Flow boiling of environmentally friendly refrigerants inside a compact enhanced tube

International Journal of Refrigeration 104 (2019) 344–355

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International Journal of Refrigeration journal homepage: www.elsevier.com/locate/ijrefrig

Flow boiling of environmentally friendly refrigerants inside a compact enhanced tube Giulia Righetti, Giovanni A. Longo, Claudio Zilio, Simone Mancin∗ Department of Management and Engineering, University of Padova, Nano Heat Transfer Lab Str.lla S. Nicola, nr. 3, I-36100 Vicenza, Italy

a r t i c l e

i n f o

Article history: Received 31 December 2018 Revised 6 May 2019 Accepted 25 May 2019 Available online 29 May 2019 Keywords: Microfin tube Flow boiling Heat transfer Pressure drops R1233zd(E) R245fa Low GWP ORC HFO

a b s t r a c t R1233zd(E) has been proposed as low GWP alternative to R245fa. This paper compares the R1233zd(E) and R245fa performance during flow boiling inside a small diameter microfin tube having internal diameter at the fin tip equal to 4.2 mm, 40 fins, 0.15 mm high, and a helix angle of 18°. The experimental measurements were carried out at a constant mean saturation temperature of 30 °C, by varying the refrigerant mass velocity between 100 kg m−2 s−1 and 300 kg m−2 s−1 , the mean vapor quality from 0.1 to 0.95, and the heat flux from 15 to 90 kW m−2 . The experimental results showed that R1233zd(E) presented almost similar flow boiling heat transfer coefficients at all the investigated mass velocities but slightly higher two-phase frictional pressure drops. The experimental measurements were compared against the calculated values of boiling heat transfer coefficients and frictional pressure drops from several models selected from the open literature. © 2019 Elsevier Ltd and IIR. All rights reserved.

Ébullition en écoulement de frigorigènes respectueux de l’environnement à l’intérieur d’un tube compact amélioré Mots-clés: Tube à micro-ailettes; Ébullition en écoulement; Transfert de chaleur; Chutes de pression; R1233zd(E); R245fa; Faible GWP; ORC; HFO

1. Introduction The subject of this paper is the new hydrochlorofluoroolefin (HCFO) molecule R1233zd(E). R1233zd(E) has recently been proposed to substitute R245fa and other low-pressure, high-GWP fluids in many applications (see for instance McLinden et al., 2014), thanks to its very low GWP (GWP<7), a near zero ODP (ODP= 0.0 0 024–0.0 0 034), a very short atmospheric lifetime (about 26 days)), a non-flammable behavior, and favorable thermophysical properties (Romeo et al., 2017) (see Table 4, where the main R1233zd(E) and R245fa physical properties involved in the heat

Abbreviations: BPHE, Brazed Plate Heat Exchanger; HTC, Heat Transfer Coefficient; ID, Internal Diameter; OD, Outer Diameter. ∗ Corresponding author. E-mail addresses: [email protected] (G. Righetti), [email protected] (G.A. Longo), [email protected] (C. Zilio), [email protected] (S. Mancin). https://doi.org/10.1016/j.ijrefrig.2019.05.036 0140-7007/© 2019 Elsevier Ltd and IIR. All rights reserved.

transfer, evaluated according to Lemmon et al., 2017, are reported and compared). Arkema (2013) proposed R1233zd(E) as blowing agent, highlighting possible insulation improvements versus other molecules (e.g. about 6% as compared to R245fa). R1233zd(E) seems to be advantageous also if applied to Organic Rankine Cycles (ORCs) working with low temperature (t<150 °C) heat sources. Many authors conducted analytical or experimental investigations, proposing R1233zd(E) as the best drop-in option to R245fa, the nowadays most used fluid in this kind of applications (Li et al., 2017 and Garg et al., 2013). Among them, Yang and Ye (2016) selected R1233zd(E) and two natural molecules (R60 0 and R60 0a) among a group of potential candidates to be the most promising fluids in ORCs that utilize low temperature geothermal sources to produce useful power. The Authors stated that the lower the refrigerant pressure, the lower the operating costs, thus R1233zd(E) should be a good candidate also under an

G. Righetti, G.A. Longo and C. Zilio et al. / International Journal of Refrigeration 104 (2019) 344–355

Nomenclature A cp,w D G h HF HTC I J k L m˙ n p PEL q r s t v V x

area [m2 ] water specific heat capacity [J kg−1 K−1 ] fin tip diameter [m] mass velocity [kg m−2 s−1 ] fin height [m] heat flux [W m−2 ] heat transfer coefficient [W m−2 K−1 ] electrical current [A] specific enthalpy [J kg−1 ] coverage factor [dimesionless] heated length [m] mass flow rate [kg s−1 ] fin number [dimesionless] pressure [Pa] electrical power [W] heat flow rate [W] latent heat of vaporization [J kg−1 ] thickness [m] temperature [°C] specific volume [m3 kg−1 ] potential [V] vapor quality [dimesionless]

Greek symbols β helix angle [°] γ apex angle [°]  variation [dimesionless] λ thermal conductivity [W m−1 K−1 ] μ dynamic viscosity [Pa s] σ surface tension [N m−1 ] Subscripts a momentum cr critical EL electrical f frictional in inlet L liquid loss loss mean mean out outlet pre.evap pre evaporator r refrigerant sat saturation sub subcooled liquid t total TS Test section V vapor w water wall wall

economic point of view. Guillaume et al. (2017) tested a small scale ORC plant working with the heat wasted by a truck through the exhaust gases at a temperature of around 180 °C. Comparing R1233zd(E) and R245fa when working on a cycle based on the same temperature levels, R1233zd(E) led to a lower turbine power output but also to a lower pump consumption, resulting in a comparable net output power. On the other hand, when the authors compared the two fluids on a cycle based on the same pressure levels, R1233zd(E) achieved better results and enabled the turbine to be operated at higher speed and/or with a lower lubrication flow rate which is also a source of losses. Molés et al. (2014, 2016) compared the performance of R1233zd(E) and R245fa

