Evaluation of a condenser based on mini-channels technology working with R410A and R32. Experimental data and performance estimate

Applied Energy 202 (2017) 112–124

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Evaluation of a condenser based on mini-channels technology working with R410A and R32. Experimental data and performance estimate Alejandro López-Belchí a,⇑, Fernando Illán-Gómez b a Engineering and Applied Technologies Department, Centro Universitario de la Defensa de San Javier (University Centre of Defence at the Spanish Air Force Academy), Ministry of Defense–Technical University of Cartagena, C/Coronel Lopez Peña, s/n, 30720, Santiago de la Ribera, Murcia, Spain b Thermal and Fluids Engineering Department, Universidad Politécnica de Cartagena, Escuela Técnica Superior de Ingeniería Industrial, C/ Doctor Fleming s/n, 30202, Cartagena, Murcia, Spain

h i g h l i g h t s
Little research is available for R32, the most appealing substitute for R410A.
Using R32 with minichannels can reduce environmental impact and improve efficiency.
Direct replacement must be carefully studied, efficiency can increase or decrease.
Operating conditions affect the efficiency, drop-in studies must consider them.
R32 refrigerant charge is lower than R410A for a fixed thermal power.

a r t i c l e

i n f o

Article history: Received 9 November 2016 Received in revised form 27 April 2017 Accepted 16 May 2017

Keywords: Two-phase flows Mini-channels R410A R32 GWP Energy efficiency

a b s t r a c t In this paper, R32 is investigated as a replacement refrigerant for R410A. The Global Warming Potential (GWP) of R32 is only 675, 32% of that of R410Awhich has a GWP of 2088. Theoretical and experimental investigations are carried out on the performance of the condensation process within a mini-channel tube. Mini-channel heat exchangers technology allows reducing refrigerant charge and lets use flammable refrigerants. Due to the aspect ratio, high heat transfer coefficients are also registered. The experimental data recorded show that, for any given saturation temperature or refrigerant mass velocity, both the heat transfer coefficient and the frictional pressure gradient are always higher for R32. So, a numerical analysis based on the experimental data was developed to determinate which refrigerant performs better. The results of this numerical analysis show that, although at high refrigerant mass velocities R410A performs better, a given heat power can be always achieved with lower mass velocities and thus with a lower compressor power input when using R32. Therefore, it can be concluded that using R32 in a mini-channel condenser reduces the environmental impact and improves the energy efficiency of the system. Ó 2017 Elsevier Ltd. All rights reserved.

1. Introduction Refrigeration systems are one of the higher energy consumption applications in present society. The emissions related to refrigerating systems are quite high. Firstly, the direct emissions related to the leakage of refrigerant fluid to the atmosphere and secondly the indirect emissions of CO2 related to the fossil fuels used in power plants to produce the electrical energy required must be considered [1]. Due to their high energy consumption and their

⇑ Corresponding author. E-mail addresses: [email protected] (A. López-Belchí), fernando. [email protected] (F. Illán-Gómez). http://dx.doi.org/10.1016/j.apenergy.2017.05.122 0306-2619/Ó 2017 Elsevier Ltd. All rights reserved.

high greenhouse gas emissions, the improvement of their energy efficiency, the replacement of high GWP fluids by environmentally friendly refrigerants, and the decrease of refrigerant charge in air conditioning equipment are focussing the efforts of many international researchers. Many researchers all over the world are trying to improve vapour compression cycles based on experimentation and modelling to reduce their energy consumption [2–5]. Others are making strong efforts in reducing the amount of refrigerant contained in refrigeration equipment. Poggi et al. [6] made an extensive review of the strategies to reduce the charge of refrigerant, and they identified the small channel technology as a solution for charge reduction while increasing the thermal performance. From then on, many research papers have been published dealing

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113

Nomenclature A COP D EER G GWP HC HCFC HFC HFO HTC k K LMTD n ODP q_ Ra RR t T u x z

area (m2) coefficient of operation hydraulic diameter (m) energy efficiency ratio mass velocity (kg m2 s1) global warming potential hydrocarbon hydrochlorofluorocarbon hydrofluorocarbon hydrofluoroolefin heat transfer coefficient (W m2 K1) thermal conductivity (W K1 m1) coverage factor logarithmic mean temperature difference number of cells ozone depletion power heat flux (W m2) roughness (mm) relative roughness thickness (m) temperature (K) uncertainty vapour quality (kg kg1) length (m)

with the topic of micro and mini-channels in refrigeration systems. Finally, others researchers are studying the replacement of the refrigerants typically used in these devices –Hydrochlorofluorocar bons (HCFC) and Hydrofluorocarbons (HFC)–by other synthetics fluids such as Hydrofluoroolefins (HFO) or natural fluids such as ammonia, hydrocarbons (HC) or carbon dioxide. Righetti et al. [7] made a critical review of the published research articles on intube two-phase heat transfer during boiling and condensation of HFOs. Since most of the alternatives to HCF are flammable or toxic fluids, many research papers study the use of HFOs or natural refrigerants in mini-channels equipment. Among many others, some of the most recent are the following: Shah [8] analysed data for condensation of carbon dioxide in macro and mini-channels; Kuang et al. [9] studied the flow boiling of ammonia in minichannels; Liu et al. [10] studied the flow condensation of propane and R1234ze in mini-channels and Mastrullo et al. [11] studied the flow boiling of R1234yf and R1234ze in mini-channels. Regarding the refrigerant fluid, European limitations of the Global Warming Potential (GWP) of refrigerant fluids began in 2006 with the Directive 40/EC [12]. Later in 2014, the European Regulation Number 517 [13] extended the previous limitations to many other systems. Although similar limitationshave been approved in most developed countries all over the world, there is still no clear solution to replace high GWP refrigerants [14]. Since it was patented in 1991, R410A has been used as the primary alternative to R22 in air-conditioning systems to eliminate the impact of refrigerant on the ozone layer. The working pressure of an R410A air conditioner is higher than that of an R22 air conditioner, and the influence of pressure drop on the performance of an R410A air conditioner is not as obvious as that of an R22 air conditioner. These characteristics lead to the diameters of tubes used in R410A air conditioners to be smaller than those in R22 air conditioners [15]. R32 was studied during the 1990s as a zero-ODP alternative to R22, but it was finally discarded because of its flammability, and it has only been used as a component of refrigerant mixtures; in fact,

Greek symbols void fraction performance density (kg m3)

a g q

Subscripts cont contraction Al aluminium dp pressure drop cell cell expn expansion f frictional g gravitational gas gas in inlet j position j liq liquid meas measured mom momentum ref refrigerant out outlet sat saturation tp two-phase w water wall wall

