate the pros and cons of each case analyzed and give some support for future experimental works. Acknowledgements To Fundação de Amparo à Pesquisa do Estado de São Paulo (FAPESP) for the financial support (Process 2014/15400-7). Dr. d’Angelo is also grateful for the support of the Center for Environmental Energy Engineering (CEEE)/University of Maryland.

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Nomenclature hi = specific enthalpy of stream i, kJ.kg-1 kij = binary interaction parameter mi = mass flow rate of stream i, kg.s-1 MM = molar mass Pc = critical pressure, MPa Pcond = condensing pressure, kPa Pevap = evaporating pressure, kPa Pi = pressure of process stream i, kPa PS1 = outlet pressure of first stage, MPa Psat = saturation pressure, kPa PVI = pressure of vapor injection line, MPa Qevap = rate of heat exchange or cooling capacity in the evaporator, kW T = absolute temperature, K T = temperature, oC Tb = normal boiling point temperature, K Tc = critical temperature, K Ti = temperature of process stream i, oC or K TS1D = first stage compressor discharge temperature, K TS2D = second stage compressor discharge temperature, K Tsat_liq = temperature of saturated liquid, K Tsat_vap = temperature of saturated vapor, K WS1 = compressor power input in the first stage, kW WS2 = compressor power input in the second stage, kW xi = mole or mass fraction of component i in liquid phase yi = mole or mass fraction of component i in vapor phase  = acentric factor Abbreviations AH = Aspen Hysys® version 7.3 CFC = chlorofluorocarbon COP = coefficient of performance CR = compression ratio FTVI = vapor injection refrigeration cycle with flash tank GWP = global warming potential HCFC = hydrochlorofluorocarbon HFC = hydrofluorocarbon HFO = hydrofluoroolefin IHX = internal heat exchanger LNG = liquefied natural gas NARM = non-azeotropic refrigerant mixture ODP = ozone depleting potential PR-EOS = Peng-Robinson Equation of State REF = Refprop© version 9.1 RTG = refrigerant temperature glide, oC or K VCC = vapor compression cycle 19 Page 19 of 34