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in a ORC for low temperature heat sources by analyzing about 160 steady-state tests. The authors measured a similar net electrical efficiency ranging from 5% to some 10%, despite a 20% lower mass flow rate circulating when R1233zd(E) is used that caused lower thermal and electrical powers. A similar experiment was run by Eyerer et al. (2016), who tested R1233zd(E) as drop in working fluid in an existing R245fa ORC plant. In this case, the use of R1233zd(E) led to higher heat transfer coefficients and pressure drops in the condenser, while performance in the evaporator was similar for both the refrigerants. The maximum gross thermal efficiency obtained with R1233zd(E) was 5.1%, while it was 4.7% with R245fa. On the other hand, the maximal gross power output was about 12% higher for R245fa. Recently, Yang et al. (2018a,b) presented a thermo-dynamic analysis to investigate the applicability of R1233zd(E) as an alternative to R245fa in an ORC plant. The maximum cycle thermal efficiency was estimated approximately 3.8% higher than R245fa and the output electrical power about 4.5% better than R245fa. Furthermore, the authors carried out an experimental comparison between the two refrigerants that confirmed what expected and finally proposed a semi-empirical method validated on the experimental data. Furthermore, other works published in literature compared the R1233zd(E) and R245fa heat transfer performance. For example, Nagata et al. (2016) experimentally tested R1233zd(E) together with other two low GWP refrigerants, R1234ze(E) and R1234ze(Z) during free convective condensation and pool boiling on a horizontal 19.12 mm OD horizontal plain single copper tube and compared the results obtained versus R245fa. The heat transfer coefficients measured during R1233zd(E) condensation were comparable to that of R245fa, while during pool boiling they were slightly lower. Moreover, Desideri et al. (2017) investigated R1233zd(E) as drop-in alternative to R245fa during vaporization inside a herringbone-type brazed plate heat exchanger at saturation temperatures of 100, 115, and 130 °C, concluding that no major variation can be experienced in the heat exchanger heat transfer coefficient and pressure drop performance. Also Lee et al. (2018) experimentally investigated the heat transfer coefficient of R1233zd(E) in a plate heat exchanger evaporator with respect to the mass flux, heat flux, saturation temperature, and vapor quality. The Authors concluded that the average heat transfer coefficient of R1233zd(E) was about 9% lower than that of R245fa because of its lower thermal conductivity. While the two refrigerants presented comparable pressure drops in the low vapor quality region. On the other hand, in the high vapor quality region due to larger velocity difference between the vapor and liquid phase the R1233zd(E) pressure drops were higher than the R245fa ones. Huang and Thome (2017) examined R1233zd(E), R245fa, and R236fa two phase pressure drops in a multi microchannel evaporator (10 mm long and 10 mm wide, having 67 parallel channels, 100 × 100 μm2 each) at mass fluxes from 1250 to 2750 kg m−2 s−1 and heat fluxes from 200 to 640 kW m−2 . R236fa exhibited the lowest pressure drops due to its lower liquid to vapor density ratio and liquid viscosity, while those of R1233zd(E) and R245fa were comparable. No heat transfer data were presented. Ju et al. (2017) proposed some zeotropic mixtures of R1233zd(E) with four hydrocarbons to replace the traditional refrigerants in heat pump water heaters. The results showed that the investigated blends were suitable drop-in candidates of R22 and R134a; moreover, the R1233zd(E)/R1270 with optimal mass fraction 16%/84% obtained the best system COP, which was 2.13% and 10.14% higher than those obtained for R22 and R134a, respectively. More recently, Righetti et al. (2018) measured the R1233zd(E) flow boiling heat transfer coefficients and two-phase frictional pressure drops inside a microfin tube having 4.3 mm internal diameter at the fin tip, 54 fins, 0.12 mm high and an helix angle of 27°.

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Fig. 1. Schematic of the experimental setup. Table 1 Uncertainty analysis. Transducer

Uncertainty (k = 2)

T-type thermocouples T-type thermopiles Electric power Coriolis mass flowmeter (refrigerant loop) Magnetic volumetric flowmeter (hot water loop) Differential pressure transducer (test section) Absolute pressure transducers Flow boiling heat transfer coefficient, HTC Frictional pressure drop, p Vapour quality, xmean

±0.1 K ±0.05 K ±0.26% of the reading ±0.10% of the reading ±0.2% of FS= 0.33 10−3 m3 s− 1 ±0.075% of 0.3 MPa ±0.065% of FS= 4 MPa ±8% ±9% ±0.03

However, no comparisons between R1233zd(E) and R245fa during two phase flow inside a tube are available in the open literature. Furthermore, just few R1233zd(E) data were published during flow boiling inside a microfin tube. At the best authors’ knowledge, only Longo et al. (2017a) and Righetti et al. (2018) presented some R1233zd(E) data collected during flow boiling inside a ID 4.3 mm microfin tube, at 30 °C of saturation temperature, mass velocities from 100 kg m−2 s−1 and 300 kg m−2 s−1 , and heat fluxes from 15 to 90 kW m−2 . This paper presents new experimental data during R1233zd(E) flow boiling inside a mini microfin tube (ID 4.2 mm) which are then compared against those collected, under the similar working conditions, in the same small diameter microfin tube for R245fa. 2. Experimental set up and test section The experimental facility is represented in Fig. 1. It consists of 3 main loops: one where the refrigerant flows, and two for cooling and heating water, respectively. The subcooled liquid is pumped in the refrigerant loop by means of a variable speed gear pump; the mass low rate is measured by means a Coriolis effect mass flow meter, The subcooled liquid is partly vaporized inside a Brazed Plate Heat Exchanger (BPHE) pre-evaporator, fed with hot water, to achieve the desired vapor quality value for the specific test condition. Then the two phase mixture flows into the test section, where it is heated by a calibrated Ni–Cr wire resistance. Finally, it is completely condensed and subcooled in a BPHE condenser fed by tap water. No oil circulates in the refrigerant loop. The saturation pressure is regulated by a damper connected to a compressed air line. The electrical power supplied to the microfin tube is indirectly measured by means of a calibrated reference resistance (shunt) and by the measurement of the effective electrical difference potential of the resistance wire inserted in the copper heater. The current can be calculated from the Ohm’s law. Preliminary tests

