Condensation of an Azeotropic Mixture Inside 2.5 Mm ID Minitubes

fluids

Article Condensation of an Azeotropic Mixture inside 2.5 mm ID Minitubes

Andrea Diani and Luisa Rossetto * Department of Industrial Engineering, University of Padova, via Venezia 1, 35131 Padova, Italy; [email protected] * Correspondence: [email protected]; Tel.: +39-049-827-6869

Received: 31 August 2020; Accepted: 30 September 2020; Published: 2 October 2020

Abstract: The ongoing miniaturization of air conditioning and refrigeration systems, in order to limit, as much as possible, the refrigerant charge, calls for smaller and smaller heat exchangers. Besides, the new environmental regulations are calling for new pure refrigerants or refrigerants mixtures with lower values of global warming potentials (GWPs). In this context, this paper analyzes the possible implementation of minitubes during condensation of the azeotropic mixture R513A. Two minitubes are tested: a smooth tube with an inner diameter of 2.5 mm, and a microfin tube with an inner diameter at the fin tip of 2.4 mm. The effects of vapor quality (varied in the range 0.10–0.99), of mass velocity 2 1 (varied in the range 200–1000 kg m− s− ), and of saturation temperature (30 ◦C and 40 ◦C) on the heat transfer coefficient are investigated. The experimental results indicate that the heat transfer coefficient increases as both vapor quality and mass velocity increase, both in the case of the smooth tube and of the microfin tube, but the slope of the heat transfer coefficient trend respect to vapor quality is higher in the case of the microfin tube. The microfin tube shows, on average, heat transfer coefficients are 79% higher than those of the smooth tube under the same working conditions. Since R513A is a possible substitute of R134a, some experimental data during condensation heat transfer are also compared against those for R134a. Finally, the experimental results are compared against values estimated by empirical correlations available in the open literature.

Keywords: condensation; R513A; heat transfer coeﬃcient; smooth tube; microﬁn tube

1. Introduction R513A is an azeotropic mixture made of R1234yf and R134a (0.56/0.44 by mass). R1234yf is an HydroFluroOlefin with a Global Warming Potential lower than 1 [1], and it is placed in the A2L flammability class. R134a is among the most used HydroFluoroCarbons, not flammable (A1 flammability class), widely implemented in existing air conditioning systems, but it must be phased out in a near future due to its high GWP, which is 1430. Therefore, the presence of R134a makes the mixture not flammable, whereas the presence of R1234yf lowers the warming impact of the mixture. Coupled with small sized tubes, R513A could be a viable solution for the next generation of air conditioning equipment. R32 and R134a local heat transfer coefficients were measured by Matkovic et al. [2]. Experimental tests were carried out at 40 ◦C saturation temperature inside a 0.96 mm ID circular minitube for mass 2 1 velocities in the range 100–1200 kg m− s− . The heat transfer coefficient increased with mass velocity and vapor quality; therefore, condensation was shear stress dominated, except some data at 100 and 2 1 200 kg m− s− , which overlapped in the graph. Zhang et al. [3] experimentally investigated the condensation of R22, R410A, and R407C inside two stainless steel tubes with inner diameters of 1.088 mm and 1.289 mm. Tests were run for mass velocities 2 1 in the range 300–600 kg m− s− , for saturation temperatures of 30 ◦C and 40 ◦C. The experimental

Fluids 2020, 5, 171; doi:10.3390/fluids5040171 www.mdpi.com/journal/fluids Fluids 2020, 5, 171 2 of 16 results indicated that the condensation heat transfer coefficient increases with mass velocity and vapor quality, increasing more rapidly in the high vapor quality region. More recently, Guo et al. [4] obtained heat transfer coefficients during the condensation of R1234ze(E), R290, R161, and R41 in a smooth horizontal tube with an inner diameter of 2 mm. 2 1 Mass velocity was varied from 200 to 400 kg m− s− , and saturation temperature from 35 ◦C to 40 ◦C. The experimental results showed that the heat transfer coefficient increases as vapor quality and mass velocity increase, and as saturation temperature decreases. Other works related to condensation in minichannels can be found in [5–7]. Adding microfins on the inner circumference of the tube has proven to enhance the two-phase heat transfer. Small-sized microfin tubes were initially studied for applications that involve carbon dioxide, due to its high working pressure [8–10]. Considering microfin tubes classified as minitubes, i.e., with diameters ranging between 1 and 3 mm [11], just a few studies investigated refrigerants’ two-phase heat transfer. Experimental measurements of the heat transfer coefficient were carried out by Diani et al. [12], who studied R1234yf and R1234ze(E) condensing inside a 2.4 mm ID microfin tube. 2 1 Mass velocities were varied from 300 to 1000 kg m− s− , for saturation temperatures of 30 ◦C and 40 ◦C. R1234ze(E) and R134a showed similar values of heat transfer coefficients, whereas R1234yf showed lower values. Bashar et al. [13] investigated the condensation heat transfer of R134a flowing inside a microfin tube having an equivalent diameter of 2.17 mm and inside a smooth tube with an inner diameter 2 1 of 2.14 mm. Data were measured for mass velocities from 50 to 300 kg m− s− , and saturation temperatures from 20 ◦C to 30 ◦C. The heat transfer coefficient was higher in the case of the microfin 2 1 tube for all the investigated mass fluxes and, considering the mass velocity 100 kg m− s− , it was 2–5 times greater than that of the smooth tube. Diani et al. [14] conducted experiments during R1234yf flow boiling inside a microfin tube with 2 1 2 an outer diameter of 3.0 mm, for mass fluxes from 375 to 940 kg m− s− , heat fluxes from 10 kW m− 2 to 50 kW m− , with a saturation temperature at the inlet of the test section of 30 ◦C. The experimental results showed that the flow boiling phenomenon is controlled by nucleate boiling, especially at low vapor qualities, and by convective boiling, especially at high vapor qualities. Jige and Inoue [15] investigated R32 flow boiling inside microfin tubes with equivalent diameters 2 1 of 2.1 mm, 2.6 mm, and 3.1 mm. Mass velocity was varied from 50 to 200 kg m− s− and heat flux 2 from 5 to 40 kW m− . The saturation temperature was 15 ◦C. The effect of tube diameter on the flow boiling heat transfer coefficient was more appreciable at higher vapor quality and lower mass velocity (the lower the tube diameter, the higher the heat transfer coefficient). Frictional pressure drops were highly affected by tube diameter: the lower the tube diameter, the higher the pressure drop. Hirose et al. [16] studied R1234ze(E) condensation inside three different microfin tubes with the same outer diameter of 4 mm. The effects of fin parameters, i.e., of helix angle, fin height, and fin 2 1 number were analyzed for mass velocities in the range 50–400 kg m− s− , at a saturation temperature of 35 ◦C. There are many candidates to replace R134a. Some of them are pure fluids, such as R1234ze(E) and R1234yf; others are azeotropic mixtures, such as R513A, R515A and R516A; others are zeotropic mixtures, such as R450A, R445A, R430A, R436A, and R456A. Heredia-Aricapa et al. [17] overviewed possible refrigerants mixtures as replacements of common HFCs like R134a, R404A, and R410A. Among the replacements of R134a, they identified R445A, R430A, R515A, R436A, R456A, and R515A. R436A and R430A are the alternatives with the greatest differences in the thermophysical properties, due to the presence of hydrocarbons in the mixtures. The comparison among fluids was given in terms of COP and cooling capacity in a vapor compression system. Kedzierski et al. [18] carried out experimental tests during pool boiling on a reentrant cavity surface for R134a, R1234yf, R513A, and R450A. At constant wall superheat and pressure, with a saturation temperature of 277.6 K, the heat flux for R1234yf and R513A was, on average, 16% and 19% 2 2 less than that for R134a, respectively, for heat fluxes between 20 kW m− and 110 kW m− , whereas the Fluids 2020, 5, 171 3 of 16

