- Email: [email protected]
Condensation of R-134a inside dimpled helically coiled tube-in-shell type heat exchanger
Applied Thermal Engineering 129 (2018) 535–548
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Condensation of R-134a inside dimpled helically coiled tube-in-shell type heat exchanger Anand Kumar Solanki ⇑, Ravi Kumar Department of Mechanical and Industrial Engineering, Indian Institute of Technology, Roorkee 247667, India
h i g h l i g h t s Dimple helically coiled tube gives a higher heat transfer coefficient. The effect of mass flux, vapor quality and saturation temperature are considered. Flow transitions for dimpled helically coiled tube occurred at lower vapor quality. Dimple helically coiled tube yields a higher pressure drop than straight tube.
a r t i c l e
i n f o
Article history: Received 1 July 2017 Revised 29 September 2017 Accepted 4 October 2017 Available online 12 October 2017 Keywords: Condensation Heat transfer coefficient Frictional pressure drop R-134a Dimpled helically coiled tube
a b s t r a c t In this study, condensation heat transfer coefficients and frictional pressure drops of R-134a inside a dimpled helically coiled tube are experimentally investigated. The inner tubes comprise of one smooth straight tube, one smooth helically coiled tube and one dimpled helically coiled tube. The experimental measurements are carried out at saturation temperature of 35°, and 45 °C with mass flux of 75, 115, 156 and 191 kg m2 s1. The experimental data of a smooth and dimpled helically coiled tube have been plotted on the mass flux versus vapor quality flow map and Tailtel and Dukler flow map. The transitions between different flow regimes have also been discussed. Moreover, the effect of mass flux, vapor quality and saturation temperature of R-134a on the heat transfer coefficients and pressure drops are examined. Comparisons between smooth straight tube, smooth helically coiled tube and dimpled helically coiled tube are also discussed. The dimple helically coiled tube produces a higher heat transfer coefficient and frictional pressure drop compared to smooth helically coiled tube and smooth straight tube. The correlations have been proposed to predict the Nusselt number and frictional pressure drop multiplier during condensation of R-134a inside horizontal dimpled helically coil tube. Ó 2017 Elsevier Ltd. All rights reserved.
1. Introduction The condenser is an essential part of any refrigeration and air conditioning unit which condenses a substance from its gaseous to liquid state by removing the latent heat. Nowadays, the effective design of the condenser is the important for the industries to obtain a maximum heat transfer with a smaller amount of pressure drop. As such as, helically coiled tube have been widely used to enhance the heat transfer by producing centrifugal effect on the flow inside the tube [1]. Nevertheless, A number of literature has been studied on two phase flow in smooth round tube [2]. Dobson and Chato [3] reported the experimental heat transfer coefficient data with refrigerants R22, R134a, R410A, R32/R125 (60/40% by
⇑ Corresponding author. E-mail address: [email protected] (A.K. Solanki). https://doi.org/10.1016/j.applthermaleng.2017.10.026 1359-4311/Ó 2017 Elsevier Ltd. All rights reserved.
mass) inside smooth round horizontal tubes with inner diameters ranging from 3.14 to 7.04 mm. They also proposed the correlation to predict the heat transfer coefficient for the stratified and annular flow pattern. Cavallini et al. [4] experimentally measured the heat transfer coefficient and pressure drop of pure HFC refrigerants (R134a, R125, R236ea, R32) and the nearly azeotropic HFC refrigerant blend R410A inside a smooth tube of diameter 8 mm at saturation temperature in the range between 30 °C and 50 °C with mass velocities varying from 100 to 750 kg m2 s1. The results show that, during condensation of pure fluids and nearly azeotropic mixtures, in the annular flow regime, the heat transfer coefficient increases with rise of mass velocity and vapor quality. Jung et al. [5] carried out the experimental heat transfer coefficient data with refrigerants R12, R22, R32, R123, R125, R134a, and R142b in a horizontal plain copper tube of 9.52 mm outside diameter and 1 m length at a fixed refrigerant saturation temperature of 40 °C with mass fluxes of 100, 200, 300 kg m2 s1 and heat flux of
536
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
Nomenclature A De Cp Re d EF D PF g
heat transfer area, m2 Dean number specific heat at constant pressure, J kg1 K1 Reynolds number tube diameter, mm enhancement factor coil mean diameter, mm penalty factor gravity, 9.81 m2 s1
Greek symbols e depth of dimple, mm a void fraction o diameter of dimple, mm v Martinelli parameter s pitch of dimple, mm h helix angle, ° p helical pitch, mm q density, kg m3 G mass velocity, kg m2 s1 l dynamic viscosity, kg m1 s1 h heat transfer coefficient, W/m2 K / two-phase multiplier h enthalpy, J/kg D difference l length, m Subscripts Nu Nusselt number a acceleration
7.3–7.7 kW m2. They found that the heat transfer coefficient of refrigerant R142b and R32 were higher than those of R22 by 8– 34% at the same mass flux while heat transfer coefficient of refrigerant of refrigerant R134a and R123 were similar to those of R22. Hossain et al. [6] conducted the experiments with refrigerants R1234ze(E), R32, R410A and a zeotropic mixture R1234ze(E)/R32 (55/45 mass%) inside a water heated double tube heat exchanger. They investigated that the experimental heat transfer coefficient of R1234ze (E) is lower than R1234ze (E)/R32 (55/45 mass %), R410A and R32 by 11%, 56% and 83%, respectively. Moreover, a few paper are also reported on the two phase-flow of refrigerant inside the dimpled straight tube during condensation process. Aroonrat and Wongwise [7] examined the condensation heat transfer coefficient and frictional pressure drop of R-134a inside the smooth and spherical dimpled straight tube at saturation temperature of 40°, 45° and 50 °C with the mass flow rate of 300, 400 and 500 kg m2 s1. They found that the heat transfer coefficient ratio for spherical dimpled straight tube varies from 1.27 to 1.37 at 300 kg m2 s1, 1.28 to 1.36 at 400 kg m2 s1 and 1.26 to 1.41 at 500 kg m2 s1, respectively, while, the two-phase friction factor ratio varies from 2.8 to 3.7 at 300 kg m2 s1, 2.7 to 4.2 at 400 kg m2 s1 and 2.6 to 4.1 at 500 kg m2 s1, respectively. Sarmadian et al. [8] investigated the convective condensation heat transfer and frictional pressure drops of R-600a inside a helically dimpled tube and a plain tube of internal diameter 8.3 mm at saturation temperature of 38 and 48 °C with mass flux varying from 114 to 368 kg m2 s1. They noticed that the heat transfer coefficients of the dimpled tube were 1.2–2 times of those in smooth tube with a pressure drop penalty ranging between 58% and 195%. Besides, a few research work have been performed on the heat transfer and pressure drop characteristics inside helical coiled
K G x f Q l T g ph Eq P exp _ m fg Fr v sat R i w o ts w tt out W in Prl T
thermal conductivity, W m1 K1 gravitational vapor quality frictional heat transfer rate, W liquid temperature, °C gas preheater equivalent pressure, kPa experimental mass flow rate, kg/s latent heat of vaporization vapor Froud number saturation condition refrigerant inner wall outer test section water turbulent–turbulent outlet water inlet Prandtl number total
tube. Kang [9] carried out the experimental study of heat transfers and pressure drops data of R-134a inside long helical coil tube. The inner diameter, outer diameter, coil diameter and number of turns of the tube were taken 12.7 mm, 21.2 mm, 177.8 mm, 34.8 mm and 10, respectively. The test runs were executed at different refrigerant mass flux varying from 100 to 400 kg m2 s1, cooling water flow Reynolds number range of 1500–9000, cooling tube wall temperature at 12 °C and 22 °C and fixed saturation temperature of 33 °C. The results revealed that the overall heat transfer coefficient raised with increase of the water mass flux. The heat transfer and pressure drop decreased with increase of tube wall temperature from 12 °C and 22 °C. Yu et al. [10] conducted an experimental investigation of heat transfer of R-134a during condensation inside horizontal, inclined and vertical helical tube at refrigerant mass flux from 100 to 400 kg m2 s1, cooling water Reynolds number between 1500 and 10000. They investigated that the helical tube orientation had significant influence on the overall heat transfer coefficients. The highest and lowest heat transfer coefficient was observed at inclined position and vertical position, respectively. Han et al. [11] reported the experimental results of condensation heat transfer and pressure drop of R-134a inside helical tube with mass flux vary from 100 to 420 kg m2 s1 at a different condensation saturation temperature 35, 40 and 46 °C. They observed that the refrigerant mass flux and the saturation temperature had a significant influence on the heat transfer coefficient. Wongwises and Polsongkram [12] presented experimental studies on two-phase flow condensation heat transfer and pressure drop of R-134a inside concentric helical coil tube in tube condenser. The experimental measurement were performed at different saturation temperature 40 and 50 °C with refrigerant mass flux between 400 and 800 kg m2 s1. The results showed that the heat transfer coefficient of the helical coiled tube were about 33–53%
