Condensing Heat Transfer Coefficients for R134a at

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condensation of R134a in an inclined smooth copper tube of inner diameter of ... heat transfer occurred at close to +40o (upward flow). In the experimental ... condensation in small/mini diameter tube approach the study .... and water flows.
Proceedings of the ASME 2013 Heat Transfer Summer Conference HT2013 July 14-19, 2013, Minneapolis, MN, USA

HT2013-17375 CONDENSING HEAT TRANSFER COEFFICIENTS FOR R134a AT DIFFERENT SATURATION TEMPERATURES IN INCLINED TUBES

Adekunle O. Adelaja Department of Mechanical and Aeronautical Engineering, University of Pretoria Pretoria, South Africa [email protected]

Jaco Dirker Department of Mechanical and Aeronautical Engineering, University of Pretoria Pretoria, South Africa [email protected]

ABSTRACT

m

This paper presents the effects of saturation temperature and inclination angle on convective heat transfer during condensation of R134a in an inclined smooth copper tube of inner diameter of 8.38 mm. Experiments were conducted for inclination angles ranging from -90° (vertical downward) to +90° (vertical upward) for mass fluxes between 100 kg/m2s and 400 kg/m2s and vapour qualities between 0.1 and 0.9 for saturation temperatures ranging between 30 °C and 50 °C. The results show that saturation temperature and inclination angles strongly influence the heat transfer coefficient. With respect to saturation temperature, an increase in saturation temperature generally leads to a decrease in heat transfer coefficient irrespective of the inclination angle. The effect of inclination angle was found to be more pronounced at mass fluxes of 100 kg/m2s and 200 kg/m2s for the range of vapour qualities considered. Within the region of influence of inclination there is an optimum angle which is between 15° and -30° (downward flow). The inclination effect corresponds to the predominance of the effect of gravity on the flow distribution.

R T x z

Q

mass flow rate (kg/s) heat transfer rate (W) thermal resistance (K/W) temperature (°C) vapour quality axial direction

Greek symbols heat transfer coefficient (W/m2K) Į ȕ inclination angle (radian) Subscripts Cu copper water H2O i inner in inlet j measurement location l liquid o outer out outlet pre pre-condenser ref refrigerant sat saturation test test-condenser v vapour w wall

NOMENCLATURE A cp d EB g G h k L

Josua P. Meyer Department of Mechanical and Aeronautical Engineering, University of Pretoria Pretoria, South Africa [email protected]

area (m2) specific heat (J/kg.K) diameter (m) energy balance gravitational acceleration (m2s) mass flux (kg/m2s) enthalpy (J/kg) thermal conductivity (W/mK) length of test section (m)

INTRODUCTION Heat exchangers in heat pumps, refrigeration and airconditioning systems play a vital role in industrial and domestic applications whether, for human comfort, preservation of farm

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inner diameter. Their study showed that the effect of annular flow boundary appear to be insignificant at the small inclination angles. Wang and Du [7] in their study of film-wise condensation in small/mini diameter tube approach the study both theoretically and experimentally. In their experiment, steam was used as the working fluid and the mass flux ranged between 10 kg/m2s and 100 kg/m2s for the four copper pipes of inner diameter of 1.94 mm, 2.8 mm, 3.95 mm and 4.98 mm. They concluded that gravity, unlike in larger tubes has a decreasing effect on flow condensation in small/mini tubes and that inclination angle mainly distributed gravity in stratifying the fluid and thinning the liquid film. No optimum inclination angle was reported. Wurfel et al. [8] examined flow condensation of n-heptane/air, water/air, and condensing nheptane inside an inclined tube and reported that the heat transfer coefficient increases with inclination, thus the consideration of the tube inclination proves to be useful for the description of the heat transfer during film condensation in turbulent two-phase flow. Akhavan-Behabadi et al. [9-11] investigated both corrugated and microfin tubes in the condensation heat transfer of R134a for low mass fluxes for the whole range of inclination angle. For microfin tubes, the mass fluxes ranged between 53 kg/m2s and 212 kg/m2s, vapour qualities between 0.2 and 0.8, average saturation temperature of 35 °C for tube of inner diameter of 8.92 mm. For the corrugated tube, the mass fluxes range was between 87 kg/m2s and 253 kg/m2s, for average condensing temperatures of between 24 °C and 38 °C for copper tube of inner diameter of 8.32 mm. Apart from the fact that the microfin and corrugated tubes had enhanced heat transfer, an optimum inclination angle of +30o (upward flow) was obtained. Theoretical investigations of the heat transfer coefficient in R134a, R141b and R11 were carried out by Safari and Naziri [12]. The governing differential equations of kinematics, dynamics and thermal thermal aspects of two-phase flow were derived from a thin film flow modelling and solved simultaneously while numerical solutions were obtained. It was reported that the optimum angle was achieved between -30o and -50° (downward flow). For R134a, the optimum angle was -30° at saturation temperature of 30 °C, for Re = 40,000. Lyulin et al. [13], in their study used pure ethanol vapour in downward flow at saturation temperature of 58 °C concluded that heat transfer coefficient reduces with growth of temperature difference between saturation and wall temperatures. The optimum inclination angle obtained was between -15° and -35° (downward flow). Lips and Meyer [14-16], in their study of convective condensation of R134a in a smooth tube of inner diameter of 8.38 mm, vapour qualities 0.1 – 0.9, mass flux of 200 – 400 kg/m2s for the whole range of inclination angle at saturation temperature of 40oC reported a significant influence of inclination angle on heat transfer especially at low mass fluxes and low vapour qualities. In their study, optimum inclination angle of -15° and -30° (downward flow) were obtained.