R410A is a mixture of 50% R32 and 50% R125 by weight. Nowadays, the current concerns and limitations about GWP have brought R32 to the attention of researchers as an alternative to R410A, mainly due to its moderate GWP, only around one-third of R410A. Although there are ultra-low GWP alternatives such as R1234yf or R1234ze, besides they are substantially more expensive than R410A, they require a larger compressor, piping, and heat exchangers and are only considered a suitable replacement to R134a but not for R410A. In contrast, R32 has pressure and pressure ratio similar to R410A, allowing a close drop-in replacement without major system redesign. Also, the cooling capacity per unit volume of R32 is higher than that of R410A. Depending on the test conditions, with the same cooling capacity, the charge amount of R32 is about 15% lower than R410A [16]. With good thermal performance and environmental characteristics, R32 has become a highly promising alternative refrigerant. Although the replacement of R410A by R32 has been extensively studied, most of the studies compare the system performance replacing R410A by R32 and its mixtures. These direct drop-in studies are interesting for evaluating the convenience of replacing R410A by R32 in existing equipment but, to assess the convenience of replacement of R410A by R32 in new facilities, the comparison must be made between systems optimised for R410A and R32 respectively. Several authors [16–21] have studied the direct drop-in substitution of R410A with R32 in simple refrigeration cycles. Although most of them [17–20] report an improvement in system efficiency, others [16,21] have found slight decreases that, according to the authors, can be explained by the reduction in compressor efficiency related to the increase in compressor discharging temperature obtained when R32 directly replaces R410A. Other authors [18,20,22] have developed similar studies in refrigeration cycles improved with vapour injection, all of them reported a better behaviour when using R32. Alternatively, other authors have compared R410A to different mixtures of R32 and other fluids. One of the first studies was conducted by Devotta

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et al. [23]. They used a vapour compression cycle design program to assess different alternatives to R22, including R410A and mixtures of R32 with other fluids (R134a and R125). According to their results, although those mixtures slightly improve the behaviour of R410A, the reduction in GWP was very low. More recently, the interest was focussed in mixtures that can lead to stronger reductions in GWP. Tian et al. [24] and He et al. [25] studied a combination of R32 and R290 (68/32% mass concentration) as an alternative to R410A. The GWP of this mixture is only 22% of that of R410A and, according to He et al. [25] its flow boiling heat transfer coefficient clearly improves that of R410A with similar pressure drop values. On the other hand, Tian et al. [24] reported slightly lower COP when replacing R410A by this mixture in a whole system, although they claim that this worse behaviour could be explained by the fact that the system was optimized for R410A. They also studied the behaviour of the system when a microchannel heat exchanger replaced the original condenser; they found a charge reduction of 35% and the COP was increased, leading to COP values for R32 very near to that obtained for R410A. Finally, Bansal [26] studied and improved a window air conditioner and replaced the original refrigerant (R410A) by a mixture of R32 and R125 (85/15% molar concentration). He found that increasing the concentration of R32, the EER of the system was improved. Although fire hazard is low when R32 leaks from a running air conditioner [27], the direct use of R32 as refrigerant is still very low because of safety reasons. Despite its flammability, the European Commission considers R32 as the only short-term climatefriendly alternative to the HFCs widely used in medium power stationary air conditioning systems [28]. Since its flammability is related to its concentration in the air which is directly related to the charge of the system, minimising that charge would lead to safer systems. In this context, the mini-channel technology can be of great help to reduce the charge of refrigerant and achieve the indoor environment safety objective. Despite the need to have a better knowledge of the performance of a refrigeration system working with R32 when refrigerant charge reduction strategies are used, there are very few publications dealing with this issue. Cavallini et al. [29] and Jige et al. [30] compared the condensation of R32 and R410A inside a multiport mini-channel tube, but these studies only compare pressure drop and heat transfer and do not compare the whole system performance. Up to the author’s knowledge, only Tian et al. [24] have studied the behaviour of a refrigeration system when a microchannel heat exchanger is used in replacement of the original condenser. So, despite existing a necessity in reducing the refrigerant charge when a flammable fluid as R32 is used in replacement of R410A, only one of the studies available in the open literature have compared system performance where compact mini-channels heat exchangers are employed. The study presented here tries to provide the tools for assessing, from an energy efficiency point of view, the replacement of R410A by R32 in new refrigeration facilities where condensers are made of mini-channels. So, previously on this analysis, it is necessary to clarify what a mini-channels heat exchanger is. One of the first researchers who studied the influence of decreasing diameters in heat transfer coefficients were Kays and London [31] in 1984. From mid 80s, researching on heat transfer with small and miniaturised geometries has been a high-interest researching topic. The transition between micro- and macro-scale flows has been a recurring theme in the literature. There is not a fixed criterion to define the transition criteria between conventional ducts or macrochannels to mini and microchannels. Kandlikar and Grande [32] provided a classification based on the channel diameter. According to these authors, all the channels with a diameter between 0.2 mm and 3 mm can be classified as mini-channels. Such type of classifi-

cation does not consider the relative magnitude of the forces acting during flow boiling or condensation like inertia, surface tension, shear stress and gravity that are responsible for macro- to microscale transition. Kew and Cornwell [33] defined a criterion based on the confinement number Co which relates the bubble detachment diameter to the characteristic channel dimension. This criterion takes into consideration the properties of the fluid, such as surface tension and density. During the phase change, the larger heat transfer coefficients in both, condensers and evaporators, are achieved when the flow follows the annular or intermittent regimen. During annular flow, the heat transfer coefficient can be improved by a thickness reduction of the annular film and so a reduction of its thermal resistance. A possibility to get that thickness reduction is to increase the core vapour velocity in two-phase flows by decreasing the tube diameter, particularly at the low vapour qualities. The corners accumulate most of the liquid getting a thinner liquid film on the flat sides and consequently reaching a low thermal resistance. That effect boosts the average heat transfer coefficient of the channel [34]. Additionally, reducing hydraulic diameters and surface tension cut off the stratified flow in mini-channels and extend the intermittent regime, improving the overall heat transfer coefficient. Flow transitions and patterns are highly dependent on tube shape and diameter [35] in mini-channels. On the other hand, this tube reduction also affects interfacial shear stress increasing pressure drop. This double effect of decreasing tube diameter may reduce the overall efficiency of the biphasic system. Nevertheless, this effect can be suppressed increasing the number of parallel tubes [36]. In this paper, R32 is evaluated as R410A alternative in a multiport mini-channels condenser using experimental data. The main objective of this paper is to experimentally compare R32 and R410A under similar conditions and perform a numerical comparison of both fluids in a multiport mini-channel condenser and compare cycle parameters. The paper is organised as described in the following lines: In Section 2, the two fluids main characteristics are exposed. Next, in Section 3, the purpose-built facility is described. Later, in Section 4 the calculation procedure of frictional pressure drop, heat transfer coefficient and their uncertainties are described. The experimental data is presented in Section 5 grouped by fluid, mass velocity and saturation temperature. Then, in Section 6 both fluids are compared and analysed under similar conditions. Finally, the conclusions of the study are summarised.

2. Properties of the refrigerants analysed R-410A is a zeotropic, but near-azeotropic mixture of difluoromethane (CH2F2, called R-32) and pentafluoroethane (CHF2-CF3, called R-125), which is typically used as a refrigerant in air conditioning applications. Difluoromethane is an organic compound of the dihalogenoalkane variety. It is based on methane, but fluorine atoms have replaced two of the four hydrogen atoms. Both fluids present null Ozone Depletion Power (ODP) while the GWP value of R32 is one-third of R410A value, which makes the use of R32 environmentally safer. The thermofluid properties of both fluids are presented in Table 1. Safety classification of both refrigerants is discussed in the next lines. R410A is classified as A1 by ASHRAE, lower toxicity, and no flame propagation, while R32 classification is A2L, lower toxicity and lower flammability. R32 is a flammable refrigerant on its own, but not when mixed with other components such as R125 to get R410A. In reality, the flammability of R32 refrigerant is very low. The burning velocity (<0.1 m s1) is too slow to cause horizontal flame propagation or explosion. R32 exhibits flame propagation with the heat of combustion lower than 19 kJ kg1.

