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R513A condensation heat transfer inside tubes: Microfin tube vs. smooth tube
International Journal of Heat and Mass Transfer 152 (2020) 119472
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R513A condensation heat transfer inside tubes: Microfin tube vs. smooth tube Andrea Diani∗, Pierfrancesco Brunello, Luisa Rossetto Department of Industrial Engineering, University of Padova, Via Venezia 1, 35131, Padova, Italy
a r t i c l e
i n f o
Article history: Received 26 November 2019 Revised 21 January 2020 Accepted 2 February 2020
Keywords: Condensation R513A Smooth tube Microfin tube Heat transfer coefficient
a b s t r a c t The imminent phase-down of the common refrigerant R134a is calling for lower GWP alternatives. Real alternatives must have a lower global warming impact and they should be not flammable. In this context, R513A (azeotropic mixture made of R1234yf and R134a at 56% and 44% by mass) has been proposed as alternative to R134a due to its similar thermodynamic and transport properties and due to its lower GWP. This paper proposes a direct comparison between the thermal performances of a 3.5 mm ID smooth tube and those of a 3.4 mm ID microfin tube, during R513A condensation under the same working conditions of vapor quality (from 0.10 to 0.99), of mass velocity (from 100 to 10 0 0 kg m−2 s−1 ), and of saturation temperature (30 °C and 40 °C). The comparison permits to highlight in which working conditions the microfin tube leads to a real heat transfer augmentation which is higher than the mere increase of heat transfer area. In the end, the experimental heat transfer coefficients, both for the smooth tube and for the microfin tube, are compared against values calculated with empirical correlations from the open literature. © 2020 Elsevier Ltd. All rights reserved.
1. Introduction The new environmental laws are calling for the use of refrigerants with always lower values of Global Warming Potential (GWP). HydroFluoroCarbons (HFCs) have high values of GWP, and, even if they are still widely used in actual HVAC and refrigeration systems, they cannot be used for the new installations. Therefore, new pure refrigerants or refrigerants mixtures with lower values of GWP, are in great demand to accomplish the new stricter and stricter environmental regulations. These new solutions should have thermodynamic and transport properties similar to those of the actual implemented refrigerants, and they should be preferable not flammable, to make a direct refrigerant drop in possible. R134a is among the most common refrigerants implemented in a wide range of refrigeration and air conditioning systems, such as domestic and industrial air conditioning, automotive air conditioning, and centrifugal chillers. In these contexts, R513A could be a direct drop in substitute of R134a. R513A is an azeotropic mixture made of R1234yf and R134a (56% and 44% in mass, respectively). R1234yf has an extremly low GWP (lower than 1 [1]), but it is classified as mildly flammable (A2L flammability class), whereas R134a
∗
Corresponding author. E-mail address: [email protected] (A. Diani).
https://doi.org/10.1016/j.ijheatmasstransfer.2020.119472 0017-9310/© 2020 Elsevier Ltd. All rights reserved.
is not flammable (A1 flammability class). Therefore, the proposed mixture R513A benefits from the not flammability of R134a and from the low GWP of R1234yf: R513A is a refrigerant which belongs to the A1 flammability class and it has a value of GWP of about 630. In order to develop HVAC and refrigeration equipment which uses R513A as working fluid, its performance during flow boiling and condensation must be studied. Besides the use of new refrigerants with lower warming impact, the refrigerant charge reduction is another key issue that the scientific community has to cope to reduce the global warming. The refrigerant charge reduction can be achieved by implementing systems with small-sized tubes. Two-phase heat transfer of new lower GWP refrigerants in small-sized tubes is a topic which deserves more studies in the literature. In particular, no studies regarding R513A condensation can be found in the literature. Among the studies relative to refrigerants condensation inside small-sized tubes, Hossain et al. [2] proposed experimental measurements of heat transfer coefficients and pressure drops of R1234ze(E), R32 and R410A in a horizontal smooth copper tube with an inner diameter of 4.25 mm. Tests were carried out for mass velocities from 150 to 400 kg m−2 s−1 , at saturation temperatures between 35 °C and 45 °C. It was found that R1234ze(E) heat transfer coefficients were about 20–45% lower than those of R32 and 10–30% higher than those of R410A for a saturation temperature of 40 °C.
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Nomenclature D e EF g G h HTC
diameter [m] error [%] Enhancement Factor [-] gravity acceleration [m s−2 ] mass velocity [kg m−2 s−1 ] specific enthalpy [J kg−1 ] heat transfer coefficient [W m−2 K−1 ] dimensionless gas velocity
G·x D·ρV ·(ρL −ρV )·g
JG = √ m˙ Rx q t x
ρ
μ
0.9 · ( V )0.5 · ( L )0.1 Xtt = ( 1−x x ) ρ μ L
V
[-] mass flow rate [kg s−1 ] enhancement area [-] heat flow rate [W] temperature [°C] thermodynamic vapor quality [-] Martinelli parameter [-]
Greek symbols T temperature difference [K] ρ density [kg m−3 ] σN standard deviation [%] Subscripts abs absolute in inlet L saturated liquid out outlet rel relative ref refrigerant sat saturation TS test section V saturated vapor Yang and Nalbandian [3] analyzed flow condensation heat transfer and pressure drop of refrigerants R1234yf and R134a in a 4.0 mm ID small circular tube for mass velocities from 200 to 1200 kg m−2 s−1 . The flow condensation heat transfer performance was affected by flow patterns at different operating conditions: at the lowest mass velocity, gravity was the major force that controls the heat transfer mechanism and the flow pattern was plug. On the contrary, at the highest mass velocity, gravity effects were negligible even at low vapor qualities. The condensation heat transfer coefficients of R134a were higher than those of R1234yf at high vapor qualities but similar at low vapor qualities. Azzolin et al. [4] investigated the condensation heat transfer coefficient and two-phase frictional pressure drop of ternary mixtures R455A (R32, R1234yf and R744 at 21.5/75.5/3.0% by mass composition) and R452B (R32, R1234yf and R125 at 67.0/26.0/7.0% by mass composition) inside a 0.96 mm ID minichannel and inside a 8.0 mm ID channel. Condensation tests were run at 40 °C of saturation temperature, at mass velocities in the range 100– 600 kg m−2 s−1 for the 8.0 mm diameter test section, and in the range 20 0–80 0 kg m−2 s−1 for the 0.96 mm diameter test section. The experimental results showed that the mixture R452B presented higher heat transfer coefficients with respect to R455A for all the working conditions. The heat transfer coefficients inside the 0.96 mm diameter minichannel were higher than those measured inside the 8.0 mm diameter channel, and the effect of the diameter was more important at high mass flux. Condensation of the zeotropic mixture R450A and of R134a was studied by Jacob et al. [5]. R450A is composed of R134a and R1234ze(E) (42/58% by mass). Experiments were carried out in a
horizontal smooth tube with an inner diameter of 4.7 mm, for a range of mass velocities from 100 to 550 kg m−2 s−1 with saturation temperatures of 45 °C and 55 °C. Distributed fiber optic temperature sensors were installed to measure the coolant axial temperature profile during condensation. Collected data showed that the heat transfer coefficients increase with increasing vapor quality and mass flux, whereas they decrease with increasing saturation temperature. Compared to R134a, R450A showed 5% lower heat transfer coefficients at high vapor quality, whereas the pressure drop of R450A was, on average, 8% higher than that of R134a. A smart option to increase the heat transfer coefficient is represented by microfin tubes, which, besides having a larger heat transfer area, promote the turbulence of the liquid film, enhancing the heat transfer mechanism. Small sized microfin tubes were initially proposed for carbon dioxide applications [6,7], due to its high working pressure. Studies regarding two-phase flow of refrigerants inside small-sized microfin tubes have begun in the last years. At the moment, no work about R513A condensation inside neither microfin tubes nor smooth tubes can be found in the literature. Among the studies relative to refrigerants condensation inside small-sized microfin tubes, Li et al. [8] studied R22 and R410A during condensation inside smooth and microfin tubes. The tested tubes had outer diameters of 5.0 mm and 9.52 mm. A refrigerant superheating of 4–7 °C was set at the inlet of the test section. The mass velocity ranged between 100 and 400 kg m−2 s−1 for the 9.52 mm OD tubes and from 300 to 550 kg m−2 s−1 for the 5.0 mm OD tubes. The heat transfer coefficient of the microfin tube was around 1.65–2.55 times that of the smooth one. The microfin tubes yielded 30% higher pressure drops compared to those of the smooth ones. Hirose et al. [9] investigated the condensation heat transfer and pressure drop characteristics of smooth and microfin smalldiameter tubes, with outer diameter of 4.0 mm, with R32, R152a and R410A as working fluids. The condensation heat transfer coefficients and pressure drops were measured in the range of mass velocities from 100 to 400 kg m−2 s−1 , at a saturation temperature of 35 °C. The frictional pressure drop of the microfin tube was approximately 1.6 times greater than that of the smooth tube for each refrigerant and mass velocity. The heat transfer coefficient of the microfin tube was approximately 2–7 times greater than that of the smooth tube for R32 at 200 kg m−2 s−1 . Condensation heat transfer of R1234yf, R1234ze(E) and R134a inside microfin tubes with outer diameter of 3.0 mm and 4.0 mm was studied by Diani et al. [10,11]. Experiments were carried out for mass velocities from 100 to 10 0 0 kg m−2 s−1 in the 4.0 mm OD microfin tube, and from 300 to 10 0 0 kg m−2 s−1 in the 3.0 mm OD microfin tube, for saturation temperature at the inlet of the test section of 30 °C and 40 °C. Generally speaking, R1234ze(E) shows slightly higher heat transfer coefficients compared to R134a, but also higher pressure drops. This study investigates R513A condensation inside a smooth tube with an inner diameter of 3.5 mm and inside a microfin tube with an inner diameter at the fin tip of 3.4 mm. A new test section to study condensation inside a 3.5 mm ID smooth tube is proposed and verified during R513A liquid forced convection tests. Present working conditions for the condensation tests are: mass velocities between 100 and 10 0 0 kg m−2 s−1 , saturation temperatures of 30 °C and 40 °C, vapor qualities from 0.1 to 0.99. This wide range of operative conditions permits to understand how each working parameter affects the thermal behavior of R513A condensing inside the smooth and microfin tubes. The comparison between thermal performances of the microfin tube and those of the smooth tube are presented, in order to highlight in which conditions the microfin tube is more effective. Finally, the experimental collected heat transfer coefficients for both smooth and microfin tubes are
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Fig. 1. Schematic of the experimental facility.
