An experimental investigation on flow boiling heat transfer of R-600a in a horizontal small tube

Accepted Manuscript Title: An experimental investigation on flow boiling heat transfer of R-600a in a horizontal small tube Author: Jeferson Diehl de Oliveira, Jacqueline Biancon Copetti, Júlio César Passos PII: DOI: Reference:

S0140-7007(16)30236-5 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.08.001 JIJR 3398

To appear in:

International Journal of Refrigeration

Received date: Revised date: Accepted date:

16-5-2016 18-7-2016 5-8-2016

Please cite this article as: Jeferson Diehl de Oliveira, Jacqueline Biancon Copetti, Júlio César Passos, An experimental investigation on flow boiling heat transfer of R-600a in a horizontal small tube, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.08.001. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

An experimental investigation on flow boiling heat transfer of R-600a in a horizontal small tube Jeferson Diehl de Oliveiraa*, Jacqueline Biancon Copettib,*, Júlio César Passosa a

Department of Mechanical Engineering, LEPTEN, Laboratory of Process of Engineering and Energy Technology, Federal University of Santa Catarina, 88010-900, Florianópolis, SC, Brazil b

Mechanical Engineering Graduate Program, LETEF, Laboratory of Thermal and Fluid Dynamic Studies, University of Vale do Rio dos Sinos, 93022 – 750, São Leopoldo, RS, Brazil *Corresponding author. Tel.:+55 51 3591 1100; fax: +55 51 3590 8172. E-mail address: [email protected] (J.D. Oliveira)

Highlights  The heat transfer coefficient is measured using R-600a.  The effect of mass velocity, heat flux and quality conditions is investigated.  The influence of flow patterns on the heat transfer coefficient is observed.  Heat transfer coefficient results are compared to several predicted correlations.

Abstract The local heat transfer coefficient and flow pattern were studied experimentally during flow boiling of R-600a inside a horizontal small tube (inner diameter ID = 1.0 mm). The effects of heat flux, mass velocity and vapor quality have been investigated. The tests were conducted at heat fluxes from 5 to 60 kW m-2, mass velocity from 240 to 480 kg m-2s-1 at 25°C saturation temperature. Results show that flow patterns are related to the heat transfer coefficient. Convective boiling was found to be predominant during the tests. Nucleate boiling was predominant only in low vapor quality region. The resultant database was compared against predicting correlation available in literature for heat transfer coefficient. It was found that mass velocity have significant influence of heat transfer coefficients.

_________________________________________________________________________________ Keywords: Heat transfer, Flow boiling, Isobutane, Minichannel, Horizontal tube, Flow patterns Nomenclature a AH b c d D g G h(z) have

angular coefficient in Eq. (A3) [°C kPa-1] heated area [m2] linear coefficient in Eq. (A3) [°C] angular coefficient in Eq. (A7) [kJ kg-1 kPa-1] angular coefficient in Eq. (A8) [kJ kg-1 kPa-1] tube diameter [m] gravitational acceleration [ms-2] mass velocity [kg m–2s–1] local heat transfer coefficient [kW m–2K–1] average heat transfer coefficient [kW m–2K–1]

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experimental heat transfer coefficient [kW m–2 K–1] predicted heat transfer coefficient [kW m–2 K–1] local enthalpy[kJ kg–1] direct current [A] liquid enthalpy [kJ kg–1] vapor enthalpy [kJ kg–1] latent heat of vaporization [kJ kg–1] pre-heater inlet enthalpy [kJ kg–1] test section inlet enthalpy [kJ kg–1] superficial liquid velocity in test section outlet [ms-1] superficial vapor velocity in test section outlet [ms-1] thermal conductivity of stainless steel [W m–1K–1] mass flow rate [kg s–1]

hexp hpred i(z) I il iv ilv ii-PH ii-TS jl(o-TS) jv (o-TS) kSS m MBE 

 h n



n

1

pred

i 1

i

 h exp i / h exp i  100

p qPH qTS q"TS

pressure [kPa] heat rate at pre-heater [W] heat rate at test section [W] heat flux at test section [W m–2] volumetric heat generation at test section [W m–3] absolute internal roughness [μm]

q TS

Ra Re

l



Re v =

G 1  x  D

Reynolds number of liquid phase [-]

l

GxD

Reynolds number of vapor phase [-]

μv

re ri RMSE 

mean bias error [%]

external radius [m] internal radius [m] 1 n

 h n

i 1

Tew Tiw Tsat Tw,bottom Tw,top U X x(i-TS) x(o-TS) zn Greek symbols α ρl ρv σ

pred

i



 h exp i / h exp i

2



 100

root mean square error [%]

external wall temperature [°C] internal wall temperature [°C] saturation temperature [°C] external wall bottom temperature [°C] external wall top temperature [°C] direct voltage [V] vapor quality [-] vapor quality in the test section inlet [%] vapor quality in the test section outlet [%] location n in test section [mm] void fraction [-] liquid density [kg m–3] vapor density [kg m–3] surface tension [N m-1]

1 Introduction Due to concerns related to energy crisis, environmental pollution, destruction of the ozone layer and greenhouse effect, intense investigations have been conducted to find replacements for fluid refrigerants such as CFC’s and other halogenated fluids. Hydrocarbons have been presented as a good choice of replacements because of their thermodynamic and transport properties, which have reduced impact on nature. Nowadays, among several applications of hydrocarbons, we can highlight the use in compact heat exchangers as refrigerators, air conditioners and electronics cooling. Shin et al. (1997) performed a study on heat transfer during flow boiling using pure refrigerants R-22, R- 32, R-134a, R-290 (propane) and R-600a (isobutane) and refrigerant mixtures of R-32/R-134a, R-290/R-600a and R-32/R-125 in a horizontal circular tube with inner diameter (ID) of 7.7 mm and saturation temperature of 12 °C. For the cases of R-290, R600a and R290/R-600a, they performed the investigation with mass velocity ranging between 265 and 583 kg m-2s-1, heat fluxes varied from 10 to 30 kW m-2. As a result, it was observed