2 heat flux for R450A was, on average, 57% less than that of R134a for heat fluxes between 30 kW m− Fluids 2020, 5, x 2 3 of 16 and 110 kW m− . Other works related to R513A can be found in Diani et al. [19,20], but no comparisons with R134a were reported. the heat flux for R450A was, on average, 57% less than that of R134a for heat fluxes between 30 kW Being a relatively new mixture, to date, no work about R513A condensation inside mini microfin m−2 and 110 kW m−2. Other works related to R513A can be found in Diani et al. [19,20], but no tubes with diameters lower than 3 mm can be found in the open literature. Two minitubes are tested in comparisons with R134a were reported. this research: the first one is a smooth tube with an inner diameter of 2.5 mm, and the second one Being a relatively new mixture, to date, no work about R513A condensation inside mini microfin is atubes microfin with tubediameters with lower an inner than diameter 3 mm can at be the found fin tipin the of 2.4open mm. literature. This work Two isminitubes aimed at are enlarging tested thein experimental this research: database the first one of heat is a transfersmooth tube coeffi withcients an during inner diameter R513A condensation of 2.5 mm, and inside the second smooth one and microfinnedis a microfin minitubes. tube with an The inner wide diameter rangeof at operativethe fin tip of conditions 2.4 mm. This will work permit is aimed us to at understand enlarging the how eachexperimental working parameter, database of i.e., heat vapor transfer quality, coefficients mass velocity, during R513A and saturation condensation temperature, inside smooth affects and the two-phasemicrofinned heat minitubes. transfer both The in wide the caserange of of the operative smooth tubeconditions and of will the permit microfin us tube,to understand highlighting how the diffeacherences working among parameter, the two tubes. i.e., vapor Finally, quality, the experimental mass velocity, condensation and saturation heat temperature, transfer coe ffiaffectscients the will be comparedtwo-phase heat against transfer values both estimated in the case by of empirical the smooth correlations. tube and of the microfin tube, highlighting the differences among the two tubes. Finally, the experimental condensation heat transfer coefficients 2. Experimentalwill be compared Facility against and values Test es Sectionstimated by empirical correlations. The experimental facility is in the Laboratory of Heat Transfer in MicroGeometries at the 2. Experimental Facility and Test Sections Department of Industrial Engineering of University of Padova, Italy. Figure1 shows a schematic of the liquidThe pumped experimental circuit. facility A magnetically is in the Laboratory coupled gear of pumpHeat Transfer was chosen in MicroGeometries to avoid any possible at the oil tracesDepartment in the refrigerant. of Industrial The Engineering pump circulates of University the refrigerant of Padova, into Italy. a CoriolisFigure 1 eshowsffect mass a schematic flowmeter, of whichthe liquid presents pumped a reading circuit. accuracy A magnetically of 0.1%. coupled The gear refrigerant pump was enters chosen an to evaporator avoid any inpossible subcooled oil ± conditions,traces in the and refrigerant. it exits in The superheated pump circulates conditions. the refrigerant The evaporator into a Coriolis is a BPHE effect mass where flowmeter, refrigerant which flows in countercurrentpresents a reading with accuracy hot water, of ±0.1%. supplied The refrigerant in the hot enters water an loop. evaporator An electric in subcooled boiler, regulatedconditions, with and a it exits in superheated conditions. The evaporator is a BPHE where refrigerant flows in countercurrent PID controller, guarantees an inlet water temperature of 60 ◦C. A tube-in-tube pre-condenser is aimed with hot water, supplied in the hot water loop. An electric boiler, regulated with a PID controller, at partially condensing the superheated vapor, which flows in the inner tube. Cold water, supplied guarantees an inlet water temperature of 60 °C. A tube-in-tube pre-condenser is aimed at partially by a water-cooled stabilized chiller, flows in the annulus. The chiller permits to set the temperature condensing the superheated vapor, which flows in the inner tube. Cold water, supplied by a water- of the water at the inlet of the pre-condenser. In addition, a valve permits variation in the water cooled stabilized chiller, flows in the annulus. The chiller permits to set the temperature of the water at flow rate, and therefore it is possible to adjust the heat flow rate exchanged in the pre-condenser the inlet of the pre-condenser. In addition, a valve permits variation in the water flow rate, and therefore to setit is thepossible refrigerant to adjust conditions the heat flow at rate the inletexchanged of the in test the section.pre-condenser The refrigerantto set the refrigerant partially conditions condenses in theat the test inlet section, of the test and section. it is later The fullyrefrigerant condensed partially and condenses subcooled in the in test the section, post-condenser, and it is later which fully is anothercondensed BPHE. and In subcooled this case, in the the refrigerant post-condenser, flows whic in counterh is another current BPHE. with In a this mixture case, the of waterrefrigerant/glycol suppliedflows in by counter an air-cooled current with chiller. a mixture The refrigerant of water/glyc passesol supplied through by a an drier air-cooled filter prior chiller. to The re-enter refrigerant into the pump.passes The through saturation a drier conditions filter prior areto re-enter controlled into bythe meanspump. The of a saturation damper, which conditions is connected are controlled to the by air pressuremeans line.of a damper, which is connected to the air pressure line.

FigureFigure 1. 1.Schematic Schematic of the test rig. rig. FluidsFluids 20202020,, 55,, 171x 4 of 16

As depicted in Figure 1, temperatures (measured by T-type thermocouples with an accuracy of ±0.05As K) depicted and pressures in Figure of1 ,the temperatures refrigerant (measured (measured by by T-type absolute thermocouples pressure transducers with an accuracy with ofan accuracy0.05 K) andof ±1950 pressures Pa) ofare the measured refrigerant in (measuredseveral locations by absolute throughout pressure the transducers circuit, so with it is anpossible accuracy to ± ofdetermine1950 Pa) its are thermodynamic measured in severalstate. The locations refrigerant throughout conditions the circuit,at the inlet so it of is the possible test section, to determine i.e., at ± itsthe thermodynamic exit of the pre-condenser, state. The derive refrigerant from an conditions energy balance at the at inlet the pre-condenser, of the test section, as will i.e., be at explained the exit ofin the next pre-condenser, section. The derive water fromflow rate an energy in the cold balance-water at loop the pre-condenser, at the pre-condenser, as will necessary be explained for the in theenergy next balance, section. is The measured water flowby a ratemagnetic in the flowme cold-waterter, with loop a at reading the pre-condenser, accuracy of ±0.25%. necessary The for water the energytemperature balance, difference is measured at the by pre-condenser a magnetic flowmeter, is measured with aby reading a T-type accuracy thermopile, of 0.25%. which The has water an ± temperatureaccuracy of ±0.03 difference K. at the pre-condenser is measured by a T-type thermopile, which has an accuracy of 0.03Two K. different minitubes made of copper were tested. The first one is a smooth tube with an ± internalTwo diameter different of minitubes 2.5 mm and made an outer of copper diameter were oftested. 4.5 mm. The The first second one one is a is smooth a microfin tube tube with with an internalan inner diameter diameter of at 2.5 the mm fin andtip of an 2.4 outer mm, diameter an outer of diameter 4.5 mm. of The 3.0 second mm, having one is a40 microfin fins, and tube each with fin anhas inner an height diameter of 0.12 at the mm fin and tip ofan 2.4 apex mm, angle an outer of 43°; diameter the helix of angle 3.0 mm, is 7°. having Both 40 the fins, test and section each finfor hasthe ansmooth height tube of 0.12 and mm the andone anfor apex the microfin angle of 43tube◦; the are helix composed angle isof 7 the◦. Both tested the tube, test sectionwhich foris horizontally the smooth tubelocated, and around the one which for the a microfin1.9 mm ID tube smooth are composed tube is wrapped. of the tested Cold tube, water, which supplied is horizontally by a thermostatic located, aroundbath and which measured a 1.9 mmby a ID magnetic smooth flowmeter tube is wrapped. with an Cold accuracy water, of supplied ±0.50% of by the a thermostatic reading, flows bath in and the measuredwrapped tube, by a magneticto remove flowmeter the heat flow with rate an necess accuracyary of for 0.50%the condensation of the reading, of the flows refrigerant in the wrapped flowing ± tube,inside to the remove horizontally the heat located flow rate tested necessary tube. forTwo the cu condensationrves, one at ofthe the inlet refrigerant and another flowing at the inside outlet, the horizontallypermit to host located two T-type tested thermocouples tube. Two curves, (accuracy one at ±0.05 the inlet K), which and another measure at the water outlet, inlet permit and to outlet host twotemperatures T-type thermocouples in the test section. (accuracy On the0.05 external K), which wallmeasure of the tested the water tubes, inlet T-type and thermocouples outlet temperatures were ± inglued the testto measure section. the On wall the externaltemperature. wall ofFigure the tested 2 shows tubes, a schematic, T-type thermocouples as well as the were geometrical glued to measuredimensions, the of wall the temperature. test section for Figure the smooth2 shows tube a schematic,. The blue dots as well represent as the geometricalthe thermocouples dimensions, which ofmeasure the test the section water for temperatures, the smooth tube. whereas The bluethe red dots do representts represent the the thermocouples thermocouples which attached measure on the waterexternal temperatures, wall of the tested whereas tube. the The red dotstested represent tubes with the the thermocouples wrapped tubes attached were onfinally theexternal inserted wall inside of thean aluminum tested tube. housing The tested with tubes a U-shape. with the This wrapped housing tubes was filled were finallywith an inserted alloy of inside tin/lead, an aluminumwhich was housingcasted inside. with aThis U-shape. alloy is This needed housing to have was a filledgood withthermal an alloycontact of between tin/lead, the which wrapped was casted tube, inside.inside Thiswhich alloy the cold is needed water toflows, have and a good the tested thermal tube, contact inside betweenwhich the the refrigerant wrapped condenses. tube, inside AF/Armaflex which the coldwas waterused as flows, thermal and insulation the tested tube,around inside the whichtest sect theion refrigerant to limit the condenses. heat losses AF through/Armaflex the was external used asambient. thermal Pressure insulation ports around were soldered the test section downstream to limit and the upstream heat losses of the through test sections: the external the upstream ambient. Pressurepressure portsport is were connected soldered to downstreaman absolute pressure and upstream transducer of the (accuracy test sections: of ±1950 the upstreamPa), whereas pressure both portupstream is connected and downstream to an absolute pressure pressure ports transducer are conne (accuracycted to of a 1950differential Pa), whereas pressure both transducer upstream ± and(accuracy downstream of ±25 Pa). pressure ports are connected to a differential pressure transducer (accuracy of 25 Pa). ±