were run to verify the heat balance between refrigerant and water sides, the results showed a misbalance always less than 2%. Hot water temperature and flow rate are adjusted independently, in order to guarantee the required heat transfer inside the evaporator. As reported by Fig. 1, water flow rate is measured by means of a magnetic flow meter, a calibrated T-type thermopile is used to measure the temperature difference in the heat exchanger, refrigerant pressure and temperature are monitored in several locations throughout the circuit, to evaluate the refrigerant properties at the inlet and outlet of each heat exchanger. The uncertainties (k = 2) of the installed instruments are listed in Table 1. When the system reached steady state conditions, signals were recorded at 2 Hz and all the data collected over a period of about 100 s were time-averaged to determine the experimental data point. The test section is represented in Fig. 2. It is realized in a copper plate, 200 mm long, 10 mm wide, and 20 mm high. The microfin tube was brazed inside a 8 mm deep guide milled on the top surface. On the bottom face of the copper plate, another guide was milled to hold a Nickel–Chrome wire resistance, which is connected to a DC current generator rated up to 900 W. On the lateral walls of the copper plate, 16 holes, equally spaced, were drilled to locate as many T-type thermocouples to measure the temperature just 1 mm below the microfin tube. The whole measurement section is placed inside an aluminum housing filled with 15 mm thick ceramic fiber blanket, to limit as much as possible the heat losses due to conduction to the ambient. The mini microfin tube under investigation has a 4.2 mm fin tip diameter, 40 fins, 0.15 mm high with an apex angle γ equal to 42°. The helix angle β is 18°, and the area enhancement with reference to the smooth tube having the inner diameter equal to the fin tip diameter is 1.62. The main geometrical characteristics of the microfin tube are listed in Table 2, while Fig. 3 reports a drawing of the tube and Fig. 4 a photo of the cross section, where the micro fins can be observed.

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Fig. 2. Drawing of the copper plate with the locations of the thermocouples.

Table 2 Microfin tube geometry.

Fig. 3. Schematic of the microfin tube.

Parameter

Value

Outer diameter Fin tip diameter, D Number of fins, n Fin height, h Apex angle, γ Helix angle, β Tube thickness, s

5 mm 4.2 mm 40 0.15 mm 42° 18° 0.25 mm

Pressure taps are located about 25 mm downstream and upstream of the copper plate, so the length for pressure drop measurements is 250 mm. In order to avoid the abrupt pressure drops caused by flow contraction and expansion, a suitable smooth connection to the refrigerant circuit having the same fin tip diameter (D = 4.2 mm) was designed and realized to join the test tube with inlet and outlet pipes. 3. Experimental measurements and data reduction The detailed data reduction process is presented in Righetti et al. (2018) and Longo et al. (2017b,c). The specific enthalpy at the inlet of the test section Jin,TS , can be easily calculated from a thermal balance at the brazed plate pre-evaporator, which is fed with hot water, as given by:

qevap = m˙ w · c p,w · (tw,



in

− tw,out ) = m˙ r · Jin,T S − JL,sub



(1)

where JL, sub is the specific enthalpy of the subcooled liquid entering the brazed plate pre-evaporator and it is evaluated from the measured values of pressure and temperature at the inlet of the pre-evaporator. Then, the vapor quality at the inlet of the test section, xin, TS , is estimated as in the below equation.

xin,T S =

Fig. 4. Photo of a cross section of the tested microfin tube.

Jin,T S − JL JV − JL

(2)

A thermal balance on the measurement section leads to the estimation of the vapor quality at the outlet xout, TS , which was always less or equal to 1. The average value between xin, TS and xout, TS is taken as reference mean vapor quality, xmean . The two-phase heat transfer coefficient HTC, referred to the nominal area A of a smooth tube having the same diameter of the

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G. Righetti, G.A. Longo and C. Zilio et al. / International Journal of Refrigeration 104 (2019) 344–355 Table 3 Operating conditions during experimental tests. Runs

HF [kW m− 2 ]

G [kg m− 2 s− 1 ]

xin [dimensionless]

xout [dimensionless]

R1233zd(E)

80

15

100, 150, 20 0, 30 0 100, 150, 20 0, 30 0 150, 200, 300 20 0, 30 0 100, 150, 20 0, 30 0 100, 150, 20 0, 30 0 150, 200, 300 20 0, 30 0

0.04–0.91

0.20–0.97

0.05–0.16

0.07–0.90

0.21–1.0

0.10–0.32

0.01–0.79 0.03–0.70 0.05–0.89

0.3–1.0 0.42–1.0 0.17–0.97

0.21–0.41 0.30–0.45 0.05–0.15

0.07–0.90

0.23–1.0

0.10–0.30

0.06–0.79 0.02–0.70

0.28–1.0 0.42–1.0

0.21–0.41 0.30–0.46

30

R245fa

60 90 15

80

30 60 90

fin tip one of the mini microfin tube, is defined as:

HT C =



qT S

A · t¯wall − t¯sat

=

π

qT S   · D · L · t¯wall − t¯sat

(3)

where D is the fin tip diameter, L is the heated length (200 mm) while t¯wall and t¯sat are evaluated as following:

t¯wall

x [dimensionless]

Fluid

16 tsat,in ( psat,in ) + tsat,out ( psat,out ) 1  = twal l ,i and t¯sat = 16 2

(4)

Table 4 Main R1233zd(E) and R245fa thermophysical properties involved in heat transfer and fluid flow at tsat = 30 °C evaluated according RefProp v.10 (Lemmon et al., 2017). Properties

R1233zd(E)

R245fa

Critical pressure, pcrit [bar] Saturation pressure, psat [bar] Density, ρ [kg m−3 ]