produce, or for cooling of heat emitting appliances or equipment. These heat exchanging systems operate at different design temperatures for example: the hot water heat pumps are designed for high temperatures of between 50oC and 60oC, industrial heat exchanging systems that are water-cooled operate at about 40oC, while the domestic heat exchange systems operate at lower than 40oC. Various researchers have investigated the performance of heat exchangers at these temperatures using tubes of different sizes, orientations and tube surfaces whether smooth or inner/ external grooved such as the microfins, corrugated and in wire inserts. Of particular interest in this study is the use of inclined smooth tubes subjected to different saturation temperatures. Extensive literature reviews have been done by Cavallini et al. [1], Daikilic and Wongwises [2] and Lips and Meyer [3]. These reviews show that limited studies have been conducted on convective condensation in inclined smooth tubes for the whole range of inclination angles. To the best of the authors’ knowledge, studies on convective condensation in inclined smooth tubes for the whole range of inclination angles at different saturation temperature are rare in the open literature. A brief overview is given here. Studies on condensation heat transfer in inclined tubes for reflux condensation are considered first. Reflux condensation are increasingly been applied to compact heat exchangers which are controlled by gravity. For this system, vapour enters the inclined or vertical condenser and flow upwards while the condensate flow downward counter currently under the influence of gravity. Fiedler and Auracher [4] investigated the heat transfer, flooding point and thickness during the reflux condensation of R134a in an inclined smooth tube of 7.0 mm both theoretically and experimentally. Inclination angle was found to have significant effects on the heat transfer and flooding point. The optimum angle where the highest flooding took place was between +45o (upward flow) and +60o (upward flow) whereas the highest heat transfer occurred at close to +40o (upward flow). In the experimental study of heat transfer during reflux condensation of the binary zeotropic mixture of R134a/ R123 in a narrow tube (7.0 mm inner diameter and 0.5 m length) and a rectangular channel (7.0 mm hydraulic diameter and 0.5 m length). Results show that the overall heat transfer coefficient is smaller during condensation of R134a. At Reynolds numbers between 50 and 80, an enhanced heat transfer of about 20% compared with +90° inclination was observed when the inclination angle was +45o. The authors did not categorically state that +45o is the optimum inclination angle, however, they acknowledged that at Reynolds numbers higher than 80 there was no noticeable effect of inclination and also that the thinning effect of the fluid thickness at inclined position was responsible for higher coefficient of heat transfer [5]. On convective condensation heat transfer, Nitheanandan and Soliman [6] investigated the effect of 10o upward and downward inclinations within the different transition lines on flow regime boundaries of condensing steam. The mass flux used was between 20 kg/m2s and 280 kg/m2s and 13.8 mm

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From the works that have been reviewed on condensation heat transfer, whether of steam or refrigerant in inclined tubes, theoretical or experimental, it is clear that various authors arrived at their conclusions based on the test apparatus and operating conditions or assumptions imposed on their investigations. These conditions seem to have significant effect on the results on which conclusions were based. However, this study is aimed at expanding knowledge on the effect of inclination angle and saturation temperature on heat transfer coefficient in inclined tubes.

fully developed, a straight calming section, 50 diameter long was situated at the entrance. At the inlet and outlet of the test section were sight glasses which enabled flow visualisation and also served as insulator against axial heat conduction. A high speed camera was installed at the outlet sight glass and used to record and document the flow patterns. A uniform Phlox backlight was positioned against the sight glass to enable good colour fidelity due to its evenly distributed light emitting diode (LED) illumination. Three pressure taps were connected between the sight glass and test section at both sides. Two of the pressure taps were connected to different Gem sensor pressure transducers for the measurement of the absolute pressure at both inlet and outlet sides of the section so that the absolute pressure recoding used on each side was the average of two pressures. The third pressure taps on each side were connected together to an FP 2000 Sensotec differential pressure transducer. Flexible hose was used to connect the test section so that it could be inclined at 90° upward and downward. On the inner-tube outer wall were drilled four holes in seven equidistant stations such that at each station four Tjunction thermocouples were soldered at the top, bottom and both sides of the tubes. The twenty eight T-type thermocouples spanned the 1.488 m length. The refrigerant temperatures were taken at six locations; two on each end of the condensers. These thermometers were arranged such that at each station three thermocouples were installed; one each at the top, bottom and on the side. The consistency of the readings at this section was checked between the saturation temperature obtained with the absolute pressure transducers and the saturation temperature acquired from the thermocouples. Similar to the refrigerant, water temperatures were obtained at six stations. All the thermocouples used were calibrated against a high-precision Pt-100 resistance temperature detector to an accuracy of 0.1 °C. Coriolis mass flow meters were used to measure the refrigerant and water flows. The temperature, pressure, mass flows and other parameter were read off and logged on a desktop computer through a data acquisition device.