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exchanger, widely used in refrigerating systems. Table 2 provides the geometric characteristics. Finally, Table 3 shows the features of the measuring devices employed.

Table 1 Tested refrigerants characteristics.

ODP GWP ASHRAE Safety Class Glide (K) Critical Pressure (kPa) Critical Temperature (K) NBP (K) Liquid densitya (kg m3) Vapour densitya (kg m3) Liquid cpa (kJ kg1 K1) Vapour cpa (kJ kg1 K1) Liquid thermal conductivitya (mW m1 K1) Vapour thermal conductivitya (mW m1 K1) Liquid viscositya (mPa s1) Vapour viscositya (mPa s1) Latent heata (kJ kg1)

R410A

Rel. dev.

0 675 A2 0 5782.0 351.26 221.50 1055.77 21.98 1.74 1.25 145.37 11.72

0 2088 A1 0.1 4901.9 344.49 221.71 1170.55 30.43 1.52 1.13 100.09 12.94

18% 2% 0% 1% 27% 14% 11% 45% 9%

150.74 11.51 315.30

165.99 12.10 221.39

9% 5% 42%

32%

4. Data reduction To endorse the experimental measuring technique, the authors followed a methodology similar to that presented by Park et al. [39]. Based on this method, a single-phase tests experimental campaign was performed to characterise the heat transfer coefficient in the water side using R134a within the tube [40–42]. Once acquired the heat transfer coefficient profile in the water side, heat flux q_ j at each location during biphasic flow can be easily calculated:

Q_ j ¼ Q_ w;j ¼ HTC w;j Aouter;j ðT w;j  T wall

Temperature 273 K.

outer;j Þ

ð1Þ

where the local water temperature, T w;j , can be approximated by:

In addition, hermetic refrigerating systems with charge lower than 0.15 kg of refrigerant (A2 or A3) can be installed with no restrictions in occupied environments without additional requirements [37]. This value of refrigerant charge makes mini-channel technology so appealing due to the high volume reduction and thus fluid charge. 3. Purpose built installation and test section The authors developed a purpose-built experimental installation to study biphasic thermofluidynamic processes within minichannels. The objective is to investigate the influence of channel geometry on pressure drop and heat transfer coefficient. Fig. 1 shows a sketch of the measuring section with the main devices employed. The facility allows controlling, measuring and varying all the variables of interest, including vapour quality at the refrigerant inlet. A detailed description of the entire experimental facility, measuring devices, control system and experimental procedure can be found in previous works by the authors [38]. The experimental installation is ATEX classified to test flammable refrigerants. A mini-channel aluminium tube of 307 mm length composes the test section. It is configured as a counter current heat exchanger; the refrigerant condensates inside the tube with water which flows in the annular chase. It has two adiabatic sectors, one at the inlet and another one at the outlet, of 24 mm length. The tested geometry presented in this paper has a hydraulic diameter of 1.16 mm. Ten square ports compose the minichannel cross section. This tube is typically used in parallel-fin parallel-flow (PF2) heat

Tw out

ðT wall

inner Þj

¼ ðT wall

outer Þj



q_ j  t kAl

ð3Þ

At each thermocouple location, the local heat transfer coefficient of the refrigerant can finally be computed using the local heat flux, the local temperatures of refrigerant and inner wall using Eq. (4).

HTC ref ;j ¼

q_ j T ref ;j  ðT wall

ð4Þ

inner Þj

Local refrigerant temperature, T ref ;j , is computed from the local saturation pressure considering saturated conditions. Since the tube is not so long, low values of pressure drop were recorded during the experimental tests. The total pressure drop of the tube was experimentally recorded. The authors used a previously developed model [43] to calculate local saturation pressures.

Pref in

Coriolis Effect Mass Flow Meter

Detail Secon A-A’ Units in mm

28

Twall9

Twall8

Twall7

Twall6

Twall5

Twall4

Twall3

Twall2

Twall1

ð2Þ

The system formed by Eqs. (1) and (2) is solved using an iterative process that starts considering a linear evolution of temperature profile in the first calculation step. The process converges rapidly to variations between successive repetitions lower than 0.5%. The local water properties were calculated on each iteration because they vary with temperature. The complete procedure is schematically represented in Fig. 2. Since the outer tube wall, T wall outer , is measured directly with thermocouples, the inner wall temperature can be computed with the known heat flux:

Test secon

Tref in

Tref out

!

1.9

Tw in

A-A’

Coriolis Effect Flow Meter

T w;j

Q_ w;j1 þ Q_ w;j ¼ T w;j1 þ _ w C p;w 2m

11

a

R32

Dp

19 Fig. 1. Installation sketch and water chase.

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Table 2 Test section geometrical parameters. Tube model

Flow area (mm2)

Outer perimeter (mm)

Inner perimeter (mm)

Ports

D (mm)

Ra (lm)

Relative roughness (–)

Square ports

12.54

40.17

43.22

10

1.16

0.226

3.89  104

Table 3 Sensors used in the experimental tests.

a

Device

Variable measured

Type

Measuring range

Accuracy

Absolute pressure transmitter Differential pressure transmitter Mass flow meter Electromagnetic flow meter

Pref DP _ ref m V_ w

Piezoelectric Piezoelectric Coriolis Electromagnetic

1–2500 kPa 0–100 kPa 0–0.03 kg s1 0–0.6 m3 h1

0.04%a 0.065%a 0.05% 0.25%

Resistive temperature detector Thermocouple

Tref & Tw Twall_i

Pt 100 1/10 DIN T Class 1

50 to 350 °C 30 to 125 °C

0.03 ± 5  104 T [°C] ±0.5 [°C]

Accuracy based on measuring range.

Twall_outer,j

HTCw,j

Tw,j (A)

(Experimentally measured)

(Imposed from single phase tests)

(Inially linear profile assumed)

and no strange effect is appreciated on heat transfer coefficient measurements with inlet vapour quality change. 4.1. Pressure drop calculation

· · Qj=Qw,j (eq. 1)

Tw,j (B) (eq. 2)

E(%)>0.5 E(%)=100*[Tw,j(A)-Tw,j(B)]/Tw,i(A)

Tw,j (A)=Tw,j (B)

E(%)<0.5

Twall_inner,j (eq. 3)

Four different terms compose the biphasic pressure gradient experimentally measured. Firstly, the gravitational pressure gradient, which takes into account the effect of gravity. In this case, this term is null because of the horizontal configuration of the measuring section. Secondly, momentum pressure drop, which considers the effect of the vapour quality change during fluid condensation. Thirdly, sudden contraction and expansion pressure losses must also be taken into account. This term includes the effect of the area change at the inlet (contraction) and the outlet (expansion) of the test section. Very different values of magnitude are acquired in the case of single and two-phase flow expansions and contractions. The models used to calculate these losses are Coleman and Krause [44], Schmidt and Friedel [45] for biphasic flows and Kays and London [46] and Abdelall et al. [47] for single phase flows. Finally, the frictional pressure gradient, which is the one used to compare the behaviour of fluids

          dp dp dp dp dp  ¼    dz biphasic dz grav dz mom dz acc dz friction

ð5Þ

HTCref,j (eq. 4)