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
higher than that of smooth round tube while pressure drop were about 29–46% higher. They also proposed the correlation for predicting the heat transfer coefficient and pressure drop. Lin and Ebadian [13] reported the effect of inclination angle (0°, 45° and 90°) on condensation heat transfers and pressure drops of R-134a flowing inside the annular section of tube. The experimental measurement was carried out at saturation temperature of 30 and 35 °C with refrigerant mass flux from 60 to 200 kg m2 s1, cooling water Reynolds number range between 3600 and 22,000 at cooling water temperature of 16, 20 and 24 °C. The Nusselt number rises with the increase of both refrigerant mass flow rate and cooling water mass flow rate. They also found that the Nusselt number of the refrigerant-side for angle 0–45° was about two times more than that of angle 0–90°. Shao et al. [14] compared the condensation heat transfer coefficient for the helical tube with straight tube at different saturation temperature 35, 40 and 45 °C, mass flux vary from 100 to 400 kg m2 s1. The heat transfer coefficients of helical tube were found to 4–13.8% higher than that of straight tube. AlHajeri et al. [15] examined the performance of heat transfers and pressure drops of R-134a inside annular helical tubes at different saturation temperature of 36, 42 and 48 °C with the mass flux varying from 50 to 680 kg m2 s1. The results showed that the heat transfer coefficients of the refrigerant-side rise with increase of the refrigerant mass flux and decline with increase of the saturation temperature of refrigerant. However, with rise of the mass flux, the pressure drops of refrigerant increased. Mosaad et al. [16] investigated the effect of refrigerant mass flux and condensation temperature difference on the heat transfer coefficients and pressure drops of R-134a inside coiled double tube. They found that, with increase of refrigerant mass flux and decrease of condensation temperature difference, the overall heat transfer coefficient increased. Similarly, pressure drop of refrigerant increases with rise of the refrigerant mass flux. Murai et al. [17] used a highspeed video system to take photographs of interfacial structure of gas-liquid flow inside the helical coiled tube. They observed that the flow transition from bubbly to plug flow for helical coiled tube compared to straight tube was obtained earlier duo to the centrifugal acceleration generating by curvature of the tube. Genic´ et al. [18] examined the shell-side heat transfer coefficient pertaining to three heat exchangers with helical coils. The geometric parameter such as winding angle, pitch and axial pitch had significant influence on shell-side heat transfer coefficient. They also proposed the correlation for shell-side heat transfer coefficient based on the shell-side hydraulic diameter. Genic´ et al. [19] also investigated experimentally the shell-side fouling factors of eight heat exchangers with parallel helical coils. The fouling resistances were slightly lesser than usual values for shell-and-tube heat exchangers with straight pipes. Gupta et al. [20] determined the effect of mass flux, vapor quality and saturation temperature on the heat transfer coefficients and pressure drop of R-134a inside horizontal helical coiled tube-in-shell type heat exchanger. The heat transfer coefficient and pressure drop of refrigerant increase with rise of mass flux and vapor quality and decrease with increase of the saturation temperature. M. Mozafari et al. [21] conducted the experimental studies of heat transfer and pressure drop of R600a inside helical tube-intube heat exchanger at different inclination angles of zero, +30°, +60° and +90° with different mass flux ranging from 155 to 265.5 kg m2 s1 at saturation temperature 38.5–47 °C. The results showed that the heat transfer coefficients were remarkably influenced by inclination angle but pressure drop were slightly affected. Maximum heat transfer coefficients and pressure drops were obtained at +30° and 0° angles, respectively. The effect of different curvature radii (43.5, 63.5 and 76.5 mm) and coil pitch (15, 25 and 35 mm) of helical coil tube-in-shell on the heat transfer coefficients of R-404A were investigated by Salimpour et al. [22]. Both curvature radii and coil pitch of helical coil tube had significant
537
influence on the heat transfer coefficient. They also found that, the effect of coil pitch on heat transfer coefficient was more pronounced at low vapor qualities. However, to the best our knowledge from the above mentioned literature, most of the research works have been done on both single and two phase flow inside smooth helically coiled tube. Besides, some experimental investigation are currently available on the condensation heat transfer and pressure drop of refrigerant inside dimpled straight tube. But, any research work inside the dimpled helically coiled tube-in-shell type heat exchanger during condensation of R134a has not been reported. In this study, the condensation heat transfer coefficients and frictional pressure drops of R134a inside dimpled helically coiled tube-in-shell with respect to vapor quality (0.1–0.84), mass flux (75, 115, 156 and 191 kg m2 s1) and saturation temperature (35 and 45 °C) have been examined. In addition, results obtained from dimpled helical coiled tube are compared with those for the straight tube and smooth helical coiled tube.
2. Experimental apparatus and procedures The simplified schematic diagram of the experimental set-up was depicted in Fig. 1. The main component of experimental setup was composed of a pre-heater, test-condenser, postcondenser, three refrigerant magnetic gear pumps, coriolis mass flow meter and data acquisition system. The test section was a counter flow type heat exchanger. The test section was connected to post condenser, where two phase quality of refrigerant coming from test section was sub-cooled to a liquid state by extracting heat to cold water circulated from the condensing unit. The municipal water was employed for cooling purpose which was circulated inside the shell of the test-section. The flow rate of the cooling water was regulated by a manually valve. The temperature of water inside the test-section was controlled by variac of 2 kW capacity. A slight glass after the post-condenser was placed to ensure the continuous liquidity of refrigerant before entering the three magnetic gear pump connected side by side. The flow rate of refrigerant was monitored by operating the speed of the magnetic gear pump through frequency inverter. The coriolis mass flow meter followed by magnetic gear pump was accommodated to measure the refrigerant flow rate. In order to control the vapor quality at the inlet of the test condenser, an appropriate preheater was designed from a 6 m long U-bend stainless steel tube. By supplying high current and low voltage through step down auto transformer, liquid refrigerant was evaporated by heating the tube. A standard filter-drier was incorporated following the pre-heater to remove any foreign particles and moisture in refrigerant loop. A number of hand shut-off valves were facilitated to provide facile installation of a component in the apparatus. Fig. 2 depicts a schematic diagram of the test section. The test section was a counter flow helical coiled tube in shell type heat exchanger in which the refrigerant flowing inside the inner tube extracts the heat to the cooling water flowing the opposite direction to refrigerant flow inside the shell. There were three type of tubes taken as inner tube, one smooth straight tube, one smooth helical coiled tube and another one dimple helical coiled tube. The inner helically coiled tube, which was made by wrapping the straight tube over the wooden pattern. As a next step, the helically coiled tube was processed to make the projections on its surface by using a round tip tool and thus a dimple helically coiled tube was obtained. The shape of dimples in the inner surface of the helical coiled tube were spherical. The Schematic diagram of the inner tubes are shown in Fig. 3. Details of the inner tubes are given in Table 1. Two quartz sight glass were mounted at the inlet and outlet of the test section to visualize the flow patterns. The pho-
538
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
Fig. 1. Schematic diagram of experimental set-ups.