EXPERIMENTAL APPARATUS AND TEST CONDITIONS Experimental Set-up The experimental rig used for the study was the same test facility on which some previous experiments have been conducted [14,17,18]. In these works, a comprehensive description and explanations have been made. However, an overview would be presented in this paper. The test rig consisted of a vapour-compression cycle and a water cycle (Figure 1). The vapour-compression cycle comprised of two high pressure lines (the test line and the bypass line) and a low pressure line through which refrigerant R134a was pumped with the aid of a hermetically sealed scroll compressor with a nominal cooling load of 10 kW. Refrigerant flow in each of the high pressure lines was controlled by electronic expansion valves (EEV). The test line consisted of three large condensers; the precondenser, test condenser, and the post condenser. The precondenser was used to regulate the inlet vapour quality into the test-condenser where test measurements were carried out whereas the post-condenser was adjusted such that it ensured that there was complete condensation i.e. liquid refrigerant reached the EEV. The bypass condenser was used to control the mass flow rate, temperature, and pressure of refrigerant flowing through the test section. The high pressure lines, after the EEVs, united and led to the scroll compressor through the low pressure line which consisted of the evaporator and suction accumulator. Cold and hot water were supplied by a 50 kW heating and 70 kW cooling dual function heat pump located about 6 m above the test section. It was thermostatically controlled such that the cold and hot water were set at 15 °C – 25 °C and 40 °C respectively. The hot and cold water were stored in two 5000 l insulated water tanks before supplied to the set-up. While hot water ran through the evaporator, the cold was supplied to the condensers. The test section was a copper tube-in-tube counterflow heat exchanger in which the refrigerant flowed in the inner tube and water in the annulus. It was 1.488 m long and had inner tube inside diameter of 8.38 mm and an annulus diameter of 15.9 mm. To ensure that the flow though the test section was

FIGURE 1 SCHEMATIC DIAGRAM OF EXPERIMENTAL SET-UP

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TABLE 1.EXPERIMENTAL VARIABLES AND UNCERTAINTIES

Parameter Tsat G x

E

Range 30oC -50oC 100 - 400 kg/m2s 0.1-0.9  90 o d E d 90 o

was used for the calculations of the fluid properties, heat transfer coefficient and other parameters of interest. The main uncertainties in the heat transfer rate come from the uncertainties in the temperature measurement on the water side while it was in the saturation and wall temperature on the refrigerant side. The uncertainties on the mean temperature measurement were approximately 0.1 K on both water and refrigerant side for each of the stations. Uncertainties in the vapour qualities were lower than 0.03 and that which resulted from the mass flow meters were negligible. However, that on the heat transfer coefficient was found to be between 4% and 9%.

Uncertainties r 0.6 oC r 5kg / m 2 s r 0.01 r 0 .1o

Data Acquisition Data from the thermocouples, pressure transducers, and Coriolis flow meters were gathered by a computerised data acquisition (DAQ) system. The DAQ system was made up of a Desktop computer on which was installed a LabVIEW programme written to automatically acquire needed data, an Analog/Digital Interface Card, shielded cable assembly, signal conditioning extensions for instrumentation (SCXI), transducer multiplexers, terminal blocks, channel multiplexers, and termination units. The LabVIEW programme displayed all measurements both primary and secondary quantities which were monitored until steady state was reached before data was captured. The refrigerant pressure was measured at the inlet of the test section by an analogue strain guage transducer with an accuracy of r 2 kPa for low mass fluxes (100 kg/m2s – 200 kg/m2s) and low vapour qualities (0.1 – 0.25) and r 12 kPa for high mass fluxes (G > 300 kg/m2s) and vapour qualities (x > 0.5). The measured value when used with the condensation curve provided by REFPROP gave the saturation temperature which was verified by direct measurement. The difference in the two values was found to be less than 0.1 °C at high mass fluxes (400 kg/m2s and above) while a higher fluctuation was observed at low mass fluxes. This, however, might be due to the non-uniformity of the flow particularly at the smooth stratified, stratified—wavy and intermittent regions. The pressure drop across the test section was measured by a differential transducer which was calibrated to an accuracy of ±0.05 kPa.