36.24 1.55 1250.6 8.5 0.081 0.011 272.1 10.2 13.91 188.52 0.0427

36.51 1.78 1324.8 10.1 0.087 0.013 374.8 10.5 12.98 188.33 0.0488

Thermal conductivity, λ [W m−1 K−1 ]

i=1

The actual heat flow rate qTS supplied to the test section is obtained by subtracting from the measured electrical power PEL , the heat loss to the surrounding qloss . Dedicated preliminary tests permitted to estimate the heat loss (qloss ) due to conduction through the test section as a function of the mean wall temperature. The tests were run under vacuum conditions on the refrigerant channel by supplying the power needed to maintain the mean wall temperature at a set value. The measurements were carried out by varying the mean wall temperature from 27 °C to 64 °C. The results demonstrated that the heat loss increases linearly as the mean wall temperature increases (R>0.99). In the tested range of wall temperature, the heat loss by conduction through the test section can be estimated by:

|qloss | = 0.1965 · t¯wall [ ◦ C] − 4.3574 [W]

(5)

The heat loss varied from 1.0% to 4.6% of the total electrical power under the investigated testing conditions; thus, the actual heat flow rate is:

qT S = PEL − |qloss | = V · I − |qloss |

(6)

where I the electrical current and V is the effective electrical difference potential of the resistance wire inserted in the copper heater. The frictional pressure gradient (− ddzp ) f is obtained from the measured total pressure gradient by subtracting the momentum and gravity pressure gradient (− ddzp )a and (− ddzp )g as:



−

dp dz





= f

−

dp dz





− tot

−

dp dz





− a

−

dp dz



(7) g

In this case, since the tube is located horizontally, the gravity pressure gradient (− ddzp )g is equal to 0 Pa m−1 . As already proposed and adopted by Righetti et al. (2018) , the model by Rouhani and Axelsson (1970) is used to estimate the void fraction to calculate the momentum pressure gradients. The void fraction is given by:

ε=

 x 1−x + ρV ρV ρL

−1 1.18(1 − x )[gσ (ρL − ρV )]0.25 + GρL0.5 x



[1 + 0.12(1 − x )]

(8)

Dynamic viscosity, μ [μPa s]

Liquid Vapor Liquid Vapor Liquid Vapor

Surface tension [mN m−1 ] Latent heat of vaporization [kJ kg−1 ] Reduced pressure, pred [dimensionless]

and the associated momentum pressure gradients are:



−

dp dz



= G2 a

d dz



vV x2 vL (1 − x )2 + ε (1 − ε )



(9)

All the thermodynamic and transport properties are estimated from RefProp v.10 (Lemmon et al., 2017). A detailed error analysis was performed in accordance with Kline and McClintock (1953) using the values of uncertainties of the instruments listed in Table 1. As listed in Table 1, the estimated uncertainty on the two-phase heat transfer coefficient is around ±8%, while the uncertainty on the vapor quality is ±0.03 for both the fluids. The pressure drops display a mean uncertainty of around 8% and 9% for R1233zd(E) and R245fa, respectively. 4. Experimental results Two sets of 80 experimental data points each were collected under the same working conditions during the vaporization process of R1233zd(E) and R245fa, respectively, at constant saturation temperature of 30 °C. Table 3 sums up the operating conditions under which the tests were taken. The range of values of heat flux q, mass flux G, inlet vapor quality xin , outlet vapor quality xout , and vapor quality change through the test section x are listed. As also proposed by Righetti et al. (2018), the value of x=0.45 was considered the maximum acceptable to allow for a proper comparison among the collected data; thus, when increasing the heat flux, one or more refrigerant mass velocities were not taken into account because they would have presented larger vapor quality changes. Table 4 lists the main thermo-physical properties of R1233zd(E) and R245fa at 30 °C of saturation temperature evaluated according RefProp v.10 (Lemmon et al., 2017). At a glance, it can be noticed that the R1233zd(E) presents similar properties as compared to R245fa, so it can be considered to be a viable low GWP alternative to R245fa. Both fluids have a critical pressure around

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Fig. 5. Effect of mass velocity on R1233zd(E) heat transfer coefficient at different heat fluxes: HF=15 kW m−2 (a) HF=30 kW m−2 (b), HF=60 kW m−2 (c), and HF=90 kW m−2 (d). G expressed in [kg m− 2 s− 1 ].

3.6 MPa, so at the present operating conditions, they have a very low reduced pressure, around 0.04. Furthermore, they present a relatively high liquid-to-vapor densities (ρ L /ρ V =147 and 131 for R1233zd(E) and R245fa, respectively), and a relatively great surface tension (around 13 mN m−1 ). These thermophysical properties affect the flow boiling heat transfer and, together with the specific working conditions (i.e., saturation temperature, mass velocity, vapor quality, heat flux), they can promote one of the two main heat transfer mechanisms: nucleate boiling and forced convective boiling. In particular, high mass velocity, vapor density, and vapor quality endorse the forced convective contribution while high heat flux and reduced pressure enforce the nucleate boiling. In this particular case, at tsat =30 °C, these fluids having very low vapor density, high surface tension, and low reduced pressure, should promote the forced convective boiling rather than the nucleate boiling. These considerations are then confirmed by the experimental results presented in what follows. Fig. 5 presents the effect of the mass velocity (ranging from 10 0 to 30 0 kg m−2 s−1 ) on R1233zd(E) flow boiling heat transfer coefficient as a function of the mean vapor quality at four different heat fluxes: HF=15 kW m−2 (a), HF=30 kW m−2 (b), HF=60 kW m−2 (c), and HF=90 kW m−2 (d). As listed in Table 3, in order to compare the experimental heat transfer coefficients, the maximum vapor quality change was kept below 0.45; for this reason, at HF=60 kW m−2 and HF=90 kW m−2 just mass velocities greater than 100 kg m−2 s−1 and 150 kg m−2 s−1 , respectively, were collected. Hence, Fig. 5(c) and (d) do not present the complete set of mass velocities.