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that the heat transfer coefficient presented a strong dependence on heat flux in the low vapor quality region and became practically independent as vapor quality increased. They also concluded that the highest heat transfer coefficients were observed in the cases of R-600a and the mixture R-290/R-600 (50/50 by mass %). Jung et al. (2000) investigated the performance of the R-290/R-600a mixture in domestic refrigerators as a substitute for CFC R-12. Their experimental results showed that the mixture 60/40 (by mass %) presented an increase in energy efficiency up to 4 % as compared with tests performed with R-12. Wen et al. (2007) investigated the flow boiling heat transfer characteristics for the R-290/R600a mixture with lubricating oil in serpentine U-tubes with 2.46 mm ID. The tests were performed for mass velocities ranging from 100 to 320 kg m-2s-1, heat flux from 15 to 25 kW m-2. As a result, it was found that the heat transfer coefficients of pure R-600a and R-290 cases increased with an increase in vapor quality for x > 0.4. In high vapor quality region (x > 0.4), the heat transfer coefficient of R-290 and R-600a with the oil decreased considerably. Following the trend in analyzing the use of hydrocarbon refrigerants in domestic refrigerators, Jwo et al. (2009) studied the efficiency of the R-290/R-600a mixture (50/50 by mass %) as a substitute for R-12 and HFC R-134a. They concluded that the mixture reduced the working time of compressor and saved 4.4 % of energy as compared with R-12 and R- 134a. Choi et al. (2009) investigated the flow boiling heat transfer characteristics for R-290 in small channels of 1.5 mm and 3.0 mm ID. The tests were performed for mass velocities ranging from 50 to 400 kg m-2s-1 and heat fluxes ranging from 5 to 20 kWm-2 at saturation temperatures of 0, 5 and 10 °C. As a result, it was found that the diameter significantly affects the heat transfer, the heat transfer coefficient increases with a decrease of inner diameter and with an increase of saturation temperature. Aprin et al. (2011) investigated flow boiling in a staggered tube buddle using R-290, R-600a and npentane (R-601) as refrigerants in different operating conditions. In their tests, the heat transfer regimes were investigated to establish important parameters related to boiling process. Copetti et al. (2013) studied the flow boiling heat transfer characteristics of R-600a in a horizontal microchannel with inner diameter of 2.6 mm with saturation temperature of 22°C, mass velocity ranging from 240 and 440 kg m-2 s-1 and heat flux between 44 and 95 kW m-2. Test results showed a significant influence of heat flux on heat transfer coefficient in low vapor quality conditions. Such an influence vanished in cases of high vapor quality conditions and high mass velocities. An investigation on influence of different saturation temperatures on heat transfer and pressure drop in a circular vertical mini-tube (1.7 mm ID) was performed by Maqbool et al. (2013) using R-290. In such a study, they investigated the vaporization with saturation temperatures of 23, 33 and 43°C, for heat fluxes ranging from 5 to 280 kW m-2 and mass velocities between 100 and 500 kg m-2 s-1. They concluded that heat transfer coefficient increased with the increases of saturation temperature and heat flux while the influence of vapor quality and mass velocity was found to be insignificant. Wang et al. (2014) studied experimentally heat transfer and pressure drop of R-290 in flow boiling in a horizontal channel (6.0 mm ID) for saturation temperatures ranging from – 35 to – 1.9°C with mass velocities between 62 and 104 kg m-2 s-1 and heat fluxes from 11.7 to 87.1 kW m-2. The local heat transfer coefficients were found to increase with heat flux and mass velocity. Del Col et al. (2014) studied the heat transfer of R-290 in a circular minichannel (0.96 mm ID). The experiments were performed at saturation temperature of 31°C. It was concluded that the heat transfer coefficient has a strong dependence of heat flux and vapor quality. They also investigated two-phase pressure drop during adiabatic flow at saturation temperature between 39.5°C and 42°C. The pressure drop increases with the increase of mass flux and with the increase of vapor quality. The study of hydrocarbons in flow boiling represents just a small fraction of the available experimental investigations in literature. Some experimental studies related to flow boiling of thirteen different refrigerants during the last years are presented in Table 1. Ong and Thome (2009), Saisorn (2010), Tibiriça and Ribatski (2010) and Keepaiboon and Wongwises (2015) also investigated the effect of flow patterns during the experiments. This present study was developed to investigate experimentally flow boiling of isobutane (R-600a) in a 1.0 mm ID tube, and analyze the local heat transfer coefficient and flow patterns dependence on heat flux and mass velocity for a specific saturation temperature. The experimental results were compared with some heat transfer coefficient correlations developed for small channel cases. 2 Experimental Setup and Methodology 2.1 Test facility and instrumentation An experimental facility was developed to investigate the flow boiling and pressure drop in a horizontal small channel. The details of such set up are shown schematically in Fig. 1. The experimental system consists of a loop that provides controlled mass velocity, and it was designed to test different fluids under a wide range of flow conditions. The main part of the loop has a Coriolis mass flow meter and pre-heater, a test section and a

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visualization section. The secondary part consists of a condenser, a refrigerant reservoir, a liquid refrigerant vessel, a dryer filter, a magnetic gear pump and a subcooler. The condenser and the subcooler have independent circuits and each uses an ethylene-glycol/water solution as the secondary refrigerant, the temperature of which is controlled by a thermostatic bath. This set up controls the refrigerant saturation temperature. The refrigerant reservoir connected to the main circuit of the bench has small volume and is partially filled with refrigerant to assist in control of pressure fluctuations that occur when there is variation in heat flux conditions, operating as a “lung”. A liquid refrigerant vessel maintains a constant static pressure at the pump suction. This procedure assures that the pump works uniformly and under immersion, avoiding cavitation. The subcooler is used to compensate the temperature rise, which usually occurs as the refrigerant passes the gear pump, and also to assure that only subcooled liquid enters the flow meter. The pre-heater establishes the experimental conditions entering the test section immediately downstream. It consists of a horizontal stainless steel tube with 1.0 mm internal diameter with a total length of 515 mm and heated length of 440 mm. This tube is uniformly heated by direct application of an electrical current in the wall (Joule effect), the intensity of which is controlled by the power supply SORENSEN model DCS 8-125E which provides both direct voltage and direct current measurements. The test section consists of the same tube and diameter with total length of 366 mm, heated length of 265 mm. As well as the pre-heater, the test section is heated by Joule effect, the intensity of which is controlled by the power supply. The absolute internal roughness (Ra) of the tube was measured with a STARRETT TM roughness tester, model SR 200, and was found to be 1.48 µm. There is a visualization section downstream of the test section with a 136 mm length glass tube with the same test section internal diameter. Both pre-heater and test section are thermally insulated. The refrigerant enters the pre-heater as a subcooled liquid and achieves the saturation condition upstream of its outlet. This condition defines the vapor quality in the test section inlet, and varies according to the heat flux imposed in the pre-heater. Two absolute pressure transducers and two 0.076 mm type-E thermocouples, in direct contact with the refrigerant, carry out the pressure and temperature measurements, at the inlet and outlet of the pre-heater. In the test section, refrigerant temperatures are measured at the inlet and outlet of the tube. The tube wall temperatures are measured by thermocouples directly fixed by a thermally conductive paste at five axial positions along the tube, with two thermocouples at each position, each one separated by 180° from the other, as shown in Fig. 1. The differential pressure transducer allows the determination of the outlet pressure. A frequency inverter is responsible for the fine adjustment during the control of the pump flow rate and a by-pass line downstream of the pump, controlled by a needle valve, is used for coarse adjustment of flow rates. Two other needle valves (downstream the preheater and upstream of the visualization section) are used to reduce instabilities pressure during testing in different flow boiling conditions. The pressure transducers, thermocouples, mass flow and power meter were connected to an acquisition data system composed of a multimeter (Agilent, model 34970A), controlled by a computer via an RS232 interface. The flow patterns observed in the visualization section during tests were recorded with a high-speed camera MotionPro model Y4-S1, using 3000 frames per second. Preliminary tests in single-phase with heating were realized to estimate the heat loss in both preheater and test section. In both sections, the heat loss was negligible (less than 0.1 %). 2.2 Data reduction The heat flux applied to the test section was given by, q " TS 