Water outlet
25mm
120 mm 100 mm 140 mm Refrigerant out
Refrigerant in
200 mm Water inlet Fig. 2. Schematic diagram of test-section.
tograph of the visualization section can be seen in Fig. 4. The inside diameter of the sight glass is identical to that of the test tube. The photographs of the flow patterns at the downstream of the test section i.e. sight glass, are captured by using high speed camera at 1000 frame per second. The sight glass with negligible surface roughness and shorter length results minor disruption in the downstream flow pattern. Hence, the redevelopment of a flow pattern in a sight glass is insignificant. The T-type thermocouples were located to measure the refrigerant temperature at the inlet of the pre-heater, inlet and outlet of the test-section. The outer wall temperature of the helical coiled tube was also measured with T-type thermocouple at six positions. In each position, four thermocouple were placed 90° apart over the top, bottom and two
sides. A special glue was used to fix thermocouples on the wall to avoid water resistance. In the present study, Presys T-25 N calibrator having temperature range 25 °C to 125 °C was used to calibrate the T-type thermocouples. The thermocouples were calibrated in the operating temperature range of 10 °C to 70 °C with an accuracy of ±0.1 °C of full scale. The pressure Transducers with accuracy 0.25% were utilized to measure the refrigerant pressure at the inlet and outlet of the test-section and at the entrance of the pre-heater. Also Pressure drop between the test-section was measured by a differential pressure drop transducer. The coriolis flow meter with an accuracy of 0.1% was employed to measure the refrigerant mass flow rate. The water flow rate in annulus tube was measured by a turbine
539
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
(a)
s θ
(b) o
p e
(c) Fig. 3. Sketch diagram of the tubes: (a) Smooth straight tube, (b) Smooth helically coiled tube and (c) Dimpled helically coiled tube.
Table 1 Dimension of inner tubes. Parameter
Smooth straight tube
Smooth helically coiled tube
Dimpled helically coiled tube
Inner diameter of tube, mm Outer diameter of tube, mm Length of tube, mm Inside tube area, mm2 Coil mean diameter, mm Coil pitch, mm Number of turns Dimple pitch (s), mm Helical pitch (p), mm Dimpled depth (e), mm Helix angle (h), ° Dimpled diameter (o), mm
8.92 9.52 1000 26,156 – – – – – – – –
8.92 9.52 2200 57,543 110 25 6 – – – – –
8.92 9.52 2200 61,041 110 25 6 8.5 10 1 49 2
100 mm
A quartz glass tube
in Table 2. The uncertainties of experimental results are evaluated by methodology which is suggested by Klein and McClintock [23] as shown in Table 3. 3. Data reduction The data reduction analysis is employed to evaluate the average heat transfer coefficient and the average frictional gradient during an experiment at steady state conditions. For calculating the thermo-physical properties of R-134a, REFPROF-9 [24] is utilized. The average heat transfer coefficient (havg) is obtained by the following equation:
hav g ¼
Qw Ai ðT sat T wi Þ
ð1Þ
where Qw, Ai and Twi are the heat transfer rate inside the test section from tube to flowing water, inner surface area on tube and average inner wall surface temperature of the tube, respectively. Tsat is the saturation temperature of refrigerant corresponding to the saturation pressure which is calculated by the inlet pressure of the test section and the total pressure drop along the test section. For pressure drop, a linear profile is assumed along the test section. The average inner wall surface temperature of the tube is given by
T wi ¼ T wo þ Q w ln
Do 2pKL Di
ð2Þ
where Two, Do, Di, K and L are the average temperature of outside tube wall, outer diameter of the tube, inner diameter of the tube, thermal conductivity of the copper and length of the tube, respectively. Fig. 4. Visualization section.
flow meter. The turbine flow meter has an accuracy of 1% of the full scale. All thermocouple and pressure transducer signal was transferred to the multichannel data acquisition system with PXI controller. The steady state condition was assumed when reading of temperatures, pressure and mass flow rate remain constant for at least 20 min. The details of the experimental condition are shown
Table 2 Experimental conditions. Parameter
Range
Refrigerant Average saturation temperature (°C) Mass flux (kg m2 s1) Inner tube material Heat flux (kW m2) Vapor quality
R-134a 35, 45 75, 115, 156, 191 Copper 1.5–14.34 0.1–0.84
540
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
For horizontal tube, the gravitational pressure gradient (dP/ dz)tp,g is considered as negligible. The acceleration pressure gradient is defined as following equation:
Table 3 Experimental uncertainties. Measurement
Uncertainty
Temperature (°C) Refrigerant mass flow rate (kg s1) Water mass flow rate (kg s1) Pressure (kPa) Pressure drop (kPa) Average vapor quality Heat transfer rate at pre-heater Heat transfer rate at test-section Heat transfer coefficient
0.1 0.050 103 1.41 103 5.201 0.125 5% 7.2% 11% 15%
( !) dP d x2 ð1 xÞ2 ¼ G2 þ dz tp;a dl qg a ql ð1 aÞ
where a is the void fraction which is calculated from Abdul-Razzak et al. correlation [25]:
a ¼ ð1 þ 0:49v0:3036 Þ tt
ð3Þ
_ w , Cpw, Tw,out and Tw,in are the mass flow rate of water inside where m the test section, specific heat of water, water temperature at outlet of test section and water temperature at inlet of test section, respectively. The average vapor quality of the refrigerant is the arithmetic mean of vapor quality at inlet and outlet of the test section which can be defined as following equation:
xav g;ts ¼
x þ x out;ts in;ts 2
ð4Þ
where inlet vapor quality ðxin Þ of the test section is evaluated by applying energy balance equation at the preheater as follows:
Q ph 1 xin ¼ hin;ph þ hf ;in;ts hfg;ts;in mR
ð5Þ
where Qph is the electrical power supplied to the pre-heater. The preheater is well insulated with ceramic wool to prevent the heat loss from the preheater to environment. The ceramic wool is also covered with the steel strips. It is assumed that, the heat loss from preheater is negligible. The hin,ph is the enthalpy at the inlet of the preheater which is calculated by measuring the temperature and pressure of sub-cooled liquid refrigerant. It is assumed that energy loss between the preheater outlet and test section inlet is neglected. Hence, the enthalpy at the test section inlet is equal to the enthalpy at the outlet of pre-heater. The enthalpy of saturated liquid (hf,in,ts) and enthalpy of vaporization (hfg,ts,in) are determined based on the temperature at the inlet of the test-section. The outlet vapor quality of the test section is given by
xout ¼
hout;ts hf ;out;ts hfg;out;ts
ð6Þ
where hf,out,ts and hfg,out,ts are the enthalpy of the saturated liquid and vaporization at the outlet of the test section, respectively, which are calculated by outlet temperature of the test section. The hout,ts is the enthalpy of the refrigerant at the outlet of the test section which is determined by energy balance equation on the test section as follows:
Q hout;ts ¼ hin;ts w mR
ð7Þ
where hin,ts and mR are the enthalpy of refrigerant at the inlet of the test section and mass flow rate of the refrigerant, respectively. The total pressure gradient (dP/dz)T is directly measured by a differential pressure transducer at inlet and outlet of the test section. The frictional pressure gradient (dP/dz)tp,f is calculated by subtracting the acceleration pressure gradient (dP/dz)tp,a and the gravitational pressure gradient (dP/dz)tp,g from the total pressure drop, as
dP dP dP dP ¼ dz tp;f dz T dz tp;a dz tp;g
1
ð10Þ
The two phase frictional pressure gradient multiplier, /2l is expressed as following equation:
The heat transfer rate (Qw) is determined from:
_ w C pw ðT w;out T W;in Þ QW ¼ m
ð9Þ
ð8Þ
/2l ¼
ðdP=dzÞtp;f ðdP=dzÞl;f
ð11Þ
The single phase liquid pressure gradient (dP/dz)l,f can be determined from:
dP G2 ð1 xÞ2 ¼ fl 2 dz l;f ðdi ql Þ
ð12Þ
The fanning friction factor fl is determined from the correlation proposed by Srinivasan et al. [26] for the coil,
( )15 1=2 2 D di fl ¼ 0:084 Rel di D
ð13Þ
where Rel is the liquid Reynolds number which is given by following equation
Rel ¼
Gð1 xÞdi
ll
ð14Þ
4. Results and discussion The experimental data points of smooth and dimpled helically coiled tube-in-shell type heat exchanger during condensation of refrigerant of R134a at the saturation temperature of 35 °C are drawn on the Taitel and dukler flow map [27] (Martinelli parameter, vtt vs vapor Froud number, Frv), as shown in Fig. 5. The vapor Froud number, Frv and Martinelli parameter vtt are given by following equation,
Gx Frv ¼ pffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi qv di gðql qv Þ
vtt ¼
0:9 0:5 0:1 1x qv ll x ql lv
ð15Þ
ð16Þ
The solid line drawn in Taitel and Dukler flow map [27] are the transition line between stratified-wavy, intermittent flow and annular flow of the air-water mixture in the straight smooth tube. Similarly, the flow pattern transitions for smooth and dimpled helically coiled tube are shown in discontinuous line. It can be found that the smooth and dimpled helical coiled tube have nearly the same transition line from stratified wavy to annular as well as stratified wavy to intermittent, while the transition criteria for the smooth straight tube comes to below that of those both tubes. The intermittent flow region falls into the annular region of the map. This may be because, the flow pattern map suggested by Taitel and Dukler was constructed for the air–water experimental data, which have different thermo-physical properties from R-134a studied in the present work. The Martinelli parameter for the transition criteria in the straight tube from intermittent to annular is 1.6, which was given by Taitel and Dukler [27]. In the current tests,
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
541
Fig. 5. Comparison of smooth and dimpled helically coiled tube experimental data on Taitel–Dukler flow map. (a) Smooth helically coiled tube. (b) Dimpled helically coiled tube.