Table. 1 Values of instruments accuracy. T-type thermocouples T-type thermopiles Coriolis effect flowmeter Volumetric flowmeters at evaporator and pre-condenser Volumetric flowmeter at test section Absolute pressure transducers
±0.05 K ±0.03 K ±0.10% of reading ±0.25% of reading ±0.50% of reading ±1950 Pa
compared against values predicted by empirical correlations available in the open literature. 2. Experimental facility The experimental set up is located in the Heat Transfer in Micro-Geometries Laboratory of the Department of Industrial Engineering of the University of Padova. As shown in Fig. 1, the experimental facility consists of five loops: the refrigerant loop, the cooling water loop at the pre-condenser, at the test section and at the post-condenser, and the hot water loop at the evaporator. The refrigerant is pumped through the circuit by means of a magnetically coupled gear pump, it is vaporized and superheated in a brazed plate heat exchanger fed with the hot water. Superheated vapor then partially condenses in a pre-condenser fed with cold water, which is supplied by a water-cooled chiller, to achieve the set vapor quality at the inlet of the test section. The refrigerant enters into the test section at a known mass velocity and vapor quality where it condenses. The fluid leaves the test section and enters into a post-condenser, which is a brazed plate heat exchanger, where it is fully condensed and subcooled. The subcooled liquid passes through a drier filter and then is sent back to the boiler by the pump. A damper connected to the compressed air line operates as pressure regulator to control the saturation conditions in the refrigerant loop. Refrigerant temperatures and pressures are measured throughout the circuit to know its thermodynamic state. The accuracies of the instruments implemented in the facility are reported in Table 1. 3. Test sections A schematic of the test section for the 3.5 mm ID smooth tube is reported in Fig. 2. A 1.9 mm ID smooth tube, where cold water flows, is wrapped around the test smooth tube, as reported in the figure. T-type thermocouples, with an accuracy of ±0.05 K, are attached on the external wall of the inlet and outlet curves, in order
Fig. 2. Schematic of the test section for smooth tube.
the measure the inlet and outlet temperatures of the water flowing inside. The cold water is supplied by a thermostatic bath and its volumetric flow rate is measured by a magnetic flowmeter having an accuracy of ±0.50% of the reading. Six T-type thermocouples are glued in as many pits milled on the external wall of the test smooth tube, to monitor the wall temperature. The so created test section (test smooth tube with the wrapped smooth tube) is inserted in a U-shaped aluminum housing, which was filled with an alloy of tin/lead. The alloy, once solidified, guarantees a thermal contact between the smooth tube under investigation and the wrapped smooth tube. The whole test section is finally insulated with AF/Armaflex, to limit as much as possible the heat losses through the ambient. The length for the thermal measurements is 250 mm. Inlet and outlet pressure ports were soldered to the tube under investigation: the inlet pressure port is connected to an absolute pressure transducer having an accuracy of ±1950 Pa, whereas both the inlet and the outlet pressure ports are connected to a differential pressure transducer having an accuracy of ±25 Pa. The same concept was applied to the test section for the microfin tube. The test section is similar, but in this case a 3.4 mm ID microfin tube is placed instead of the 3.5 mm ID smooth tube.
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Details of the test section for the microfin tube can be found in Diani et al. [12]. The microfin tube under investigation has an inner diameter at the fin tip of 3.4 mm and the outer diameter is 4.0 mm. It has 40 internal fins, each fin is 0.12 mm high with an apex angle of 43°. The helix angle is 18°. From its geometrical characteristics, the ratio between its total heat transfer area and the area of an equivalent smooth tube is 1.69. Both the test sections are horizontally located in the experimental facility previously shown. 4. Data analysis The thermal performances of both the smooth tube and of the microfin tube are given in terms of heat transfer coefficient HTC, calculated as:
HTC =
qTS
AD · t sat − t wall
(1)
where AD is the inner surface area of the smooth tube, or the inner surface area of an equivalent smooth tube having an inner diameter equal to the diameter at the fin tip in the case of the microfin tube. t¯sat is the average saturation temperature, calculated as arithmetic average between inlet and outlet saturation temperatures. Inlet and outlet saturation temperatures are calculated from the values of inlet and outlet pressures with the software Refprop 10 [13], which implements new specific correlations for the azeotropic mixture R513A. t¯wall is the wall temperature. Results will also be referred to the mean vapor quality xmean , arithmetic average between inlet and outlet vapor quality. The inlet vapor quality xin is calculated as:
xin =
hT S,in − hL hV − hL
(2)
with hL and hV the specific enthalpies of the saturated liquid and vapor, respectively, calculated from the knowledge of the inlet pressure. hTS,in derives from an energy balance at the precondenser. The outlet vapor quality xout is similarly calculated, and in this case the specific enthalpy at the outlet of the test section derives from an energy balance at the test section itself. More details about the data analysis can be found in Diani et al. [12]. By following the procedure for the calculation of the uncertainty as suggested by Kline and McKlintock [14], the mean, maximum and minimum uncertainties on the heat transfer coefficient for the smooth tube are 6.2%, 7.6% and 5.0%, respectively, whereas for the microfin tube are 4.0%, 5.2% and 2.8%, respectively. The mean uncertainty on the mean vapor quality is ±0.03. 5. Experimental results This section presents the experimental results during R513A single phase flow inside the smooth tube and during R513A condensation inside both the smooth tube and the microfin tube. Single phase tests were necessary in order to verify the new test section with the 3.5 mm ID smooth tube, i.e. to verify the heat balance between the heat flow rate calculated on the refrigerant side and the one on the cooling water side. The test section with the 3.4 mm ID microfin tube was previously verified by Diani et al. [12]. 5.1. Single phase tests inside smooth tube Preliminary liquid forced convection tests were carried out inside the 3.5 mm ID smooth tube. These tests were run for mass velocities from 400 to 10 0 0 kg m−2 s−1 . The mass velocity is defined as:
G=
4 · m˙ re f
π · D2i
(3)
Fig. 3. Experimental and predicted single phase heat transfer coefficients versus mass velocity for smooth tube.
where Di is the inner diameter. During these tests, R513A pressure was at about 11.1 bar with a subcooling at the inlet of the test section between 16 K and 21 K, to ensure subcooled conditions. Reynolds number was between 7664 and 20,403. The difference between the measured heat flow rate on the water side and the one on the refrigerant side is within ±3 W. Fig. 3 shows the single phase heat transfer coefficients plotted against the mass velocity. Single phase heat transfer coefficients are calculated using Eq. (1) where t¯sat is replaced by t¯re f . t¯re f is the arithmetic averaged value between two refrigerant temperatures measured by two T-type thermocouples located on the external wall of the tested smooth tube, upstream and downstream of the aluminum housing. As it appears from the figure, the single phase heat transfer coefficient increases with increasing mass velocity. The experimental values are compared against the values predicted by two empirical correlations: the correlation of Petukhov and Popov [15] is able to estimate the experimental values of heat transfer coefficients with, on average, a relative deviation of 11.6%, whereas the correlation of Dittus and Boelter [16] is able to estimate the experimental values with, on average, a relative deviation of −0.3%. Therefore, since the heat flow rates calculated on refrigerant and water sides are within ±3 W, and since the empirical correlations well estimate the experimental single phase heat transfer coefficients, the test section for smooth tube is deemed verified. 5.2. Condensation tests inside smooth tube Condensation tests inside the smooth tube were carried out for mass velocities, calculated as reported in Eq. 3, from 100 to 10 0 0 kg m−2 s−1 for a saturation temperature at the inlet of the test section of 30 °C, and from 200 to 10 0 0 kg m−2 s−1 for a saturation temperature at the inlet of the test section of 40 °C. The heat transfer coefficients were calculated from the experimental tests as reported in Eq. (1). Every single experimental point is the average value of 100 data recorded in steady state conditions with a frequency of 1 Hz. Fig. 4 reports the condensation heat transfer coefficients for the smooth tube plotted against the mean vapor quality at a saturation temperature of 30 °C (top) and at 40 °C (bottom). Generally speaking, the heat transfer coefficient increases as both vapor quality and mass velocity increase. More in detail, the heat transfer coefficient at the mass velocity G = 100 kg m−2 s−1 is weakly affected by vapor quality, passing from a value of 1280 W m−2 K−1 at xmean
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Fig. 5. Data for smooth tube plotted in the flow regime map developed by Cavallini et al. [17]. G expressed in (kg m−2 s−1 ).