U I

D i L h

(1)

where Di is the internal diameter, Lh is the heated length and U and I represented the voltage and current applied by power supply in test section, respectively. The external wall temperature was assumed to be the average of two local temperatures measured for each axial position z along the test section, as Eq. (2): T ew ( z ) 

T w , top ( z )  T w , bottom ( z )

(2)

2

The local average temperature of internal wall, T iw ( z ) , is calculated assuming radial conduction through the wall of the tube, subject to internal heat generation, according to T iw ( z )  T ew ( z ) 

q TS 4 k ss

r

e

2

 ri

2



q TS 2 k ss

r 2 re ln  e  ri

   

(3)

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where T ew ( z ) corresponds to the local average temperature measured by thermocouples on external wall in each position z, q TS represents the volumetric heat generated along the tube in test section, kss is the thermal conductivity of stainless steel, and re e ri are the external and internal radius of test section, respectively. The local heat transfer coefficient at any location was calculated by Eq. (4). h( z) 

q " TS

(4)

T iw ( z )  T sat ( z )

The local saturation temperature, T sat ( z ) , was calculated as a function of pressure inlet and pressure drop in test section, considering the pressure drop as a linear variation along the test section. The vapor quality in the test section inlet was calculated from energy balance in pre-heater, according to Eq. (5).

x ( i  TS ) 

 q PH   i i  PH   i l  p sat   m 



(5)

i lv ( p sat ) where il and ilv represent the liquid enthalpy and latent heat of vaporization as function of saturation pressure. Since the local vapor quality along the test section was calculated as a function of local enthalpy and the enthalpy of test section inlet, as Eq. (6). x( z) 

i ( z )  i l ( p sat ) i lv  p sat

(6)



In such case, local enthalpy, i(z), is  q TS ( z )  i( z )    i i TS   m 

(7)

All thermodynamics and transport properties used in data reduction and heat transfer correlations were obtained from REFPROP v. 9.1 (Lemmon et al., 2013). The experimental conditions are summarized in Table 2. 3 Experimental results 3.1 Flow pattern The flow patterns were identified by images collected with high-speed camera. For each test, six hundred images were obtained to ensure a consistent pattern analysis, thus obtaining more than 234,000 images as result of all tests. The predominant patterns observed were churn, wavy-annular and smooth-annular flows Figure 2 presents some patterns observed in specific conditions. In all images, parameters (related to test section outlet o-TS) as vapor quality x(o-TS), Reynolds number of liquid phase Rel(o-TS), Reynolds number of vapor phase Rev(o-TS), superficial velocity of liquid phase jl(o-TS), superficial velocity of vapor phase jv(o-TS) and void fraction α(o-TS) are also presented. The void fraction was evaluated with the relation developed by Rouhani and Axelsson (1970) and by homogeneous model (Carey, 1992) given by Eq. (8) and (9), respectively. The difference between both void fractions models was smaller than 5 %.  x x  1 x      1  0 . 12 1  x   v  l  v

 1 x   1    x 

 1 . 18 1  x  g    l   v   0 .5  G l 

  v         l  

0 . 25

    

1

(8)

1

(9)

where ρl , ρv , g and σ represent liquid density, vapor density, gravitational acceleration and surface tension, respectively. Patterns as bubbly/plug, plug and slug were observed only in particular cases involving both low heat flux and low vapor quality. In fact, bubbles were observed just in G = 480 kg m-2 s-1 and for minimum heat condition in both pre-heater and test section. Tests showed that it was not possible to observe bubbly flow for x(oTS) > 0.03.

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Slug flow is characterized as “oblong nose” that protruded into the core. That was the second pattern observed for low heat conditions but for all mass velocities. For the tests involving slug flow it was found that α(o- TS) < 0.8 and jv(o-TS)/jl(o-TS) ≤ 2. Churn flow can be considered the transition between slug and annular flows and it was observed in all mass velocities, but for a limited heat condition. For G = 240 kg m-2 s-1, it was observed during tests with q”TS = 10 kW m-2 and α(o-TS) = 0.83 (Fig. 2) and 20 kW m-2, α(o-TS) = 0.88. For G = 320 and 400 kg m-2 s-1, it was observed in q”ST = 5, 10 and 20 kW m-2. But during the tests with G = 480 kg m-2 s-1, this pattern was verified only for q”ST = 10 kW m-2 (Fig. 2). Pettersen (2004) found similar patterns in an experiment involving flow boiling of CO2 in microchannel tubes with 0.8mm ID. The most predominant patterns observed during this investigation were wavy-annular and smoothannular flows. The wavy-annular flow pattern is characterized by the presence of liquid front wave sliding on the liquid-vapor interface in annular flow. Tibiriçá and Ribatski (2010) observed such pattern during flow boiling with R134a and R245fa. Increasing the heat flux or mass velocity, such fronts disappear and smooth-annular flow is observed. Both annular flow patterns were observed during tests with x(o-TS) ≥ 0.2 and jv(o-TS)/jl(o-TS) ≥ 2.66 (Fig. 2). The ratio of Reynolds numbers of phases (Rev(o-TS)/ Rel(o-TS)) increases considerably with the increase of mass velocity. The increase of heat flux for each G, influences the increase of the vapor quality. But the ratio of Reynolds numbers was less affected by the heat flux. The maximum Reynolds number for vapor phase was 40,051 for G = 480 kg m-2 s-1 and q”ST = 60 kW m-1. Situations where the ratio Rev(o-TS )/ Rel(o-TS) < 1 were observed for plug flow pattern for all mass velocities where x(o-TS) < 0.05.