DATA REDUCTION AND EXPERIMENTAL PROCEDURE The energy balance, Eq. (1), was the criterion used to ascertain that the system had reached steady state. Once the temperature, pressure and mass flows of the system reached steady state and the energy balance was stable within 3% data was acquired.

EB

Qref  QH 2O

(1)

Qref

Temperature and pressure measurements were used to determine the properties of the refrigerants at the entrance of the pre-condenser and at the exit of the post condenser. The thermo-physical properties of the condensing fluid were obtained by employing data from a refrigerant property database [19]. The inlet quality xin at the test condenser was calculated from the enthalpy of the refrigerant at the inlet of the test section and the enthalpies of the liquid hl and htest,in vapour hv at the same condition of temperature and pressure as:

x in Test Conditions and Methods

htest ,in  hl

(2)

h v  hl

The enthalpy of the refrigerant at the inlet of the test section htest,in however was calculated from the enthalpy at the inlet of the pre-condenser hpre,in (obtained using the temperature and the pressure condition at the inlet of the pre-condenser), the heat  ref in the transfer rate Qpre and the refrigerant mass flow rate m

The test conditions described in Table 1 give the range of experimental variables used in this study and their uncertainties. Tests were conducted for each saturation temperature for a particular mass flux and vapour quality while the angle of inclination of the test section was varied. In most cases, tests were conducted for various mass fluxes (200 – 400 kg/m2s), for different qualities (0.1 – 0.9). During the experiment, the test section heat transfer was maintained at between 230 W and 270 W. After steady state was reached, the different sensor signals were recorded continuously through the DAQ system for about a period of 5 minutes (201 points). In order to avoid noise measurement, the average of the points

pre-condenser.

htest ,in

4

h pre ,in 

Q pre m ref

(3)

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Thome Two-Phase Flow Pattern Map 700

The heat transfer through the pre-condenser was calculated from the water mass flow rate, m H 2O, pre , the specific heat

600

capacity cP and the inlet temperature, Tpre,in, and outlet temperature, Tpre,out :

m H 2O , pre c p T pre ,in  T pre ,out

(4)

The vapour quality at the exit of the test section was calculated by replacing the enthalpy at the inlet, htest,in , by the enthalpy at the outlet, htest,out , in equation (2) at the refrigerant saturation temperature and pressure. The enthalpy at the outlet of the test section is however calculated as follows:

htest ,in 

htest ,out



Qtest

300

200

Stratified-wavy

0.1

0.2

0.3

0.4 0.5 0.6 Vapour Quality

0.7

0.8

0.9

1

FIGURE 2 EXPERIMENTAL TEST POINTS (HORIZONTAL) ON THE THOME-EL HAJAL-CAVALLINI [20] FLOW PATTERN MAP FOR R134A (d = 8.38 mm).

Where Twj, o is the average temperature at the jth location of the seven different stations and (zj+1 - zj) is the distance between the measurement locations. Figure 2 summarises the matrix of the 42 experimental conditions for horizontal orientation on the Thome- El Hajal Cavallini [20] flow pattern map. The conditions basically represent annular and stratified-wavy flow regime with some test conditions also being in the intermittent regime. The appropriate application of the flow pattern model should involve the use of the actual mass velocity and saturation temperature. However, Figure 2 was constructed using mass flux of 300 kg/m2s at 40 °C saturation temperature for simplicity for all the experimental conditions. A similar test matrix was used for all other inclination angles.

(6)



T sat=50 oC

400

0

m ref

A Tw,i  Tsat

T sat=45 oC

Statified

The heat flow Qtest through the test condenser was obtained by replacing the mass flow rate, specific heat capacity, and the inlet and outlet temperatures in equation (4) by these same parameters at the test section conditions. The mean vapour quality of the test section was obtained from the arithmetic mean of the inlet and outlet qualities. The heat transfer coefficient in the test condenser was calculated from the Newton’s law of condensation as follows:

D

T sat=40 oC

100

(5)

Qtest

T sat=35 oC

Annular

500 Mass Velocity(kg/m2s)

Q pre

T sat=30 oC

Intermittent

where A is the inner surface area of the inner tube of the test condenser, Tsat is the mean of the saturation temperature at the inlet and outlet of the section. Tw ,i is the mean inner wall

EFFECT OF INCLINATION ANGLE ON HEAT TRANSFER COEFFICIENT

temperature and it is related to the mean outer wall temperature, Tw,o , of the tube through the thermal resistance of the wall of the copper tube, Rw: Tw,o  Qtest R w

Tw , i

with Rw

The condensation heat transfer coefficient in the test section was obtained for a range of saturation temperatures of 30 °C – 50°C, mass fluxes 100 kg/m2s – 600 kg/m2s, quality value 0.1 – 0.9 and inclination angles -90° to +90°. It was imperative to establish and test the integrity of the experimental apparatus and hence the data. 546 data points of the condensation of 134a inside a smooth incline tube was compared with six different well- established convective heat transfer correlation models. This was done by comparing the experimental data points of the horizontal flow only (286 data points) with these correlations. This was necessary because most of the earlier studies on heat transfer experiments and correlations were developed for this orientation.