A maximum change of vapour quality of 0.2 was achieved in the campaign. Therefore, several tests were made to measure the entire range of vapour quality. The authors fixed values of mass velocity, then saturation temperature and vapour quality were changed. The inlet vapour quality in the first test was set to 0.95 and exit value of vapour quality registered was 0.75. To guarantee a minimum overlapping between two successive experimental measurements, the inlet vapour quality in the second test was set to 0.8. The process was repeated until measuring the whole range of vapour quality. Fig. 3 shows several tests performed by the authors at a fixed value of saturation temperature and mass velocity. In those tests, the authors varied the entrance vapour quality. The experimental results depicted in that image show that the experimental records of heat transfer coefficients are not affected by two-phase flow distribution at the inlet of the test section. The trend is monotonic,

Heat Transfer Coefficient [Wm-2K-1]

Fig. 2. HTC calculation process.

Tsat = 40ºC - G = 470 kg m-2 s-1

12000

Xin= 0.94 Xin= 0.83 Xin= 0.58

10000

Xin= 0.42

8000 6000 4000 2000 0

0

0.2

0.4

0.6

0.8

1

1- x [-] Fig. 3. R134a condensing heat transfer coefficient measured with different vapour quality entrance.

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The momentum pressure term for condensing flow is computed using the Eq. (6)

" #   dp d x2 ð1  xÞ2 2  ¼G  þ dz mom dz a  qv ð1  aÞ  qf

ð6Þ

Different void fraction equations are presented in Winkler et al. [48]. The momentum pressure gradient value is almost negligible, reaching in the highest case values lower than 6% of experimental records. 4.2. Uncertainty analysis Taylor and Kuyatt [49] methodology was followed to estimate the uncertainty propagation of the derived variables such as heat transfer coefficient and pressure drop. Fluid properties used in Eqs. (1)–(6) were computed with IMST ART [50]. The uncertainty calculation of the heat transfer coefficient was performed from Eq. (4). The same procedure was used to get the frictional pressure drop uncertainty

uHTC ref ;j

vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi !2 u   u @HTC 2 @HTC ref ;j 2 2 @HTC ref ;j ref ;j ¼t u2q_ ij þ uT ref ;j þ u2T ðwall innerÞ j @ q_ j @T ref ;j @T ðwall innerÞj

values can also be found in a previous publication of the authors [38]. Fluid properties can explain the differences in the experimentally measured pressure drop for both refrigerants. Pressure drop is directly proportional to friction coefficient and inversely proportional to density. Whereas the liquid viscosity of R32 is quite close to that of R410A, the vapour viscosity is clearly lower. So, in almost all vapour quality range, the biphasic viscosity of R32 is lower than R410A and therefore, for a fixed value of mass flow rate, the Reynolds number of R32 is higher, which leads to higher friction coefficients. At the same time, the density of R32 is lower than that of R410A and therefore the pressure drop is clearly higher for R32. Since the differences in fluid properties are clearer as vapour quality increases, the differences in pressure drop are more accentuated for high vapour quality values. Fluid properties also explain the influence of temperature. Both fluids follow a similar behaviour: as temperature decreases, liquid density and viscosity increase, whereas vapour density and vapour viscosity decrease. So, at high vapour quality values, pressure drop clearly increases as temperature decreases being this effect less clear as vapour quality decreases. Experimentally recorded data of heat transfer coefficient are depicted in Fig. 5. The tendency of all experimental data is similar:

ð7Þ

Table 4 summarises the expanded uncertainty results with a coverage factor K = 2, which represents a level of confidence of 95%. From the process described, water to wall and wall to refrigerant temperature differences uncertainty is the larger contribution to the heat transfer coefficient uncertainty. 5. Experimental results Multiple parameters were varied during the experimental campaign to capture their influence on heat transfer coefficient and frictional pressure drop. The parameters modified during tests were: the entrance vapour quality [from 0.95 to 0.1], the mass flow rate [from 350 to 710 kg m2 s1], the inlet saturation temperature [from 30 to 50 °C] and the condensing heat flux [from 0.77 to 4.69 kW m2]. Fig. 4 shows the mass velocity variation effect and refrigerant temperature on the frictional pressure gradient for both fluids. Uncertainty data is also plotted in that Figures for both fluids and one value of mass velocity to show the nature of the measurements. Based on the experimental results, the conventional theory explains the biphasic frictional pressure gradient. Thus, the frictional pressure drop decreases with decreasing values of vapour quality and mass velocity. Similar results were registered at different saturation temperatures. The analysis of other mass velocity

Table 4 Expanded uncertainties of measured variables. Parameter

Expanded uncertainty (%)

Heat Flux Vapour quality Heat transfer coefficient Water heat transfer coefficient Inlet saturation pressure Frictional pressure drop

3.4–4.5 2.2–12.5 5.6–21.7 3.2–11.3 1.6–3.4 2.2–11.9

Pressure drop comparison at 45ºC 100

Frictional Pressure Drop [kPam-1]

ð8Þ

a

G = 350 kgm-2s-1 G = 470 kgm-2s-1 G = 590 kgm-2s-1 G =710 kgm-2s-1 R410A R32

90 80 70 60 50 40 30 20 10 0

0

0.2

0.4

0.6

0.8

1

1−x [−]

b 100

Frictional Pressure Drop [kPam-1]

udpf

vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi ffi !2 u  2  2 u @dp 2 @dpf @dpf @dpf f t 2 2 2 2 ¼ udpmeas þ udpexpn þ udpcont þ udpmom @dpmeas @dpexpn @dpcont d@pmom

Tsat effect on pressure drop at G = 710 kg m-2s -1 Tsat = 30ºC Tsat = 40ºC Tsat = 50 ºC R410A R32

90 80 70 60 50 40 30 20 10 0

0

0.2

0.4

0.6

0.8

1

1−x [−] Fig. 4. Mass velocity and refrigerant temperature influence on pressure drop.