References Almeida, I.M.G., Barbosa, C.R.F., Fontes, F.A.O., 2010. Thermodynamic and Thermophysical Assessment of Hydrocarbons Application in Household Refrigerator. Eng. Térmica (Thermal Eng. 9, 19–27. doi:10.1016/S0196-8904(02)00065-1 Arcaklıoğlu, E., Çavuşoğlu, A., Erişen, A., 2006. Thermodynamic analysis of refrigerant mixtures for possible replacements for CFCs by an algorithm compiling property data. Appl. Therm. Eng. 26, 430–439. doi:10.1016/j.applthermaleng.2005.06.005 AspenTech, 2015. Aspen HYSYS Help Section. Aspen Technology, Inc., Burlington, MA, USA. https://www.aspentech.com/ Calm, J.M., 2008. The next generation of refrigerants - Historical review, considerations, and outlook. Int. J. Refrig. 31, 1123–1133. doi:10.1016/j.ijrefrig.2008.01.013 Chen, J., Havtun, H., Palm, B., 2015. Conventional and advanced exergy analysis of an ejector refrigeration system. Appl. Energy 144, 139–151. doi:10.1016/j.apenergy.2015.01.139 Chen, X., Omer, S., Worall, M., Riffat, S., 2013. Recent developments in ejector refrigeration technologies. Renew. Sustain. Energy Rev. 19, 629–651. doi:10.1016/j.rser.2012.11.028 Cho, H., Baek, C., Park, C., Kim, Y., 2009. Performance evaluation of a two-stage CO2 cycle with gas injection in the cooling mode operation. Int. J. Refrig. 32, 40–46. doi:10.1016/j.ijrefrig.2008.07.008 Chunnanond, K., Aphornratana, S., 2004. Ejectors: applications in refrigeration technology. Renew. Sustain. Energy Rev. 8, 129–155. doi:10.1016/j.rser.2003.10.001 Cimsit, C., Ozturk, I.T., 2012. Analysis of compression-absorption cascade refrigeration cycles. Appl. Therm. Eng. 40, 311–317. doi:10.1016/j.applthermaleng.2012.02.035 Cimsit, C., Ozturk, I.T., Kincay, O., 2015. Thermoeconomic optimization of LiBr/H2O-R134a compression-absorption cascade refrigeration cycle. Appl. Therm. Eng. 76, 105–115. doi:10.1016/j.applthermaleng.2014.10.094 Dalkilic, a. S., Wongwises, S., 2010. A performance comparison of vapour-compression refrigeration system using various alternative refrigerants. Int. Commun. Heat Mass Transf. 37, 1340–1349. doi:10.1016/j.icheatmasstransfer.2010.07.006 Granryd, E., 2001. Hydrocarbons as refrigerants — an overview. Int. J. Refrig. 24, 15–24. doi:10.1016/S0140-7007(00)00065-7 Hammad, M., Alsaad, M., 1999. The use of hydrocarbon mixtures as refrigerants in domestic refrigerators. Appl. Therm. Eng. 19, 1181–1189. doi:10.1016/S1359-4311(98)00116-1 He, M.-G., Song, X.-Z., Liu, H., Zhang, Y., 2014. Application of natural refrigerant propane and propane/isobutane in large capacity chest freezer. Appl. Therm. Eng. 70, 732–736. doi:10.1016/j.applthermaleng.2014.05.097 Högberg, M., Vamling, L., Berntsson, T., 1993. Calculation methods for comparing the performance of pure and mixed working fluids in heat pump applications. Int. J. Refrig. 16, 403–413. doi:10.1016/0140-7007(93)90057-F Jwo, C.-S., Ting, C.-C., Wang, W.-R., 2009. Efficiency analysis of home refrigerators by replacing hydrocarbon refrigerants. Measurement 42, 697–701. doi:10.1016/j.measurement.2008.11.006 Kruse, H., 1981. The advantages non-azeotropic refrigerant mixtures for heat pump application. Int. J. Refrig. 4, 119–125. doi:10.1016/0140-7007(81)90102-X 20 Page 20 of 34

Kyoto Protocol to the United Nations Framework Convention on Climate Change, 1997. United Nations (UN), New York, NY, USA. http://unfccc.int/resource/docs/publications/08_unfccc_kp_ref_manual.pdf Lee, H., Hwang, Y., Radermacher, R., Chun, H., 2013. Potential benefits of saturation cycle with two-phase refrigerant injection. Appl. Therm. Eng. 56, 27–37. doi:10.1016/j.applthermaleng.2013.03.030 Lee, M.-Y., Lee, D.-Y., Kim, Y., 2008. Performance characteristics of a small-capacity directly cooled refrigerator using R290/R600a (55/45). Int. J. Refrig. 31, 734–741. doi:10.1016/j.ijrefrig.2007.11.014 Lemmon, E.W., Huber, M.L., McLinden, M.O. NIST Standard Reference Database 23: Reference Fluid Thermodynamic and Transport Properties-REFPROP, Version 9.1, National Institute of Standards and Technology, Standard Reference Data Program, Gaithersburg, 2013. Lim, J.S., Ho, Q.N., Park, J.-Y., Lee, B.G., 2004. Measurement of Vapor−Liquid Equilibria for the Binary Mixture of Propane (R-290) + Isobutane (R-600a). J. Chem. Eng. Data 49, 192–198. doi:10.1021/je030106k Ma, G.Y., Zhao, H.X., 2008. Experimental study of a heat pump system with flash-tank coupled with scroll compressor. Energy Build. 40, 697–701. doi:10.1016/j.enbuild.2007.05.003 Makhnatcha, P., Khodabandeha, R., 2014. The role of environmental metrics (GWP, TEWI, LCCP) in the selection of low GWP refrigerant. Energy Procedia, 61, 2460-2463. doi:10.1016/j.egypro.2014.12.023 Mani, K., Selladurai, V., 2008. Experimental analysis of a new refrigerant mixture as drop-in replacement for CFC12 and HFC134a. Int. J. Therm. Sci. 47, 1490–1495. doi:10.1016/j.ijthermalsci.2007.11.008 Mathison, M.M., Braun, J.E., Groll, E. a, 2011. Performance limit for economized cycles with continuous refrigerant injection. Int. J. Refrig. 34, 234–242. doi:10.1016/j.ijrefrig.2010.09.006 McLinden, M.. O., Radermacher, R., 1987. Methods for comparing the performance of pure and mixed refrigerants in the vapour compression cycle. Int. J. Refrig. 10, 318–325. doi:http://dx.doi.org/10.1016/0140-7007(87)90117-4 Miyara, A., Koyama, S., Fujii, T., 1992. Consideration of the performance of a vapour-compression heat pump cycle using non-azeotropic refrigerant mixtures. Int. J. Refrig. 15, 35–40. doi:10.1016/0140-7007(92)90065-3 Mohanraj, M., Jayaraj, S., Muraleedharan, C., 2007. Improved energy efficiency for HFC134a domestic refrigerator retrofitted with hydrocarbon mixture (HC290/HC600a) as drop-in substitute. Energy Sustain. Dev. 11, 29–33. doi:10.1016/s0973-0826(08)60407-x Mohanraj, M., Jayaraj, S., Muraleedharan, C., Chandrasekar, P., 2009. Experimental investigation of R290/R600a mixture as an alternative to R134a in a domestic refrigerator. Int. J. Therm. Sci. 48, 1036–1042. doi:10.1016/j.ijthermalsci.2008.08.001 Montreal Protocol on Substances That Deplete the Ozone Layer, 1987. United Nations (UN), New York, NY, USA (1987 with subsequent amendments). http://ozone.unep.org/en/handbookmontreal-protocol-substances-deplete-ozone-layer/5 Peng, D.Y., Robinson, D.B., 1976. A new two-constant equation of state. Ind. Eng. Chem. Fundam., 15, 59-64. Sarbu, I., 2014. A review on substitution strategy of non-ecological refrigerants from vapour 21 Page 21 of 34