(7)

ln d o d i 2SkCu L .

The average outer wall temperature was calculated using the trapezoidal numerical integration:

Tw,o

>



@

1 6 ¦ Twj,o  Twj,o1 z j 1  z j L j1

(8)

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Figure 3 displays the comparison of test data with the six notable heat transfer correlations for the horizontal flow. The correlation of Shah [21] was aimed at predicting the heat transfer coefficient irrespective of the tube orientation. It was based on a two-phase multiplier for condensation in single phase proposed by Dittus-Boelter [22] and it is capable of predicting flow in horizontal, vertical and inclined tubes. Dobson and Chato [23] improved on Chato [24] by including a stratified-wavy model. Jung et al. [25] on the other hand modified the correlation of Dobson and Chato [23] and included the heat and mass flux ratio to arrive at their model. Furthermore, the model of Thome et al. [20] was based on flow pattern map while the model by Cavallini et al. [26] was established on two different flow regimes depending on whether the heat transfer coefficient is dependent on the condensation heat flux (i.e. ȴT- dependent flow regime) or otherwise (i.e. ȴT- independent flow regime). Here ȴT refers to the temperature difference between the wall and the saturated fluid. Finally, the model of Kim and Ghajar [27] employed a two-phase heat transfer model developed for vertical pipe flow with modified constants to predict air-water, silicon-air, waterhelium and water-Freon 12 data for horizontal orientation. The models of Shah [21], Kim and Ghajar [27] and Jung et al. [25] under estimated the horizontal experimental data for low mass fluxes and low qualities whereas for high mass fluxes and/ or high vapour qualities, Dobson and Chato [23], and, Kim and Ghajar [27] over-predicted the data. However, Thome et al. [20] and Cavallini et al. [26] are in very good agreements (within r 25%) with horizontal tube data. For the Cavallini correlation this is shown in Figure 4 for different saturation temperatures.

2

[W/m[W/m K]2K] D Heat Transfer Coefficient Predicted cond,correlations

6000 6000

D cond,correlations [W/m K]

Predicted Heat Transfer Coefficient [W/m2K] 2

5000 -25%

3000

0

Shah (1979) Dobson & Chato (1998) Kim & Ghajar (2002) Jung et al (2003) Thome et al (2003) Cavallini et al (2006) 0

1000

2000

3000

4000

5000

6000

o TTsat ==30 30°C C

3000 3000

sat

o TTsat ==35 35°C C sat

2000 2000

o TTsat ==40 40°C C sat

o TTsat ==45 45°C C

1000 1000

sat

o TTsat ==50 50°C C sat

0 00

1000 1000

2000 2000

3000 3000

4000 4000

5000 5000

6000 6000

Evaluating the deviations of the six correlations against the present data show that the models of Shah [21], Jung et al. [25], Thome et al. [20] and Cavallini et al. [26] display mean deviation of 21.4%, 10.7%, 15.7% and 10.5% respectively with average deviation of less than 6.5%. Conversely, the models of Dobson and Chato [23] and Kim and Ghajar [27] showed larger average and mean deviations. While Kim and Ghajar gave an average deviation less than 13% and mean less than 22.5%, for Dobson and Chato [23] they were greater. The effects of inclination on the heat transfer coefficient for saturation temperatures (30 °C – 50 °C), mass flux of 100 kg/m2s – 400 kg/m2s and quality of x = 0.5 are shown in Figures 5–9. Figure 5 shows that for x = 0.5 and Tsat = 40 °C, the effect of inclination decreases as mass flux increases; however, there is a general increase in the coefficient of heat transfer. The effect of inclination is pronounced between -90o (vertical downward) and +15° (slightly inclined upward) for G = 100 kg/m2s and 200 kg/m2s with the highest heat transfer occurring at -15° (slightly inclined downward). This agrees with the findings of Lips and Meyer. Moreover, because annular flow exists for higher mass fluxes, they are independent of inclination for this case. Since inclination effect is prominent at low mass flux and or low vapour quality, it is reasonable to consider mass flux of 200 kg/m2s or less for the study of the effect of inclination angle on heat transfer rate for different saturation temperatures.