In Fig. 5(a) and (b) the heat transfer coefficient increases with the vapor quality at each investigated mass velocity, meaning that forced convection is mainly affecting the phase-change process at these low heat fluxes (HF ≤ 30 kW m−2 ). It is worth pointing out that at xmean >0.5, the values heat transfer coefficient measured at G = 200 kg m−2 s−1 and HF=15 kW m−2 are greater than those measured at higher mass velocity; this can be linked to a particular effect the helical microfins that can enhance the flow boiling heat transfer at certain operating conditions, which basically depends upon the combination of fluid (i.e. its thermophysical properties), operating conditions, and tube geometry. Similar results have already been found by Righetti et al. (2018), Longo et al. (2017a,b,c), Jige et al. (2016), and Mancin et al. (2015). Moreover, it can be stated that the slope tends to be reduced by increasing the vapor quality before the dryout. This behavior can also be affected by the saturation temperature drop related to the high two-phase pressure drop, which increases with both mass velocity and vapor quality. In fact, the helical microfins, which, at G = 200 kg m−2 s−1 , play a fundamental role in enhancing the heat transfer without introducing a noticeable fluid dynamic penalization, at G = 300 kg m−2 s−1 lead to great pressure drops. For example, this is clear when one observes the data collected at 300 kg m−2 s−1 and HF=15 kW m−2 : for vapor qualities lower than 0.5, the heat transfer coefficient increase, passing from about 7500 W m−2 K−1 at x = 0.23to about 10,0 0 0 W m−2 K−1 at x = 0.41. At higher vapor qualities, the saturation temperature drop leads to a boiling penalization and the heat transfer coefficient remains almost constant at around 10,0 0 0 W m−2 K−1 . Besides, Fig. 5(c) and (d) present slightly different results. The higher

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Fig. 6. Effect of heat flux on R1233zd(E) heat transfer coefficient at two different mass velocities: G = 200 kg m−2 s−1 (a) and G = 300 kg m−2 s−1 (b). HF in [kW m−2 ].

heat fluxes tend to promote nucleate boiling instead of forced convective boiling mechanism. In fact, even at 90 kW m−2 , only at the maximum mass velocity (G = 300 kg m−2 s−1 ) heat transfer coefficient remains slightly influenced by vapor quality. At the lower mass velocities, the heat transfer coefficients are almost constant, showing a plateau before the dryout. Fig. 6 shows the effect of heat flux on heat transfer coefficient at two mass velocities: 200 kg m−2 s−1 (a) and 300 kg m−2 s−1 (b); these two mass velocities were considered because they present the complete set of investigated heat fluxes. Considering Fig. 6(a), it can be stated

that at heat flux lower than 30 kW m−2 , there is not any noticeable effect of heat flux on the boiling heat transfer, and the heat transfer coefficients increase with vapor quality. Differently, when increasing the heat flux to 60 and then to 90 kW m−2 , the heat transfer coefficient tends to present a reduced slope as a function of the vapor quality. At 90 kW m−2 , the heat transfer coefficient shows a plateau around 11,0 0 0 W m−2 K−1 where it can be considered almost constant, and then it slightly decreases up to the onset of the dryout. In fact, an incipient instability related to the continuous dewetting-rewetting phenomenon is recorded by the wall

Fig. 7. Effect of mass velocity on R245fa heat transfer coefficient at different heat fluxes: HF=15 kW m−2 (a) HF=30 kW m−2 (b), HF=60 kW m−2 (c), and HF=90 kW m−2 (d). G expressed in [kg m−2 s−1 ].

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351

Fig. 8. Effect of heat flux on R245fa heat transfer coefficient at two different mass velocities: G = 200 kg m−2 s−1 (a) and G = 300 kg m−2 s−1 (b). HF expressed in [kW m−2 ].

thermocouples, which penalizes the boiling heat transfer and reduces the heat transfer coefficient. Similar consideration can be drawn when analyzing Fig. 6(b). At heat flux lower than 60 kW m−2 , there is not any noticeable effect of heat flux on the boiling heat transfer coefficient, and the data increases with the vapor quality underlining a prevalence of the convective boiling mechanism. At HF= 90 kW m−2 , the heat transfer coefficient increases, reaching values around 12,500 W m−2 K−1 . In this case, it seems that the nucleate boiling and the two-phase forced convective boiling combine their contributions leading to an enhanced overall performance. Figs. 7 and 8 report the data recorded during R245fa flow boiling inside the same microfin tube. No substantial phenomenological differences from R1233zd(E) can be appreciated. Fig. 7 shows the heat transfer coefficient as a function of the vapor quality at different mass velocities and at four different heat fluxes: HF=15 kW m−2 (a), HF=30 kW m−2 (b), HF=60 kW m−2 (c), and HF=90 kW m−2 (d). At HF=15 kW m−2 and HF=30 kW m−2 (Fig. 7(a) and (b)), it clearly appears that at all the investigated mass velocities, the heat transfer coefficient increases almost linearly with the vapor quality, meaning that the two-phase forced convection is the predominant heat transfer mechanism. Once more, when increasing the heat flux to HF=60 kW m−2 (Fig. 7(c)) heat transfer coefficients measured at the three investigated mass velocities present comparable values and they seem to exhibit similar behavior up to the respective vapor qualities at the onset of dryout. Their slope is limited, indicating a lower forced convective contribution as compared to the lower heat fluxes. Finally, as already highlighted for R1233zd(E) data (Fig. 5(d)), in Fig. 7(d), at HF=90 kW m−2 and G = 200 kg m−2 s−1 , the heat transfer coefficient tends to gradually decrease with vapor quality, while at G = 300 kg m−2 s−1 the heat transfer coefficient slightly increases with vapor quality. Fig. 8 reports the heat transfer coefficient trends as a function of the heat flux for two mass velocities: 200 kg m−2 s−1 (a) and 300 kg m−2 s−1 (b). As previously stated, |x| =0.45 was chosen as the maximum acceptable vapor quality change; thus, just three mass velocities were investigated at HF=60 kW m−2 and two at HF=90 kW m−2 . In general, the data series collected at low heat fluxes show higher slopes as a function of the vapor quality, due to a stronger convective boiling contribution to heat transfer. Furthermore, at low vapor qualities (i.e., x<0.5) the higher the heat flux, the higher the heat transfer coefficient. While at vapor qualities greater than 0.5, thanks to a stronger forced convective boiling contribution, the heat transfer coefficients are less influenced by heat flux, except to the series collected at HF=90 W m−2 and