3.2 Heat transfer The influence of heat flux, mass velocity and vapor quality on heat transfer coefficient has been investigated. Flow patterns are also considered in order to analyze the mechanisms of heat transfer during flow boiling. 3.2.1 Effect of heat flux Figure 3 shows local heat transfer coefficient as a function of vapor quality for mass velocities considered in this study. In general, the higher the vapor quality, larger is the heat transfer coefficient. For all mass velocities, the heat transfer coefficient presents strong dependence on heat flux at low quality region. For the case of G = 240 kg m-2 s-1 with q”TS = 5 kW m-2 and 10 kW m-2 (Fig. 3a), the dominance of nucleate boling can be related with slug flow (0.02 ≤ x(o-TS) ≤ 0.08 and Fig. 2) and churn flow (0.08 < x(o-TS) ≤ 0.23 and Fig. 2). For higher vapor qualities, only annular flow was observed, independently of the heat flux. For G =400 kg m-2s-1 and G = 480 kg m-2 s-1 (Fig. 3c and Fig. 3d) the dependence on heat flux at low vapor quality is stronger. Such dependence is also due the dominance of nucleate boiling heat transfer characterized by plug, slug and churn flows. As the vapor quality increases the dependence of heat transfer coefficient on heat flux vanishes and the convective boiling heat transfer becomes the dominant process. Sempértegui-Tapia and Ribatski (2015) reported similar phenomenon. It was also observed that the dominance of nucleate boiling decreases with increasing the mass velocity. In low vapor quality region, the effect of heat flux on heat transfer coefficient decreases with increasing mass velocity and with increasing vapor quality. That happens because with increasing mass velocity, the surface velocities of both phases increase and consequently it occurs the hastiness of the flow patterns related to convective boiling. On the other hand, the effect of the heat flux on heat transfer coefficient stands out with increasing mass velocity in low quality region. In this region the difference of the heat transfer coefficients as functions of the different heat fluxes increases with increasing mass velocity. This phenomenon is related to the increase of the ratio of the contribution between heat flow and local latent heat as a function of saturation pressure. 3.2.2 Effect of mass velocity Mass velocity is one of the parameters that the most affects the process of heat trasfer, mainly in convective boiling. The effect of mass velocity on local heat transfer coefficient is presented in Figs. 4a-b. The Fig. 4a shows the behavior of the local heat transfer coefficient for the lowest and highest heat fluxes (10 kW m-2 and 60 kWm- 2) and mass velocities (240 kg m-2s-1 and 480 kg m-2 s-1) considered in this study. The experimental data show that local heat transfer coefficient increases with increasing the mass velocity. And as the vapor quality increases, the effect of mass velocity becomes more evident. In Fig. 4b it is presented the local heat transfer coefficient for mass velocities equal to 240 kg m-2s-1 and 320 kg m-2s-1 for heat fluxes of 20 and 60 kW

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m- 2. In this case, the nucleate boiling contribution is observed for x < 0.15. But in both graphics it is evident the predominance of convective boiling on heat transfer coefficient due the absence of influence of heat flux. The average heat transfer coefficient versus mass velocity, for different heat fluxes, is presented in Fig. 5. In this case, the average heat transfer coefficient is considered to be the average of all local heat transfer coefficients for each heat flux condition. Results show that as mass velocity increases the coefficient becomes higher but it is nearly not affected by heat fluxes. The Reynolds number and superficial velocity in each phase also increase, which leads to the appearance of annular flow. A stronger contribution of nucleate boiling can be seen only for G = 240 kg m-2s-1 and q”ST = 5 kW m-2. In this case, according to the images analyzed, slug and churn flows were predominant (Fig. 2). Another way to analyze the influence of mass velocities on heat transfer coefficient is by the boiling curve for different mass velocities. The heat flux as function of excess temperature, Tiw - Tsat, and the outlet vapor quality, x(o-TS), of 0.15, is shown in Fig. 6. The effect caused by mass velocities can be seen in higher excess temperatures for G = 320, 400 and 480 kg m-2s-1. For G = 240 kg m-2s-1 such effect can be seen even for lower excess temperatures. As the mass velocity increases the slope of each curve (qTS”/(Tiw – Tsat)) increases, thus indicating the dependence of heat transfer coefficient on the mass velocity. It is also evident that the influence of mass velocity increases as the excess temperature becomes higher. 3.3 Comparison of experimental results with correlation for heat transfer Experimental results have been compared against six models for the flow boiling heat transfer coefficient developed for small channels: Kew and Cornwell (1997), Kandlikar and Balasubramanian (2004), Bertsch et al. (2009), Sun and Mishima (2009), Li and Wu (2010), Kim and Mudawar (2013). The performance of the prediction models was evaluated using both mean bias error (MBE), which indicates the variation of the calculated values with respect to measured ones, and root mean square error (RMSE), which indicates the dispersion of the regression. Comparisons of experimental heat transfer coefficient with predictions are presented in Fig. 7. Kew and Cornwell (1997) studied flow boiling heat transfer of R-141b in tubes with inner diameter between 1.39 mm and 3.69 mm in saturation temperature of 32°C. In fact, Kew and Cornwell (1997) proposed a modification in the correlation of Lazarek and Black (1982) due the increase in heat transfer coefficient with vapor quality during their tests. On their correlation, Fig. 7a, it was found that the fraction of data within ± 35% error band is 56 %. MBE and RMSE are –25 % and 46.9 %, respectively. Kandlikar and Balasubramanian (2004) modified a correlation proposed by Kandlikar (1990). The new correlation was based on flow boiling of R-113, R-134a, R-123, R-141b and water taking into account effects related to inner diameter ranging from 0.19 mm to 2.59 mm. For each fluid studied in their investigation, the authors suggest a specific value for the fluid-surface parameter. For fluids not included in the correlation, they suggest the use of the fluid-surface parameter equal to one. So, the comparison, shown in Fig. 7b, resulted in MBE equal to -35%, RMSE equal to 44.2 % and only 40.2 % of data agree with measurements to within ± 35%. However, for x ≤ 15 % (nucleate boiling domain in most cases), nearly 81 % of data points are found within ± 35 % error band. Sun and Mishima (2009) proposed a correlation based on correlation of Lazarek and Black. It is based on a database with 2505 data points with eleven different working fluids for mass velocities ranging from 44 to 1500 kg m-2 s-1 and hydraulic diameters between 0.21 and 6.5 mm. Taking into account the trends in the experimental data, it was observed that the heat transfer coefficient was strongly dependent on Weber number, which is the ratio of fluid’s inertia to the surface tension. Sun and Mishima correlation captured 35.6 % of the data within ± 35% error band, as can be seen in Fig.7c, with a MBE of -33.8 % and RMSE equal to 38.9 %. Bertsch et al. (2009) proposed a correlation based on Chen’s model (1966) for flow boiling in mini and microchannels. Their correlation was developed based on a database about experiments with hydraulic diameters ranging from 0.16 mm to 2.92 mm, heat flux between 4 and 1150 kW m -2, mass velocity ranging from 20 to 3000 kg m-2 s-1. Twelve different refrigerants were used, with saturation temperature ranging from – 194°C to 97°C. One of the main characteristics of such correlation is that the enhancement factor F is strongly influenced by the confinement of bubbles in small channels. According to the comparison, Fig. 7d, the correlation shown a MBE of -24% and RMSE equal to 41.2 %. The fraction of data within ± 35% error band is 55 %. The correlation proposed by Li and Wu (2010) covered experimental studies related to saturation flow boiling heat transfer in micro and mini-channels using thirteen different refrigerant fluids including propane for mass velocities between 20.3 and 3570 kg m-2 s-1 and hydraulic diameters ranging from 0.16 to 3.0 mm. To consider the effects of small hydraulic diameter, the authors introduced both the Bond number and Reynolds number of liquid phase. According to them, the relationship of two nondimensional parameters presents the influence of body force, surface tension, viscous force and inertia force in saturation flow boiling in reduce channels. As result, the correlation presents a MBE of – 36.1 % and RMSE of 38 % with 33.1 % of data within ± 35% error band (Fig. 7e). Kim and Mudawar (2013) developed a correlation for pre-dryout saturated flow boiling in mini and micro-channels in both single- and multi-channel configurations. Their correlation is based on a database of