mass velocity, highlighting the higher and higher convective contribution as mass velocity increases. The comparison between the figure on the top and the one on the bottom of Fig. 4 reveals the effect of the saturation temperature on the heat transfer behavior. At constant mass velocity, a higher saturation temperature implies a higher vapor density, with a consequent lower vapor velocity. This is reflected on the experimental heat transfer coefficients: at fixed mass velocity and mean vapor quality, the higher the saturation temperature, the lower the heat transfer coefficient, due to a lower convective contribution. 5.3. Condensation tests inside microfin tube
Fig. 4. Experimental condensation heat transfer coefficient versus mean vapor quality for smooth tube. Data at tsat = 30 °C (top) and at tsat = 40 °C (bottom). G expressed in (kg m−2 s−1 ).
of 0.31 to a value of 2070 W m−2 K−1 at xmean of 0.79, whereas, for instance, the heat transfer coefficient at the mass velocity G = 400 kg m−2 s−1 in the same mean vapor quality region passes from about 2500 W m−2 K−1 to 50 0 0 W m−2 K−1 : this can be explained considering the different flow regimes that occur at the different mass velocities. At G = 100 kg m−2 s−1 , the flow seems to be stratified, with consequent low heat transfer coefficient. On the contrary, at G = 400 kg m−2 s−1 , the flow seems to be annular, due to the higher contribute of mass velocity and of vapor quality. This is reflected in Fig. 5, which reports the experimental data for the smooth tube in the flow regime map developed by Cavallini et al. [17]. The dashed line represents the transition from Tdependent flow regime (below the line) to T-independent flow regime (above the line). The T-dependent regime is associated to gravity controlled flow patterns, i.e. wavy and smooth stratified regimes, whereas the T-independent regime is associated to the shear stress forces controlled regimes, i.e. annular flow. The flow pattern map reveals that all the data at G = 100 kg m−2 s−1 are in stratified flow regime, whereas a transition from stratified to annular flow occurs at G = 200 kg m−2 s−1 , and, as a result, the heat transfer coefficients at G = 200 kg m−2 s−1 are higher than those at G = 100 kg m−2 s−1 especially at high mean vapor quality. Almost all the other experimental data fall above the transition line. The slope of the heat transfer coefficient curve increases with the
Condensation tests inside the microfin tube were carried for mass velocities, calculated as reported in Eq. (3) where Di represents, in this case, the inner diameter at the fin tip of the microfin tube, from 100 to 10 0 0 kg m−2 s−1 for a saturation temperature at the inlet of the test section of 30 °C, and from 200 to 10 0 0 kg m−2 s−1 for a saturation temperature at the inlet of the test section of 40 °C. The heat transfer coefficients were calculated from the experimental tests as reported in Eq. (1). Every single experimental point is the average value of 100 data recorded in steady state conditions with a frequency of 1 Hz. Fig. 6 reports the condensation heat transfer coefficients for the microfin tube plotted against the mean vapor quality at a saturation temperature of 30 °C (top) and at 40 °C (bottom). Generally speaking, the heat transfer coefficient increases as both vapor quality and mass velocity increase, but in this case the effect of vapor quality seems to be stronger compared to the case of the smooth tube. Considering the mass velocity G = 100 kg m−2 s−1 , there is a sudden change of the slope of the heat transfer coefficient curve at a mean vapor quality of about 0.6. This can be linked to a transition of the flow regime from stratified to annular flow. Fig. 7 reports the experimental data for the microfin tube plotted in the flow regime map developed by Doretti et al. [18] and suggested by Cavallini et al. [19]. The black continuous line represents the transition from T-dependent regime (below the line) to the T-independent regime (above the line) for microfin tube. As comparison, the transition for the case of smooth tube is also reported. The transition in case of microfin tube occurs at lower dimensionless gas velocity at constant Martinelli parameter compared to the case of the smooth tube. This is reflected on the heat transfer coefficients reported in Fig. 6. At G = 100 kg m−2 s−1 for mean vapor qualities higher than 0.6, the flow is annular, and thus
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Fig. 7. Data for microfin tube plotted in the flow regime map developed by Doretti et al. [18]. G expressed in (kg m−2 s−1 ).
Fig. 6. Experimental condensation heat transfer coefficient versus mean vapor quality for microfin tube. Data at tsat = 30 °C (top) and at tsat = 40 °C (bottom). G expressed in (kg m−2 s−1 ). Fig. 8. Experimental condensation frictional pressure drop versus mean vapor quality for microfin tube. Data at tsat = 30 °C. G expressed in (kg m−2 s−1 ).
the heat transfer mechanism is enhanced, whereas all the data at G = 100 kg m−2 s−1 are in stratified flow regime for smooth tube. In the case of the microfin tube, the flow can be considered annular for all the data from the mass velocity G = 200 kg m−2 s−1 . Conclusions similar to those for the case of the smooth tube can be drawn for the effect of the saturation temperature: the higher the saturation temperature, the lower the heat transfer coefficient, due to a lower convective effect. The comparison between Figs. 4 and 6 reveals that the low mass velocities are especially enhanced by the microfins, because the microfins tend to thin the liquid film thickness around them, with a consequent lower thermal resistance, and because the microfin tube makes the transition to annular flow happen at lower mass velocities. Therefore, the microfin tube is effective particularly at low mass velocities. Fig. 8 reports the frictional pressure drop for microfin tube plotted against the mean vapor quality at different mass velocities for the saturation temperature at the inlet of the test section of 30 °C. The model of Rouhani and Axelsson [20] was considered to estimate the void fraction, necessary for the calculation of momentum pressure drop. As it appears from the figure, at constant mean vapor quality, the higher the mass velocity, the higher the frictional pressure gradient. At constant mass velocity, the frictional pressure gradient increases as mean vapor quality increases up to a maximum value, after which it slightly decreases.
Taking for instance mass velocities of 400 and 800 kg m−2 s−1 , at a mean vapor quality of 0.8, the frictional pressure gradient at G = 800 kg m−2 s−1 is more than 3 times the values at G = 400 kg m−2 s−1 . Indeed, considering Fig. 6 (top), the heat transfer coefficient at xmean = 0.8 at G = 800 kg m−2 s−1 is only 1.3 times the value at G = 400 kg m−2 s−1 . Thus, in real applications, lower mass velocities could be preferable to higher mass velocities. 5.4. Comparison between microfin and smooth tube In this paragraph, the thermal performances of the microfin tube are systematically compared against those of the smooth tube. The comparison is based on the parameter called Enhancement Factor, defined as the ratio between the heat transfer coefficient of the microfin tube and that of the smooth tube under the same working conditions, i.e.:
EF =
HT Cmicro f in HT Csmooth
(4)
Fig. 9 shows the enhancement factor plotted against the mass velocity. The dotted horizontal line represents Rx, which is the ratio between the total heat transfer area of the microfin tube and that of an equivalent smooth tube with an inner diameter equal
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Fig. 9. Enhancement factor versus mass velocity.
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Fig. 10. Calculated and experimental heat transfer coefficients for smooth tube. Model by Dobson and Chato. [21].
to the diameter at the fin tip of the microfin tube under investigation. When the Enhancement Factor is approximately equal to the value of Rx, it means that the heat transfer augmentation due to the microfins is mainly due to the increased heat transfer area. On the contrary, when the Enhancement Factor is higher than Rx, it means that the heat transfer augmentation is not only due to the increased heat transfer area, but also due to higher convective effects led to the microfins. As it appears from Fig. 9, EF tends to decrease as mass velocity increases, and thus the microfin tube is effective in the augmentation of the heat transfer mechanism especially at low mass velocities, due to the presence of the microfins which tend to thin the liquid film thickness. Considering all the experimental data, the mean Enhancement Factor is 2.0, and it reaches values higher than 4 at G = 200 kg m−2 s−1 . 6. Empirical models 6.1. Models for smooth tube In this section, the experimental results of the heat transfer coefficients collected during R513A condensation inside the 3.5 mm ID smooth copper tube are compared against values predicted by empirical models available in the open literature. Dobson and Chato [21] presented a model to estimate the heat transfer coefficient during condensation inside smooth tubes. The experimental data from which the model was developed comprises data during R12, R22, R134a and near-azeotropic blends of R32/R125 at 50/50% and 60/40% compositions, which condense inside smooth tubes with diameters ranging between 3.14 mm and 7.04 mm and mass velocities from 53 to 807 kg m−2 s−1 . Fig. 10 shows a comparison between calculated and experimental heat transfer coefficients for smooth tube. The model is able to estimate the experimental values with a relative, absolute and standard deviation of 36.6%, 36.3% and 10.0%, respectively. The model of Shah [22] was developed for predicting heat transfer coefficients during condensation inside plain tubes. The data used for the correlation’s validation include 22 fluids, among which water, halocarbon refrigerants, hydrocarbon refrigerants and organic fluids, condensing in horizontal, vertical and downwardinclined tubes. The range of parameters includes tube diameters from 2 to 49 mm, reduced pressures from 0.0 0 08 to 0.9, and mass velocities from 4 to 820 kg m−2 s−1 . Fig. 11 shows a comparison between calculated and experimental condensation heat transfer coefficients for smooth tubes. The model is able to estimate the
Fig. 11. Calculated and experimental heat transfer coefficients for smooth tube. Model by Shah [22]. Table. 2 Deviations between calculated and experimental heat transfer coefficients. Models for smooth tube.