Table 5 Critical values of vapor quality at the onset of the dryout. Operating conditions

R1233zd(E)

G [kg m−2 s−1 ]

HF [kW m−2 ]

xcr [dimensionless]

100

15 30 15 30 60 15 30 60 90 15 30 60 90

0.78 0.60 0.85 0.82 0.58 0.91 0.85 0.72 0.64 – 0.97 0.90 0.85

150

200

300

R245fa

0.80 0.73 0.85 0.84 0.52 0.90 0.87 0.85 0.61 – 0.92 0.90 0.81

G = 200 kg m−2 s−1 (Fig. 8(a)) where the heat transfer coefficient decreases with vapor quality even before the dryout because of the presence of a persistent heat transfer instability due to the dewetting-rewetting phenomenon. The experimental frictional pressure gradients are plotted in Fig. 9 for R1233zd(E) (a) and R245fa (b), respectively as a function of the mean vapor quality. For the sake of clarity, only data sets relative to an imposed heat flux of 30 kW m−2 are reported. As mentioned before, the Rouhani and Axelsson (1970) model was used to estimate the momentum pressure drops, which were subtracted from the total measured pressure drops. The results show that, at constant mass velocity, the frictional pressure gradient increases with vapor quality. Furthermore, at constant vapor quality, the frictional pressure gradient increases as the mass velocity increases. Finally, the vapor quality at the onset of the dryout was estimated, as described by Righetti et al. (2018), from the analysis of the values of the standard deviation of the temperature readings of the wall thermocouples. Table 5 lists the calculated values. In general, it can be stated that the onset vapor quality at the dryout increases as the mass velocity increases and as the heat flux decreases. In fact, as expected, the higher the heat flux, the earlier the onset of the dryout occurs. 5. Fluids comparison As discussed in the previous section, when looking at the main thermophysical properties involved in the heat transfer and

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Fig. 9. Experimental frictional pressure gradients: (a) R1233zd(E) and (b) R245fa.

Fig. 10. Comparison between R1233zd(E) and R245fa heat transfer coefficient at different mass velocities: G = 100 kg m−2 s−1 (a), G = 150 kg m−2 s−1 (b), G = 200 kg m−2 s−1 (c) and G = 300 kg m−2 s−1 (d) HF=15 kW m−2 , mean saturation temperature= 30 °C.

fluid flow (Table 4), the R1233zd(E) seems to be a proper alternative low-GWP refrigerant for the replacement of R245fa. It is well-known that the flow boiling heat transfer coefficient is mainly related to reduced pressure (0.043 vs. 0.049 for R1233zd(E) and R245fa, respectively), liquid thermal conductivity (0.081 vs. 0.087 W m−1 K−1 for R1233zd(E) and R245fa, respectively), and latent heat (almost similar for the two fluids). So, it can be inferred that the two refrigerants should perform almost similarly. On the other hand, pressure drops are related to the liq-

uid and vapor density ratio, liquid density and viscosity, and the reduced pressure. Again, the two refrigerants should present similar values, but the R1233zd(E) could exhibit slightly higher pressure drops due to lower liquid and vapor densities (i.e., 1250 vs. 1324 kg m−3 and 8.5 vs. 10.1 kg m−3 for liquid and vapor, R1233zd(E) and R245fa, respectively) and a lower reduced pressure. Fig. 10 presents a comparison between R1233zd(E) and R245fa heat transfer coefficients collected at the same mean saturation

G. Righetti, G.A. Longo and C. Zilio et al. / International Journal of Refrigeration 104 (2019) 344–355

353

Fig. 11. Comparison between R1233zd(E) and R245fa heat transfer coefficient at different mass velocities: G = 200 kg m−2 s−1 (a), G = 350 kg m−2 s−1 (b). HF=90 kW m−2 , mean saturation temperature tsat = 30 °C.

Fig. 12. Comparison between R1233zd(E) and R245fa frictional pressure gradients at different mass velocities, HF=15 kW m−2 , mean saturation temperature= 30 °C.

temperature (tsat =30 °C) and heat flux (HF=15 W m−2 ) but at different mass velocities: G = 100 kg m−2 s−1 (a), G = 150 kg m−2 s−1 (b), G = 200 kg m−2 s−1 (c), and G = 300 kg m−2 s−1 (d). Similarly, Fig. 11 shows a comparison between R1233zd(E) and R245fa heat transfer coefficient at the same mean saturation temperature (30 °C), at different mass velocities: G = 200 kg m−2 s−1 (a), G = 300 kg m−2 s−1 (b), at a fixed heat flux equal to 90 W m−2 . Both the figures present similar heat transfer behavior; as expected from the analysis of the thermophysical properties, no noticeable differences can be appreciated between the refrigerants in terms of absolute values, trend, and dryout inception. In general, it can be stated that for these low pressure refrigerants, the two-phase forced convection seems to be very effective as compared to the nucleate boiling. Similar results were also found by Righetti et al. (2018), Longo et al. (2017a), and by Kondou et al. (2014) during flow boiling of R245fa and R1234ze(Z), respectively, inside a different small diameter microfin tubes. On the other hand, Fig. 12 shows the comparison between R1233zd(E) and R245fa frictional pressure gradient at HF=15 W m−2 and tsat = 30 °C. From this standpoint, R1233zd(E) exhibits higher pressure drops (i.e. on average +16%), which can be explained considering the higher shear stress due to its lower vapor density. 6. Models assessment In this final section, the collected experimental measurements were compared against several two-phase heat transfer and