Figure 2. Schematic of the test section for the 2.5 mm ID smooth tube. Figure 2. Schematic of the test section for the 2.5 mm ID smooth tube. 3. Data Analysis 3. Data Analysis The vapor quality at the inlet of the test section, hTS,in, can be extrapolated from the speciﬁc The vapor quality at the inlet of the test section, hTS,in, can be extrapolated from the specific enthalpy at the exit of the pre-condenser, hPC,out, calculated as: enthalpy at the exit of the pre-condenser, hPC,out, calculated as: . m cp,w ∆T 𝑚w ,,PC ∙𝑐, ∙∆𝑇w,PC, h = h = hvs · . · (1) ℎTS,,in =ℎPC,,out =ℎ−− (1) mre𝑚 f

where hvs is the specific enthalpy of the refrigerant in superheated vapor conditions at the exit of the

evaporator, which is calculated from the knowledge of its temperature and pressure. 𝑚 , is the Fluids 2020, 5, 171 5 of 16

where hvs is the specific enthalpy of the refrigerant in superheated vapor conditions at the exit of the . evaporator, which is calculated from the knowledge of its temperature and pressure. mw,pc is the mass flow rate of the water flowing in the pre-condenser, cp,w is the water specific heat, ∆Tw,pc is the water . temperature difference between outlet and inlet of the pre-condenser, and mre f is the refrigerant mass flow rate. The inlet vapor quality can be calculated from the knowledge of hTS,in as:

hTS,in hL xin = − (2) hV hL − where hV and hL are the speciﬁc enthalpies of saturated vapor and liquid, respectively, calculated from the inlet pressure. REFPROP 10 (Lemmon et al. [21]) was used to calculate all the refrigerant thermodynamic properties, since it implements updated equations for R513A. In a similar way, the refrigerant speciﬁc enthalpy at the outlet of the test section, hTS,out, can be derived from an energy balance, as:

. qw,TS mw,TS cp,w (tw,TS,out tw,TS,in) hTS,out = hTS,in . = hTS,in · · . − (3) − mre f − mre f

. where qw,TS is the heat flow rate calculated in the water side, and, in this case, mw,TS is the mass flow rate of the water flowing in the test section, tw,TS,out and tw,TS,in are the water temperature at the outlet and inlet of the test section, respectively. Once hTS,out is known, the outlet vapor quality is calculated as:

hTS,out hL xout = − (4) hV hL − where hV and hL are the specific enthalpy of saturated vapor and liquid, respectively, calculated at the outlet pressure. Experimental results, discussed in the next section, will be referred to the mean vapor quality, which is the arithmetic average between inlet vapor quality and outlet vapor quality. The experimental tests permitted to calculate, both for the smooth tube and for the microfin tube, the heat transfer coefficient (HTC), as:

qw,TS HTC = (5) AD tsat t · − wall where tsat and twall are the mean saturation and wall temperatures. AD is equal to the inner surface area in the case of the smooth tube, whereas it is equal to the inner surface area of an equivalent smooth with the inner diameter at the fin tip in the case of the microfin tube. Following the procedure suggested by Kline and McKlintock [22], considering a level of confidence of 95%, the mean uncertainty on the heat transfer coefficient is 4.2%, whereas on the mean vapor ± quality is 0.027. ± 4. Experimental Results

4.1. Smooth Tube

4.1.1. Single-Phase Results The test section developed for the smooth tube was first verified during single phase flow involving the forced convection of subcooled liquid. The main purposes of these tests were to check the heat flow rate calculated on the water side and the one on the refrigerant side and to verify if the measuring technique was suitable by calculating the single phase heat transfer coefficients and by comparing these results against values predicted by well-known correlations from the literature. Fluids 2020, 5, 171 6 of 16

These tests were run with an inlet subcooling level between 17 K and 21 K, at a pressure of about 2 1 11.5 bar, for mass velocities from 400 to 1000 kg m− s− . Mass velocity is deﬁned as:

. 4 mre f G = · (6) Fluids 2020, 5, x π D2 6 of 16 · i where Di is the inner diameter. The heat transfer coefficient can be calculated as in Equation (5) but where Di is the inner diameter. The heat transfer coefficient can be calculated as in Equation (5) but in in this case, 𝑡̅ is used instead of 𝑡̅ . Two T-type thermocouples, downstream and upstream of this case, tre f is used instead of tsat. Two T-type thermocouples, downstream and upstream of the test the test section, were used to measure the inlet and outlet refrigerant temperatures. section, were used to measure the inlet and outlet refrigerant temperatures. The experimental results demonstrate that the heat flow rates calculated on the water side and The experimental results demonstrate that the heat flow rates calculated on the water side and on on the refrigerant side are, on average, within ±4%. Figure 3 reports the heat transfer coefficients the refrigerant side are, on average, within 4%. Figure3 reports the heat transfer coe fficients plotted plotted against the mass velocity. As expected,± the heat transfer coefficient increases as the mass againstvelocity the mass increases. velocity. Figure As 3 expected,also reports the the heat comparison transfer between coefficient the experimental increases as values the mass and velocitythe increases.values Figurepredicted3 also by the reports correlations the comparison of Dittus and between Boelter [23] the and experimental of Petukhov valuesand Popov and [24]. the The values predictedcorrelation by the of correlations Petukhov ofand Dittus Popov and [24] Boelter aptly [23predicts] and ofthe Petukhov experimental and Popovvalues, [ 24with]. The a relative correlation of Petukhovdeviation and of Popov3.9%, an [24 absolute] aptly deviation predicts theof 4.0% experimental and a standard values, deviation with a of relative 4.0%. The deviation correlation of 3.9%, of an absoluteDittus deviation and Boelter of 4.0% [23] andshows a standard a relative deviation deviation of of 4.0%. −0.9%, The an correlationabsolute deviation of Dittus of and7.8% Boelter and a [23] showsstandard a relative deviation deviation of 9.1%. of 0.9%, an absolute deviation of 7.8% and a standard deviation of 9.1%. −

Figure 3. Single-phase heat transfer coeﬃcient for the smooth tube vs. mass velocity. Experimental Figure 3. Single-phase heat transfer coefficient for the smooth tube vs. mass velocity. Experimental data anddata estimationsand estimations by Dittusby Dittus and and Boelter Boelter [ 23[23]] and and byby Petukhov and and Popov Popov [24]. [24 ].