Model of Dobson and Chato [21] Model of Shah [22] Model of Cavallini et al. [17]
erel
eabs
σ std
36.2% 22.3% 13.5%
36.3% 23.0% 13.5%
10.0% 10.2% 4.5%
experimental values with a relative, absolute and standard deviation of 22.3%, 23.0% and 10.2%, respectively. The model of Cavallini et al. [17] was developed to estimate the heat transfer coefficients during condensation of refrigerants flowing inside horizontal smooth tubes with inner diameters larger than 3 mm. For the validation, the model was compared against heat transfer coefficients collected during condensation of HCFCs, HFCs, HCs, carbon dioxide, ammonia, water, pure fluids or near azeotropic mixtures, for mass velocities from 24 to 2240 kg m−2 s−1 , and saturation temperatures from −15 °C to 302 °C. Fig. 12 shows a comparison between calculated and experimental heat transfer coefficients. The model well estimates the experimental values of heat transfer coefficients, with a relative, absolute and standard deviation of 13.5%, 13.5% and 4.5%, respectively. Table 2 summarizes the deviations of each model.
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Fig. 12. Calculated and experimental heat transfer coefficients for smooth tube. Model by Cavallini et al. [17].
Fig. 14. Calculated and experimental heat transfer coefficients for microfin tube. Model by Cavallini et al. [19]. Table. 3 Deviations between calculated and experimental heat transfer coefficients. Models for microfin tube.
Model of Kedzierski and Goncalves [23] Model of Cavallini et al. [19]
erel
eabs
σ std
15.0% −4.4%
19.0% 14.0%
29.5% 19.3%
duced pressure between 0.1 and 0.67, mass velocities between 80 and 890 kg m−2 s−1 . The microfin tubes had inside diameters between 5.9 mm and 14.2 mm, fin heights from 0.15 mm to 0.38 mm, apex angle from 15° to 90°, helix angle from 7° to 30°, number of fins from 21 to 82. Fig. 14 shows a comparison between calculated and experimental values of transfer coefficient for microfin tube. The model estimates the experimental values with a relative, absolute and standard deviation of −4.4%, 14.0% and 19.3%, respectively. Table 3 summarizes the deviations of each model. Fig. 13. Calculated and experimental heat transfer coefficients for microfin tube. Model by Kedzierski and Goncalves [23].
7. Conclusions 6.2. Models for microfin tube Kedzierski and Goncalves [23] proposed a model for the estimation of condensation heat transfer coefficients for microfin tubes. The model was developed from experimental data of local convective condensation measurements for four refrigerants (R134a, R410A, R125 and R32) inside a microfin tube with fin tip diameter of 8.51 mm, 60 fins, each 0.2 mm high. Mass velocities were in the range 57–552 kg m−2 s−1 . A comparison between calculated and experimental heat transfer coefficients for microfin tube is reported in Fig. 13. The model estimates the experimental values with a relative, absolute and standard deviation of 15.0%, 19.0% and 29.5%, respectively. Cavallini et al. [19] developed a model for the estimation of heat transfer coefficients during condensation in microfin tubes of halogenated and natural refrigerants, pure fluids and nearly azeotropic mixtures. The model considers the geometrical characteristics of the microfin tube, i.e. number of fins, fin height, apex and helix angles, as well as flow patterns and thermophysical properties of the refrigerant. It was validated with data considering R22, R134a, R123, R410A and CO2 . The working conditions of the data from which the model was validated were the following: saturation temperatures between −15 °C and 70 °C, re-
This paper proposed experimental heat transfer coefficients during R513A condensation inside a smooth tube with an inner diameter of 3.5 mm and inside a microfin tube with an inner diameter at the fin tip of 3.4 mm. A new test section for smooth tube was developed and verified during liquid forced convection tests. Condensation tests were carried out for mass velocities from 100 to 10 0 0 kg m−2 s−1 , at saturation temperatures at the inlet of the test section of 30 °C and 40 °C. Generally speaking, the higher the mass velocity and vapor quality, the higher the heat transfer coefficient. Most of the data fall in the T-independent flow regime, and the microfin tube was proved to make the transition from T-dependent to T-independent happen at lower mass velocities compared to the case of smooth tube. The microfin tube is effective in heat transfer augmentation especially at low mass velocities, with an average Enhancement Factor of 2.0. The experimental values of heat transfer coefficients were then compared against values predicted by empirical models. The model of Cavallini et al. [17] was able to estimate the condensation heat transfer coefficients for smooth tube with a relative, absolute and standard deviation of 13.5%, 13.5% and 4.5%, respectively. The model of Cavallini et al. [19] was able to estimate the condensation heat transfer coefficients for microfin tube with a relative, absolute and standard deviation of −4.4%, 14.0% and 19.3%, respectively.
A. Diani, P. Brunello and L. Rossetto / International Journal of Heat and Mass Transfer 152 (2020) 119472
Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. CRediT authorship contribution statement Andrea Diani: Conceptualization, Validation, Investigation, Writing - original draft, Visualization. Pierfrancesco Brunello: Writing - review & editing. Luisa Rossetto: Supervision, Funding acquisition. Acknowledgments The MS student Luca Vettorato is gratefully acknowledged. The supports of the MIUR through the PRIN Project 2017F7KZWS_005 and of the Università di Padova (Project CPDA107382 and Project BIRD172935/17) on this research are gratefully acknowledged. References [1] O. Hodnebrog, M. Etminiam, J.S. Fuglestvedt, G. Marston, G. Myhre, C.J. Nielsen, K.P. Shine, T.J. Wallington, Global warming potentials and radiative efficiencies of halocarbons and related compounds: a comprehensive review, Rev. Geophys. 51 (2013) 300–378. [2] M.A. Hossain, Y. Onaka, A. Miyara, Experimental study on condensation heat transfer and pressure drop in horizontal smooth tube for R1234ze(E), R32 and R410A, Int. J. Refrig. 35 (2012) 927–938. [3] C.Y. Yang, H. Nalbandian, Condensation heat transfer and pressure of refrigerants HFO-1234yf and HFC-134a in small circular tube, Int. J. Heat Mass Tran. 127 (2018) 218–227. [4] M. Azzolin, A. Berto, S. Bortolin, L. Moro, D. Del Col, Condensation of ternary low GWP zeotropic mixtures inside channels, Int. J. Refrig. 103 (2019) 77–90. [5] T.A. Jacob, E.P. Matty, B.M. Fonk, Experimental investigation of in-tube condensation of low GWP refrigerant R450A using a fiber optic distributed sensor, Int. J. Refrig. 103 (2019) 274–286. [6] Y.J. Kim, J.M. Cho, M.S. Kim, 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. 31 (2008) 771–779.
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[7] C. Dang, N. Haraguchi, E. Nihara, Flow boiling heat transfer of carbon dioxide inside a small-sized microfin tube, Int. J. Refrig. 33 (2010) 655–663. [8] G. Li, L. Huang, L. Tao, Experimental investigation or refrigerant condensation heat transfer characteristics in the horizontal microfin tubes, Appl. Therm. Eng. 123 (2017) 1484–1493. [9] M. Hirose, J. Ichinose, N. Inoue, Development of the general correlation for condensation heat transfer and pressure drop inside horizontal 4 mm small-diameter smooth and microfin tubes, Int. J. Refrig. 90 (2018) 238– 248. [10] A. Diani, M. Campanale, L. Rossetto, Experimental study on heat transfer condensation of R1234ze(E) and R134a inside a 4.0 mm OD microfin tube, Int. J. Heat Mass Tran. 126 (2018) 1316–1325. [11] A. Diani, M. Campanale, A. Cavallini, L Rossetto, Low GWP refrigerants condensation inside 2.4 mm ID microfin tube, Int. J. Refrig. (2018) 312–321. [12] A. Diani, A. Cavallini, L. Rossetto, R1234yf condensation inside a 3.4 mm ID horizontal microfin tube, Int. J. Refrig. 75 (2017) 178–189. [13] E.W. Lemmon, M.L. Huber, M.O. McLinden, NIST standard reference database 23, nist reference fluid thermodynamic and transport properties, REFPROP, version 10.0, 2018, Gaithersburg, MD: Standard Reference Data Program, National Institute of Standards and Technology. [14] S.J. Kline, F.A. McClintock, Describing the uncertainties in single-sample experiments, Mech. Eng. 75 (1953) 3–8. [15] B.S. Petukhov, V.N. Popov, Theoretical calculations of heat exchange and frictional resistance in turbulent flow in tubes of an incompressible fluid with variable physical properties, Trans. High Temp. 1 (1963) 69–83. [16] F.W. Dittus, L.M.K. Boelter, Heat transfer in automotive radiators of the annular type, Uni. Calif. Publ. Eng. 2 (1930) 443–461. [17] A. Cavallini, D. Del Col, L. Doretti, M. Matkovic, L. Rossetto, C. Zilio, G. Censi, Condensation in horizontal smooth tubes: a new heat transfer model for heat exchanger design, Heat Tran. Eng. 27 (2006) 31–38. [18] L. Doretti, F. Fantini, C. Zilio, Flow patterns during condensation of three refrigerants: microfin tube vs. smooth tube, in: Proc. of IIR International Conference on Thermophysical Properties and Transfer Processes of Refrigerants, 2, Vicenza, Italy, 2005 August 31-September. [19] A. Cavallini, D. Del Col, S. Mancin, L. Rossetto, Condensation of pure and nearazeotropic azeotropic refrigerants in microfin tubes: a new computational procedure, Int. J. Refrig. 29 (2009) 162–174. [20] S.Z. Rouhani, E. Axelsson, Calculation of void fraction in the subcooled and quality boiling regions, Int. J. Heat Mass Trans. 13 (1970) 383–393. [21] M.K. Dobson, J.C. Chato, Condensation inside smooth horizontal tubes, J. Heat Tran.—Trans. ASME 120 (1998) 193–213. [22] M.M. Shah, An improved and extended general correlation for heat transfer during condensation in plain tubes, HVAC&R Res. 15 (2009) 889–913. [23] M.A. Kedzierski, J.M. Goncalves, Horizontal convective condensation of alternative refrigerants within a micro-fin tube, J. Enhanc. Heat Tran. 6 (1998) 161–178.
Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/hmt
R513A condensation heat transfer inside tubes: Microfin tube vs. smooth tube Andrea Diani∗, Pierfrancesco Brunello, Luisa Rossetto Department of Industrial Engineering, University of Padova, Via Venezia 1, 35131, Padova, Italy
a r t i c l e
i n f o
Article history: Received 26 November 2019 Revised 21 January 2020 Accepted 2 February 2020
Keywords: Condensation R513A Smooth tube Microfin tube Heat transfer coefficient
a b s t r a c t The imminent phase-down of the common refrigerant R134a is calling for lower GWP alternatives. Real alternatives must have a lower global warming impact and they should be not flammable. In this context, R513A (azeotropic mixture made of R1234yf and R134a at 56% and 44% by mass) has been proposed as alternative to R134a due to its similar thermodynamic and transport properties and due to its lower GWP. This paper proposes a direct comparison between the thermal performances of a 3.5 mm ID smooth tube and those of a 3.4 mm ID microfin tube, during R513A condensation under the same working conditions of vapor quality (from 0.10 to 0.99), of mass velocity (from 100 to 10 0 0 kg m−2 s−1 ), and of saturation temperature (30 °C and 40 °C). The comparison permits to highlight in which working conditions the microfin tube leads to a real heat transfer augmentation which is higher than the mere increase of heat transfer area. In the end, the experimental heat transfer coefficients, both for the smooth tube and for the microfin tube, are compared against values calculated with empirical correlations from the open literature. © 2020 Elsevier Ltd. All rights reserved.
1. Introduction The new environmental laws are calling for the use of refrigerants with always lower values of Global Warming Potential (GWP). HydroFluoroCarbons (HFCs) have high values of GWP, and, even if they are still widely used in actual HVAC and refrigeration systems, they cannot be used for the new installations. Therefore, new pure refrigerants or refrigerants mixtures with lower values of GWP, are in great demand to accomplish the new stricter and stricter environmental regulations. These new solutions should have thermodynamic and transport properties similar to those of the actual implemented refrigerants, and they should be preferable not flammable, to make a direct refrigerant drop in possible. R134a is among the most common refrigerants implemented in a wide range of refrigeration and air conditioning systems, such as domestic and industrial air conditioning, automotive air conditioning, and centrifugal chillers. In these contexts, R513A could be a direct drop in substitute of R134a. R513A is an azeotropic mixture made of R1234yf and R134a (56% and 44% in mass, respectively). R1234yf has an extremly low GWP (lower than 1 [1]), but it is classified as mildly flammable (A2L flammability class), whereas R134a
∗
Corresponding author. E-mail address: [email protected] (A. Diani).
https://doi.org/10.1016/j.ijheatmasstransfer.2020.119472 0017-9310/© 2020 Elsevier Ltd. All rights reserved.
is not flammable (A1 flammability class). Therefore, the proposed mixture R513A benefits from the not flammability of R134a and from the low GWP of R1234yf: R513A is a refrigerant which belongs to the A1 flammability class and it has a value of GWP of about 630. In order to develop HVAC and refrigeration equipment which uses R513A as working fluid, its performance during flow boiling and condensation must be studied. Besides the use of new refrigerants with lower warming impact, the refrigerant charge reduction is another key issue that the scientific community has to cope to reduce the global warming. The refrigerant charge reduction can be achieved by implementing systems with small-sized tubes. Two-phase heat transfer of new lower GWP refrigerants in small-sized tubes is a topic which deserves more studies in the literature. In particular, no studies regarding R513A condensation can be found in the literature. Among the studies relative to refrigerants condensation inside small-sized tubes, Hossain et al. [2] proposed experimental measurements of heat transfer coefficients and pressure drops of R1234ze(E), R32 and R410A in a horizontal smooth copper tube with an inner diameter of 4.25 mm. Tests were carried out for mass velocities from 150 to 400 kg m−2 s−1 , at saturation temperatures between 35 °C and 45 °C. It was found that R1234ze(E) heat transfer coefficients were about 20–45% lower than those of R32 and 10–30% higher than those of R410A for a saturation temperature of 40 °C.
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Nomenclature D e EF g G h HTC
diameter [m] error [%] Enhancement Factor [-] gravity acceleration [m s−2 ] mass velocity [kg m−2 s−1 ] specific enthalpy [J kg−1 ] heat transfer coefficient [W m−2 K−1 ] dimensionless gas velocity
G·x D·ρV ·(ρL −ρV )·g
JG = √ m˙ Rx q t x
ρ
μ
0.9 · ( V )0.5 · ( L )0.1 Xtt = ( 1−x x ) ρ μ L
V
[-] mass flow rate [kg s−1 ] enhancement area [-] heat flow rate [W] temperature [°C] thermodynamic vapor quality [-] Martinelli parameter [-]
Greek symbols T temperature difference [K] ρ density [kg m−3 ] σN standard deviation [%] Subscripts abs absolute in inlet L saturated liquid out outlet rel relative ref refrigerant sat saturation TS test section V saturated vapor Yang and Nalbandian [3] analyzed flow condensation heat transfer and pressure drop of refrigerants R1234yf and R134a in a 4.0 mm ID small circular tube for mass velocities from 200 to 1200 kg m−2 s−1 . The flow condensation heat transfer performance was affected by flow patterns at different operating conditions: at the lowest mass velocity, gravity was the major force that controls the heat transfer mechanism and the flow pattern was plug. On the contrary, at the highest mass velocity, gravity effects were negligible even at low vapor qualities. The condensation heat transfer coefficients of R134a were higher than those of R1234yf at high vapor qualities but similar at low vapor qualities. Azzolin et al. [4] investigated the condensation heat transfer coefficient and two-phase frictional pressure drop of ternary mixtures R455A (R32, R1234yf and R744 at 21.5/75.5/3.0% by mass composition) and R452B (R32, R1234yf and R125 at 67.0/26.0/7.0% by mass composition) inside a 0.96 mm ID minichannel and inside a 8.0 mm ID channel. Condensation tests were run at 40 °C of saturation temperature, at mass velocities in the range 100– 600 kg m−2 s−1 for the 8.0 mm diameter test section, and in the range 20 0–80 0 kg m−2 s−1 for the 0.96 mm diameter test section. The experimental results showed that the mixture R452B presented higher heat transfer coefficients with respect to R455A for all the working conditions. The heat transfer coefficients inside the 0.96 mm diameter minichannel were higher than those measured inside the 8.0 mm diameter channel, and the effect of the diameter was more important at high mass flux. Condensation of the zeotropic mixture R450A and of R134a was studied by Jacob et al. [5]. R450A is composed of R134a and R1234ze(E) (42/58% by mass). Experiments were carried out in a
horizontal smooth tube with an inner diameter of 4.7 mm, for a range of mass velocities from 100 to 550 kg m−2 s−1 with saturation temperatures of 45 °C and 55 °C. Distributed fiber optic temperature sensors were installed to measure the coolant axial temperature profile during condensation. Collected data showed that the heat transfer coefficients increase with increasing vapor quality and mass flux, whereas they decrease with increasing saturation temperature. Compared to R134a, R450A showed 5% lower heat transfer coefficients at high vapor quality, whereas the pressure drop of R450A was, on average, 8% higher than that of R134a. A smart option to increase the heat transfer coefficient is represented by microfin tubes, which, besides having a larger heat transfer area, promote the turbulence of the liquid film, enhancing the heat transfer mechanism. Small sized microfin tubes were initially proposed for carbon dioxide applications [6,7], due to its high working pressure. Studies regarding two-phase flow of refrigerants inside small-sized microfin tubes have begun in the last years. At the moment, no work about R513A condensation inside neither microfin tubes nor smooth tubes can be found in the literature. Among the studies relative to refrigerants condensation inside small-sized microfin tubes, Li et al. [8] studied R22 and R410A during condensation inside smooth and microfin tubes. The tested tubes had outer diameters of 5.0 mm and 9.52 mm. A refrigerant superheating of 4–7 °C was set at the inlet of the test section. The mass velocity ranged between 100 and 400 kg m−2 s−1 for the 9.52 mm OD tubes and from 300 to 550 kg m−2 s−1 for the 5.0 mm OD tubes. The heat transfer coefficient of the microfin tube was around 1.65–2.55 times that of the smooth one. The microfin tubes yielded 30% higher pressure drops compared to those of the smooth ones. Hirose et al. [9] investigated the condensation heat transfer and pressure drop characteristics of smooth and microfin smalldiameter