the transition criteria from intermittent to annular flow takes place at Martinelli parameter around 0.51 for smooth helical coiled tube and 0.61 for the dimpled helical coiled tube. To better understand the flow patterns of R-134a inside the smooth and dimpled helically coiled tube, the axis of the Taitel and Dukler flow pattern map have been converted to mass flux versus vapor quality. It can be clearly seen form Fig. 6(a) and (b) that the stratified-wavy flow occurs at low mass flux (75 kg m2 s1) for both tube. As the mass flux increases from 75 to 115 kg m2 s1, for dimpled helically coiled tube, the intermittent flow takes place at small range of vapor quality, while annular flow occurs at large range of vapor quality. Similarly, for smooth helically coiled tube, the stratified-wavy exists at low vapor quality, then transfers to annular flow at high vapor quality without the presence of any intermittent flow. Yang et al. [28] was also observed similar flow characteristics. Moreover, at mass flux greater than 115 kg m2 s1, the intermittent and annular flow exist for both tube. The flow pattern transition from intermittent to annular flow occurred at a vapor quality in the range of 0.32–0.33 for the smooth helically coiled tube, and 0.27 for the dimpled helically coiled tube. Cui et al. [29] proposed the transition from intermittent to annular flow at vapor quality range 0.2–0.3 for micro-fin helically coiled tube during boiling of R-134a. In Fig. 7, it can be observed that the flow pattern inside the dimpled helically coiled tube at G = 156 kg m2 s1 and x = 0.1310 prevails intermittent flow, whereas in the smooth helical coiled tube at G = 156 kg m2 s1 and x = 0.1028, the stratified flow exists. It should be observed that the difference is that the waves produced in stratified-wavy flow touch the upper portion of the tube wall and wet the upper portion of the tube, intermittently. Furthermore, vapor quality increases, the flow pattern inside the smooth helically coiled tube at x = 0.2158 becomes intermittent. In the dimpled helically coiled tube at x = 0.2728, the waves on the interface of the liquid-vapor push the liquid film and cover the top portion of the tube wall as a thin liquid film experiences in the bottom portion of the tube wall. The thin liquid film and more turbulence in the interface of the liquid-vapor are observed in the dimpled helically coiled tube. It is also observed that the dimpled helical coil quickens the transition from intermittent to annular flow compared to that of smooth helical coiled tube. 4.1. Condensation heat transfer coefficient Firstly, as shown in Fig. 8, the experimental heat transfer coefficient of smooth straight tube were compared with the data pre-
Fig. 6. Mass flux (G) vs average vapor quality (x) flow map for (a) Smooth helically coiled tube (SHLCT) (b) Dimpled helically coiled tube (DHLCT).
dicted by the renowned correlation given by Dobson and Chato [3] and Shah [30]. It was apparently seen that the data was well predicted by Dobson and Chato [3], followed by Shah [30] correlation within ±30% error band. Additionally, in Fig. 9, experimental frictional pressure drop data of smooth straight tube was compared with well-known correlation proposed by Friedel [31] and Muller-Steinhagen and Heck [32]. It was found that most of experimental data fall into an error band of ±30%. This suitable agreement of experimental heat transfer coefficients and frictional pressure drop with predicted data certifies the experimental setup. Form Fig. 10, the average heat transfer coefficient data of dimple helically coiled tube was compared to smooth helical coiled and straight tube at saturation temperature of 35 °C with the mass flux of 75, 115, 156 and 191 kg m2 s1. It can be seen that dimple helically coiled tube gives higher heat transfer coefficient compared to smooth helical coiled tube and smooth straight tube in the percentage of 18–32% and 51–61%, respectively, while, the smooth helical coiled tube produces 37–48% higher heat transfer coefficient than that of smooth straight tube at all range of vapor quality. This is because, helically coiled tube yields the centrifugal effect on two-phase flow and produces secondary flow which influences the flow of the vapor and liquid film. As vapor quality increases, the vapor velocity rises, consequently, secondary flow becomes stronger. The vapor flowing from center core of the helically tube pushes the liquid film from the outer to inner tube wall, hence, vapor flows at center core of tube again. These produces a
542
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
(a) Smooth helically coiled tube
x = 0.1028
x = 0.2158
x = 0.3318
(b) Dimpled helically coiled tube
x = 0.1310
x = 0.2728
x = 0.3410
Fig. 7. Flow pattern of R-134a at a mass flux of 156 kg m2 s1 in (a) smooth helically coiled tube and (b) dimpled helically coiled tube.
Fig. 8. Comparison of two-phase heat transfer coefficient data for smooth straight tube with existing correlation.
Fig. 9. Condensation frictional pressure drop data of the smooth round tube compared with existing correlations.
chaotic waves on the liquid film and more liquid entrainment, thereby, heat transfer surface area increases. As a results, the heat transfer mechanism of helically coiled was found to be different
from that of straight tube. The similar behaviors for smooth helically coiled tube were reported by Wongwises et al. [12], previously. It can be also seen that dimpled helically coiled tube gives higher heat transfer coefficient. This is because, dimpled helically coiled tube provides a protrusions inside the tube which provides more heat transfer surface area and turbulence on the fluid flow. Therefore, secondary flow becomes more effective and entrainment of liquid droplets also increased which tends to a lower thermal resistance. As a results, higher heat transfer coefficient ensures for dimpled helical coiled tube. The same phenomenon investigated for dimpled helically straight tube were given by Aroonrat et al. [7]. They found that protrusions inside the tube enhance the heat transfer by proper mixing of fluid flow and increase an amount of turbulence on the two phase flow. In Fig. 11, the effect of mass flux on average heat transfer coefficient of dimpled helically coiled tube is revealed. It can be seen that the mass flux has a remarkable influence on the heat transfer coefficient. This is because, mass flux enhances the flow velocity which increases the shearing effect between the layer of vapor and liquid. Moreover, the experimental results also gives, as the mass flux increases from 75 to 191 kg m2 s1, the heat transfer coefficient increases up to 113–187% in the all range of vapor quality. In Fig. 12, saturation temperature has a small effect on the heat transfer coefficient. It was found that the heat transfer coefficient for saturation temperature of 45 °C at same mass flux lowers than that of saturation temperature of 35 °C. This is because, a higher saturation temperature results lower specific volume, leading to the lower vapor velocity of refrigerant. Another reason is that, the thermal conductivity of refrigerant decreases with an increase in saturation temperature, which in turn higher liquid thermal resistance. Hence, the heat transfer coefficient decreased. Fig. 13 illustrates the effect of mass flux (75, 115, 156 and 191 kg m2 s1) on heat transfer enhancement factor for smooth and dimpled helically coiled tubes, which is defined as the ratio of heat transfer coefficient of helically coiled tube (smooth or dimpled helically coiled tube) to that of smooth straight tube. The enhancement factor for smooth helical coiled tube varies from 1.67 to 1.91 at 75 kg m2 s1, 1.61 to 1.90 at 115 kg m2 s1, 1.66 to 1.77 at 156 kg m2 s1 and 1.62 to 1.72 at 191 kg m2 s1, Similarly, for dimpled helical coiled tube, the enhancement factor ranges from 2.10 to 2.43, 2.07 to 2.36, 2.03 to 2.39 and 1.94 to 2.13 for the mass flux of 75, 115, 156 and 191 kg m2 s1, respectively. Moreover, Table 4 reveals that the effect of mass flux and saturation temperature on the enhancement factor. The mass flux and saturation temperature have fewer effect on the enhancement factor. It can
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
543
Fig. 10. Comparison of the heat transfer coefficient for tubes at different mass flux.