Therefore,Therefore, since since the the percentage percentage di differencefference between between heat flow flow rate rate on on the the refrigerant refrigerant side side and the and the one onone the on water the water side isside within is within4%, ±4%, and and since since empirical empirical correlations correlations taken taken from from thethe literatureliterature are able ± to estimateable to theestimate experimental the experimental values va oflues liquid-phase of liquid-phase heat heat transfer transfer coe coefficients,fficients, the thetest test section for for the 2.5 mmtheID 2.5smooth mm ID smooth tube is tube deemed is deemed to be to verified. be verified.

4.1.2.4.1.2. Condensation Condensation Results Results The operatingThe operating conditions conditions during during the the condensing condensing tests tests for for the the 2.5 2.5 mm mm ID IDsmooth smooth tube tube are the are the following: mass velocity in the range 200–1000 kg m−2 s2−1 and1 saturation temperature at the inlet of 30 following: mass velocity in the range 200–1000 kg m− s− and saturation temperature at the inlet of °C and 40 °C. Experimental tests involve partial condensation inside the test tube, with a vapor 30 C and 40 C. Experimental tests involve partial condensation inside the test tube, with a vapor ◦ quality difference◦ between inlet and outlet from 0.08 to 0.40, depending on the mass velocity. quality difference between inlet and outlet from 0.08 to 0.40, depending on the mass velocity. Therefore, Therefore, in order to investigate vapor qualities from 0.1 to 0.99, tests were run by varying the inlet in ordervapor to investigatequality. vapor qualities from 0.1 to 0.99, tests were run by varying the inlet vapor quality. The collectedThe collected heat transferheat transfer coeffi coefficientscients during duri condensationng condensation inside inside the smooththe smooth tube attube a saturation at a temperaturesaturation at thetemperature inlet of the at testthe inlet section of the of 40test◦C sectio are reportedn of 40 °C in are Figure reported4. Only in Figure data which 4. Only involve data an exchangedwhich heatinvolve flow an rateexchanged higher heat than flow 30 Wrate are higher reported, than 30 in W order are reported, to avoid in data order with to avoid high data uncertainties. with Data arehigh plotted uncertainties. against Data the meanare plotted vapor against quality the in me thean test vapor section quality at diinff theerent test mass section velocities. at different A clear mass velocities. A clear effect of both mass velocity and vapor quality is visible: the higher the mass velocity and the vapor quality, the higher the heat transfer coefficient, due to the diminishing liquid film thickness. According to the flow regime map developed by Doretti et al. [25] and reported by Fluids 2020, 5, 171 7 of 16 effect of both mass velocity and vapor quality is visible: the higher the mass velocity and the vapor quality, the higher the heat transfer coefficient, due to the diminishing liquid film thickness. According Fluids 2020,, 5,, xx 7 of 16 to the flow regime map developed by Doretti et al. [25] and reported by Cavallini et al. [26], all the 2 1 experimentalCavallini dataet al. fall [26], in all the the∆T -independentexperimental data zone, fall i.e., in annularthe ∆T-independent-independent flow, except zone,zone, data ati.e.,i.e.,G annularannular= 200 kg flow,flow, m− s− up toexcept a mean data vapor at G == quality 200200 kgkg m ofm−−22 0.5, ss−−11 upup which toto aa meanmean fall into vaporvapor the qualityquality∆T-dependent ofof 0.5,0.5, whichwhich zone, fallfall intointo i.e., thethe they ∆T are-dependent-dependent in stratified flow regime.zone, i.e., Data they are at higher in stratified mass flow velocity regime. shows Data aathigher higherhigher massmass slope velocityvelocity of the heat showsshows transfer aa higherhigher coe slopeslopefficient ofof thethe trend, highlightingheat transfer a higher coefficient convective trend, high effectlightinglighting as the aa mass higherhigher velocity convectiveconvective increases. effecteffect asas thethe massmass velocityvelocity increases.increases.

Figure 4. Heat transfer coefficient during condensation for the smooth tube vs. mean vapor quality at Figure 4. Heat transfer coefficient during condensation for the smooth tube vs. mean vapor quality at 2 1 tsat = 40 ◦C. G expressed in [kg m−−2 s−1 ]. ttsatsat == 4040 °C.°C. G expressedexpressed inin [kg[kg mm−2 ss−1].]. Figure5 reports the data collected during condensation inside the smooth tube for a saturation Figure 5 reports the data collected during condensation inside the smooth tube for a saturation temperaturetemperaturetemperature at the atat inlet thethe inletinlet of the ofof testthethe testtest section sectionsection of 30ofof ◦3030C. °C.°C. All AllAll these thesethese data datadata fall fallfall in inin the thethe∆ T∆T-independent-independent-independent zonezone and, as a result,and, as the a result, heat transfer the heat coetransferfficient coefficient increases increases as both as the both vapor the vapor quality quality and mass and mass velocity velocity increase. The comparisonincrease.increase. TheThewith comparisoncomparison Figure 4withwith reveals FigureFigure the 44 erevealsrevealsffect of thethe the effect saturation of the saturation temperature. temperature. A higher A higher saturation temperaturesaturation implies temperature a higher implies vapor a higher density; vapor thus, density; at constant thus,thus, atat constantconstant mass velocity, massmass velocity,velocity, it also itit impliesalsoalso impliesimplies a lower vapora velocity,lower vapor with velocity, consequent with consequent lower convective lower convective effects. effects. This This aspect aspect can can be be observed observed in inthe the heat transferheat coe transferfficients coefficients reported reported in Figures in Figures4 and 54: and the 5: heat thethe heat transferheat transfertransfer coe coefficientcoefficientfficient tends tendstends to toto decreasedecrease asas as the the saturation temperature increases, due to lower convective effects. saturationthe saturation temperature temperature increases, increases, due to due lower to lower convective convective effects. effects.