Page 7 of 17

10,805 experimental data points with eighteen different working fluids and hydraulic diameter ranging from 0.19 to 6.5 mm and mass velocities between 19 and 1608 kg m-2 s-1. In this case it was found (Fig. 7f) the lowest variation between experimental and predicted values, and lowest dispersion of the data, MBE = 4.4 % and RMSE = 33.6 %, respectively, with 68 % datapoints within ± 35 %. Due to the high values of experimental heat transfer coefficient, under predictions were observed in all heat transfer correlations. It is important to emphasize that none of the above correlations were developed based on flow boiling experimental database with isobutane. Some singular thermodynamic properties of isobutane can also be taken into account. Higher thermal conductivity and lower viscosity of R-600a contribute considerably to the heat transfer coefficient being higher than those of predicted by the correlations (Shin et al., 1997 and Copetti et al., 2013). 4 Conclusion The experimental investigation on the flow pattern and boiling heat transfer of R-600a was conducted with a horizontal circular tube of 1.0 mm in diameter for various mass velocities and heat fluxes and constant saturation temperature of 25°C. The main findings are summarized as follows:       

The two-phase flow pattern observations have shown a dominance of intermittent churn, wavy-annular and smooth-annular flow regimes. Plug and slug flow were observed only during lower vapor quality conditions, indicating that such patterns are related to nucleate boiling dominance. The heat transfer coefficient is highly dependent on the mass velocity while heat flux is found to have no significant effect. Only in low vapor quality region, the influence of heat flux on the heat transfer coefficient was observed for all mass velocities. The heat transfer coefficient increases with increasing vapor quality. It was not observed the occurrence of dryout, even for high vapor quality. Among the correlations used to predict the heat transfer coefficient in this investigation, Kim and Mudawar (2013) correlation satisfactorily presented the best fit when compared with experimental data. Finally, the present investigation provides a basis of knowledge about heat transfer and flow patterns during flow boiling of R-600a in a minichannel, which is an important foundation for the development of new technologies related to compact heat exchangers.

Acknowledgements This research project was financially supported by Brazilian National Council of Research (CNPq), through the project 162620/2015-2. __________________________________________________________________________________________ Appendix A Uncertainty analysis in the calculation All experimental uncertainty is basically made up by two different components. The first component is derived from repeated observations (type A uncertainty) and the second component is obtained from manufacturer’s specifications and the instruments calibrations (type B uncertainty), according to Del Col et al. (2014). The combined standard uncertainty was used to establish the uncertainty of each parameter measured (external temperature, mass flow, absolute and differential pressures) and calculated (heat flux, internal and saturation temperatures, mass velocity, heat transfer coefficient, vapor quality such as thermodynamics and transport properties). A methodology used to calibrate and stablish the combined standard uncertainty of thermocouples, proposed by Oliveira et al. (2015), was implemented in this experiment. Experimental uncertainties measured and calculated in this work are show in Table A1. The combined standard uncertainty in heat flux, based on Eq. (1), is given by Eq. (A1) 2

u q " 

2

 q"   q"   q"  dU    dI    dA H    U   I   A H 

2

(A1)

where dU, dI and dAH represent the uncertainties of voltage, current and heated area presented in Table A1. The combined standard uncertainty in internal wall is given by Eq.(A2), based on Eqs. (2) and (3), ignoring uncertainty related to thermal conductivity of stainless steel.

Page 8 of 17

u T iw ( z )  

2

  T iw ( z )    T iw ( z )  d T ew ( z )    d q TS   q  T ( z )  TS ew   

2

2

   T iw ( z )    T iw ( z )    dr e    dr i    r   r  e i     

2

(A2)

Since the saturation temperature is assumed to be a linear function of saturation pressure, it can be describe as Eq. (A3). T sat ( z )  a  p sat  b

(A3)

In fact, the coefficients a and b were obtained considering a linear regression between 24 °C and 26 °C (340.54 kPa and 361.02 kPa). So, the uncertainty in saturation temperature is given by Eq. (A4). dT sat ( z )  u T sat ( z )   a  dp sat

(A4)

where dpsat is assumed to be equal to the uncertainty of absolute pressure transducer in the inlet of test section. Then, combined standard uncertainty in local heat transfer coefficient is defined according to Eq. (A5). 2

2

u h ( z )  

 h ( z )   h  h ( z )   dq "    d T iw ( z )    dT sat    q"    T sat   T iw ( z ) 

   

2

(A5)

Li et al. (2012), Del Col et al. (2013) and Charnay et al. (2015) present similar uncertainty analysis of heat transfer coefficient. The vapor quality in each location has its combined standard uncertainty related to local enthalpy and the enthalpy of liquid and vapor phases, shown in Eq. (A6). 2

u x( z) 

2

 x ( z )   x ( z )   x ( z )   di ( z )    di l    di v   i( z )   il   iv 

2

(A6)

where di(z) is obtained by solving the combined standard uncertainty from Eq. (7). Similarly, to saturation temperature, dil and div are considered as functions of saturation pressure and are given by Eqs. (A7) and (A8). di l  u i l   c  dp sat