be seen that enhancement factor declines with the rise of both the mass flux and saturation temperature. Fig. 14 presents the comparison of the experimental Nusselt number for smooth helically coiled tube with renowned correlations given by the Wongwise and Polsongkram [12] and Gupta et al. [20]. The correlations for smooth helical coiled tube are listed in the Table 5. As shown in Fig. 14, these correlations well predict the experimental data for smooth helically coiled tube. It is revealed that Wongwise and Polsongkram [12] overestimated the experimental results, however, the data exists within an error band of ±30%. At present, there is no correlation for prediction of the two phase flow Nusselt number in a dimpled helically coiled tube. In order to develop the correlation, the effect of equivalent Dean number, DeEq, Prandtl number, Prl, Martinelli parameter,vtt and reduced pressure, pr are considered. The new developed correlation to predict the two phase flow Nusselt number in a dimpled helically coiled tube is:
ðNutp ÞDT ¼ 0:0208ðDeEq Þ1:05 ðPr l Þ0:65 ðvtt Þ0:113 ðpr Þ0:169 Fig. 11. Effect of mass flux on condensation heat transfer coefficients in the dimpled helically tube.
where an equivalent Dean number, DeEq is given by equation:
ð17Þ
544
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
Prl ¼
Cpl ll kl
ð19Þ
Reduced pressure, pr is evaluated from
pr ¼
ps pcrit
ð20Þ
The proposed correlation gives prediction of Nusselt number in the range of non-dimensional numbers as follows: 1564 DeEq 112,666; 3.19 Pr l 3.29; 0.062 vtt 2.01; 0.21 pr 0.28. In Fig. 15, the experimental two phase Nusselt number were compared with proposed correlation. As shown in Fig. 14, the correlation is well predicted the experimental data within the error bar ±20%. 4.2. Frictional pressure drop
Fig. 12. Effect of saturation temperature on condensation heat transfer coefficient for tubes.
Fig. 13. Enhancement factor as a function of mass flux at Tsat = 35 °C.
(
l DeEq ¼ Rel þ Rev v ll
ql qv
0:5 )
di 2R
0:5
Prandtl number, Prl, is determined from
ð18Þ
From Fig. 16, the frictional pressure drop of dimpled helically coiled tube was compared to smooth helically coiled and straight tube at saturation temperature of 35 °C with the mass flux of 75 and 115 kg m2 s1. It can be seen that dimpled helically coiled tube offers higher frictional pressure drop compared to smooth helically coiled and smooth straight tube in the percentage of 47–61% and 61–75%, respectively, while, the smooth helically coiled tube gives 25–43% higher frictional pressure drop than that of smooth straight tube at all range of vapor quality. This is because, vapor velocity increases with rise of vapor quality and mass flux and enhances the shear stress between vapor-liquid interfaces, hence, secondary flow inside the helically coiled tube turn out to be more effective which cause the higher pressure drop inside helically coiled tube compared to that of straight tube. From Fig. 16, it is also seen that dimpled helically coiled tube produces a higher frictional pressure drop compared to that of both, smooth helically coiled tube and straight tube. The pressure drop growth was yielded by the protrusions provided inside the dimpled helically coiled tube which rises the turbulence level on the twophase flow, and thereby, secondary flow becomes stronger compared to that of smooth helically coiled tube. Consequently, higher frictional pressure drop occurs in the dimpled helically coiled tube. In Fig. 17, frictional pressure drop data of dimpled helically coiled tube data are revealed. It can be seen that the frictional pressure drop increases when mass flux increases. This is because, mass flux of refrigerant rises the flow velocity, which causes higher interfacial shear stress on the liquid film, hence, the frictional pressure drop raised. The experimental results gives, as the mass flux increases from 75 to 191 kg m2 s1, the frictional pressure drop increases up to 261–310% in the all range of vapor quality. Fig. 18 presents the effect of saturation temperature on frictional pressure gradient inside the tubes at mass flux of 156 kg m2 s1. It can be seen that the frictional pressure drop decreases with the rise of saturation temperature of refrigerant. This is because, at high saturation temperature, the specific volume of vapor decreases, hence the vapor velocity falls, resulting lower interfacial shear stress between the vapor and liquid film. Consequently, lower value of frictional pressure drop takes place at a lower saturation temperature. Fig. 19 illustrates the effect of mass flux (75, 115, 156 and 191 kg m2 s1) on pressure drop penalty factor for smooth and dimpled helically coiled tubes, which is defined as the ratio of frictional pressure drop of helical coiled tube (smooth or dimpled helical coiled tube) to that of smooth straight tube. The penalty factor for smooth helically coiled tube varies from 1.35 to 1.75 at 75 kg m2 s1, 1.38 to 1.70 at 115 kg m2 s1, 1.44 to 1.75 at 156 kg m2 s1 and 1.39 to 1.74 at 191 kg m2 s1, Similarly, for the dimpled helically coiled tube, the penalty factor ranges from 2.57 to 3.92, 2.76 to 4.03, 2.99 to 4.60 and 2.92 to 4.01 for the mass flux
545
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548 Table 4 Heat transfer enhancement factor (EF) and pressure drop penalty factor (PF) of the smooth and dimpled helically coiled tube. Tube
T (°C)
G (kg m2 s1)
x
EF
PF
Smooth helical coiled tube
35
75 115 156 191
0.23 0.25 0.21 0.23
1.78 1.75 1.74 1.70
1.47 1.48 1.54 1.51
45
75 115 156 191
0.24 0.22 0.20 0.25
1.77 1.71 1.72 1.66
1.50 1.49 1.66 1.62
35
75 115 156 191
0.20 0.24 0.19 0.22
2.26 2.16 2.14 2.08
3.16 3.23 3.35 3.53
45
75 115 156 191
0.23 0.26 0.23 0.20
2.18 2.12 2.06 1.92
3.20 3.25 3.43 3.41
Dimpled helical coiled tube
little effect of the penalty factor. It can be seen that penalty factor increases with the rise of both the mass flux and saturation temperature. In Fig. 20, the experimental liquid two- phase frictional pressure drop multiplier value is compared to well-known correlations proposed by the Wongwises and Polsongkram [12] and Gupta et al. [20]. The correlations for smooth helically coiled tube are listed in the Table 5. As shown in Fig. 20, these correlation well predicted the experimental data for the smooth helically coiled tube. At present, there is no correlation for predicting the two phase flow frictional pressure drop in a dimpled helical coiled tube. In order to develop the correlation, a two-phase frictional multiplier is used to correlate the frictional pressure drop of two-phase flow. A newly correlation is developed which is given by following Eq. (21),
DPtp;f 4:8671 1 ¼ ð/2l ÞDT ¼ 4:5033 1 þ 1:4919 þ 2 p0:321 DPl;f DT vtt vtt r Fig. 14. Comparison of experimental Nusselt number for smooth helically tube with existing correlation.
of 75, 115, 156 and 191 kg m2 s1, respectively. Moreover, Table 4 reveals that the effect of mass flux and saturation temperature on the penalty factor. The mass flux and saturation temperature have
ð21Þ
where pr and vtt are the reduced pressure of the refrigerant, defined in Eq. (19) and Martinelli parameter, respectively. The proposed correlation gives prediction of two-phase frictional multiplier in the range of non-dimensional numbers as follows: 0.062 vtt 2.01; 0.21 pr 0.28. In Fig. 21, the experimental two phase frictional pressure drop were compared with proposed correlation. As shown in Fig. 21, the experimental frictional pressure drop for dim-
Table 5 Wongwise and Polsongkram [12] and Gupta et al. [20] correlations for smooth helically coiled tube. Correlation Wongwise and Polsongkram [12]
Two phase flow Nusselt number (Nutp) ðNutp Þ ¼ 0:1352ðDeEq Þ0:7654 ðPr l Þ0:8144 ðvtt Þ0:0432 ðpr Þ0:3356 ðBo 104 Þ 0:5 0:5 C l di DeEq ¼ Rel þ Rev llv qql , Pr l ¼ plk l , pr ¼ pps 2R l
v
l
0:112
crit
q i i Rel ¼ Gð1xÞd , Rev ¼ Gxd l l , Bo ¼ Gilv ; v
l
The two-phase frictional pressure gradient multiplier ð/2l Þ 1 ð/2l Þ ¼ 1 þ v5:569 1:496 þ v2 tt
Gupta et al. [20]
tt
Two phase flow Nusselt number (Nutp) 0:63 ¼ 0:4 1 þ 3:024 v0:8 pr
Nutp Nul
tt
Nul is the single phase Nusselt number which is proposed by Mori and Nakayama [33] (
16 ) 1=12 2:5 5=6 di 1 Rel Pr0:4 1 þ 0:061 Rel dDi Nul ¼ 41 l D The two-phase frictional pressure gradient multiplier ð/2l Þ 1 ð/2l Þ ¼ 2:76 1 þ v7:094 p0:7 1:378 þ v2 r tt
tt
546
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
Fig. 15. Experimental Nusselt number of dimpled helically coiled tube compared with proposed correlation.