compression-based refrigeration, air-conditioning and heat pump systems. Int. J. Refrig. 46, 123– 141. doi:10.1016/j.ijrefrig.2014.04.023 Sekhar, S.J., Lal, D.M., 2005. HFC134a/HC600a/HC290 mixture a retrofit for CFC12 systems. Int. J. Refrig. 28, 735–743. doi:10.1016/j.ijrefrig.2004.12.005 Sumeru, K., Nasution, H., Ani, F.N., 2012. A review on two-phase ejector as an expansion device in vapor compression refrigeration cycle. Renew. Sustain. Energy Rev. 16, 4927–4937. doi:10.1016/j.rser.2012.04.058 Swinney, J., Jones, W., Wilson, J., 1998. The impact of mixed non-azeotropic working fluids on refrigeration system performance. Int. J. Refrig. 21, 607–616. doi:10.1016/S0140-7007(98)00039-5 Tan, Y., Wang, L., Liang, K., 2015. Thermodynamic performance of an auto-cascade ejector refrigeration cycle with mixed refrigerant R32 + R236fa. Appl. Therm. Eng. 84, 268–275. doi:10.1016/j.applthermaleng.2015.03.047 Wallington, T.J., Sulbaek Andersen, M.P., Nielsen, O.J., 2010. Estimated photochemical ozone creation potentials (POCPs) of CF3CF{double bond, long}CH2 (HFO-1234yf) and related hydrofluoroolefins (HFOs). Atmos. Environ. 44, 1478–1481. doi:10.1016/j.atmosenv.2010.01.040 Wang, B., Shi, W., Han, L., Li, X., 2009a. Optimization of refrigeration system with gas-injected scroll compressor. Int. J. Refrig. 32, 1544–1554. doi:10.1016/j.ijrefrig.2009.06.008 Wang, X., Hwang, Y., Radermacher, R., 2009b. Two-stage heat pump system with vapor-injected scroll compressor using R410A as a refrigerant. Int. J. Refrig. 32, 1442–1451. doi:10.1016/j.ijrefrig.2009.03.004 Wongwises, S., Chimres, N., 2005. Experimental study of hydrocarbon mixtures to replace HFC134a in a domestic refrigerator. Energy Convers. Manag. 46, 85–100. doi:10.1016/j.enconman.2004.02.011 Xu, X., Hwang, Y., Radermacher, R., 2013. Performance comparison of R410A and R32 in vapor injection cycles. Int. J. Refrig. 36, 892–903. doi:10.1016/j.ijrefrig.2012.12.010 Xu, X., Hwang, Y., Radermacher, R., 2011. Refrigerant injection for heat pumping/air conditioning systems: Literature review and challenges discussions. Int. J. Refrig. 34, 402–415. doi:10.1016/j.ijrefrig.2010.09.015 Yan, G., Cui, C., Yu, J., 2015. Energy and exergy analysis of zeotropic mixture R290/R600a vaporcompression refrigeration cycle with separation condensation. Int. J. Refrig. 53, 155–162. doi:10.1016/j.ijrefrig.2015.01.007