(A7)

di v  u i v   d  dp sat

(A8)

REFERENCES __________________________________________________________________________________________ Anwar, Z., Palm, B., Khodabandeh, R., 2014. Flow boiling heat transfer and dryout characteristics of R152a in a vertical mini-channel. Exp. Therm. Fluid Sci. 53, 207 – 217. Aprin, L., Mercier, P., Tadrist, L., 2011. Local heat transfer analysis for boiling of hydrocarbons in complex geometries: A new approach for heat transfer prediction in staggered tube bundle. Int. J. Heat Mass Transf. 54, 4203 – 4219. Bertsch, S. S., Groll, E. A., Garimella, S. V., 2009. A composite heat transfer correlation for saturated flow boiling in small channels. Int. J. Heat Mass Transf. 52, 2110 – 2118. Carey, V. P., 1992. Liquid-Vapor Phase-Change Phenomena: An Introduction to the Thermophysics of Vaporization and Condensation Processes in Heat Transfer Equipment, 1st ed., Taylor & Francis, New York. Charnay, R., Revellin, R., Bonjour, J., 2015. Flow boiling heat transfer in minichannels at high saturation temperatures: Part I – Experimental investigation and analysis of the heat transfer mechanisms. Int. J. Heat Mass Transf. 87, 636 – 652. Choi, K-. I., Pamitran, A.S., Oh, C-. Y., Oh, J-.T., 2007. Boiling heat transfer of R-22, R-134a, and CO2 in horizontal smooth minichannels. Int. J. Refrigeration 30, 1336 – 1346. Choi, K-. I., Pamitran, A.S., Oh, J-.T., Saito, K., 2009. Pressure drop and heat transfer during two-phase flow vaporization of propane in horizontal smooth minichannels. Int. J. Refrigeration 32, 837 - 845. Copetti, J.B., Macagnan, M.H., Zinani, F., 2013. Experimental study on R-600a boiling in 2.6 mm tube. Int. J. Refrigeration 36, 325 - 334. Del Col, D., Bortolin, S., Torresin, D., Cavallini, A., 2013. Flow boiling of R1234yf in a 1 mm diameter channel. Int. J. Refrigeration 36, 353 – 362.

Page 9 of 17

Del Col, D., Bortolato, M., Bortolin, S., 2014. Comprehensive experimental investigation of two-phase heat transfer and pressure drop with propane in a minichannel. Int. J. Refrigeration 47, 66 – 84. Ducoulombier, M., Colasson, S., Bonjour, J., Haberschill, P., 2011b. Carbon dioxide flow boiling in a single microchannel – Part II: Heat transfer. Exp. Therm. Fluid Sci. 35, 597 – 611. Jung, D., Kim, C-.B., Song, K., Park, B., 2000. Testing of propane/isobutane mixture in domestic refrigerators. Int. J. Refrigeration 23, 517 – 527. Jwo, C-.S., Ting, C-.C., Wang, W-.R., 2009. Efficiency analysis of home refrigerators by replacing hydrocarbon refrigerants. Measurement 42, 697 – 701. Kandlikar, S. G., 1990. A general correlation for two-phase flow boiling heat transfer coefficient inside horizontal and vertical tube. J. Heat Transfer 102, 219 – 228. Kandlikar, S. G., Balasubramanian, P., 2004. An extension of the flow boiling correlation to transition, laminar, and deep laminar flows in minichannels and microchannels. Heat Transf. Eng. 25(3), 86 – 93. Keepaiboon, C., Wongwises, S., 2015. Two-phase flow patterns and heat transfer characteristics of R134a refrigerant during flow boiling in a single rectangular micro-channel. Exp. Therm. Fluid Sci. 66, 36 – 45. Kew, P. A., Cornwell, K., 1997. Correlations for the prediction of boiling heat transfer in small-diameter channels. Applied Thermal Eng. 17(8 – 10), 705 – 715. Kim, S-.M., Mudawar, I., 2013. Universal approach to predicting saturated flow boiling heat transfer in mini/micro-channels – Part II. Two-phase heat transfer coefficient. Int. J. Heat Mass Transf. 64, 1239 – 1256. Kundu, A., Kumar, R., Gupta, A., 2014. Comparative experimental study on flow boiling heat transfer characteristics of pure and mixed refrigerants. Int. J. Refrigeration 45, 136 – 147. Lazarek, G.M., Black, S.H., 1982. Evaporative heat transfer, pressure drop and critical heat flux in a small vertical tube with R-113. Int. J. Heat Mass Transfer 25, 945 – 960. Lemmon, E. W., Hube, M. L., Mclinden, M. O., Physics and chemical properties division, REFPROP 9.1, NIST Standard Reference Database 23, Version 9.1. Li, W., Wu, Z., 2010. A general correlation for evaporative heat transfer in micro/mini-channels. Int. J. Heat Mass Transf. 53, 1778 – 1787. Li, M., Dang, C., Hihara, E., 2012. Flow boiling heat transfer of HFO1234yf and R32 refrigerant mixtures in a smooth horizontal tube: Part I. Experimental investigation. Int. J. Heat Mass Transf. 55, 3437 – 3446. Lin, S., Kew, P. A., Cornwell, K., 2001. Two-phase heat transfer to a refrigerant in a 1 mm diameter tube. Int. J. Refrigeration 24, 51 – 56. Liu, Z. Bi, Q., Guo, Y., Su, Q., 2012. Heat transfer characteristics during subcooled flow bboiling of a kerosene kind hydrocarbon fuel in a 1 mm diameter channel.. Int. J. Heat Mass Transf. 55, 4987 – 4995. Maqbool, M. H., Palm, B., Khodabandeh, 2013. Investigattion of two phase heat transfer and pressure drop of propane in a vertical circular minichannel. Exp. Thermal Fluid Sci. 46, 120 – 130. Oliveira, J. D., Zanette, G. P., Coan, P. W., Passos, J. C., Güths, S., Copetti, J. B., 2015. Uncertainty analysis during vapor flow inside MPE microchannels. In: 23rd ABCM International Congress of Mechanical Engineering. Rio de Janeiro, Brazil. Ong, C. L., Thome, J. R., 2009. Flow boiling heat transfer of R134a, R236fa and R245fa in a horizontal 1.030 mm circular channel. Int. J. Thermal Fluid Sci. 33, 651 – 663. Ozawa, M., Ami, T., Ishihara, I., Umekawa, H., Matsumoto, R., Tanaka, Y., Yamamoto, T., Ueda, Y., 2009. Flow pattern and boiling heat transfer of CO2 in horizontal small-bore tubes. Int. J. Multiphase Flow 35, 699 – 709. Park, C. Y., jang, Y., Kim, B., Kim, Y., 2012. Flow boiling heat transfer coefficients and pressure drop of FC-72 in microchannels. Int. J. Multiphase Flow 39, 45 – 54. Pettersen, J., 2004. Flow vaporization of CO2 in microchannel tubes. Exp. Thermal Fluid Sci. 28, 111 – 121. Qi, S. L., Zhang, P., Wang, R. Z., Xu, L. X., 2006. Single-phase pressure drop and heat transfer characteristics of turbulent liquid nitrogen flow in micro-tubes. Int. J. Heat Mass Transf. 50, 1993 – 2001. Rouhani, S.Z., Axelsson, E., 1970. Calculations of void volume fraction in the subcooled and quality boiling region. Int. J. Heat Mass Transf. 13, 383 – 393. Saisorn, S., K-Oh, J., Wongwises, S., 2010. Flow pattern and heat transfer characteristics of R-134a refrigerant during flow boiling in a horizontal circular mini-channel. Int. J. Heat Mass Transf. 53, 4023 – 4038. Saitoh, S., Daiguji, H., Hihara, E., 2005. Effect of tube diameter on boiling heat transfer of R-134a in horizontal small-diameter tubes. . Int. J. Heat Mass Transf. 48, 4973 – 4984. Saraceno, L., Celata, G. P., Furrer, M., Mariani, A., Zummo, G., 2012. Flow boiling heat transfer of refrigerant FC -72 in microchannels. Int. J. Thermal Sci. 53, 35 – 41. Sempértegui-Tapia, D. F., Ribatski, G. 2015. Flow boiling heat transfer and two-phase pressure drop of isobutane in a 1.1 mm diameter tube. In: 23rd ABCM International Congress of Mechanical Engineering. Rio de Janeiro, Brazil. Shiferaw, D. Huo, X., Karayiannis, T. G., Kenning, D. B. R., 2007. Examination of heat transfer correlations and a model for flow boiling of R134a in small diameter tubes. Int. J. Heat Mass Transf. 50, 5177 – 5193.