Figure 5. Heat transfer coeﬃcient during condensation for the smooth tube vs. mean vapor quality at 2 1 tsat = 30 ◦C. G expressed in [kg m− s− ]. Fluids 2020, 5, x 8 of 16

Figure 5. Heat transfer coefficient during condensation for the smooth tube vs. mean vapor quality at Fluids 2020tsat =, 305, 171 °C. G expressed in [kg m−2 s−1]. 8 of 16

4.2. Microfin Tube 4.2. Microfin Tube 4.2.1. Single-Phase Results 4.2.1. Single-Phase Results In a similar way to the validation procedure depicted in Section 4.1.1, single-phase tests were carriedIn aout similar during way R1234yf to the liquid validation flow procedureto validate depicted the test section in Section for 4.1.1the 2.4, single-phase mm ID microfin tests weretube, carriedand the outexperimental during R1234yf results liquid revealed flow that to validatethe heat flow the test rate section calculated for the on 2.4the mmwater ID side microfin was within tube, and±1.5 theW experimentalthe one on resultsthe refrigerant revealed thatside, the and heat that flow the rate experimental calculated on single the water-phase side heat was transfer within 1.5 W the one on the refrigerant side, and that the experimental single-phase heat transfer coefficients ±coefficients were aptly estimated by empirical correlations from the literature. Therefore, the test weresection aptly for estimatedthe microfin by tube empirical was deemed correlations verified. from Mo there literature. details can Therefore, be found in the Diani test sectionet al. [12]. for the microfin tube was deemed verified. More details can be found in Diani et al. [12]. 4.2.2. Condensation Results 4.2.2. Condensation Results The operating conditions during the condensing tests for the 2.4 mm ID microfin tube are the The operating conditions during the condensing tests for the 2.4 mm ID microfin tube are the following: mass velocity in the range 200–1000 kg m−2 s−1, saturation temperature at the inlet of 30 °C following: mass velocity in the range 200–1000 kg m 2 s 1, saturation temperature at the inlet of 30 C and 40 °C. In this case, the mass velocity is calculated− − as reported in Equation (6) but the inner◦ and 40 ◦C. In this case, the mass velocity is calculated as reported in Equation (6) but the inner diameter diameter at the fin tip is used as Di. at the fin tip is used as D . Figure 6 shows thei condensation heat transfer coefficients plotted against the mean vapor Figure6 shows the condensation heat transfer coe fficients plotted against the mean vapor quality quality for the microfin tube with a saturation temperature at the inlet of the test section of 40 °C. for the microfin tube with a saturation temperature at the inlet of the test section of 40 C. Again, Again, only data which involve an exchanged heat flow rate higher than 30 W are◦ reported. only data which involve an exchanged heat flow rate higher than 30 W are reported. According to According to the flow regime map developed by Doretti et al. [25] and reported by Cavallini et al. the flow regime map developed by Doretti et al. [25] and reported by Cavallini et al. [26], all the [26], all the experimental data fall in the annular flow regime. Therefore, the higher the vapor quality experimental data fall in the annular flow regime. Therefore, the higher the vapor quality and mass and mass velocity, the higher the heat transfer coefficient will be. Compared to the results during velocity, the higher the heat transfer coefficient will be. Compared to the results during condensation condensation inside the smooth tube reported in Figure 4, the condensation heat transfer coefficients inside the smooth tube reported in Figure4, the condensation heat transfer coe fficients for the microfin for the microfin tube are higher, and the comparison reveals that data at mass velocities of 200, 300, tube are higher, and the comparison reveals that data at mass velocities of 200, 300, and 400 kg m 2 s 1 and 400 kg m−2 s−1 are particularly enhanced. Indeed, the slope of the heat transfer coefficients− with− are particularly enhanced. Indeed, the slope of the heat transfer coefficients with mean vapor quality mean vapor quality is now much higher for these low mass velocities: taking for instance2 1G = 400 kg is now−2 −1 much higher for these low mass velocities: taking for instance G −=2 400−1 kg m− s− , the heat m s , the heat transfer coefficient passes from approximately2 20001 W m K at xmean = 0.3 to a value transfer coefficient− passes2 −1 from approximately 2000 W m− K− at xmean = 0.3 to a value of about of about 50002 W1 m K at xmean = 0.8 for the smooth tube, i.e., it is 2.5 times higher, whereas it passes 5000 W m− K− at xmean = 0.8 for the smooth tube, i.e., it is 2.5 times higher, whereas it passes from from approximately 3500 W m−2 K−1 at xmean = 0.3 to a value of about 12500 W m−2 K−1 at xmean = 0.8 for approximately 3500 W m 2 K 1 at x = 0.3 to a value of about 12500 W m 2 K 1 at x = 0.8 for the microfin tube, i.e., it is− 3.5 −times meanhigher. − − mean the microfin tube, i.e., it is 3.5 times higher.

Figure 6. Heat transfer coeﬃcient during condensation for the microﬁn tube vs. mean vapor quality at Figure 6. Heat transfer coefficient 2during1 condensation for the microfin tube vs. mean vapor quality tsat = 40 ◦C. G expressed in [kg m− s− ]. at tsat = 40 °C. G expressed in [kg m−2 s−1]. Fluids 2020, 5, 171 9 of 16 Fluids 2020, 5, x 9 of 16

ResultsResults with an imposedimposed saturation temperature at the inlet of the test section of 3030 ◦°CC forfor thethe microfinmicrofin tubetube are are reported reported in in Figure Figure7. Again, 7. Again, all the all experimental the experimental data fall data in thefall annularin the annular flow regime, flow andregime, the heatand transferthe heat coe transferfficient increasescoefficient with increases mass velocity with andmass vapor velocity quality. and The vapor comparison quality. withThe Figurecomparison6 reveals with the Figure e ffect 6 reveals of saturation the effect temperature: of saturation as temperature: stated for the as smooth stated tube,for the the smooth higher tube, the saturationthe higher the temperature, saturation thetemperature, lower the the heat lower transfer the heat coeffi transfercient, due coefficient, to lower due convective to lower econvectiveffects that resulteffects from that theresult lower from vapor the lower velocity. vapor velocity.

Figure 7. Heat transfer coefficient during condensation for the microfin tube vs. mean vapor quality at Figure 7. Heat transfer coefficient 2during1 condensation for the microfin tube vs. mean vapor quality tsat = 30 ◦C. G expressed in [kg m− s− ]. at tsat = 30 °C. G expressed in [kg m−2 s−1]. 4.3. Comparison among the Tubes 4.3. Comparison among the Tubes The heat transfer coefficients of smooth tube and microfin tube are here compared under the same workingThe conditions.heat transfer To coefficients better explain of smooth the comparison, tube and the microfin parameter tube called are here the compared Enhancement under Factor the (EF)same is working introduced: conditions. To better explain the comparison, the parameter called the Enhancement Factor (EF) is introduced: HTCmicro f in EF = (7) HTC 𝐻𝑇𝐶smooth 𝐸𝐹 = (7) where HTCmicrofin and HTCsmooth represent the heat𝐻𝑇𝐶 transfer coefficient of the microfin tube and of the smooth tube, respectively, under the same operative conditions. where HTCmicrofin and HTCsmooth represent the heat transfer coefficient of the microfin tube and of the The Enhancement Factor is plotted versus the mean vapor quality in Figure8, considering a smooth tube, respectively, under the same operative conditions. saturation temperature of 40 C. Adding microfins on the internal area of the microfin tube improves the The Enhancement Factor◦ is plotted versus the mean vapor quality in Figure 8, considering a thermal behavior during condensation; indeed, the EF is always higher than 1. In addition, the benefits saturation temperature of 40 °C. Adding microfins on the internal area of the microfin tube improves of the microfins increase as mass velocity decreases, i.e., the highest EF is shown at the lowest mass the thermal behavior during condensation; indeed, the EF is always higher than 1. In addition, the velocities. At these lowest mass velocities, the EF increases along with the increasing vapor quality. benefits of the microfins increase as mass velocity decreases, i.e., the highest EF is shown at the lowest Similar conclusions are made at t = 30 C. mass velocities. At these lowest satmass velocities,◦ the EF increases along with the increasing vapor The area enhancement ratio, i.e., the ratio between the total inner surface area of the microfin quality. Similar conclusions are made at tsat = 30 °C. tube and the inner surface of a smooth tube having the fin tip diameter as inner diameter, is 1.89. 2 1 It is worth highlighting that, at G = 300 and 400 kg m− s− for mean vapor qualities higher than 0.4, the EF is higher than the mere increase of the heat transfer area, and thus the microfins have a positive effect on the heat transfer especially at low mass velocities, increasing the turbulence of the liquid film. Considering all the experimental data, the average Enhancement Factor is 1.79. Fluids 2020,, 5,, 171x 10 of 16