Fig. 17. Effect of mass flux on frictional pressure drop in the dimpled helically tube.
Fig. 18. Effect of saturation temperature on frictional pressure drop for tubes.
Fig. 16. Comparison of the frictional pressure drop for tubes at mass flux of 75 & 115 kg m2 s1.
Fig. 19. Penalty factor as function of mass flux at Tsat of 35 °C.
pled helical coiled tube is well predicted by new proposed correlation, almost, within an error band of ±20%. Furthermore, the maximum heat transfer rate is concern, it is always preferable to provide the more surface area and the dim-
pled helically coiled tube absolutely fulfill the same. The design of a heat exchanger comprising dimpled helically coiled tube in context to condensation need a prediction tool. Therefore, a non-dimensional correlations predicting the Nusselt number and
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
2.
3.
4.
Fig. 20. Comparison of experimental liquid two-phase frictional multiplier for smooth helically tube with Wongwises and Polsongkram [12] and Abhinav Gupta et al. [20] corrrelations.
5.
547
effect on the heat transfer coefficient. On the other hand side, the frictional pressure drop for dimpled helically coiled tube also increases with the rise of mass flux and vapor quality, but, it decreases with the increase of saturation temperature. In a dimpled helically coiled tube, the transition from intermittent to annular flow occurred at lower vapor quality (x = 0.27) than the smooth helical coiled tube (x = 0.31–0.33). Dimple helically coiled tube gives a higher heat transfer coefficient compared to smooth helically coiled tube and smooth straight tube in the percentage of 18–32% and 51–61%, respectively, while, the smooth helically coiled tube produces 37–48% higher heat transfer coefficient than that of smooth straight tube at all range of vapor quality. The frictional pressure drops of dimpled helically coiled tube were about 2.57–4.01 times higher than those of smooth straight tube, while, for smooth helically coiled tube, the frictional pressure drops were almost 1.35–1.75 times higher than those of smooth straight tube. Experimental Nusselt number and frictional pressure drop of dimpled helically coiled tube are not adequately predicted by correlation, hence, a newly correlation is developed which predicts the experimental results of dimpled helical coiled tube in the mass flux range of 75–191 kg m2 s1.
Appendix A. Supplementary material Supplementary data associated with this article can be found, in the online version, at https://doi.org/10.1016/j.applthermaleng. 2017.10.026. References
Fig. 21. Comparison of the two-phase frictional pressure drop data for dimpled helically tube by proposed correlation.
frictional pressure drop, were developed and best suitable to small scale heat exchanger. Since the effect of dimples was not considered while developing the proposed correlation, the applicability is limited to specified geometry i.e., pitch and depth of the dimple. However, in order to generalize the correlation in context to effect of dimples, further research would be done i.e. multiple tubes of varying pitch and depth of dimple. 5. Conclusion In this experiment, the effect of mass flux (75, 115, 156 and 191 kg m2 s1), vapor quality in the range of 0.1–0.8 and saturation temperature (35 & 45 °C) on the average heat transfer coefficient and average frictional pressure drop inside a smooth and dimpled helically coiled tube have been examined. The results obtained from smooth and dimpled helically coiled tube were also compared with results from those smooth straight tube. The following results are obtained. 1. The average heat transfer coefficient for dimpled helically coiled tube increases, as mass flux or vapor quality of refrigerant increases, while, the saturation temperature has a small
[1] D.A. Reay, Heat transfer enhancement-a review of techniques and their possible impact on energy efficiency in the U.K., Heat Recover. Syst. CHP. 11 (1991) 1–40. [2] J.G. Collier, J.R. Thome, in: Convective Boiling and Condensation, third ed., Oxford Univ. Press, Berlin, 1994, pp. 99–280. [3] M.K. Dobson, J.C. Chato, Condensation in smooth horizontal tubes, J. Heat Transf. 120 (1998) 193. [4] A. Cavallini, G. Censi, D. Del Col, L. Doretti, G.A. Longo, L. Rossetto, Experimental investigation on condensation heat transfer and pressure drop of new HFC refrigerants (R134a, R125, R32, R410A, R236ea) in a horizontal smooth tube, Int. J. Refrig. 24 (2001) 73–87. [5] D. Jung, K.H. Song, Y. Cho, S.J. Kim, Flow condensation heat transfer coefficients of pure refrigerants, Int. J. Refrig. 26 (2003) 4–11. [6] M. Anowar Hossain, Y. Onaka, H.M.M. Afroz, A. Miyara, Heat transfer during evaporation of R1234ze(E), R32, R410A and a mixture of R1234ze(E) and R32 inside a horizontal smooth tube, Int. J. Refrig. 36 (2013) 465–477. [7] K. Aroonrat, S. Wongwises, Experimental study on two-phase condensation heat transfer and pressure drop of R-134a flowing in a dimpled tube, Int. J. Heat Mass Transf. 106 (2017) 437–448. [8] A. Sarmadian, M. Shafaee, H. Mashouf, S.G. Mohseni, Condensation heat transfer and pressure drop characteristics of R-600a in horizontal smooth and helically dimpled tubes, Exp. Therm. Fluid Sci. 86 (2017) 54–62. [9] H.J. Kang, C.X. Lin, M.A. Ebadian, Condensation of R134a flowing inside helicoidal pipe, Int. J. Heat Mass Transf. 43 (2000) 2553–2564. [10] B. Yu, J.T. Han, H.J. Kang, C.X. Lin, A. Awwad, M.A. Ebadian, Condensation heat transfer of R-134A flow inside helical pipes at different orientations, Int. Commun. Heat Mass Transf. 30 (2003) 745–754. [11] J.T. Han, C.X. Lin, M.A. Ebadian, Condensation heat transfer and pressure drop characteristics of R-134a in an annular helical pipe, Int. Commun. Heat Mass Transf. 32 (2005) 1307–1316. [12] S. Wongwises, M. Polsongkram, Condensation heat transfer and pressure drop of HFC-134a in a helically coiled concentric tube-in-tube heat exchanger, Int. J. Heat Mass Transf. 49 (2006) 4386–4398. [13] C.X. Lin, M.A. Ebadian, Condensation heat transfer and pressure drop of R134a in annular helicoidal pipe at different orientations, Int. J. Heat Mass Transf. 50 (2007) 4256–4264. [14] L. Shao, J.Tian Han, G.Ping Su, J.Hong Pan, Condensation heat transfer of R134A in horizontal straight and helically coiled tube-in-tube heat exchangers, J. Hydrodyn. 19 (2007) 677–682. [15] M.H. Al-Hajeri, A.M. Koluib, M. Mosaad, S. Al-Kulaib, Heat transfer performance during condensation of R-134a inside helicoidal tubes, Energy Convers. Manage. 48 (2007) 2309–2315.