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Fig. 1. Progression of refrigerants (adapted from Calm, 2008).

Fig. 2. Schematic of the vapor injection refrigeration cycle with a flash tank and P-h diagram.

Saturation pressure/kPa

2500 2000

1500

R290-AH R290-REF R600a-AH R600a-REF 90%R290-AH 90%R290-REF 50%R290-AH 50%R290-REF

1000 500 0 -70 -60 -50 -40 -30 -20 -10 0 10 20 30 40 50 60 70

Temperature/°C 23 Page 23 of 34

Fig. 3. Saturation pressure (in kPa) as a function of temperature (in oC) for pure fluids and mixtures using two different data banks: Refprop (REF) and Aspen Hysys (AH).

Fig. 4. Comparison between COP obtained in this work and from Swinney et al. (1998).

3.00 2.75

COP

2.50 2.25 2.00 1.75

70%

50%

30%

VCC

1.50 0

10

20

30

40

50

60

70

80

90 100

R290 mass fraction in stream #5 /wt% Fig. 5. COP for FTVI and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case A).

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3.00 70%

50%

30%

VCC

2.75

COP

2.50 2.25 2.00 1.75 1.50 0

10

20

30

40

50

60

70

80

90

100

R290 mass fraction in stream #5 /wt% Fig. 6. COP for FTVI and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case B).

35

COPgain / %

30 25 20 70%_A 70%_B 50%_A 50%_B 30%_A 30%_B

15 10

0

10 20 30 40 50 60 70 80 90 100

R290 mass fraction in stream #5 /wt% Fig. 7. Increase in COP (%) from VCC to FTVI cycles (Cases A and B) as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve.

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Mixture refrigerant mass flow rate in stream #5 x 10³ / kg.sˉ¹

5.40 5.20 5.00 4.80 4.60 4.40

70% 50%

4.20

30% VCC

4.00 0

10

20

30

40

50

60

70

80

90 100

R290 mass fraction in stream #5 /wt% Fig. 8. Mixture refrigerant mass flow rate in stream #5 (kg.s-1) for FTVI and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case A).

Mixture refrigerant mass flow rate in stream #5 x 10³ / kg.sˉ¹

5.40 5.20 5.00 4.80 4.60 70%

4.40

50% 30%

4.20

VCC

4.00 0

10

20

30

40

50

60

70

80

90 100

R290 mass fraction in stream #5 /wt% Fig. 9. Mixture refrigerant mass flow rate in stream #5 (kg.s-1) for FTVI and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case B).

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Temperature - stream #8 / K

252.0 250.0 248.0 246.0 244.0

70% 50%

242.0

30% VCC

240.0 0

10

20

30

40

50

60

70

80

90 100

R290 mass fraction in stream #5 /wt% Fig. 10. Temperature at stream #8 (K) for FTVI and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case B).

Fig. 11. Temperature at stream #5 (K) for FTVI and VCC as a function of R290 mass fraction (wt%) (Case A).

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Refrigerant temperature glide in the evaporator/ K

Fig. 12. Refrigerant temperature glide (K) in the condenser for both FTVI cases as a function of R290 mass fraction (wt%).