2 1 Figure 8. Enhancement Factor (EF) vs. mean vapor quality at tsat = 40 ◦C. G expressed in [kg m− s− ]. Figure 8. Enhancement Factor (EF) vs. mean vapor quality at tsat = 40 °C. G expressed in [kg m−2 s−1]. 4.4. Comparison among Refrigerants The area enhancement ratio, i.e., the ratio between the total inner surface area of the microfin Being one of the possible direct drop in replacements of R134a in the near future, the performance tube and the inner surface of a smooth tube having the fin tip diameter as inner diameter, is 1.89. It of R513A is compared against that of R134a. Heat transfer coefficients during R134a condensation is worth highlighting that, at G = 300 and 400 kg m−2 s−1 for mean vapor qualities higher than 0.4, the inside a microfin tube with an inner diameter at the fin tip of 2.4 mm are borrowed by Diani et al. [12]. EF is higher than the mere increase of the heat transfer area, and thus the microfins have a positive The comparison can start by comparing the different thermophysical and transport properties of the effect on the heat transfer especially at low mass velocities, increasing the turbulence of the liquid two fluids. Table1 reports the main properties of R134a and R513A. film. Considering all the experimental data, the average Enhancement Factor is 1.79. Table 1. Thermophysical and transport properties of R134a and R513A. Values from REFPROP 10 4.4. Comparison(Lemmon et among al. [21]). Refrigerants

Being one of the possible directProperties drop in re R134aplacements R513A of R134a in the near future, the performance of R513A is compared against that of R134a. Heat transfer coefficients during R134a t [ C] 30 30 condensation inside a microfin tube withsat ◦ an inner diameter at the fin tip of 2.4 mm are borrowed by psat [bar] 7.702 8.211 Diani et al. [12]. The comparison can start by3 comparing the different thermophysical and transport ρL [kg m− ] 1188 1115 3 properties of the two fluids. Table ρ1V report[kg m−s the] main37.54 properties 43.50 of R134a and R513A. 1 1 λL [W m− K− ] 0.079 0.068 1 1 Table 1. Thermophysical and λtransportV [W m− propertiesK− ] of0.014 R134a and 0.015 R513A. Values from REFPROP 10 (Lemmon et al. [21]). µL [µPa s] 183 156 µV [µPa s] 11.9 11.8 1 1 cp,L Properties[J kg− K− ] R134a1447 R513A 1436 1 1 cp,V [Jt kgsat −[°C]K− ] 1066 30 30 1092 Pr [-] 3.35 3.29 psatL [bar] 7.702 8.211 Pr [-] 0.885 0.888 V −3 ρL [kg m1 ] 1188 1115 σ [N m− ] 0.0074 0.0061 V −3 ρ p r[kg[-] m ] 37.54 0.190 43.50 0.225 λL [W m−1 K−1] 0.079 0.068 λV [W m−1 K−1] 0.014 0.015 Among the listed properties, R134a shows a lower vapor density than R513A, i.e., at constant μL [μPa s] 183 156 mass velocity the vapor velocity is higher, with consequent higher convective effects. Furthermore, μV [μPa s] 11.9 11.8 the liquid thermal conductivity of R134a is higher than that of R513A, with consequently better thermal cp,L [J kg−1 K−1] 1447 1436 characteristics of the liquid film thickness. Considering the pressure drop, the higher vapor velocity at cp,V [J kg−1 K−1] 1066 1092 constant mass velocity for R134a will also be reflected in the hydraulic behavior. PrL [-] 3.35 3.29 Figure9 shows a comparison between R134a and R513A heat transfer coe fficient during PrV [-] 0.885 0.888 condensation inside the microfin tube. As it appears from the figure, the heat transfer coefficients σ [N m−1] 0.0074 0.0061 of R134a are higher than those of R513A, especially in the high vapor quality region, due to the pr [-] 0.190 0.225 aforementioned motivations. Similar conclusions can be drawn at a saturation temperature of 40 ◦C. Fluids 2020, 5, x 11 of 16

Among the listed properties, R134a shows a lower vapor density than R513A, i.e., at constant mass velocity the vapor velocity is higher, with consequent higher convective effects. Furthermore, the liquid thermal conductivity of R134a is higher than that of R513A, with consequently better thermal characteristics of the liquid film thickness. Considering the pressure drop, the higher vapor velocity at constant mass velocity for R134a will also be reflected in the hydraulic behavior. Figure 9 shows a comparison between R134a and R513A heat transfer coefficient during condensation inside the microfin tube. As it appears from the figure, the heat transfer coefficients of R134a are higher than those of R513A, especially in the high vapor quality region, due to the Fluidsaforementioned2020, 5, 171 motivations. Similar conclusions can be drawn at a saturation temperature of 1140 of °C. 16

Figure 9. R134a and R513A heat transfer coefficient during condensation for the microfin tube vs. mean Figure 9. R134a and R513A heat transfer coefficient during condensation for the microfin tube vs. 2 1 vapor quality at tsat = 30 ◦C. G expressed in [kg m− s− ]. mean vapor quality at tsat = 30 °C. G expressed in [kg m−2 s−1]. Frictional pressure gradients can be evaluated from the total measured pressure drop by subtracting Frictional pressure gradients can be evaluated from the total measured pressure drop by the momentum pressure gradient. The model of Rouhani and Axelsson [27] is applied to estimate subtracting the momentum pressure gradient. The model of Rouhani and Axelsson [27] is applied to the void fraction, needed for the calculation of the momentum term. Figure 10 shows the comparison estimate the void fraction, needed for the calculation of the momentum term. Figure 10 shows the of frictional pressure gradients at tsat = 30 ◦C. The frictional pressure gradient increases as the mass comparison of frictional pressure gradients at tsat = 30 °C. The frictional pressure gradient increases velocity increases and as the mean vapor quality increases, until a maximum value is reached at mean as the mass velocity increases and as the mean vapor quality increases, until a maximum value is vapor qualities in the range 0.8–0.9, and then it slightly decreases. The comparison between the two reached at mean vapor qualities in the range 0.8–0.9, and then it slightly decreases. The comparison fluids reveals that R134a shows a slightly higher frictional pressure gradient, especially at high vapor between the two fluids reveals that R134a shows a slightly higher frictional pressure gradient, qualities, due to its lower vapor density with the consequent higher vapor velocity at constant mass especially at high vapor qualities, due to its lower vapor density with the consequent higher vapor velocity.Fluids 2020,A 5, similarx conclusion can be drawn at a saturation temperature of 40 C. 12 of 16 velocity at constant mass velocity. A similar conclusion can be drawn at a saturation◦ temperature of 40 °C.

Figure 10. R134a and R513A frictional pressure gradient for the microﬁn tube vs. mean vapor quality Figure 10. R134a and R513A frictional pressure gradient for the microfin tube vs. mean vapor quality 2 1 at tsat = 30 ◦C. G expressed in [kg m− s− ]. at tsat = 30 °C. G expressed in [kg m−2 s−1].

5. Empirical Modelling

5.1. Smooth Tube This section provides a comparison between the experimental condensation heat transfer coefficients collected for the smooth tube and the values estimated by empirical correlations available in the literature. Figure 11 shows a comparison between the experimental heat transfer coefficients for the smooth tube and the values estimated by the correlations developed by Dobson and Chato [28], Shah [29], Cavallini et al. [30], and Dorao and Fernandino [31]. The correlation of Dobson and Chato [28] is able to estimate the experimental values with a relative deviation of 29.6%, an absolute deviation of 29.6%, and a standard deviation of 16.2%; the correlation of Shah [29] with a relative deviation of 16.3%, an absolute deviation of 18.0%, and a standard deviation of 18.4%; the correlation of Cavallini et al. [30] with a relative deviation of 5.9%, an absolute deviation of 9.7%, and a standard deviation of 13.6%; the correlation of Dorao and Fernandino [31] with a relative deviation of 2.5%, an absolute deviation of 15.3%, and a standard deviation of 22.8%. Therefore, the comparison reveals that the correlation of Cavallini et al. [30], even if it was developed for internal diameters larger than 3 mm, as well as that by Dorao and Fernandino [31], are the ones that better predict the experimental values. Fluids 2020, 5, 171 12 of 16