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a clear increase of heat transfer coefficient as vapour quality or mass velocity increases. Heat transfer coefficient is also affected by refrigerant temperature, enlarging their values with decreasing saturation temperatures. Fluid properties can also explain the differences in the behaviour of heat transfer coefficient between both fluids. R32 seems to present higher heat transfer coefficient than R410A because of its higher thermal conductivity. This effect is not such clear due to the higher uncertainty values recorded at higher vapour qualities. The flow pattern in mini-channels is mostly annular or intermittent, so liquid phase is displaced to the corners with the thin liquid film on the walls. The thin liquid film of R32 presents higher liquid thermal conductivity than R410A that leads to lower thermal resistance and thus to higher heat transfer coefficient. This effect is clearer as saturation temperature increases. The data presented in Figs. 4 and 5 are also compared against two models available in the open literature. Experimental data of heat transfer coefficient are compared against Kim et al. model [51] in Fig. 6a. In Fig. 6b, the readers can find the same comparison with experimental frictional pressure gradient data using LópezBelchí et al. model [43].

a

6. R32 and R410A condenser assessment From the results presented in Fig. 4 it is clear that the pressure drop for R32 is slightly higher than for R410A but, on the other side, the heat transfer coefficient is higher for R32. Additionally, with the saturation temperatures tested, latent heat for R32 is between 40 and 50% higher than for R410A so, for a given power demand, mass velocities of R32 are typically much lower than R410A values. Since from an energy efficiency analysis all these effects must be taken into account, a deeper analysis is necessary to assess which fluid performs better in a condenser. Based on the experimental records provided in this paper and other results previously published [38], a simulation of refrigerant condensation within a multiport mini-channel tube was performed (see Fig. 7). The simulation was done with a tube of 1-meterlength, and the same geometrical characteristics as the one experimentally tested included in Table 2. The numerical setup was configured as a counter current heat exchanger with one mini-channel multiport tube cooled with water at different temperatures. In a first step, the water inlet temperature and water and refrigerant mass flow rates are known; the refrigerant enters into the heat exchanger as saturated vapour and leaves as saturated liquid,

HTC Comparison Tsat = 45ºC

10

Experimental HTC [kWm-2K-1]

8 7 6 5 4 3

R410A R32

2

5000 - 20%

4000 3000 2000 1000

1 0

+ 20%

6000

Calculated HTC [Wm-2K-1]

G = 350 kgm s G = 470 kgm-2s-1 G = 590 kgm-2s-1 G = 710 kgm-2s-1 R410A R32

9

Kim et al. model (2000)

7000

-2 -1

0

0

0.2

0.4

0.6

0.8

0

1000

1

2000

3000

4000

5000

6000

7000

Experimental HTC [Wm-2K-1]

1-x [-] Fig. 6a. Calculated vs experimental for HTC measurements.

10

Tsat = 30ºC Tsat = 40ºC Tsat = 50ºC R410A R32

Experimental HTC [kWm-2K-1]

9 8 7 6 5 4 3 2 1 0

0

0.2

0.4

0.6

0.8

1

1−x [−]

Calculated Frictional Pressure Drop [kPa]

b

Tsat effect on heat transfer coefficient at G = 710 kg m-2s-1 López-Belchí et al. model (2014)

140 R410A R32

120

+ 20%

100 - 20%

80 60 40 20

0

0

20

40

60

80

100

120

140

Experimental Frictional Pressure Drop [kPa] Fig. 5. Mass velocity and refrigerant temperature influence on heat transfer coefficient.

Fig. 6b. Calculated vs experimental for frictional pressure drop measurements.

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Tw

out

¼ Tw

in

þ

_ ref  ½href m

in ðP ref ; x

¼ 1Þ  href out ðPref ; x ¼ 0Þ _ w  C p;w ðT w in Þ m

ð9Þ

In a second step, the tube is divided into one hundred cells and an iterative calculation process is performed. The objective is to obtain the inlet refrigerant conditions in the tube to get the saturated liquid state at the outlet. The heat transferred from the refrigerant to the water in each cell ‘‘j” can be calculated solving the following system of equations

(a)

4000

Heat Transfer Coefficient [Wm-2K-1]

HTC ΔTref-Tw

30

R32

3500

28

3000

26

2500

24

2000

22

1500

1000

20

R410A

0

0.2

0.4

0.6

ΔTref-Tw

assuming that, in this first step, the saturation pressure remains constant (P ref in ¼ P ref out ¼ P refsat ðT refin Þ ¼ Pref ). In that way, an approximate value of water outlet temperature can be obtained as

18 1

0.8

Tube length [m]

(b)

1

X Qaggregated

0.9

800 700

0.8 600

R32

500

X [-]

0.6

400

0.5 0.4

300

Qaggregated [W]

0.7

0.3 200 R410A

0.2

100

0.1 0

0

0.2

0.4

0.6

0.8

0 1

Tube length [m]

(c)

Pressure drop [Pa]

100

60 Pressure drop Tube temperature

R32

59 58

90

57

80

56 70 55 60 54 50

53

40

52 R410A

30 20

Tube Saturation Temperature [ºC]

110

51 0

0.2

0.4

0.6

0.8

1

50

Tube length [m]

Fig. 7. Block diagram of the computational process.

Fig. 8. Evolution of biphasic refrigerant flow while condensing within multiport mini-channel tube. Numerical test at G = 350 kg m2 s1and Tw = 30 °C.

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Q_ j ¼ Q_ w;j ¼ HTC w;j Aouter;j ðT w;j  T wall

outer;j Þ

Q_ j ¼ Q_ ref ;j ¼ HTC ref ;j Ainner;j ðT ref ;j  T wall kAl Q_ j ¼ Q_ cond;j ¼ Aint;j ðT wall t

inner;j

ð10Þ

inner;j Þ

 T wall

ð11Þ

outer;j Þ

ð12Þ

that cell, the pressure drop, the heat transferred from wall to water and thus from the refrigerant to water are computed, obtaining the outlet values which are the inlet values to the next cell. This process is done over until the end of the tube. Then the total heat transferred to the refrigerant side is compared to the total heat transferred to the water side and, if the difference exceeds the

where the equation employed in the simulation to predict the HTC in the water side is an adjustment to the results obtained in previous tests developed using R134a within the tube [40–42]. Kim et al. model [51] was used to evaluate the HTC in the refrigerant side. In the same way, the pressure drop can be easily computed in each cell. The model used for frictional pressure drop calculation was the model developed by the authors [43]. Since that model was not developed for R410A, the accuracy of the predictions obtained for this fluid are worse than for R32, as shown in Fig. 6b. In order to improve this accuracy for R410A predictions, a new value for the ‘‘C” parameter was employed only for R410A

p pcrit

7:1748

 Re0:4387 l

ql qv

5:7189

X 0:4243 tt

Heat Power/Compressor Power

C ¼ 3:7056  105



5

ð13Þ

The calculation is performed iteratively. From the initial calculation, water outlet conditions and refrigerant inlet conditions were obtained. Inasmuch as the tube is configured as a countercurrent heat exchanger, water and refrigerant conditions are known in the first cell. Therefore, in the first cell, the refrigerant temperature is known and the vapour quality is fixed to unity. In

R32 R410A Tw = 20ºC Tw = 25ºC Tw = 30ºC Tw = 35ºC

4.5

4

3.5

3

2.5

2 300

350

400

450

500

550

600

650

700

750

Mass velocity [kgm-2s-1] Fig. 10. Evolution of heat to power ratio at different mass velocities and water temperatures.

4

1800

Power [W]

1600

3.5 R32 R410A Tw = 20ºC Tw = 25ºC Tw = 30ºC Tw = 35ºC

x 10

3 Pressure drop [Pa]

2000

1400 1200

R32 R410A Tw = 20ºC Tw = 25ºC Tw = 30ºC Tw = 35ºC

2.5 2 1.5

1000

1

800 600 300

350

400

450

500

550

600

650

700

0.5 300

750

350

400

450

550

600

650

700

750

(b)

(a) 60

5000

55 50 45 40

R32 R410A Tw = 20ºC Tw = 25ºC Tw = 30ºC Tw = 35ºC

35 30 300

350

400

450

500

550

600

650

700

750

Average HTC [Wm-2K-1]

Saturation Temperature [ºC]

500

Mass velocity [kgm-2s-1]

Mass velocity [kgm-2s-1]

4500

4000

3500 R32 R410A Tw = 20ºC Tw = 25ºC Tw = 30ºC Tw = 35ºC

3000

2500 300

350

400

450

500

550

600

650

Mass velocity [kgm s ]

Mass velocity [kgm-2s-1]

(c)

(d)

-2 -1

700

750

Fig. 9. (a) Heat power variation with mass velocity at different water temperatures. (b) Saturation temperature needed for complete condensation. (c) Pressure drop computed with different water temperatures. (d) Average HTC computed with different water temperatures.