10 9 8 7 6 5 4 3 2 1 0

70%

0

10

20

30

50%

40

50

30%

60

VCC

70

80

90 100

R290 mass fraction in stream #5 /wt% Fig. 13. Refrigerant temperature glide (K) in the evaporator for FTVI and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case A).

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Compressor power / kW

0.50 0.45 0.40 0.35 0.30 0.25 0.20 0.15 0.10 0.05 0.00

S1_70% S2_70% S1_50% S2_50% S1_30% S2_30% Wtotal_30% Wtotal_50% Wtotal_70%

0

10 20 30 40 50 60 70 80 90 100

R290 mass fraction in stream #5 /wt%

Compressor power / kW

Fig. 14. Compressor power (in kW) for FTVI as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case A).

0.50 0.45 0.40 0.35 0.30 0.25 0.20 0.15 0.10 0.05 0.00

S1_70% S2_70% S1-50% S2-50% S1_30% S2_30% Wtotal_70% Wtotal_50% Wtotal_30%

0 10 20 30 40 50 60 70 80 90 100

R290 mass fraction in stream #5 /wt% Fig. 15. Compressor power (in kW) for FTVI as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case B).

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15.0

70%_A 70%_B 50%_A 50%_B 30%_A 30%_B VCC_A VCC_B

Compression ratio

14.0 13.0 12.0 11.0 10.0 9.0 8.0 0

10 20 30 40 50 60 70 80 90 100

R290 mass fraction in stream #5 /wt%

R290 mass fraction (y = vapor, x = liquid)

Fig. 16. Compressor ratio for FTVI (both cases) and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve.

1.00 0.90 0.80 0.70 0.60 0.50 0.40 0.30 0.20 0.10

x_70% x_50% x_30% y_70% y_50% y_30%

0.00 0 10 20 30 40 50 60 70 80 90 100

R290 mass fraction in stream #5 /wt% Fig. 17. R290 mass fraction in vapor (y) and liquid (x) flash tank outlet streams as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Case A).

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Mass flow ratio V/F in the flash tank

0.45 0.40 0.35 0.30 0.25 0.20 0.15 0.10 70%_A 50%_B

0.05

70%_B 30%_A

50%_A 30%_B

0.00 0

10

20

30

40

50

60

70

80

90 100

R290 mass fraction in stream #5 /wt% Fig. 18. Mass flow ratio (both cases) between vapor (V, stream #9) and feed stream (F, stream #6) in the flash tank as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve.

Process stream pressure / kPa

2000

P5_A_VCC P5_B_VCC P6_A_30% P6_B_30% P6_A_50% P6_B_50% P6_A_70% P6_B_70%

1800 1600 1400 1200 1000 800 600 400 200 0 0

10

20

30

40

50

60

70

80

90 100

R290 mass fraction in stream #5 /wt% Fig. 19. Pressure at process streams #5 and #6 (in kPa) for FTVI (both cases) and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve.

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Temperature at stream #6 / K

315.0 310.0 305.0 300.0 295.0 70%_#6_A 70%_#6_B 50%_#6_A 50%_#6_B 30%_#6_A 30%_#6_B

290.0 285.0 280.0 275.0 0

10

20

30

40

50

60

70

80

90 100

R290 mass fraction in stream #5 /wt%

Vapor fraction in stream #8 - mass basis

Fig. 20. Temperature (K) at stream #6 as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve (Cases A and B).

0.55 0.50 0.45 0.40 0.35 0.30 0.25 0.20 0.15 0.10

70%_A

70%_B

50%_A

50%_B

0.05

30%_A

30%_B

VCC_A

VCC_B

0.00 0

10

20

30

40

50

60

70

80

90

100

R290 mass fraction in stream #5 /wt% Fig. 21. Vapor fraction (mass basis) in stream #8 for FTVI (both cases) and VCC as a function of R290 mass fraction (wt%) and ER (%) in upper-stage valve.