tubes, with outer diameter of 4.0 mm, with R32, R152a and R410A as working fluids. The condensation heat transfer coefficients and pressure drops were measured in the range of mass velocities from 100 to 400 kg m−2 s−1 , at a saturation temperature of 35 °C. The frictional pressure drop of the microfin tube was approximately 1.6 times greater than that of the smooth tube for each refrigerant and mass velocity. The heat transfer coefficient of the microfin tube was approximately 2–7 times greater than that of the smooth tube for R32 at 200 kg m−2 s−1 . Condensation heat transfer of R1234yf, R1234ze(E) and R134a inside microfin tubes with outer diameter of 3.0 mm and 4.0 mm was studied by Diani et al. [10,11]. Experiments were carried out for mass velocities from 100 to 10 0 0 kg m−2 s−1 in the 4.0 mm OD microfin tube, and from 300 to 10 0 0 kg m−2 s−1 in the 3.0 mm OD microfin tube, for saturation temperature at the inlet of the test section of 30 °C and 40 °C. Generally speaking, R1234ze(E) shows slightly higher heat transfer coefficients compared to R134a, but also higher pressure drops. This study investigates R513A condensation inside a smooth tube with an inner diameter of 3.5 mm and inside a microfin tube with an inner diameter at the fin tip of 3.4 mm. A new test section to study condensation inside a 3.5 mm ID smooth tube is proposed and verified during R513A liquid forced convection tests. Present working conditions for the condensation tests are: mass velocities between 100 and 10 0 0 kg m−2 s−1 , saturation temperatures of 30 °C and 40 °C, vapor qualities from 0.1 to 0.99. This wide range of operative conditions permits to understand how each working parameter affects the thermal behavior of R513A condensing inside the smooth and microfin tubes. The comparison between thermal performances of the microfin tube and those of the smooth tube are presented, in order to highlight in which conditions the microfin tube is more effective. Finally, the experimental collected heat transfer coefficients for both smooth and microfin tubes are
A. Diani, P. Brunello and L. Rossetto / International Journal of Heat and Mass Transfer 152 (2020) 119472
3
Fig. 1. Schematic of the experimental facility.
Table. 1 Values of instruments accuracy. T-type thermocouples T-type thermopiles Coriolis effect flowmeter Volumetric flowmeters at evaporator and pre-condenser Volumetric flowmeter at test section Absolute pressure transducers
±0.05 K ±0.03 K ±0.10% of reading ±0.25% of reading ±0.50% of reading ±1950 Pa
compared against values predicted by empirical correlations available in the open literature. 2. Experimental facility The experimental set up is located in the Heat Transfer in Micro-Geometries Laboratory of the Department of Industrial Engineering of the University of Padova. As shown in Fig. 1, the experimental facility consists of five loops: the refrigerant loop, the cooling water loop at the pre-condenser, at the test section and at the post-condenser, and the hot water loop at the evaporator. The refrigerant is pumped through the circuit by means of a magnetically coupled gear pump, it is vaporized and superheated in a brazed plate heat exchanger fed with the hot water. Superheated vapor then partially condenses in a pre-condenser fed with cold water, which is supplied by a water-cooled chiller, to achieve the set vapor quality at the inlet of the test section. The refrigerant enters into the test section at a known mass velocity and vapor quality where it condenses. The fluid leaves the test section and enters into a post-condenser, which is a brazed plate heat exchanger, where it is fully condensed and subcooled. The subcooled liquid passes through a drier filter and then is sent back to the boiler by the pump. A damper connected to the compressed air line operates as pressure regulator to control the saturation conditions in the refrigerant loop. Refrigerant temperatures and pressures are measured throughout the circuit to know its thermodynamic state. The accuracies of the instruments implemented in the facility are reported in Table 1. 3. Test sections A schematic of the test section for the 3.5 mm ID smooth tube is reported in Fig. 2. A 1.9 mm ID smooth tube, where cold water flows, is wrapped around the test smooth tube, as reported in the figure. T-type thermocouples, with an accuracy of ±0.05 K, are attached on the external wall of the inlet and outlet curves, in order
Fig. 2. Schematic of the test section for smooth tube.
the measure the inlet and outlet temperatures of the water flowing inside. The cold water is supplied by a thermostatic bath and its volumetric flow rate is measured by a magnetic flowmeter having an accuracy of ±0.50% of the reading. Six T-type thermocouples are glued in as many pits milled on the external wall of the test smooth tube, to monitor the wall temperature. The so created test section (test smooth tube with the wrapped smooth tube) is inserted in a U-shaped aluminum housing, which was filled with an alloy of tin/lead. The alloy, once solidified, guarantees a thermal contact between the smooth tube under investigation and the wrapped smooth tube. The whole test section is finally insulated with AF/Armaflex, to limit as much as possible the heat losses through the ambient. The length for the thermal measurements is 250 mm. Inlet and outlet pressure ports were soldered to the tube under investigation: the inlet pressure port is connected to an absolute pressure transducer having an accuracy of ±1950 Pa, whereas both the inlet and the outlet pressure ports are connected to a differential pressure transducer having an accuracy of ±25 Pa. The same concept was applied to the test section for the microfin tube. The test section is similar, but in this case a 3.4 mm ID microfin tube is placed instead of the 3.5 mm ID smooth tube.
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Details of the test section for the microfin tube can be found in Diani et al. [12]. The microfin tube under investigation has an inner diameter at the fin tip of 3.4 mm and the outer diameter is 4.0 mm. It has 40 internal fins, each fin is 0.12 mm high with an apex angle of 43°. The helix angle is 18°. From its geometrical characteristics, the ratio between its total heat transfer area and the area of an equivalent smooth tube is 1.69. Both the test sections are horizontally located in the experimental facility previously shown. 4. Data analysis The thermal performances of both the smooth tube and of the microfin tube are given in terms of heat transfer coefficient HTC, calculated as:
HTC =
qTS
AD · t sat − t wall
(1)
where AD is the inner surface area of the smooth tube, or the inner surface area of an equivalent smooth tube having an inner diameter equal to the diameter at the fin tip in the case of the microfin tube. t¯sat is the average saturation temperature, calculated as arithmetic average between inlet and outlet saturation temperatures. Inlet and outlet saturation temperatures are calculated from the values of inlet and outlet pressures with the software Refprop 10 [13], which implements new specific correlations for the azeotropic mixture R513A. t¯wall is the wall temperature. Results will also be referred to the mean vapor quality xmean , arithmetic average between inlet and outlet vapor quality. The inlet vapor quality xin is calculated as:
xin =
hT S,in − hL hV − hL
(2)
with hL and hV the specific enthalpies of the saturated liquid and vapor, respectively, calculated from the knowledge of the inlet pressure. hTS,in derives from an energy balance at the precondenser. The outlet vapor quality xout is similarly calculated, and in this case the specific enthalpy at the outlet of the test section derives from an energy balance at the test section itself. More details about the data analysis can be found in Diani et al. [12]. By following the procedure for the calculation of the uncertainty as suggested by Kline and McKlintock [14], the mean, maximum and minimum uncertainties on the heat transfer coefficient for the smooth tube are 6.2%, 7.6% and 5.0%, respectively, whereas for the microfin tube are 4.0%, 5.2% and 2.8%, respectively. The mean uncertainty on the mean vapor quality is ±0.03. 5. Experimental results This section presents the experimental results during R513A single phase flow inside the smooth tube and during R513A condensation inside both the smooth tube and the microfin tube. Single phase tests were necessary in order to verify the new test section with the 3.5 mm ID smooth tube, i.e. to verify the heat balance between the heat flow rate calculated on the refrigerant side and the one on the cooling water side. The test section with the 3.4 mm ID microfin tube was previously verified by Diani et al. [12]. 5.1. Single phase tests inside smooth tube Preliminary liquid forced convection tests were carried out inside the 3.5 mm ID smooth tube. These tests were run for mass velocities from 400 to 10 0 0 kg m−2 s−1 . The mass velocity is defined as:
G=
4 · m˙ re f
π · D2i
(3)
Fig. 3. Experimental and predicted single phase heat transfer coefficients versus mass velocity for smooth tube.