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analysis. This Figure shows the ratio between the power exchanged in the condenser and the power consumed by the compressor assuming ideal compressor (g = 1), an evaporation temperature of 10 °C and 0 °C of superheating at the exit of the evaporator. According to this Figure, for both fluids, the increasing rate of the total thermal power dissipated is lower than the increasing rate of the compressor power consumption. Therefore, the ratio heat power to compressor power decreases as mass velocity increases. Many other aspects have an influence on the system performance (compressor technology, lubrication, refrigerant charge and its influence over subcooling and superheating) that have not been considered in the analysis shown in Fig. 10. From Fig. 10 it is clear that, for low mass velocities, it is always better to use R32 than R410A because it presents higher ratio between the heat dissipated and the power consumption. As mass velocity increases and the efficiency of the system decreases, R410A starts to behave better than R32. Initially only at low water temperatures and finally, when the mass velocity is very high, R410A behaves always better. On the other hand, Fig. 11 shows that, independently of water temperature, a given heat power demand can always be achieved with lower mass velocity using R32 (Fig. 11a) and then the heat to power ratio is better using R32 (Fig. 11b). Therefore, for a given heat power demand, R32 per1500

Heat Power [W]

1400 1300 1200 1100 1000 R32 R410A Tw = 20ºC Tw = 35ºC

900 800 200

300

400

500

600

700

800

900 1000 1100

Mass velocity [kgm-2s-1]

(a) 5

Heat Power/Compressor Power

maximum error allowed, a new value for the water outlet temperature is imposed and the whole procedure is repeated until both values match. Fig. 8 summarises the process mentioned above. The results of the computation are depicted in Fig. 8, which provides extensive knowledge about the evolution of the condensing refrigerant and water temperatures and also about heat transfer coefficient, pressure drop, heat power and vapour quality profiles along the tube for both refrigerants. The images obtained from the simulation show that the mean HTC for R32 is around 31% higher than for R410A, which together with its higher heat of vaporisation, leads to a total heat power that is around 38% higher when using R32. So, despite its higher HTC, the LMTD must be higher for R32, and therefore, the saturation temperature is also higher. Although this higher condensing temperature for R32 has a positive effect on pressure drop, R32 clearly presents higher frictional pressure drop than R410A. In both fluids, the total pressure drop is quite small and its influence on saturation temperature is almost negligible. According to Fig. 9, the same conclusions obtained at G = 350 kg m2s1 and Tw = 30 °C are valid for the whole range of refrigerant mass velocities and water temperatures experimentally studied. In all cases, for a given mass velocity and water temperature, the average HTC and the heat transfer rate are higher for R32. R32 also presents higher pressure drops and higher saturation temperatures. All the parameters analysed in Fig. 9 (heat power, saturation temperature, pressure drop and HTC) increase as mass velocity increases. In all cases, the differences between the refrigerants analysed are clearer at higher mass velocity values. In general terms, an increase in water inlet temperature has an adverse impact on system behaviour, being the influence different depending on the parameter analysed. As expected, for both fluids the saturation temperature (condensing temperature) increases as water inlet temperature increases. In that latter case, the differences between these refrigerants are almost independent of water inlet temperature and refrigerant mass velocity. As a consequence of the increase in condensing temperature, in both fluids the enthalpy of vaporisation decreases and therefore total heat power decreases, being the differences between the two refrigerants lower as water inlet temperature increases at any refrigerant mass velocity. On the other hand, as condensing temperature increases liquid viscosity decreases for both fluids, and then, despite the increase in vapour viscosity, pressure drop decreases. The differences in pressure drop between both refrigerants decrease as water inlet temperature increases at any refrigerant mass velocity. Finally, the heat transfer coefficient presents a different behaviour. Whereas at low refrigerant mass velocities, the HTC increases as water inlet temperature increases, as mass velocity increases this tendency is reversed, being the HTC almost independent of water inlet temperature at mass velocities around 460 kg m2 s1 and then increasing the HTC as water temperature increases for higher refrigerant mass velocities. Most of the conclusions obtained for R290 in previous works [52] are also applicable for R32 and R410Abut, for the two fluids studied in this work, the variation of the heat transfer rate is almost linear but not proportional. Thus, so as to increase the thermal power exchanged in the system it is preferable to increase the number of tubes better than increase mass velocity. Whether done by keeping the number of tubes and increasing the mass velocity or maintaining the mass velocity and increasing the number of tubes, increased power means to increase the mass flow rate through the compressor and thus an enlargement in the compression power, which also must be taken into account. Since the compressor power consumption depends, in addition to fluid properties, on the compressor features, the pressure ratio –which depends on condensation and evaporation pressure–, it cannot be done a universal analysis. Fig. 10 attempts to depict this sort of

R32 R410A Tw = 20ºC Tw = 35ºC

4.5 4 3.5 3 2.5 2 800

900

1000

1100

1200

1300

1400

1500

Heat Power [W]

(b) Fig. 11. (a) Variation of mass velocity for a given heat power at different water temperatures. (b) Heat to power ratio for a given heat power at different water temperatures.

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form better than R410A but, in order to take advantage of this better behaviour, the system must be designed taking into account the differences between both fluids. A direct drop-in replacement of R410A by R32 could lead to both, an increase or a decrease of system efficiency depending on system design and operating conditions. Therefore, based on the numerical model previously presented, the replacement of R410A by R32 is analysed in the following lines. The system is originally designed to dissipate 1000 W working with R410A, maintaining a condensation temperature of 45 °C and assuming a water inlet temperature of 25 °C. With these conditions, the length of tube necessary is 2.38 m, the refrigerant mass velocity must be 570 kg m2 s1, the heat power/compressor power ratio is 3.16 and the total amount of refrigerant contained in the tube is 0.0094 kg. In a direct drop-in replacement, due to the lower density of R32 at compressor inlet (see Table 1), the mass velocity decreases around 27% but, on the other hand, and according to Table 1, the latent heat of R32 increases around 40%. As Fig. 12a shows, despite its lower mass velocity, the heat transfer coefficient of R32 is clearly higher and the heat dissipated increases, reaching a total accumulated value of 1085 W. On the other hand, as Fig. 12b shows, due to the higher amount of heat dissipated, the average temperature of water increases and then the refrigerant temperature increases despite the increase in the heat transfer coefficient. This higher refrigerant temperature leads to a lower heat power/compressor power ratio, which decreases to 3.10. The variation of vapour quality in each cell is directly proportional to the heat dissipated in that cell and inversely proportional to the product of the