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Table 1 Summary of some studies involving vapor injection cycles with flash tank for cooling applications. Reference Cho et al., 2009

Refrigerant CO2

Comments experimental

Wang et al., 2009b

R410a

theoretical approach

Mathison et al., 2011

R410a

theoretical approach

Xu et al., 2011

-

review of technology

Xu et al., 2013 Lee et al., 2013

R410a and R32 12 pure and two mixed (R32/R134a and R32/R152a)

experimental theoretical approach

Table 2 Summary of some recent works involving mixture refrigerant R290/R600a. Reference Wongwises and Chimres (2005) Sekhar and Lal (2005)

Approach Exp. Theor.   

Arcaklioğlu et al. (2006) Mohanraj et al. (2007)



Lee et al. (2008)



Mani and Selladurai (2008)



Jwo et al. (2009)



Mohanraj et al. (2009)



Dalkilic and Wongwises (2010)



Almeida et al. (2010)

 

He et al. (2014)



Yan et al. (2015a)

Application Replace R134a in domestic refrigerators. Ternary mixture of R134/R290/R600a to replace R12 in domestic refrigerators. System with subcooling and superheating heat exchangers. R134a domestic refrigerator retrofit. Replacement of R134a by a mixture 55/45 wt% R290/R600a in refrigerators. Analysis of mixture 68/32 wt% R290/R600a as drop-in replacement for R12 and R134. Evaluate the replacement of R134a and R12 by a mixture 50/50 wt% R290/R600a. Investigation of mixture refrigerant with different compositions as alternative to R134a. Comparison of a vapor compression refrigeration system using different compositions of mixture refrigerant. Performance evaluation of a mixture refrigerant 60/40 and 50/50 wt% R290/R600a. Performance evaluation applied to large capacity chest freezer. Vapor compression cycle with internal HX for subcooling and superheating.

Exp. = experimental, Theor. = theoretical

Table 3 Reference temperatures and refrigerant temperature glide (RTG) for cases A and B. Case A B Equipment Evaporator Condenser Evaporator Condenser Saturated vapor at condenser pressure 327.55 K (54.4 oC)

Outlet stream #1 is 249.85 K (-23.3 oC)

Initial at pressure of temperature stream #8

T of saturated liquid

T of saturated liquid at pressure of stream #5 (P5 = P4)

T of saturated liquid at pressure of stream #8

Final T of saturated vapor temperature at pressure of

T of saturated vapor, fixed as 54.4

T of saturated vapor fixed as -23.3 oC, at

Reference Inlet stream #8 is o temperature 249.85 K (-23.3 C)

Outlet stream #5 is 327.55 K (54.4 oC) T of saturated liquid, fixed as 54.4 oC, at pressure of stream #5 (P5 = P4) T of saturated vapor at pressure of 33 Page 33 of 34

stream #8 (P8 = P1)

o

C, at pressure of stream #4

pressure of stream #8 (P8 = P1)

stream #5 (P5 = P4)

RTG = Final temperature – Initial temperature = Tsat_vap – Tsat_liq (at constant pressure) Table 4 Comparison of thermodynamic properties of pure refrigerants from two different data banks: Refprop© 9.1 and Aspen Hysys® 7.3. Refrigerant Data base MM Tb /K Tc /K Pc /MPa 

Refprop v. 9.1 44.096 231.04 369.89 4.2512 0.1521

R290 Aspen Hysys v. 7.3 44.100 231.07 369.82 4.2420 0.1488

Refprop v. 9.1 58.122 261.40 407.81 3.6290 0.1840

R600a Aspen Hysys v. 7.3 58.120 261.36 407.98 3.6550 0.1812

Table 5 Comparison between simulated cycles using refrigerants R290, R600a and 90/10 wt% R290/R600a. Refrigerant Compressor power (W) Outputs Pcond (kPa) Pevap (kPa) COP

R290

R600a

R290/R600a (90/10 wt%)

162.4

145.1

155.5

Ref.

t.w.

1882.6 217.6 1.62

1893.0 218.3 1.61

Deviation (%) 0.55 0.32 0.62

Ref.

t.w.

762.32 62.43 1.70

765.1 64.12 1.68

Deviation (%) 0.36 2.71 1.18

Ref.

t.w.

1700.2 204.03 1.69

1799 189 1.68

Deviation (%) 5.81 7.37 0.59

Ref. = He et al. (2014), t.w. = this work

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