where Di is the inner diameter. During these tests, R513A pressure was at about 11.1 bar with a subcooling at the inlet of the test section between 16 K and 21 K, to ensure subcooled conditions. Reynolds number was between 7664 and 20,403. The difference between the measured heat flow rate on the water side and the one on the refrigerant side is within ±3 W. Fig. 3 shows the single phase heat transfer coefficients plotted against the mass velocity. Single phase heat transfer coefficients are calculated using Eq. (1) where t¯sat is replaced by t¯re f . t¯re f is the arithmetic averaged value between two refrigerant temperatures measured by two T-type thermocouples located on the external wall of the tested smooth tube, upstream and downstream of the aluminum housing. As it appears from the figure, the single phase heat transfer coefficient increases with increasing mass velocity. The experimental values are compared against the values predicted by two empirical correlations: the correlation of Petukhov and Popov [15] is able to estimate the experimental values of heat transfer coefficients with, on average, a relative deviation of 11.6%, whereas the correlation of Dittus and Boelter [16] is able to estimate the experimental values with, on average, a relative deviation of −0.3%. Therefore, since the heat flow rates calculated on refrigerant and water sides are within ±3 W, and since the empirical correlations well estimate the experimental single phase heat transfer coefficients, the test section for smooth tube is deemed verified. 5.2. Condensation tests inside smooth tube Condensation tests inside the smooth tube were carried out for mass velocities, calculated as reported in Eq. 3, from 100 to 10 0 0 kg m−2 s−1 for a saturation temperature at the inlet of the test section of 30 °C, and from 200 to 10 0 0 kg m−2 s−1 for a saturation temperature at the inlet of the test section of 40 °C. The heat transfer coefficients were calculated from the experimental tests as reported in Eq. (1). Every single experimental point is the average value of 100 data recorded in steady state conditions with a frequency of 1 Hz. Fig. 4 reports the condensation heat transfer coefficients for the smooth tube plotted against the mean vapor quality at a saturation temperature of 30 °C (top) and at 40 °C (bottom). Generally speaking, the heat transfer coefficient increases as both vapor quality and mass velocity increase. More in detail, the heat transfer coefficient at the mass velocity G = 100 kg m−2 s−1 is weakly affected by vapor quality, passing from a value of 1280 W m−2 K−1 at xmean
A. Diani, P. Brunello and L. Rossetto / International Journal of Heat and Mass Transfer 152 (2020) 119472
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Fig. 5. Data for smooth tube plotted in the flow regime map developed by Cavallini et al. [17]. G expressed in (kg m−2 s−1 ).
mass velocity, highlighting the higher and higher convective contribution as mass velocity increases. The comparison between the figure on the top and the one on the bottom of Fig. 4 reveals the effect of the saturation temperature on the heat transfer behavior. At constant mass velocity, a higher saturation temperature implies a higher vapor density, with a consequent lower vapor velocity. This is reflected on the experimental heat transfer coefficients: at fixed mass velocity and mean vapor quality, the higher the saturation temperature, the lower the heat transfer coefficient, due to a lower convective contribution. 5.3. Condensation tests inside microfin tube
Fig. 4. Experimental condensation heat transfer coefficient versus mean vapor quality for smooth tube. Data at tsat = 30 °C (top) and at tsat = 40 °C (bottom). G expressed in (kg m−2 s−1 ).
of 0.31 to a value of 2070 W m−2 K−1 at xmean of 0.79, whereas, for instance, the heat transfer coefficient at the mass velocity G = 400 kg m−2 s−1 in the same mean vapor quality region passes from about 2500 W m−2 K−1 to 50 0 0 W m−2 K−1 : this can be explained considering the different flow regimes that occur at the different mass velocities. At G = 100 kg m−2 s−1 , the flow seems to be stratified, with consequent low heat transfer coefficient. On the contrary, at G = 400 kg m−2 s−1 , the flow seems to be annular, due to the higher contribute of mass velocity and of vapor quality. This is reflected in Fig. 5, which reports the experimental data for the smooth tube in the flow regime map developed by Cavallini et al. [17]. The dashed line represents the transition from Tdependent flow regime (below the line) to T-independent flow regime (above the line). The T-dependent regime is associated to gravity controlled flow patterns, i.e. wavy and smooth stratified regimes, whereas the T-independent regime is associated to the shear stress forces controlled regimes, i.e. annular flow. The flow pattern map reveals that all the data at G = 100 kg m−2 s−1 are in stratified flow regime, whereas a transition from stratified to annular flow occurs at G = 200 kg m−2 s−1 , and, as a result, the heat transfer coefficients at G = 200 kg m−2 s−1 are higher than those at G = 100 kg m−2 s−1 especially at high mean vapor quality. Almost all the other experimental data fall above the transition line. The slope of the heat transfer coefficient curve increases with the
Condensation tests inside the microfin tube were carried for mass velocities, calculated as reported in Eq. (3) where Di represents, in this case, the inner diameter at the fin tip of the microfin tube, from 100 to 10 0 0 kg m−2 s−1 for a saturation temperature at the inlet of the test section of 30 °C, and from 200 to 10 0 0 kg m−2 s−1 for a saturation temperature at the inlet of the test section of 40 °C. The heat transfer coefficients were calculated from the experimental tests as reported in Eq. (1). Every single experimental point is the average value of 100 data recorded in steady state conditions with a frequency of 1 Hz. Fig. 6 reports the condensation heat transfer coefficients for the microfin tube plotted against the mean vapor quality at a saturation temperature of 30 °C (top) and at 40 °C (bottom). Generally speaking, the heat transfer coefficient increases as both vapor quality and mass velocity increase, but in this case the effect of vapor quality seems to be stronger compared to the case of the smooth tube. Considering the mass velocity G = 100 kg m−2 s−1 , there is a sudden change of the slope of the heat transfer coefficient curve at a mean vapor quality of about 0.6. This can be linked to a transition of the flow regime from stratified to annular flow. Fig. 7 reports the experimental data for the microfin tube plotted in the flow regime map developed by Doretti et al. [18] and suggested by Cavallini et al. [19]. The black continuous line represents the transition from T-dependent regime (below the line) to the T-independent regime (above the line) for microfin tube. As comparison, the transition for the case of smooth tube is also reported. The transition in case of microfin tube occurs at lower dimensionless gas velocity at constant Martinelli parameter compared to the case of the smooth tube. This is reflected on the heat transfer coefficients reported in Fig. 6. At G = 100 kg m−2 s−1 for mean vapor qualities higher than 0.6, the flow is annular, and thus
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A. Diani, P. Brunello and L. Rossetto / International Journal of Heat and Mass Transfer 152 (2020) 119472
Fig. 7. Data for microfin tube plotted in the flow regime map developed by Doretti et al. [18]. G expressed in (kg m−2 s−1 ).
Fig. 6. Experimental condensation heat transfer coefficient versus mean vapor quality for microfin tube. Data at tsat = 30 °C (top) and at tsat = 40 °C (bottom). G expressed in (kg m−2 s−1 ). Fig. 8. Experimental condensation frictional pressure drop versus mean vapor quality for microfin tube. Data at tsat = 30 °C. G expressed in (kg m−2 s−1 ).
the heat transfer mechanism is enhanced, whereas all the data at G = 100 kg m−2 s−1 are in stratified flow regime for smooth tube. In the case of the microfin tube, the flow can be considered annular for all the data from the mass velocity G = 200 kg m−2 s−1 . Conclusions similar to those for the case of the smooth tube can be drawn for the effect of the saturation temperature: the higher the saturation temperature, the lower the heat transfer coefficient, due to a lower convective effect. The comparison between Figs. 4 and 6 reveals that the low mass velocities are especially enhanced by the microfins, because the microfins tend to thin the liquid film thickness around them, with a consequent lower thermal resistance, and because the microfin tube makes the transition to annular flow happen at lower mass velocities. Therefore, the microfin tube is effective particularly at low mass velocities. Fig. 8 reports the frictional pressure drop for microfin tube plotted against the mean vapor quality at different mass velocities for the saturation temperature at the inlet of the test section of 30 °C. The model of Rouhani and Axelsson [20] was considered to estimate the void fraction, necessary for the calculation of momentum pressure drop. As it appears from the figure, at constant mean vapor quality, the higher the mass velocity, the higher the frictional pressure gradient. At constant mass velocity, the frictional pressure gradient increases as mean vapor quality increases up to a maximum value, after which it slightly decreases.
Taking for instance mass velocities of 400 and 800 kg m−2 s−1 , at a mean vapor quality of 0.8, the frictional pressure gradient at G = 800 kg m−2 s−1 is more than 3 times the values at G = 400 kg m−2 s−1 . Indeed, considering Fig. 6 (top), the heat transfer coefficient at xmean = 0.8 at G = 800 kg m−2 s−1 is only 1.3 times the value at G = 400 kg m−2 s−1 . Thus, in real applications, lower mass velocities could be preferable to higher mass velocities. 5.4. Comparison between microfin and smooth tube In this paragraph, the thermal performances of the microfin tube are systematically compared against those of the smooth tube. The comparison is based on the parameter called Enhancement Factor, defined as the ratio between the heat transfer coefficient of the microfin tube and that of the smooth tube under the same working conditions, i.e.:
EF =
HT Cmicro f in HT Csmooth
(4)
Fig. 9 shows the enhancement factor plotted against the mass velocity. The dotted horizontal line represents Rx, which is the ratio between the total heat transfer area of the microfin tube and that of an equivalent smooth tube with an inner diameter equal
A. Diani, P. Brunello and L. Rossetto / International Journal of Heat and Mass Transfer 152 (2020) 119472
Fig. 9. Enhancement factor versus mass velocity.