latent heat and the mass flow of refrigerant. According to Fig. 12c, the increase in the heat dissipated is compensated by the decrease in the mass flow and the augment in the latent heat, so the evolution of quality is almost the same for both fluids. According to Zivi’s model, the void fraction is a function of the quality and the density ratio (ratio between vapour and liquid density). As Fig. 12.c shows, the quality is almost the same for both fluids, but the density ratio is clearly higher for R410A, leading to lower void fractions in the case of R410A. Thanks to its lower density and its higher void fraction, the total amount of refrigerant contained in each cell is lower for R32 as Fig. 12d shows. In fact, the accumulated value for all the tube is 0.0075 kg. If the length of the tube is kept constant and the compressor is changed so as to obtain the same dissipation rate, the refrigerant mass velocity decreases to 374 kg m2 s1. As Fig. 13a shows, the heat dissipation rate is the same for both fluids and the heat transfer coefficient is clearly higher for R32. Therefore the condensation temperature is lower for R32 (Fig. 13 b). A lower condensation temperature increases the heat power/compressor power ratio up to 3.23. The rest of variables remain stable, except the density ratio of R32, which due to the lower condensation temperature decreases slightly. Thus the total amount of refrigerant contained in the tube is reduced up to 0.0074 kg. Finally, if the system is redesigned to maintain the same dissipation rate with the same condensation temperature, the tube length decreases to 2.3 m and the refrigerant mass velocity increases slightly (375 kg m2 s1). In this case, the heat power/compressor power ratio decreases to 3.20, and the total amount of refrigerant contained in the tube decreases to 0.0072 kg.

(b)

(a)

(c)

(d) Fig. 12. Analysis of direct drop-in of R410A by R32.

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(b)

(a)

(c)

(d) Fig. 13. Analysis of fluid and compressor sustitution.

According to these results, a direct drop-in replacement of R410A by R32 could not always be positive. It will lead to lower direct emissions (lower amount of refrigerant and lower GWP) but higher indirect emission (lower energy efficiency) so the Total Equivalent Warming Impact (TEWI) or the Life Cycle Climate Performance (LCCP) could increase or decrease depending on the specific operating conditions. On the other hand, if the system is specifically designed for R32 or modified to work with R32, both, the direct and the indirect emissions will decrease. 7. Conclusions A pressure drop and heat transfer study of a multiport minichannel aluminium tube with R32 and R410A was completed successfully. The experimentally measured heat transfer coefficients and frictional pressure gradients were analysed under different conditions of mass velocity and saturation temperature. New experimental measurements of the most important variables such as pressure drop and heat transfer in a mini-channel tube are presented. The experimental data were contrasted with models accepted by the scientific community. Besides, a theoretical analysis comparing both refrigerants in a single tube condenser was also performed focusing on the performance of each fluid and the different energy consumption required by the compressor in each case. To summarise the experimental measurements:
Up to the authors’ knowledge, no experimental studies were found working with the two fluids presented here under similar

conditions in multiport tubes. R32 is a flammable refrigerant. Thus, extensive safety precautions must be taken.
The conventional theory explains the experimental frictional pressure gradient. Frictional pressure drops increase with increasing mass velocities and decreasing saturation temperatures. As expected, higher vapour qualities lead to higher frictional pressure gradients.
The heat transfer coefficient experimental measurements augment with the increase of mass velocity and vapour quality. The effect of the saturation temperature is not clearly appreciated. To summarise the numerical analysis performed:
R32 presents higher latent heat, HTC and pressure drop than R410A so, for the same thermal power, R32 mass velocity can be lower than R410A, presenting lower frictional pressure drop and higher heat to compressor power rates than R410A.
For a fixed heat power demand R32 performs better than R410A thanks to its lower mass velocity required and better heat to power ratio.
In the range covered by the experimental tests developed, the direct drop-in replacement of R410A by R32 could be positive or negative depending on operating conditions since it always leads to lower refrigerant charge and lower energy efficiency.
If the replacement is performedin a system specifically designed for R32 or modified to operate with R32 the refrigerant charge will decrease, and the energy efficiency will increase, decreasing the environmental impact of the whole system.

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Acknowledgements This study has been carried out within the framework of two research projects; one financed by the Spanish Ministry of Science and Innovation (DPI2011-26771-C02-02) and the other by the ‘‘F undaciónSéneca” (ref: 19501/PI/14).

References [1] Pedro Dinis, G, Pedro Dinho da S. Handbook of Research on Advances and Applications in Refrigeration Systems and Technologies (2 Volumes). In: IGI Global, Hershey, PA, USA, 2015, pp. 1–924. [2] Wu M, Yuan XR, Xu YJ, Qiao XG, Han XH, Chen GM. Cycle performance study of ethyl fluoride in the refrigeration system of HFC-134a. Appl Energy 2014;136:1004–9. [3] Li H, Cao F, Bu X, Wang L, Wang X. Performance characteristics of R1234yf ejector-expansion refrigeration cycle. Appl Energy 2014;121:96–103. [4] Nunes TK, Vargas JVC, Ordonez JC, Shah D, Martinho LCS. Modeling, simulation and optimization of a vapor compression refrigeration system dynamic and steady state response. Appl Energy 2015;158:540–55. [5] Rasmussen LB, Bacher P, Madsen H, Nielsen HA, Heerup C, Green T. Load forecasting of supermarket refrigeration. Appl Energy 2016;163:32–40. [6] Poggi F, Macchi-Tejeda H, Leducq D, Bontemps A. Refrigerant charge in refrigerating systems and strategies of charge reduction. Int J Refrig 2008;31:353–70. [7] Righetti G, Zilio C, Mancin S, Longo GA. A review on in-tube two-phase heat transfer of hydro-fluoro-olefines refrigerants. Sci Technol Built Environ 2016;22:1191–225. [8] Shah MM. Prediction of heat transfer during condensation of carbon dioxide in channels. Appl Therm Eng 2016;93:192–9. [9] Kuang YW, Wang W, Miao JY, Yu XG, Zhang HX, Zhuan R. Flow boiling of ammonia and flow instabilities in mini-channels. Appl Therm Eng 2017;113:831–42. [10] Liu N, Xiao H, Li J. Experimental investigation of condensation heat transfer and pressure drop of propane, R1234ze(E) and R22 in minichannels. Appl Therm Eng 2016;102:63–72. [11] Mastrullo R, Mauro AW, Thome JR, Vanoli GP, Viscito L. Critical heat flux: Performance of R1234yf, R1234ze and R134a in an aluminum multiminichannel heat sink at high saturation temperatures. Int J Therm Sci 2016;106:1–17. [12] European Parliament and the Council of the European Union. Directive 2006/ 40/EC of The European Parliament and of the Council of 17 May 2006 relating to emissions from air conditioning systems in motor vehicles and amending Council Directive 70/156/EC. Off J Europ Union, vol. 161; 2006. p. 1–11 [2006]. [13] European Parliament and the Council of the European Union. Regulation (EU) No 517/2014 of the European Parliament and the Council of 16 April 2014 on fluorinated greenhouse gases and repealing Regulation (EC) No 842/2006. Off J Europ Union, vol. 150; 2014. p. 195–230 [2014]. [14] Calm JM. The next generation of refrigerants – historical review, considerations, and outlook. Int J Refrig 2008;31:1123–33. [15] Hu H-T, Ding G-L, Huang X-C, Deng B, Gao Y-F. Pressure drop during horizontal flow boiling of R410A/oil mixture in 5 mm and 3 mm smooth tubes. Appl Therm Eng 2009;29:3353–65. [16] Pham HM, Rajendran R. R32 and HFOs as low-GWP refrigerants for air conditioning; 2012. [17] Yajima R, Kita K, Taira S, Domyo N. R32 as a solution for energy conservation and low emission; 2000. [18] Cho IY, Seo H, Kim D, Kim Y. Performance comparison between R410A and R32 multi-heat pumps with a sub-cooler vapor injection in the heating and cooling modes. Energy 2016;112:179–87. [19] Cheng S, Wang S, Liu Z. Cycle performance of alternative refrigerants for domestic air-conditioning system based on a small finned tube heat exchanger. Appl Therm Eng 2014;64:83–92. [20] Xu X, Hwang Y, Radermacher R. Performance comparison of R410A and R32 in vapor injection cycles. Int J Refrig 2013;36:892–903. [21] Alabdulkarem A, Eldeeb R, Hwang Y, Aute V, Radermacher R. Testing, simulation and soft-optimization of R410A low-GWP alternatives in heat pump system. Int J Refrig 2015;60:106–17. [22] Qv D, Dong B, Cao L, Ni L, Wang J, Shang R, et al. An experimental and theoretical study on an injection-assisted air-conditioner using R32 in the refrigeration cycle, Applied Energy, 185. Part 2017;1:791–804. [23] Devotta S, Waghmare A, Sawant N, Domkundwar B. Alternatives to HCFC-22 for air conditioners. Appl Therm Eng 2001;21:703–15. [24] Tian Q, Cai D, Ren L, Tang W, Xie Y, He G, et al. An experimental investigation of refrigerant mixture R32/R290 as drop-in replacement for HFC410A in household air conditioners. Int J Refrig 2015;57:216–28.