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Shin, J. Y., Kim, M. S., Ro, S. T., 1997. Experimental study on forced convective boiling heat transfer of pure refrigerants and refrigerant mixtures in a horizontal tube. Int. J. Refrigeration 20, 267 – 275. Sun, L., Mishima, K., 2009. An evaluation of prediction methods for saturated flow boiling heat transfer in minichannels. Int. J. Heat Mass Transf. 52, 5323 – 5329. Tibiriçá, C. B., Ribatski, G., 2010. Flow boiling heat transfer of R134a and R245fa in a 2.3 mm tube. Int. J. Heat Mass Transf. 53, 2459 – 2468. Wang, S., Gong, M. Q., Chen, G. F., Sun, Z. H., Wu, J. F., 2014. Two-phase heat transfer and pressure drop of propane during saturated flow boiling inside a horizontal tube. Int. J. Refrigeration 41, 200 – 209. Wen, M-Y., Ho, C-Y., Jang, J-K., 2007. Boiling heat transfer of refrigerant R-600a/R-290-oil mixtures in the serpentine small-diameter U-tubes. Applied Thermal Eng. 27, 2353 – 2362.

Figure 1 – Experimental test rig and location of each thermocouple on external wall in test section.

Page 11 of 17

Flow direction

Plug flow

G [kgm-2s-1] 480

q"ST [kWm-2] 5

G [kgm-2s-1] 480

q"ST [kWm-2] 10

x(i-TS) [-] 0.0

G [kgm-2s-1] 320

q"ST [kWm-2] 5

x(i-TS) [-] 0.02

G [kgm-2s-1] 480

q"ST [kWm-2] 20

x(i-TS) [-] 0.0

G [kgm-2s-1] 240

G [kgm-2s-1] 240

q"ST [kWm-2] 5

G [kgm-2s-1] 240

q"ST [kWm-2] 10

x(i-TS) [-] 0.34

x(i-TS) [-] 0.47

jl(o-TS) jv(o-TS) [ms-1] [ms-1] 1.1 1.5 Slug flow

jl(o-TS) jv(o-TS) [ms-1] [ms-1] 2.17 3.47 Slug flow

Rel(o-TS) [-] 2903

Rel(o-TS) [-] 2430

x(o-TS) jl(o-TS) jv(o-TS) [-] [ms-1] [ms-1] 0.07 1.94 3.7 Churn flow

x(o-TS) jl(o-TS) jv(o-TS) [-] [ms-1] [ms-1] 0.11 3.4 7.1 Churn flow

x(i-TS) [-] 0.05

q"ST [kWm-2] 10

q"ST [kWm-2] 60

x(o-TS) [-] 0.05

x(i-TS) [-] 0.0

G q"ST [kgm-2s-1] [kWm-2] 400 10 Wave liquid front

G [kgm-2s-1] 320

x(o-TS) [-] 0.01

x(o-TS) jl(o-TS) [-] [ms-1] 0.14 3.26 Wavy-annular flow

jv(o-TS) [ms-1] 7.77

x(o-TS) jl(o-TS) jv(o-TS) [-] [ms-1] [ms-1] 0.4 3.79 10.1 Smooth-annular flow

x(o-TS) [-] 0.92

jl(o-TS) jv(o-TS) [ms-1] [ms-1] 3.26 9.3 slug flow

Rev(o-TS) [-] 541.7

α(o-TS) [-] 0.64

Rev(o-TS) [-] 2581

Rel(o-TS) [-] 1973

Rev(o-TS) [-] 3181

Rel(o-TS) [-] 2686

Rel(o-TS) [-] 2228

Rel(o-TS) [-] 1282

Rel(o-TS) [-] 1631

α(o-TS) [-] 0.30

Rev(o-TS) [-] 6623

Rev(o-TS) [-] 7429

Rev(o-TS) [-] 17430

Rev(o-TS) [-] 33342

α(o-TS) [-] 0.73

α(o-TS) [-] 0.79

α(o-TS) [-] 0.82

α(o-TS) [-] 0.91

α(o-TS) [-] 0.96

Rel(o-TS) [-] 1556

Rev(o-TS) [-] 3479

α(o-TS) [-] 0.77

x(i-TS) x(o-TS) jl(o-TS) jv(o-TS) Rel(o-TS) [-] [-] [ms-1] [ms-1] [-] 0.04 0.16 2.3 6.16 1425 Figure 2 – Two phase flow patterns observed in tests.