pressure drop correlations available in the open literature and proposed for flow boiling in microfin tubes. The models, listed in chronological order, by Padovan et al. (2011), Diani et al. (2014), and Rollmann and Spindler (2016) were chosen for the flow boiling heat transfer coefficient assessment, while those proposed by Haraguchi et al. (1993), Kedzierski and Goncalves (1999), Cavallini et al. (20 0 0), Miyara et al. (20 0 0), Bandarra Filho et al. (2004), Oliver et al. (2004), Diani et al. (2014), and Rollmann and Spindler (2016) were selected to be compared against the experimental two-phase frictional pressure gradients. Tables 6 and 7 show the relative and absolute deviations of the selected models for two-phase heat transfer coefficient and frictional pressure gradient calculations, respectively. The thermophysical properties required in the correlations were evaluated according to Lemmon et al. (2017). The heat transfer models here applied were also implemented by Righetti et al. (2018) to assess R1233zd(E) in a small microfin tube; in particular, Padovan et al. (2011) model is one of the most used for microfin tubes in literature, the Diani et al. (2014) one was proposed for small diameter tubes, while the Rollmann and Spindler (2016) model is one of the most recently published. Padovan et al. (2011) collected some experimental measures inside a commercial copper microfin tube heated by hot water, having a 7.69 mm diameter at the fin tip (40 fins, 0.23 mm fins height, 13° helix angle and 43° apex angle) by using R134a and R410A as working fluids. A wide range of working conditions was assessed: mass flux from 80 to 600 kg m−2 s−1 , heat flux from 14 to 83.5 kW m−2 , saturation temperature of 30 and 40 °C, and vapor quality from 0.1 to 0.99. The Authors modified the correlation initially proposed by Cavallini et al. (1999) based on a large database of refrigerants (including the low pressure refrigerant R123) and microfin tube geometries. Diani et al. (2014) tested R1234ze(E) inside a mini microfin tube electrically heated with internal diameter at the fin tip of 3.4 mm. The experimental measurements were carried out at constant saturation temperature of 30 °C, and three different heat fluxes: 10, 25, and 50 kW m−2 by varying the refrigerant mass velocity between 190 and 940 kg m−2 s−1 , The Authors modified the Padovan et al. (2011) correlation, since it was not able to satisfactorily estimate the flow boiling heat transfer and pressure drop behaviors of R1234ze(E) at all the investigated operating conditions. So, Diani et al. (2014) proposed a new model version valid for D > 3.4 mm, x < xcr and 150 < G < 940 kg m−2 s−1 . Finally, Rollmann and Spindler (2016) collected 1164 new data on R407C flow boiling in a horizontal copper tube electrically heated having a total fin number of 55 and a helix angle of 15°. The fin height was 0.24 mm and the inner tube diameter at fin

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G. Righetti, G.A. Longo and C. Zilio et al. / International Journal of Refrigeration 104 (2019) 344–355 Table 6 Relative and absolute deviations of the selected models for flow boiling heat transfer coefficient calculation evaluated according to Refprop v.10 (Lemmon et al., 2017). R1233zd(E) Padovan et al. (2011)

Diani et al. (2014)

Rollmann and Spindler (2016)

HF

Rel. Dev. %

Abs. Dev. %

Rel. Dev. %

Abs. Dev. %

Rel. Dev. %

Abs. Dev. %

15 30 60 90 All

−40.2 −36.2 −22.9 −17.5 −30.2

40.2 36.2 23.5 22.2 31.4

−22.7 −29.4 −27.2 −30.3 −27.4

25.6 29.5 28.1 32.0 28.7

−9.7 −7.6 8.6 14.2 0.1

18.0 15.7 23.0 24.8 19.9

20.9 20.7 21.6 25.8 21.8

−18.5 −6.7 6.2 2.7 −5.6

19.6 16.8 24.4 21.8 20.3

R245fa 15 30 60 90 All

−45.3 −35.4 −25.5 −20.3 −33.4

45.3 35.4 25.5 20.8 33.5

−20.9 −19.5 −20.6 −25.6 −21.2

Table 7 Relative and absolute deviations of the selected models for two-phase frictional pressure gradient calculation, evaluated according to RefProp v.10 (Lemmon et al., 2017). Relative deviation Haraguchi et al. (1993)

Kedzierski and Goncalves (1999)

Cavallini et al. (20 0 0)

Miyara et al. (20 0 0)

Bandarra Filho et al. (2004)

Oliver et al. (2004)

Diani et al. (2014)

Rollmann and Spindler (2016)

R1233zd(E) R245fa

−4.1 −8.6

38.0 19.8

21.8 15.3

−29.9 −33.8

−32.6 −30.0

18.8 20.4

19.1 11.7

25.7 4.1

R1233zd(E) R245fa

26.7 29.4

40.4 31.5

26.0 19.8

56.5 60.4

24.9 28.0

21.1 18.7

28.3 18.8

Absolute deviation 31.3 36.4

root is 8.95 mm. The investigated saturation temperatures ranged between −30 and 10 °C. The heat flux was varied between 1 and 20 kW m−2 and the mass velocity from 25 to 300 kg m−2 s−1 . Since the Authors did not find any accurate correlation as compared to their experimental database, they proposed a new procedure. The Rollmann and Spindler (2016) model presents the best agreement when implemented to assess the heat transfer coefficients of both the fluids investigated in the present paper. The relative deviation is around 0.1% and −5.6% for R1233zd(E) and R245fa, respectively. On the other hand, the models by Padovan et al. (2011), Diani et al. (2014) underestimate the experimental values. The heat flux influences the accuracy of the prediction capabilities of the Padovan et al. (2011) and the Rollmann and Spindler (2016) models. Low heat flux data points tend to be underestimated. For instance, the Rollmann and Spindler (2016) model predicts the R1233zd(E) HF=15 W m−2 data within −9.7%, while the HF=90 W m−2 ones within + 14.2%. On the other hand, the Diani et al. (2014) correlation accuracy is less affected by the heat flux, since this model predicts all the data with a similar relative deviation. As far as frictional pressure gradient is concerned, the models by Haraguchi et al. (1993), Oliver et al. (2004), Diani et al. (2014), and Rollmann and Spindler (2016) give the best agreement with the experimental data. The Haraguchi et al. (1993) model is based on experimental tests during condensation of pure refrigerants (R134a, R123, and R22) in a horizontal microfin tube having an inner diameter equal to 8.37 mm. Oliver et al. (2004) collected some experimental data points in a horizontal helical microfin tube with an inner diameter equal to 8.94 mm during the condensation process at 40 °C of R22, R407C, and R134a with mass flux set at 40 0, 60 0, and 800 kg m−2 s−1 . The Authors modified an existing correlation pro-