+25%

1000

- 25%

cond,experimental FIGURE 4 COMPARISON OF EXPERIMENTAL HEAT TRANSFER DATA WITH THE CORRELATION OF CAVALLINI et al. [26] FOR HORIZONTAL FLOW (ȕ = 0°C) AT DIFFERENT SATURATION TEMPERATURES

6000

2000

4000 4000

2 Experimental Heat Transfer Coefficient [W/m2[W/m K] K] D

7000

4000

+25%

5000 5000

7000

[W/m2K]

D cond,experimental Experimental Heat Transfer Coefficient [W/m2K]

FIGURE 3 A COMPARISON BETWEEN EXPERIMENTAL HEAT TRANSFER RESULTS AND DIFFERENT CORRELATIONS FOR HORIZONTAL FLOW AT Tsat=40°C

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3500

3000

2 Heat Transfer Coefficient [W/m K] 2

Heat transfer coefficient [W/m K]

2

2 Heat Transfer [W/m Heat transfer Coefficient coefficient [W/m K] K]

2900 3000

2500

2000

1500

100 kg/m22ss G=100kg/m 200 kg/m22ss G=200kg/m 300 kg/m22ss G=300kg/m 400 kg/m22ss G=400kg/m

1000

500 -90

-60

-30

0

30

60

2800 2700 2600 2500 2400

30°CoC Tsat=30

2300

35°CoC Tsat=35

2200

40°CoC Tsat=40 Tsat=45 45°CoC

2100

o Tsat=50 50°C C

2000 -90

90

o Inclination angle[ ] [°] Inclination Angle

FIGURE 5 INCLINATION EFFECT ON HEAT TRANSFER COEFFICIENT FOR DIFFERENT MASS FLUXES FOR CASE (Tsat = 40°C, x = 0.5)

-60

-30

0 30 Inclination Angle Inclination angle[ R] [°]

90

FIGURE 7 INCLINATION EFFECT ON HEAT TRANSFER COEFFICIENT FOR (G = 300 kg/m2s, Tsat = 30°C – 50°C, x = 0.5).

2

2500

Heattransfercoefficient[W/m K]

2 Heat Transfer Coefficient [W/m 2 K]

2 200 kg/m s 200 kg/m s [exp]

2

Heat Transfer Coefficient Heat transfer coefficient [W/m [W/m2K]K]

60

2000

30°CoC Tsat=30

1500

35°CoC Tsat=35 40°CoC Tsat=40 45°CoC Tsat=45

[exp] 200 kg/m s 300 kg/m2 s [exp] [corl] 300 kg/m2 s [corl] 2 300 kg/m s 400 kg/m2 s [exp] [exp] 2 2 400 kg/m s [corl] 300 kg/m s 200 kg/m2 s [corl] 2

4000

3500

3000

2500

Tsat=50 50°CoC 1000 -90

2000

-60

-30

0 30 Inclination angle[ R] [°] Inclination Angle

60

90

30

35

40

45

50

o Saturation Temperature [°C] Saturationtemperature C

FIGURE 8 COMPARISON OF THE EFFECT OF SATURATION TEMPERATURE ON EXPERIMENTAL HEAT TRANSFER COEFFICIENT FOR DIFFERENT MASS FLUXES WITH CAVALLINI et al. [26] FOR (x = 0.5, ȕ = 0°).

FIGURE 6 INCLINATION EFFECT ON HEAT TRANSFER COEFFICIENT FOR (G = 200 kg/m2s, Tsat = 30°C – 50°C, x = 0.5).

Figure 6 shows that irrespective of the saturation temperature, the effects of inclination angle on the heat transfer coefficient are similar and except at saturation temperature of 40oC, an optimum inclination angle of -30o (downward flow) was obtained. Inclination effect tends to increase with saturation temperature. Compared with the vertical downward flow, the improvement in heat transfer coefficient for ȕ = -30° was 21.4%, 25.7%, 38.6%, 43.8% and 48.7% for saturation temperatures of 30oC, 35oC, 40oC, 45oC and 50oC respectively. These indicate that there is gain in heat transfer rate as saturation temperature increases. However, the heat transfer coefficient increases with decreasing saturation temperature. For the case when the mass flux was 300 kg/m2s, Figure 7 shows that the effect of inclination was not observed at lower saturation temperatures (30 °C – 40 °C), it only appears at saturation temperature of 45 °C and the effect was greater for 50 °C.

Therefore, it can be inferred that up to mass flux of 2 s, saturation temperature in the range tested here 300 kg/m 3500 increases the tendency of refrigerant to respond to gravitational effect (i.e. to stratify). At saturation temperatures below 45 °C for the 3000 case when the mass flux was 300 kg/m2s and at all higher mass fluxes, annular flow predominated. Figure 8 shows the experimental and Cavallini et al. [26] 2500 correlation heat transfer coefficient dependence on the refrigerant saturation temperature for x = 0.5 in a horizontal tube (ȕ = 0°). It is obvious that the local heat transfer generally 2000 30 35 45 50 decreases with increasing saturation40temperature for mass flux Saturation Temperature [°C] 2 200-400 kg/m s. When compared with the correlation of Cavallini et al. [26], identified by the ’corl’ – tag in Figure 8, a similar trend was observed though, it over-predicted the experimental data.