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Fig. 10. Calculated and experimental heat transfer coefficients for smooth tube. Model by Dobson and Chato. [21].
to the diameter at the fin tip of the microfin tube under investigation. When the Enhancement Factor is approximately equal to the value of Rx, it means that the heat transfer augmentation due to the microfins is mainly due to the increased heat transfer area. On the contrary, when the Enhancement Factor is higher than Rx, it means that the heat transfer augmentation is not only due to the increased heat transfer area, but also due to higher convective effects led to the microfins. As it appears from Fig. 9, EF tends to decrease as mass velocity increases, and thus the microfin tube is effective in the augmentation of the heat transfer mechanism especially at low mass velocities, due to the presence of the microfins which tend to thin the liquid film thickness. Considering all the experimental data, the mean Enhancement Factor is 2.0, and it reaches values higher than 4 at G = 200 kg m−2 s−1 . 6. Empirical models 6.1. Models for smooth tube In this section, the experimental results of the heat transfer coefficients collected during R513A condensation inside the 3.5 mm ID smooth copper tube are compared against values predicted by empirical models available in the open literature. Dobson and Chato [21] presented a model to estimate the heat transfer coefficient during condensation inside smooth tubes. The experimental data from which the model was developed comprises data during R12, R22, R134a and near-azeotropic blends of R32/R125 at 50/50% and 60/40% compositions, which condense inside smooth tubes with diameters ranging between 3.14 mm and 7.04 mm and mass velocities from 53 to 807 kg m−2 s−1 . Fig. 10 shows a comparison between calculated and experimental heat transfer coefficients for smooth tube. The model is able to estimate the experimental values with a relative, absolute and standard deviation of 36.6%, 36.3% and 10.0%, respectively. The model of Shah [22] was developed for predicting heat transfer coefficients during condensation inside plain tubes. The data used for the correlation’s validation include 22 fluids, among which water, halocarbon refrigerants, hydrocarbon refrigerants and organic fluids, condensing in horizontal, vertical and downwardinclined tubes. The range of parameters includes tube diameters from 2 to 49 mm, reduced pressures from 0.0 0 08 to 0.9, and mass velocities from 4 to 820 kg m−2 s−1 . Fig. 11 shows a comparison between calculated and experimental condensation heat transfer coefficients for smooth tubes. The model is able to estimate the
Fig. 11. Calculated and experimental heat transfer coefficients for smooth tube. Model by Shah [22]. Table. 2 Deviations between calculated and experimental heat transfer coefficients. Models for smooth tube.
Model of Dobson and Chato [21] Model of Shah [22] Model of Cavallini et al. [17]
erel
eabs
σ std
36.2% 22.3% 13.5%
36.3% 23.0% 13.5%
10.0% 10.2% 4.5%
experimental values with a relative, absolute and standard deviation of 22.3%, 23.0% and 10.2%, respectively. The model of Cavallini et al. [17] was developed to estimate the heat transfer coefficients during condensation of refrigerants flowing inside horizontal smooth tubes with inner diameters larger than 3 mm. For the validation, the model was compared against heat transfer coefficients collected during condensation of HCFCs, HFCs, HCs, carbon dioxide, ammonia, water, pure fluids or near azeotropic mixtures, for mass velocities from 24 to 2240 kg m−2 s−1 , and saturation temperatures from −15 °C to 302 °C. Fig. 12 shows a comparison between calculated and experimental heat transfer coefficients. The model well estimates the experimental values of heat transfer coefficients, with a relative, absolute and standard deviation of 13.5%, 13.5% and 4.5%, respectively. Table 2 summarizes the deviations of each model.
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A. Diani, P. Brunello and L. Rossetto / International Journal of Heat and Mass Transfer 152 (2020) 119472
Fig. 12. Calculated and experimental heat transfer coefficients for smooth tube. Model by Cavallini et al. [17].
Fig. 14. Calculated and experimental heat transfer coefficients for microfin tube. Model by Cavallini et al. [19]. Table. 3 Deviations between calculated and experimental heat transfer coefficients. Models for microfin tube.
Model of Kedzierski and Goncalves [23] Model of Cavallini et al. [19]
erel
eabs
σ std
15.0% −4.4%
19.0% 14.0%
29.5% 19.3%
duced pressure between 0.1 and 0.67, mass velocities between 80 and 890 kg m−2 s−1 . The microfin tubes had inside diameters between 5.9 mm and 14.2 mm, fin heights from 0.15 mm to 0.38 mm, apex angle from 15° to 90°, helix angle from 7° to 30°, number of fins from 21 to 82. Fig. 14 shows a comparison between calculated and experimental values of transfer coefficient for microfin tube. The model estimates the experimental values with a relative, absolute and standard deviation of −4.4%, 14.0% and 19.3%, respectively. Table 3 summarizes the deviations of each model. Fig. 13. Calculated and experimental heat transfer coefficients for microfin tube. Model by Kedzierski and Goncalves [23].
7. Conclusions 6.2. Models for microfin tube Kedzierski and Goncalves [23] proposed a model for the estimation of condensation heat transfer coefficients for microfin tubes. The model was developed from experimental data of local convective condensation measurements for four refrigerants (R134a, R410A, R125 and R32) inside a microfin tube with fin tip diameter of 8.51 mm, 60 fins, each 0.2 mm high. Mass velocities were in the range 57–552 kg m−2 s−1 . A comparison between calculated and experimental heat transfer coefficients for microfin tube is reported in Fig. 13. The model estimates the experimental values with a relative, absolute and standard deviation of 15.0%, 19.0% and 29.5%, respectively. Cavallini et al. [19] developed a model for the estimation of heat transfer coefficients during condensation in microfin tubes of halogenated and natural refrigerants, pure fluids and nearly azeotropic mixtures. The model considers the geometrical characteristics of the microfin tube, i.e. number of fins, fin height, apex and helix angles, as well as flow patterns and thermophysical properties of the refrigerant. It was validated with data considering R22, R134a, R123, R410A and CO2 . The working conditions of the data from which the model was validated were the following: saturation temperatures between −15 °C and 70 °C, re-
This paper proposed experimental heat transfer coefficients during R513A condensation inside a smooth tube with an inner diameter of 3.5 mm and inside a microfin tube with an inner diameter at the fin tip of 3.4 mm. A new test section for smooth tube was developed and verified during liquid forced convection tests. Condensation tests were carried out for mass velocities from 100 to 10 0 0 kg m−2 s−1 , at saturation temperatures at the inlet of the test section of 30 °C and 40 °C. Generally speaking, the higher the mass velocity and vapor quality, the higher the heat transfer coefficient. Most of the data fall in the T-independent flow regime, and the microfin tube was proved to make the transition from T-dependent to T-independent happen at lower mass velocities compared to the case of smooth tube. The microfin tube is effective in heat transfer augmentation especially at low mass velocities, with an average Enhancement Factor of 2.0. The experimental values of heat transfer coefficients were then compared against values predicted by empirical models. The model of Cavallini et al. [17] was able to estimate the condensation heat transfer coefficients for smooth tube with a relative, absolute and standard deviation of 13.5%, 13.5% and 4.5%, respectively. The model of Cavallini et al. [19] was able to estimate the condensation heat transfer coefficients for microfin tube with a relative, absolute and standard deviation of −4.4%, 14.0% and 19.3%, respectively.
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Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. CRediT authorship contribution statement Andrea Diani: Conceptualization, Validation, Investigation, Writing - original draft, Visualization. Pierfrancesco Brunello: Writing - review & editing. Luisa Rossetto: Supervision, Funding acquisition. Acknowledgments The MS student Luca Vettorato is gratefully acknowledged. The supports of the MIUR through the PRIN Project 2017F7KZWS_005 and of the Università di Padova (Project CPDA107382 and Project BIRD172935/17) on this research are gratefully acknowledged. References [1] O. Hodnebrog, M. Etminiam, J.S. Fuglestvedt, G. Marston, G. Myhre, C.J. Nielsen, K.P. Shine, T.J. Wallington, Global warming potentials and radiative efficiencies of halocarbons and related compounds: a comprehensive review, Rev. Geophys. 51 (2013) 300–378. [2] M.A. Hossain, Y. Onaka, A. Miyara, Experimental study on condensation heat transfer and pressure drop in horizontal smooth tube for R1234ze(E), R32 and R410A, Int. J. Refrig. 35 (2012) 927–938. [3] C.Y. Yang, H. Nalbandian, Condensation heat transfer and pressure of refrigerants HFO-1234yf and HFC-134a in small circular tube, Int. J. Heat Mass Tran. 127 (2018) 218–227. [4] M. Azzolin, A. Berto, S. Bortolin, L. Moro, D. Del Col, Condensation of ternary low GWP zeotropic mixtures inside channels, Int. J. Refrig. 103 (2019) 77–90. [5] T.A. Jacob, E.P. Matty, B.M. Fonk, Experimental investigation of in-tube condensation of low GWP refrigerant R450A using a fiber optic distributed sensor, Int. J. Refrig. 103 (2019) 274–286. [6] Y.J. Kim, J.M. Cho, M.S. Kim, 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. 31 (2008) 771–779.
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