[25] He G, Liu F, Cai D, Jiang J. Experimental investigation on flow boiling heat transfer performance of a new near azeotropic refrigerant mixture R290/R32 in horizontal tubes. Int J Heat Mass Transf 2016;102:561–73. [26] Bansal P. High efficiency novel window air conditioner. Appl Energy 2015;156:311–20. [27] Jia L, Jin W, Zhang Y. Analysis of indoor environment safety with R32 leaking from a running air conditioner. Procedia Eng 2015;121:1605–12. [28] E. Commission, Climate-friendly alternatives to HFCs and HCFCs, in: Climate Action; 2017. . [29] Cavallini A, Doretti L, Matkovic M, Rossetto L. Update on condensation heat transfer and pressure drop inside minichannels. Heat Transfer Eng 2006;27:74–87. [30] Jige D, Inoue N, Koyama S. Condensation of refrigerants in a multiport tube with rectangular minichannels. Int J Refrig 2016;67:202–13. [31] Kays WM, London AL. Compact heat exchangers; 1984. [32] Kandlikar SG, Grande WJ. Evolution of microchannel flow passagesthermohydraulic performance and fabrication technology. Heat Transfer Eng 2003;24:3–17. [33] Kew PA, Cornwell K. Correlations for the prediction of boiling heat transfer in small-diameter channels. Appl Therm Eng 1997;17:705–15. [34] Bortolin S, Da Riva E, Del Col D. Condensation in a square minichannel: application of the VOF method. Heat Transfer Eng 2014;35:193–203. [35] Coleman JW, Garimella S. Characterization of two-phase flow patterns in small diameter round and rectangular tubes. Int J Heat Mass Transf 1999;42:2869–81. [36] Fernando P, Palm B, Lundqvist P, Granryd E. Propane heat pump with low refrigerant charge: design and laboratory tests. Int J Refrig 2004;27:761–73. [37] B. British Standard, EN 378–1. Refrigerating Systems and Heat Pumps e Safety and Environmental Requirements. BSI, 68; 2008. [38] López-Belchí A, Illán-Gómez F, Cascales JRG, García FV. R32 and R410A condensation heat transfer coefficient and pressure drop within minichannel multiport tube. Experimental technique and measurements. Appl Therm Eng 2016;105:118–31. [39] Park JE, Vakili-Farahani F, Consolini L, Thome JR. Experimental study on condensation heat transfer in vertical minichannels for new refrigerant R1234ze(E) versus R134a and R236fa. Exp Thermal Fluid Sci 2011;35:442–54. [40] Ramírez-Rivera F, López-Belchí A, Vera-García F, García-Cascales J, IllánGómez F. Two phase flow pressure drop in multiport mini-channel tubes using R134a and R32 as working fluids. Int J Therm Sci 2015;92:17–33. [41] López Belchí A, Illán Gómez F, García Cascales JR, Vera García F. Heat transfer coefficient during condensation inside a minichannel multiport tube with R32 and R410A as working fluids, Science and Technology for the. Built Environ 2015;21:535–44. [42] Illán-Gómez F, López-Belchí A, García-Cascales J, Vera-García F. Experimental two-phase heat transfer coefficient and frictional pressure drop inside minichannels during condensation with R1234yf and R134a. Int J Refrig 2015;51:12–23. [43] López-Belchí A, Illán-Gómez F, Vera-García F, García-Cascales JR. Experimental condensing two-phase frictional pressure drop inside mini-channels. Comparisons and new model development. Int J Heat Mass Transfer 2014;75:581–91. [44] Coleman JW, Krause PE. Two phase pressure losses of R134a in microchannel tube headers with large free flow area ratios. In: Experimental Thermal and Fluid Science The International Symposium on Compact Heat Exchangers, vol. 28; 2004. p. 123–30. [45] Schmidt J, Friedel L. Two-phase pressure drop across sudden contractions in duct areas. Int J Multiph Flow 1997;23:283–99. [46] Kays WM, London AL. Compact heat exchangers. Krieger Publishing Company; 1984. [47] Abdelall FF, Hahn G, Ghiaasiaan SM, Abdel-Khalik SI, Jeter SS, Yoda M, et al. Pressure drop caused by abrupt flow area changes in small channels. Exp Thermal Fluid Sci 2005;29:425–34. [48] Winkler J, Killion J, Garimella S, Fronk BM. Void fractions for condensing refrigerant flow in small channels: Part I literature review. Int J Refrig 2012;35:219–45. [49] Taylor BN, Kuyatt CE. Guidelines for evaluating and expressing the uncertainty of NIST measurement results. NIST TEchnical Note 1927, United States Department of Commerce Technology Administration. National Institute of Standards and Technology; 2009. [50] Corberán JM. IMST-ART. A simulation tool to assist the selection, design and optimization of refrigeration equipment and components, Computer Software, Investigación y Modelado de Sistemas Térmicos, Instituto de Ingeniería Energética, Universitat Politécnica de Valencia, Valencia, Spain; 2009. [51] Kim N-H, Cho J-P, Kim J-O. R-22 Condensation in flat aluminum multi-channel tubes. Enhanced Heat Transfer 2000;7:427–38. [52] López-Belchí A, Illán-Gómez F, García-Cascales JR, Vera-García F. Condensing two-phase pressure drop and heat transfer coefficient of propane in a horizontal multiport mini-channel tube: experimental measurements. Int J Refrig 2016;68:59–75.