posed by Carnavos (1980) for herringbone type microfin tubes to have a better fitting of the experimental values. The new correlation gave similar results to the Miyara et al. (20 0 0) one at low mass fluxes, and some lower values at high mass fluxes. When compared against the present experimental results, the Haraguchi et al. (1993) correlation obtained the lower relative deviation for R1233zd(E), while the Rollmann and Spindler (2016) was the best for R245fa. The Diani et al. (2014) model, the only one developed for small diameter tubes, presented the lower absolute deviations (i.e., 21.1% and 18.7% for R1233zd(E) and R245fa, respectively). 7. Conclusions In this paper, some new experimental heat transfer coefficient and pressure drop data measured during R1233zd(E) and R245fa flow boiling inside a small diameter microfin tube have been presented. The microfin tube under investigation has a 4.2 mm fin tip root diameter, 40 fins, 0.15 mm high with an apex angle γ equal to 42°. The helix angle β is 18° and the area enhancement with reference to the smooth tube having the same fin tip diameter is equal to 1.62. The tube was brazed inside a copper plate in a horizontal position and electrically heated from the bottom. Tests were run at a constant mean saturation temperature of 30 °C, by varying the mean vapor quality from 0.2 to 0.95, the mass velocity from 100 to 300 kg m−2 s−1 , and the heat flux from 15 to 90 kW m−2 . The results were critically discussed. First, it was observed that the heat transfer coefficient is influenced by heat flux, mass velocity, and vapor quality. Under the investigated working conditions, the two-phase forced convection seems to be the prevailing heat transfer mechanism that affects the heat transfer coefficient, since both the fluids can be classified as low pressure refrigerants. In general, both nucleate boiling and forced convection mechanisms con-

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tribute in the flow boiling heat transfer, and the prevailing one depends upon the actual operating test condition. As far as pressure drops are concerned, they increase with both mass velocity and vapor quality. On the base of the thermophysical properties, R1233zd(E) was proposed as a low-GWP, drop-in substitute to R245fa. The performance of both the fluids was compared under similar working conditions, and it can be stated that these two refrigerants display similar heat transfer coefficients, but R1233zd(E) shows up to 16% higher frictional pressure drops. All the data collected were finally compared against some models available in the literature for two-phase heat transfer coefficient and frictional pressure gradient in microfin tubes. In terms of heat transfer coefficient, the Rollmann and Spindler (2016) model presented the best agreement with both the fluids. The relative deviations are around 0.1% and −5.6% for R1233zd(E) and R245fa, respectively. Besides, in terms of frictional pressure gradient, the Haraguchi et al. (1993) and the Rollmann and Spindler (2016) correlations obtained the lower relative deviations, while the Diani et al. (2014) model presented the lower absolute deviations (i.e., 21.1% and 18.7% for R1233zd(E) and R245fa, respectively). Acknowledgment This research project was partially funded by: CariVerona Foundation, Verona, Italy, Ricerca Scientifica e Tecnologica 2016– 2019: “Sostenibilità e autenticazione nutrizionale di filiere lattierocasearie a tutela del consumatore”. The supports of CENTRAL GLASS CO. ltd for donating the fluid and of Wieland-Werke AG (Christoph Walther) for donating the tube samples are gratefully acknowledged. Generalmeccanica Snc and Dr. Damiano Soprana are gratefully acknowledged for their valuable help in the manufacturing of the test section. References Arkema, 2013. Arkema reveals new low-GWP blowing agent for polyurethane foams. Addit. Polym. 11, 3. Bandarra Filho, E.P., Saiz Jabardo, J.M., Barbieri, P.E.L, 2004. Convective boiling pressure drop of refrigerant R-134a in horizontal smooth and microfin tubes. Int. J. Refrig. 27, 895–903. Carnavos, T.C., 1980. Heat transfer performance of internally finned tubes in turbulent flow. Heat Transf. Eng. 4, 32–37. Cavallini, A., Del Col, D., Doretti, L., Longo, G.A., Rossetto, L., 1999. A new computational procedure for heat transfer and pressure drop during refrigerant condensation inside enhanced tubes. J. Enhanc. Heat Transf. 6, 441–456. Cavallini, A., Del Col, D., Doretti, L., Longo, G.A., Rossetto, L.., 20 0 0. Heat transfer and pressure drop during condensation of refrigerants inside horizontal enhanced tubes. Int. J. Refrig. 23, 4–25. Desideri, A., Zhang, J., Kærn, M.R., Ommen, T.S., Wronski, J., Lemort, V., Haglind, F., 2017. An experimental analysis of flow boiling and pressure drop in a brazed plate heat exchanger for organic Rankine cycle power systems. Int. J. Heat Mass Transf. 113, 6–21. Diani, A., Mancin, S., Rossetto, L., 2014. R1234ze(E) flow boiling inside a 3.4 mm ID microfin tube. Int. J. Refrig. 47 (1), 05–119. Eyerer, S., Wieland, C., Vandersickel, A., Spliethoff, H., 2016. Experimental study of an ORC (Organic Rankine Cycle) and analysis of R1233zd-E as a drop-in replacement for R245fa for low temperature heat utilization. Energy 103, 660–671. Garg, P., Kumar, P., Srinivasan, K., Dutta, P., 2013. Evaluation of isopentane, R-245fa and their mixtures as working fluids for organic Rankine cycles. Appl. Therm. Eng. 51, 292–300. Guillaume, L., Legros, A., Desideri, A., Lemort, V., 2017. Performance of a radial-inflow turbine integrated in an ORC system and designed for a WHR on truck application: an experimental comparison between R245fa and R1233zd. Appl. Energy 186, 408–422. Haraguchi, H., Koyama, S., Esaki, J., Fujii, T.., 1993. Condensation heat transfer of refrigerants HFC134a, HCFC123 and HCFC22 in a horizontal smooth tube and a horizontal microfin tube. In: Proceedings of the Thirtieth National Symposium of Japan, Yokohama, pp. 343–345.

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