5. Empirical Modelling

5.1. Smooth Tube This section provides a comparison between the experimental condensation heat transfer coeﬃcients collected for the smooth tube and the values estimated by empirical correlations available in the literature. Figure 11 shows a comparison between the experimental heat transfer coeﬃcients for the smooth tube and the values estimated by the correlations developed by Dobson and Chato [28], Shah [29], Cavallini et al. [30], and Dorao and Fernandino [31]. The correlation of Dobson and Chato [28] is able to estimate the experimental values with a relative deviation of 29.6%, an absolute deviation of 29.6%, and a standard deviation of 16.2%; the correlation of Shah [29] with a relative deviation of 16.3%, an absolute deviation of 18.0%, and a standard deviation of 18.4%; the correlation of Cavallini et al.[30] with a relative deviation of 5.9%, an absolute deviation of 9.7%, and a standard deviation of 13.6%; the correlation of Dorao and Fernandino [31] with a relative deviation of 2.5%, an absolute deviation of 15.3%, and a standard deviation of 22.8%. Therefore, the comparison reveals that the correlation of Cavallini et al. [30], even if it was developed for internal diameters larger than 3 mm, as well as that by DoraoFluids 2020 and, 5, Fernandino x [31], are the ones that better predict the experimental values. 13 of 16

Figure 11. Experimental and calculated heat transfer coeﬃcients for the 2.5 mm ID smooth tube. Models Figure 11. Experimental and calculated heat transfer coefficients for the 2.5 mm ID smooth tube. developed by Dobson and Chato [28], Shah [29], Cavallini et al. [30], and Dorao and Fernandino [31]. Models developed by Dobson and Chato [28], Shah [29], Cavallini et al. [30], and Dorao and 5.2. MicroﬁnFernandino Tube [31].

5.2. MicrofinThis section Tube shows the comparison between the condensation heat transfer coefficients for the microfin tube against the values predicted by the correlation developed by Cavallini et al. [26]. A comparisonThis section between shows experimental the comparison and calculatedbetween the condensation condensation heat heat transfer transfer coeffi coefficientscients is reported for the inmicrofin Figure 12tube. The against correlation the values can predict predicted the experimental by the correlation values withdevel aoped relative, by absolute,Cavallini andet al. standard [26]. A deviationcomparison of 3.5%,between 21.2%, experimental and 31.1%, and respectively. calculated This con correlationdensation heat was transfer developed coefficients from data is obtainedreported in Figure 12. The correlation can predict the experimental values2 with1 a relative, absolute, and during condensation, for mass velocity in the range 80–890 kg m− s− , saturation temperature in standard deviation of 3.5%, 21.2%, and 31.1%, respectively. This correlation was developed from data the range 15 ◦C–70 ◦C, inside microfin tubes with inner diameters at the fin tip from 5.9 mm and − −2 −1 14.2obtained mm. Thisduring aspect condensation, may partly explainfor mass the velocity reason why in the correlationrange 80–890 tends kg to m underestimate s , saturation the experimentaltemperature in values the range at the −15 lowest °C–70 mass °C, inside velocities microfin in the tubes high with vapor inner quality diameters zone, atwhose the fin trendtip from is peculiar5.9 mm ofand small-sized 14.2 mm. microfin This aspect tubes. may partly explain the reason why the correlation tends to underestimate the experimental values at the lowest mass velocities in the high vapor quality zone, whose trend is peculiar of small-sized microfin tubes.

Figure 12. Experimental and calculated heat transfer coefficients for the 2.4 mm ID microfin tube. Model developed by Cavallini et al. [26]. Fluids 2020, 5, x 13 of 16

Figure 11. Experimental and calculated heat transfer coefficients for the 2.5 mm ID smooth tube. Models developed by Dobson and Chato [28], Shah [29], Cavallini et al. [30], and Dorao and Fernandino [31].

5.2. Microfin Tube This section shows the comparison between the condensation heat transfer coefficients for the microfin tube against the values predicted by the correlation developed by Cavallini et al. [26]. A comparison between experimental and calculated condensation heat transfer coefficients is reported in Figure 12. The correlation can predict the experimental values with a relative, absolute, and standard deviation of 3.5%, 21.2%, and 31.1%, respectively. This correlation was developed from data obtained during condensation, for mass velocity in the range 80–890 kg m−2 s−1, saturation temperature in the range −15 °C–70 °C, inside microfin tubes with inner diameters at the fin tip from 5.9 mm and 14.2 mm. This aspect may partly explain the reason why the correlation tends to underestimate the experimental values at the lowest mass velocities in the high vapor quality zone, Fluidswhose2020 trend, 5, 171 is peculiar of small-sized microfin tubes. 13 of 16

FigureFigure 12.12. ExperimentalExperimental andand calculatedcalculated heatheat transfertransfer coecoefficientsfficients forfor thethe 2.42.4 mmmm IDID microfinmicrofin tube.tube. ModelModel developed developed by by Cavallini Cavallini et et al. al. [ 26[26].]. 6. Conclusions The paper presented experimental data during condensation of R513A inside a smooth minitube with an inner diameter of 2.5 mm and inside a mini microfin tube with an inner diameter at the fin tip of 2.4 mm. The following working conditions were investigated: mass velocities in the range 2 1 200–1000 kg m− s− and saturation temperatures at the inlet of the test section of 40 ◦C and 30 ◦C. The test section for the smooth tube was verified during single-phase tests, involving liquid forced convection: once the heat flow rate calculated on the water side and the one on the refrigerant side were verified to be within 4%, and once the experimental heat transfer coefficients were found to be ± well estimated by correlations from the literature, the new test section was deemed verified. The condensation tests revealed that the heat transfer coefficient increases as both vapor quality and mass velocity increases, both for the smooth tube and for the microfin tube. However, the increase of the heat transfer coefficient with vapor quality is more evident with the microfin tube at mass 2 1 velocities up to 400 kg m− s− , highlighting its better convective properties, in particular at the lowest mass velocities. Concerning the saturation temperature, the higher the saturation temperature, the lower the heat transfer coefficient, for both the smooth tube and the microfin tube. The average Enhancement Factor, i.e., the ratio between the heat transfer coefficient of the smooth tube and the one of the microfin tube under the same operative conditions, is 1.79, with the highest 2 1 values shown at G = 300 kg m− s− at high vapor quality. Experimental R513A heat transfer coefficients and frictional pressure drops obtained for the microfin tube were compared against the counterpart with R134a. R134a shows higher values of condensation heat transfer coefficient, particularly at high vapor qualities, but also slightly higher values of frictional pressure gradients. The comparison against values predicted by empirical correlations revealed that the correlation of Cavallini et al. [30] and Dorao and Fernandino [31] are the most suitable to predict the condensation heat transfer coefficient inside the smooth tube, and the correlation of Cavallini et al. [26] is suitable to predict the condensation values for the microfin tube.

Author Contributions: Conceptualization, A.D.; validation, A.D., investigation, A.D.; writing—original draft preparation, A.D.; visualization, A.D.; supervision, L.R.; funding acquisition, L.R. All authors have read and agreed to the published version of the manuscript. Funding: This research was funded by MIUR through PRIN Project 2017F7KZWS_005 and by Università degli Studi di Padova through Project CPDA107382 and Project BIRD172935/17. Fluids 2020, 5, 171 14 of 16

Acknowledgments: Jacopo Carnio is gratefully acknowledged. Conﬂicts of Interest: The authors declare no conﬂict of interest.