Rev(o-TS) [-] 5633

α(o-TS) [-] 0.83

x(i-TS) [-] 0.04

x(o-TS) [-] 0.08

jl(o-TS) [ms-1] 1.88 churn flow

jv(o-TS) [ms-1] 4.07

Page 12 of 17

Figure 3 – Local heat transfer coefficients versus vapor quality at constant mass velocities: a) 240 kg m-2 s-1; b) 320 kg m-2 s-1; c) 400 kg m-2 s-1; d) 480 kg m-2 s-1.

Figure 4 – Effect of mass velocity on local heat transfer coefficient: (a) For G = 240 kg m-2 s-1 and 480 kg m-2 s-1, q”ST = 10 kW m-2 and 60 kW m-2. (b) For G = 240 kg m-2 s-1 and 320 kg m-2 s-1, q”ST = 20 kW m-2 and 60 kW m-2.

Page 13 of 17

Figure 5 – Effect of mass velocity on average heat transfer coefficient for different heat fluxes for x(i-TS) = 0.15.

Figure 6 – Boiling curves for x(i-TS) = 0.15 and different mass velocities, G.

Page 14 of 17

Figure 7 – Comparison of predictions of heat transfer coefficient correlation for flow boiling with experimental database for (a) Kew and Cornwell (1997), (b) Kandlikar Balasubramanian (2004), (c) Sun and Mishima (2009), (d) Bertsch et al. (2009), (e) Li and Wu (2010) and (f) Kim and Mudawar (2013).

Page 15 of 17

Table 1 - Summary of some experimental flow boiling heat transfer studies in microchannels. Author Lin et al. (2001)

Channel C

Saitoh et al. (2005)

C

Qi et al. (2006)

C

0.531, 0.834, 1.042, and 1.931

C C C

1.03

Ozawa et al. (2009)

C

1.0, 2.0 and 3.0

CO2

Saisorn (2010) Tibiriça and Ribatski (2010) Ducoulombier et al. (2011)

C

1.75

R-134a

C

2.3

C

0.529

Liu et al. (2012)

C

1.0

Park et al. (2012) Saraceno et al. (2012) Anwar et al. (2014)

R

0.061 and 0.278

A kerosene kind hydrocarbon fuel FC-72

C

1.0

FC-72

1025 – 2000

-

10 – 150

-0.5 – 0.3

C

1.6

R152a

100 – 500

27 – 32

5 - 245

0-1

Kundu et al. (2014)

C

7.0

R-134a, R-410a and R-407C

100 - 400

5-9

3 – 10

0.1 – 09

Including quasi-azeotropic mixture (R410a) and zeotropic mixture (R-407C)

3.0

R-245fa

100 - 1500

100 - 120

10 – 50

0–1

-

0.68

R-134a

600 - 1400

23 - 31

7.63 – 49.46

-

-

Shiferaw et al. (2007) Choi et al. (2007) Chin and Thome (2009)

Charnay et al. C (2015) Keepaiboon and R Wongwises (2015) C = circular; R = rectangular

Dh [mm] 1.0 0.51, 1.12 and 3.1

Fluids R-141b

G [kg/m²s] 300 – 2000

Tsat [°C] 39 – 56

q" [kW/m²] 10 – 1150

x[-] 0–1

Observation -

R-134a

150 – 450

5 – 15

5 – 39

0 – 0.2

-

N2

1000 - 10000

-

19 – 35.4

-

-

4.26 and 2.01

R-134a

100 – 500

-

13 – 150

0 – 0.9

-

1.5 and 3.0

R-134a, R-22 and CO2 R-134a, R-236fa and R-245fa

200 – 600

10

10 - 40

0–1

-

200 – 1600

31

2.3 – 250

0–1

-

0.05 – 0.9

-

5.3 – 26.8

5 – 50

200 – 1000

200 – 700

35, 43 and 52

1 – 83

-

-

R-134a and R-245fa

50 – 700

22, 31 and 41

5 – 55

0.1 – 1

-

CO2

200 - 1200

-10, -5 and 0

10 - 30

0–1

-

1260, 1680 and 2100 188 – 1539

25 – 40 52.8 – 84.4

20 – 1500 0.6 – 45.1

0–1

Pressure 3 – 7 bar and ΔTsub in = 8 – 33 K -

Page 16 of 17

Table 2 - Special features of present study. Parameter Values Heat flux in test section, q”ST 5, 10, 20, 40, 60 Heating power in pre-heater, PPH 5, 15, 25, 35 Mass velocity, G 240, 320, 400, 480 Saturation temperature 25

Unit kW m-2 W kg m-2 s-1 °C

Table A1 - Characteristics of instruments, geometric parameters and uncertainty. Parameter Instrument Range Uncertainty Temperature

Type – E Thermocouple

- 200 °C – +900 °C

u(T) [± 0.25 °C]

Absolute and saturation pressure

Piezoelectric transducer

0 MPa – 1 MPa

u(p) [± 1 kPa]

Differential pressure

Piezoelectric transducer

0 kPa – 75 kPa

u(dp) [± 0.1 kPa]

Mass flow

Coriolis mass flow meter

0 gs-1 – 10 g s-1

 ) [± 0.001 gs-1] u( m

Heat flux

DC power supply

I [0 A- 125 A] U [0 V - 8 V]

u(I) [500 mA] u(U) [8mV]

Heated area

Electronic microscope

-

u(AH ) [4.17x10-6 m2]

Internal and external diameters

Electronic telescope

-

u(re ) and u(ri ) [2.5x10–5 m]

Internal wall temperature

Eq. (A2)

-

u T iw  [± 0.26 °C]

Saturation Temperature

Eq. (A4)

-

u T sat ( z )  [± 0.1 °C]

Heat transfer coefficient

Eq. (A5)

2.282 kWm-2 K-1– 25.2 kW m-2 °K-1

Vapor quality

Eq. (A6)

0 % – 92 %

For q”ST = 5 kW m-2, u(h(z))/h(z) q”ST = 10 kW m-2, u(h(z))/h(z) q”ST = 20 kW m-2, u(h(z))/h(z) q”ST = 40 kW m-2, u(h(z))/h(z) q”ST = 60 kW m-2, u(h(z))/h(z) For q”ST = 5 kW m-2, u(x(z))/x(z) q”ST = 10 kW m-2, u(x(z))/x(z) q”ST = 20 kW m-2, u(x(z))/x(z) q”ST = 40 kW m-2, u(x(z))/x(z) q”ST = 60 kW m-2, u(x(z))/x(z)

[17.9% – 34.3%] [12.1% - 33.2%] [8.2% – 19.2%] [5.8% – 13.1%] [4.9% – 9.8%] [0.4% – 9.0%] [0.6% – 10%] [0.8% – 13%] [1.1% – 15%] [1.2% – 17%]

Page 17 of 17