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Heat Transfer Coefficient [W/m2K]

Heattransfercoefficient[W/m2K]

3200 3200

Conversely, Figure 9, (for x = 0.5) shows that the condensation heat transfer coefficient increases as mass flux increases for -30° downward flow. Generally, for the mass fluxes considered increase in saturation temperature results in reduction in heat transfer coefficient. For G = 300 kg/m2s at Tsat = 45 oC and 50 °C, the inclination angle influences tne heat transfer coefficient. In Figures 10 and 11, the effect of vapour qualities on the heat transfer coefficient for saturation temperatures of 40 °C and 50 °C for mass fluxes of 200 kg/m2s and 300 kg/m2s respectively are given at the optimum inclination angle of ȕ = -30°. The heat transfer coefficients for ȕ = 0° are also included for comparative purposes. Figure 10 shows that for Tsat = 40 °C, the heat transfer coefficient increases with vapour quality. For Tsat = 50 °C, there was increase in heat transfer coefficient between x = 0.25 and x = 0.5 after which a zigzag profile was observed, visible for both ȕ = -30° and ȕ = 0°. In all cases it was observed that there is an enhancement in heat transfer at ȕ = -30° as compared with the horizontal flow. As much as 13.9% heat enhancement was observed at x = 0.25 for Tsat = 50 °C. As vapour quality increases these heat gain decreases. This is because as quality increases, the flow tends to be annular. A critical observation of Figure 11 (where G = 300 kg/m2s) shows that inclining the tube at ȕ = -30° does not have any effect when the vapour qualities are higher than x = 0.55 for Tsat =40 °C, and higher than x = 0.65 for Tsat =50 °C. The study clearly shows that saturation temperature has significant influence on the heat transfer coefficient, even, on inclination effect and mass flux. In Earlier studies, Lips and Meyer [14] show that for G = 300 kg/m2s at Tsat = 40 °C, inclination effect can only be observed up to x = 0.25. In the current study, it has been shown that at higher saturation temperatures especially for the case where G = 300 kg/m2s the effect of inclination can be pronounced up to x = 0.5. Therefore, flow pattern map/ distribution and other flow parameters are subject to change with saturation temperature.

30°C o Tsat=30 3000 3000

2800 2800 2600 2600

35°C o Tsat=35 40°C o Tsat=40 45°C o Tsat=45 50°C o Tsat=50

2400 2400 2200 2200 2000

2000 200 200

250 250

300 300

350 350

400 400

Mass Flux [kg/m2s]2

massflux[kg/m s] FIGURE 9 EFFECT OF SATURATION TEMPERATURE ON HEAT TRANSFER COEFFICIENT FOR DIFFERENT MASS FLUXES (x=0.5, ȕ = -30°)

2300 2300

2

2 Heat Transfer Coefficient [W/m K] Heattransfercoefficient[W/m K]

2400 2400

2200 2200 2100 2100 2000 2000 1900 1900

40°C, = E-30° T =40ȕoC, =Ͳ30o sat

1800 1800

40°C, = E0° T =40ȕoC, =0o sat

1700 1700

50°C, = E-30° T =50ȕoC, =Ͳ30o sat

1600 1600 1500 1500 0.2 0.2

o 50°C, ȕ =oC,0° E =0 T =50 sat

0.3 0.3

0.4 0.4

0.5 0.5

0.6

0.7 0.7

0.8 0.8

0.9 0.9

1

Vapour Quality [-] Vapourquality[Ͳ]

FIGURE 10 COMPARING THE EFFECT OF SATURATION TEMPERATURES AND VAPOUR QUALITIES ON HEAT TRANSFER COEFFICIENT FOR OPTIMUM INCLINATION AND HORIZONTAL FLOWS FOR G = 200 kg/m2s.

2

Heat Transfer Coefficient [W/m K]

Heattransfercoefficient[W/m2K]

3400 3400

3200 3200

CONCLUSIONS

3000 3000 2800 2800

In this study, experiments in convective condensation were performed in a smooth inclined tube of angle varying between -90o (vertical downward) and +90o (vertical upward) using R134a as the working fluid. Experiments were conducted over a range of saturation temperature (between 30 °C and 50 °C) for vapour qualities of 0.1-0.9, mass flux 100 kg/m2s – 600 kg/m2s and heat transfer rate at the water side was maintained at 230 W – 270 W. The effects of inclination angles and saturation temperature were investigated with respect to heat transfer coefficient. Generally, heat transfer rate increases with mass flux and vapour quality and the effect of inclination is more pronounced at low mass fluxes, low vapour qualities and high

2600 2600 2400 2400

40°C, T =40ȕoC,=E-30° =Ͳ30o sat

2200 2200

40°C, T =40ȕoC,=E0° =0o sat

2000 2000

T =50ȕoC,=E-30° =Ͳ30o 50°C, sat

1800 1800

o o E =0 T =50 50°C, ȕ =C,0° sat

1600 1600 0.1 0.1

0.2 0.2

0.3 0.3

0.4 0.4

0.5 0.5

0.6 0.6

0.7 0.7

0.8 0.8

0.9 0.9

Vapour Quality [-] Vapourquality[Ͳ]

FIGURE 11 COMPARING THE EFFECT OF SATURATION TEMPERATURES AND VAPOUR QUALITIES ON HEAT TRANSFER COEFFICIENT FOR OPTIMUM INCLINATION AND HORIZONTAL FLOWS FOR G = 300 kg/m2s.