548
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
[16] M. El-Sayed Mosaad, M. Al-Hajeri, R. Al-Ajmi, A.M. Koliub, Heat transfer and pressure drop of R-134a condensation in a coiled, double tube, Heat Mass Transf. 45 (2009) 1107–1115. [17] Y. Murai, S. Yoshikawa, S.I. Toda, M.A. Ishikawa, F. Yamamoto, Structure of airwater two-phase flow in helically coiled tubes, Nucl. Eng. Des. 236 (2006) 94– 106. [18] S.B. Genic´, B.M. Jac´imovic´, M.S. Jaric´, N.J. Budimir, M.M. Dobrnjac, Research on the shell-side thermal performances of heat exchangers with helical tube coils, Int. J. Heat Mass Transf. 55 (2012) 4295–4300. [19] S.B. Genic´, B.M. Jac´imovic´, M.S. Jaric´, N.J. Budimir, Analysis of fouling factor in district heating heat exchangers with parallel helical tube coils, Int. J. Heat Mass Transf. 57 (2013) 9–15. [20] A. Gupta, R. Kumar, A. Gupta, Condensation of R-134a inside a helically coiled tube-in-shell heat exchanger, Exp. Therm. Fluid Sci. 54 (2014) 279–289. [21] M. Mozafari, M.A. Akhavan-Behabadi, H. Qobadi-Arfaee, M. Fakoor-Pakdaman, Condensation and pressure drop characteristics of R600a in a helical tube-intube heat exchanger at different inclination angles, Appl. Therm. Eng. 90 (2015) 571–578. [22] M.R. Salimpour, A. Shahmoradi, D. Khoeini, Experimental study of condensation heat transfer of R-404A in helically coiled tubes, Int. J. Refrig. 74 (2017) 584–591. [23] S.J. Kline, F.A. McClintock, Describing uncertainties in single-sample experiments, Mech. Eng. 75 (1953) 3–8. [24] E.W.W. Lemmon, M.L.L. Huber, M.O.O. McLinden, NIST Standard Reference Database 23: Reference Fluid Thermodynamic and Transport Properties (REFPROP), Version 9.0, Phys.
[25] A. Abdul-Razzak, M. Shoukri, J.S. Chang, Characteristics of refrigerant R-134A liquid-vapor two-phase flow in a horizontal pipe, ASHRAE Trans. (1995). [26] P.S. Srinivasan, S.S. Nandapurkar, F.A. Holland, Friction factor for coils, Trans. Inst. Chem. Eng. 48 (1970) T156–T161. [27] Y. Taitel, A.E. Dukler, A model for predicting flow regime transitions in horizontal and near horizontal gas-liquid flow, AIChE J. 22 (1976) 47–55. [28] Z. Yang, M. Gong, G. Chen, X. Zou, J. Shen, Two-phase flow patterns, heat transfer and pressure drop characteristics of R600a during flow boiling inside a horizontal tube, Appl. Therm. Eng. 120 (2017) 654–671. [29] W. Cui, L. Li, M. Xin, T.C. Jen, Q. Liao, Q. Chen, An experimental study of flow pattern and pressure drop for flow boiling inside microfinned helically coiled tube, Int. J. Heat Mass Transf. 51 (2008) 169–175. [30] M.M. Shah, An Improved and Extended General Correlation for Heat Transfer During Condensation, HVAC&R Res. 15 (2009) 37–41. [31] L. Friedel, Improved friction pressure drop correlations for horizontal and vertical two phase pipe flow, in: Eur. Two-Phase Flow Gr. Meet., 1979, pp. 485–491. [32] H. Müller-Steinhagen, K. Heck, H. Müller-Steinhagen, K. Heck, A simple pressure drop correlation for two-phase flow in pipes, Chem. Eng. Prog. 20 (1986) 297–308. [33] Y. Mori, W. Nakayawa, Study on forced convective heat transfer in curved pipes: 3rd report, theoretical analysis of uniform wall temperature and practical formulae, Trans. Jpn. Soc. Mech. Eng. 33 (1967) 104–112.
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Condensation of R-134a inside dimpled helically coiled tube-in-shell type heat exchanger Anand Kumar Solanki ⇑, Ravi Kumar Department of Mechanical and Industrial Engineering, Indian Institute of Technology, Roorkee 247667, India
h i g h l i g h t s Dimple helically coiled tube gives a higher heat transfer coefficient. The effect of mass flux, vapor quality and saturation temperature are considered. Flow transitions for dimpled helically coiled tube occurred at lower vapor quality. Dimple helically coiled tube yields a higher pressure drop than straight tube.
a r t i c l e
i n f o
Article history: Received 1 July 2017 Revised 29 September 2017 Accepted 4 October 2017 Available online 12 October 2017 Keywords: Condensation Heat transfer coefficient Frictional pressure drop R-134a Dimpled helically coiled tube
a b s t r a c t In this study, condensation heat transfer coefficients and frictional pressure drops of R-134a inside a dimpled helically coiled tube are experimentally investigated. The inner tubes comprise of one smooth straight tube, one smooth helically coiled tube and one dimpled helically coiled tube. The experimental measurements are carried out at saturation temperature of 35°, and 45 °C with mass flux of 75, 115, 156 and 191 kg m2 s1. The experimental data of a smooth and dimpled helically coiled tube have been plotted on the mass flux versus vapor quality flow map and Tailtel and Dukler flow map. The transitions between different flow regimes have also been discussed. Moreover, the effect of mass flux, vapor quality and saturation temperature of R-134a on the heat transfer coefficients and pressure drops are examined. Comparisons between smooth straight tube, smooth helically coiled tube and dimpled helically coiled tube are also discussed. The dimple helically coiled tube produces a higher heat transfer coefficient and frictional pressure drop compared to smooth helically coiled tube and smooth straight tube. The correlations have been proposed to predict the Nusselt number and frictional pressure drop multiplier during condensation of R-134a inside horizontal dimpled helically coil tube. Ó 2017 Elsevier Ltd. All rights reserved.
1. Introduction The condenser is an essential part of any refrigeration and air conditioning unit which condenses a substance from its gaseous to liquid state by removing the latent heat. Nowadays, the effective design of the condenser is the important for the industries to obtain a maximum heat transfer with a smaller amount of pressure drop. As such as, helically coiled tube have been widely used to enhance the heat transfer by producing centrifugal effect on the flow inside the tube [1]. Nevertheless, A number of literature has been studied on two phase flow in smooth round tube [2]. Dobson and Chato [3] reported the experimental heat transfer coefficient data with refrigerants R22, R134a, R410A, R32/R125 (60/40% by
⇑ Corresponding author. E-mail address: [email protected] (A.K. Solanki). https://doi.org/10.1016/j.applthermaleng.2017.10.026 1359-4311/Ó 2017 Elsevier Ltd. All rights reserved.