Nomenclature

A area [m2] 1 1 c specific heat [J kg− K− ] D diameter [m] 2 1 G mass velocity [kg m− s− ] 1 h specific enthalpy [J kg− ] 2 1 HTC heat transfer coefficient [W m− K− ] L length [m] . 1 m mass flow rate [kg s− ] p pressure [bar] q heat flow rate [W] Pr Prandtl number [-] t temperature [◦C] x thermodynamic vapor quality [-] Greek symbols ∆p pressure drop [Pa] ∆T temperature difference [K] 1 1 λ thermal conductivity [W m− K− ] µ dynamic viscosity [Pa s] 3 ρ density [kg m− ] 1 σ surface tension [N m− ] Subscripts f frictional i inner in at the inlet L of the saturated liquid out at the outlet PC at the pre-condenser r reduced ref of the refrigerant sat saturated TS in the test section V of the saturated vapor vs of the superheated vapor w of the water wall at the wall

References

1. Hodnebrog, O.; Etminiam, M.; Fuglestvedt, J.S.; Marston, G.; Myhre, G.; Nielsen, C.J.; Shine, K.P.; Wallington, T.J. Global warming potentials and radiative efficiencies of halocarbons and related compounds: A comprehensive review. Rev. Geophys. 2013, 51, 300–378. [CrossRef] 2. Matkovic, M.; Cavallini, A.; Del Col, D.; Rossetto, L. Experimental study on condensation heat transfer inside a single circular minichannel. Int. J. Heat Mass Transf. 2009, 52, 2311–2323. [CrossRef] 3. Zhang, H.Y.; Li, M.J.; Liu, N.; Wang, B.X. Experimental investigation of condensation heat transfer and pressure drop of R22, R410A and R407C in mini.tubes. Int. J. Heat Mass Transf. 2012, 55, 3522–3532. [CrossRef] 4. Guo, Q.; Li, M.; Gu, H. Condensation heat transfer characteristics of low-GWP refrigerants in a smooth horizontal mini tube. Int. J. Heat Mass Transf. 2018, 126, 26–38. [CrossRef] 5. Azzolin, M.; Bortolin, S.; Del Col, D. Convective condensation at low mass flux: Effect of turbulence and tube orientation on the heat transfer. Int. J. Heat Mass Transf. 2019, 144, 118646. [CrossRef] Fluids 2020, 5, 171 15 of 16

6. Azzolin, M.; Berto, A.; Bortolin, S.; Moro, L.; Del Col, D. Condensation of ternary low GWP zeotropic mixtures inside channels. Int. J. Refrig. 2019, 103, 77–90. [CrossRef] 7. Murphy, D.L.; Macdonald, M.P.; Mahvi, A.J.; Garimella, S. Condensation of propane in vertical minichannels. Int. J. Heat Mass Transf. 2019, 137, 1154–1166. [CrossRef] 8. Gao, L.; Honda, T.; Koyama, S. Experiments on flow boiling heat transfer of almost pure CO2 and CO2-oil mixtures in horizontal smooth and microfin tube. HVAC&R 2007, 13, 415–425. 9. Kim, Y.J.; Cho, J.M.; Kim, M.S. Experimental study on the evaporative heat transfer and pressure drop of CO2 flowing upward in vertical smooth and micro-fin tubes with the diameter of 5 mm. Int. J. Refrig. 2008, 31, 771–779. [CrossRef] 10. Dang, C.; Haraguchi, N.; Hihara, E. Flow boiling heat transfer of carbon dioxide inside a small-sized microfin tube. Int. J. Refrig. 2010, 33, 655–663. [CrossRef] 11. Qu, W.; Mudawar, I. Measurement and correlation of critical heat flux in two-phase micro.channel heat sinks. Int. J. Heat Mass Transf. 2004, 47, 2045–2059. [CrossRef] 12. Diani, A.; Campanale, M.; Cavallini, A.; Rossetto, L. Low GWP refrigerants condensation inside a 2.4 mm ID microfin tube. Int. J. Refrig. 2018, 86, 312–321. [CrossRef] 13. Bashar, M.K.; Nakamura, K.; Kariya, K.; Miyara, A. Experimental study of condensation heat transfer and pressure drops inside a small diameter microfin and smooth tube at low mass flux condition. Appl. Sci. 2018, 8, 2146. [CrossRef] 14. Diani, A.; Cavallini, A.; Rossetto, L. R1234yf flow boiling heat transfer inside a 2.4 mm microfin tube. Heat Transf. Eng. 2017, 38, 303–312. [CrossRef] 15. Jige, D.; Inoue, N. Flow boiling heat transfer and pressure drop of R32 inside 2.1 mm, 2.6 mm and 3.1 mm microfin tubes. Int. J. Heat Mass Transf. 2019, 123, 566–573. [CrossRef] 16. Hirose, M.; Jige, D.; Inore, N. Optimum fin geometries on condensation heat transfer and pressure drop of R1234ze(E) in 4 mm outside diameter horizontal microfin tubes. Sci. Technol. Built Environ. 2019, 25, 1271–1280. [CrossRef] 17. Heredia-Aricapa, Y.; Belman-Flores, J.M.; Mota-Babiloni, A.; Serrano-Arellano, J.; Garcìa-Pablon, J.J. Overview of low GWP mixtures for replacement of HFC refrigerants: R134a, R404A and R410A. Int. J. Refrig. 2020, 111, 113–123. [CrossRef] 18. Kedzierski, M.A.; Lin, L.; Kang, D. Pool boiling of low-global warming potential replacements for R134a on a reentrant cavity surface. J. Heat Transf. 2018, 140, 121502. [CrossRef] 19. Diani, A.; Rossetto, L. R513A flow boiling heat transfer inside horizontal smooth tube and microfin tube. Int. J. Refrig. 2019, 107, 301–314. [CrossRef] 20. Diani, A.; Rossetto, L. Characteristics of R513A evaporation heat transfer inside small-diameter smooth and microfin tubes. Int. J. Heat Mass Transf. 2020, 162, 120402. [CrossRef] 21. Lemmon, E.W.; Huber, M.L.; McLinden, M.O. NIST Standard Reference Database 23, NIST Reference Fluid Thermodynamic and Transport Properties, REFPROP, Version 10.0, Standard Reference Data Program; National Institute of Standards and Technology: Gaithersburg, MD, USA, 2018. 22. Kline, S.J.; McKlintock, F.A. Describing the uncertainties ins single-sample experiments. Mech. Eng. 1953, 75, 3–8. 23. Dittus, F.W.; Boelter, L.M.K. Heat transfer in automotive radiators of the annular type. Univ. Calif. Publ. Eng. 1930, 2, 443–461. 24. Petukhov, B.S.; Popov, V.N. Theoretical calculation of heat exchange and frictional resistance in turbulent flow in tubes of an incompressible fluid with variable physical-properties. High Temper. 1963, 1, 69–83. 25. Doretti, L.; Fantini, F.; Zilio, C. Flow patterns during condensation of three refrigerants: Microfin tube vs. smooth tube. In Proceeding of the IIR International Conference on Thermophysical Properties and Transfer Processes pf Refrigerant, Vicenza, Italy, 31 August–2 September 2005. 26. Cavalini, A.; Del Col, D.; Mancin, S.; Rossetto, L. Condensation of pure and near-azeotropic refrigerants in microfin tubes: A new computational procedure. Int. J. Refrig. 2009, 32, 162–174. [CrossRef] 27. Rouhani, S.Z.; Axelsson, E. Calculation of void fraction in the subcooled and boiling region. Int. J. Heat Mass Transf. 1970, 13, 383–393. [CrossRef] 28. Dobson, M.K.; Chato, J.C. Condensation heat transfer inside smooth horizontal tubes. J. Heat Transf. ASME 1998, 120, 193–213. [CrossRef] Fluids 2020, 5, 171 16 of 16

29. Shah, M.M. An improved and extended general correlation for heat transfer during condensation in plain tubes. HVAC&RRes. 2009, 15, 889–913. 30. Cavallini, A.; Del Col, D.; Doretti, L.; Matkovic, M.; Rossetto, L.; Censi, G. Condensation in horizontal smooth tubes: A new heat transfer model for heat exchanger design. Heat Transf. Eng. 2006, 27, 31–38. [CrossRef] 31. Dorao, C.A.; Fernandino, M. On the heat transfer deterioration during condensation of binary mixtures. Appl. Phys. Lett. 2019, 114, 171902. [CrossRef]

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