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[12] Saffari, H., and Naziri, V., 2010, “Theoretical Modelling and Numerical Solutions of Stratified Condensation in Inclined Tubes,” J. Mech. Sci Technol., 24, pp. 2587-2596. [13] Lyulin, Y., Marchuk, I., Chikov, S., and Kabov, O., 2011, “Experimental Study of Laminar Convective Condensation of Pure Vapour Inside an Inclined Circular Tube,” Microgravity Sci. Technol., 23, pp. 439-445. [14] Lips, S., and Meyer, P. J., 2012a, “Experimental Study of Convective Condensation in an Inclined Smooth Tube. Part 1: Inclination Effect on Flow Pattern and Heat Transfer Coefficient,” Int. J. Heat and Mass Transfer, 55, pp. 395-404. [15] Lips, S., and Meyer, P. J., 2012b, “Experimental Study of Convective Condensation in an Inclined Smooth Tube. Part II: Inclination Effect on Pressure Drops and Void Fractions,” Int. J. Heat and Mass Transfer, 55, pp. 405-412. [16] Lips, S., and Meyer, P. J., 2012c, “Stratified Flow Model for Convective Condensation in an Inclined Tube,” Int. J. Heat and Fluid Flow, 36, pp. 83-91. [17] Van-Rooyen, E., Christian, M, Liebenberg, L., and Meyer, J. P., 2010, “Probabilistic Flow Pattern-Based Heat Transfer Correlation for Condensing Intermittent Flow in Smooth Horizontal Tubes,” Int. J. Heat and Mass Transfer, 53, pp. 14461460. [18] Suliman, R. Liebenberg, L., and Meyer, J. P., 2009, “Improved Flow Pattern Map for Accurate Prediction of the Heat Transfer Coefficient during Condensation of R134a in Smooth Horizontal Tubes and within the Low-Mass Flux Range,” Int. J. Heat and Mass Transfer, 52, pp. 5701--5711. [19] REFPROP., 2005, “NIST Thermodynamic Properties of Refrigerants and Refrigerant Mixtures (REFPROP),” version 8.0, NIST Standard Reference Database 23, National Institute of Standards and Technology, Gaithersbury, MD. [20] Thome, J. R. El Hajal, J., and Cavallini, A. 2003, “Condensation in Horizontal Tubes. Part II: New Heat Transfer Model Based on Flow Regimes,” Int. J. Heat and Mass Transfer, 46, pp. 3365--3387. [21] Shah, M., 1979, “A general Correlation for Heat Transfer during Film Condensation Inside Pipes,” Int. J. Heat and Mass Transfer, 22, pp. 547—556. [22] Dittus, F. W., and Boelter, L. M. K., 1930, “Heat Transfer for Automobile Radiators of the Tubular Type,” Publications in Engineering, University of California, Berkeley, 2, pp. 443. [23] Dobson, M., and Chato, J., 1998, “Condensation in Smooth Horizontal Tubes,” ASME J. Heat Transfer, 120, pp. 193-213. [24] Chato, J. C., 1960, “Laminar Condensation inside Horizontal and Inclined Tubes,” Massachusetts Inst. Technol. [25] Jung, D., Song, K., Cho, Y., and Kim, S., 2003, “Flow Condensation Heat Transfer Coefficients of Pure Refrigerant,” Int. J. Refrigeration, 26, pp. 4-11. [26] Cavallini, A., Del Col, D., Doretti, L., Matkovic, M., Rossetto, L., Zilio, C., Censi, G.., 2006, “Condensation in Horizontal Smooth Tubes: A New Heat Transfer Model for Heat Exchanger Design,” Heat Transfer Eng., 27, pp. 31-38. [27] Kim, D., and Ghajar, A.J., 2002, “Heat Transfer Measurement and Correlations for Air-Water Flow of Different Flow Patterns in a Horizontal Pipe,” Exp. Thermal Fluid Sci., 25, pp. 659-676.

saturation temperature. With respect to the saturation temperature, heat transfer usually reduces irrespective of the inclination angles and mass fluxes. Furthermore, inclination is more pronounced at higher saturation temperatures.

ACKNOWLEDGMENTS The funding obtained from the NRF, TESP, University of Stellenbosch/ University of Pretoria, SANERI/SANEDI, CSIR, EEDSM Hub and NAC is acknowledged and duly appreciated.

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