mass) inside smooth round horizontal tubes with inner diameters ranging from 3.14 to 7.04 mm. They also proposed the correlation to predict the heat transfer coefficient for the stratified and annular flow pattern. Cavallini et al. [4] experimentally measured the heat transfer coefficient and pressure drop of pure HFC refrigerants (R134a, R125, R236ea, R32) and the nearly azeotropic HFC refrigerant blend R410A inside a smooth tube of diameter 8 mm at saturation temperature in the range between 30 °C and 50 °C with mass velocities varying from 100 to 750 kg m2 s1. The results show that, during condensation of pure fluids and nearly azeotropic mixtures, in the annular flow regime, the heat transfer coefficient increases with rise of mass velocity and vapor quality. Jung et al. [5] carried out the experimental heat transfer coefficient data with refrigerants R12, R22, R32, R123, R125, R134a, and R142b in a horizontal plain copper tube of 9.52 mm outside diameter and 1 m length at a fixed refrigerant saturation temperature of 40 °C with mass fluxes of 100, 200, 300 kg m2 s1 and heat flux of
536
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
Nomenclature A De Cp Re d EF D PF g
heat transfer area, m2 Dean number specific heat at constant pressure, J kg1 K1 Reynolds number tube diameter, mm enhancement factor coil mean diameter, mm penalty factor gravity, 9.81 m2 s1
Greek symbols e depth of dimple, mm a void fraction o diameter of dimple, mm v Martinelli parameter s pitch of dimple, mm h helix angle, ° p helical pitch, mm q density, kg m3 G mass velocity, kg m2 s1 l dynamic viscosity, kg m1 s1 h heat transfer coefficient, W/m2 K / two-phase multiplier h enthalpy, J/kg D difference l length, m Subscripts Nu Nusselt number a acceleration
7.3–7.7 kW m2. They found that the heat transfer coefficient of refrigerant R142b and R32 were higher than those of R22 by 8– 34% at the same mass flux while heat transfer coefficient of refrigerant of refrigerant R134a and R123 were similar to those of R22. Hossain et al. [6] conducted the experiments with refrigerants R1234ze(E), R32, R410A and a zeotropic mixture R1234ze(E)/R32 (55/45 mass%) inside a water heated double tube heat exchanger. They investigated that the experimental heat transfer coefficient of R1234ze (E) is lower than R1234ze (E)/R32 (55/45 mass %), R410A and R32 by 11%, 56% and 83%, respectively. Moreover, a few paper are also reported on the two phase-flow of refrigerant inside the dimpled straight tube during condensation process. Aroonrat and Wongwise [7] examined the condensation heat transfer coefficient and frictional pressure drop of R-134a inside the smooth and spherical dimpled straight tube at saturation temperature of 40°, 45° and 50 °C with the mass flow rate of 300, 400 and 500 kg m2 s1. They found that the heat transfer coefficient ratio for spherical dimpled straight tube varies from 1.27 to 1.37 at 300 kg m2 s1, 1.28 to 1.36 at 400 kg m2 s1 and 1.26 to 1.41 at 500 kg m2 s1, respectively, while, the two-phase friction factor ratio varies from 2.8 to 3.7 at 300 kg m2 s1, 2.7 to 4.2 at 400 kg m2 s1 and 2.6 to 4.1 at 500 kg m2 s1, respectively. Sarmadian et al. [8] investigated the convective condensation heat transfer and frictional pressure drops of R-600a inside a helically dimpled tube and a plain tube of internal diameter 8.3 mm at saturation temperature of 38 and 48 °C with mass flux varying from 114 to 368 kg m2 s1. They noticed that the heat transfer coefficients of the dimpled tube were 1.2–2 times of those in smooth tube with a pressure drop penalty ranging between 58% and 195%. Besides, a few research work have been performed on the heat transfer and pressure drop characteristics inside helical coiled
K G x f Q l T g ph Eq P exp _ m fg Fr v sat R i w o ts w tt out W in Prl T
thermal conductivity, W m1 K1 gravitational vapor quality frictional heat transfer rate, W liquid temperature, °C gas preheater equivalent pressure, kPa experimental mass flow rate, kg/s latent heat of vaporization vapor Froud number saturation condition refrigerant inner wall outer test section water turbulent–turbulent outlet water inlet Prandtl number total
tube. Kang [9] carried out the experimental study of heat transfers and pressure drops data of R-134a inside long helical coil tube. The inner diameter, outer diameter, coil diameter and number of turns of the tube were taken 12.7 mm, 21.2 mm, 177.8 mm, 34.8 mm and 10, respectively. The test runs were executed at different refrigerant mass flux varying from 100 to 400 kg m2 s1, cooling water flow Reynolds number range of 1500–9000, cooling tube wall temperature at 12 °C and 22 °C and fixed saturation temperature of 33 °C. The results revealed that the overall heat transfer coefficient raised with increase of the water mass flux. The heat transfer and pressure drop decreased with increase of tube wall temperature from 12 °C and 22 °C. Yu et al. [10] conducted an experimental investigation of heat transfer of R-134a during condensation inside horizontal, inclined and vertical helical tube at refrigerant mass flux from 100 to 400 kg m2 s1, cooling water Reynolds number between 1500 and 10000. They investigated that the helical tube orientation had significant influence on the overall heat transfer coefficients. The highest and lowest heat transfer coefficient was observed at inclined position and vertical position, respectively. Han et al. [11] reported the experimental results of condensation heat transfer and pressure drop of R-134a inside helical tube with mass flux vary from 100 to 420 kg m2 s1 at a different condensation saturation temperature 35, 40 and 46 °C. They observed that the refrigerant mass flux and the saturation temperature had a significant influence on the heat transfer coefficient. Wongwises and Polsongkram [12] presented experimental studies on two-phase flow condensation heat transfer and pressure drop of R-134a inside concentric helical coil tube in tube condenser. The experimental measurement were performed at different saturation temperature 40 and 50 °C with refrigerant mass flux between 400 and 800 kg m2 s1. The results showed that the heat transfer coefficient of the helical coiled tube were about 33–53%
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
higher than that of smooth round tube while pressure drop were about 29–46% higher. They also proposed the correlation for predicting the heat transfer coefficient and pressure drop. Lin and Ebadian [13] reported the effect of inclination angle (0°, 45° and 90°) on condensation heat transfers and pressure drops of R-134a flowing inside the annular section of tube. The experimental measurement was carried out at saturation temperature of 30 and 35 °C with refrigerant mass flux from 60 to 200 kg m2 s1, cooling water Reynolds number range between 3600 and 22,000 at cooling water temperature of 16, 20 and 24 °C. The Nusselt number rises with the increase of both refrigerant mass flow rate and cooling water mass flow rate. They also found that the Nusselt number of the refrigerant-side for angle 0–45° was about two times more than that of angle 0–90°. Shao et al. [14] compared the condensation heat transfer coefficient for the helical tube with straight tube at different saturation temperature 35, 40 and 45 °C, mass flux vary from 100 to 400 kg m2 s1. The heat transfer coefficients of helical tube were found to 4–13.8% higher than that of straight tube. AlHajeri et al. [15] examined the performance of heat transfers and pressure drops of R-134a inside annular helical tubes at different saturation temperature of 36, 42 and 48 °C with the mass flux varying from 50 to 680 kg m2 s1. The results showed that the heat transfer coefficients of the refrigerant-side rise with increase of the refrigerant mass flux and decline with increase of the saturation temperature of refrigerant. However, with rise of the mass flux, the pressure drops of refrigerant increased. Mosaad et al. [16] investigated the effect of refrigerant mass flux and condensation temperature difference on the heat transfer coefficients and pressure drops of R-134a inside coiled double tube. They found that, with increase of refrigerant mass flux and decrease of condensation temperature difference, the overall heat transfer coefficient increased. Similarly, pressure drop of refrigerant increases with rise of the refrigerant mass flux. Murai et al. [17] used a highspeed video system to take photographs of interfacial structure of gas-liquid flow inside the helical coiled tube. They observed that the flow transition from bubbly to plug flow for helical coiled tube compared to straight tube was obtained earlier duo to the centrifugal acceleration generating by curvature of the tube. Genic´ et al. [18] examined the shell-side heat transfer coefficient pertaining to three heat exchangers with helical coils. The geometric parameter such as winding angle, pitch and axial pitch had significant influence on shell-side heat transfer coefficient. They also proposed the correlation for shell-side heat transfer coefficient based on the shell-side hydraulic diameter. Genic´ et al. [19] also investigated experimentally the shell-side fouling factors of eight heat exchangers with parallel helical coils. The fouling resistances were slightly lesser than usual values for shell-and-tube heat exchangers with straight pipes. Gupta et al. [20] determined the effect of mass flux, vapor quality and saturation temperature on the heat transfer coefficients and pressure drop of R-134a inside horizontal helical coiled tube-in-shell type heat exchanger. The heat transfer coefficient and pressure drop of refrigerant increase with rise of mass flux and vapor quality and decrease with increase of the saturation temperature. M. Mozafari et al. [21] conducted the experimental studies of heat transfer and pressure drop of R600a inside helical tube-intube heat exchanger at different inclination angles of zero, +30°, +60° and +90° with different mass flux ranging from 155 to 265.5 kg m2 s1 at saturation temperature 38.5–47 °C. The results showed that the heat transfer coefficients were remarkably influenced by inclination angle but pressure drop were slightly affected. Maximum heat transfer coefficients and pressure drops were obtained at +30° and 0° angles, respectively. The effect of different curvature radii (43.5, 63.5 and 76.5 mm) and coil pitch (15, 25 and 35 mm) of helical coil tube-in-shell on the heat transfer coefficients of R-404A were investigated by Salimpour et al. [22]. Both curvature radii and coil pitch of helical coil tube had significant
537
influence on the heat transfer coefficient. They also found that, the effect of coil pitch on heat transfer coefficient was more pronounced at low vapor qualities. However, to the best our knowledge from the above mentioned literature, most of the research works have been done on both single and two phase flow inside smooth helically coiled tube. Besides, some experimental investigation are currently available on the condensation heat transfer and pressure drop of refrigerant inside dimpled straight tube. But, any research work inside the dimpled helically coiled tube-in-shell type heat exchanger during condensation of R134a has not been reported. In this study, the condensation heat transfer coefficients and frictional pressure drops of R134a inside dimpled helically coiled tube-in-shell with respect to vapor quality (0.1–0.84), mass flux (75, 115, 156 and 191 kg m2 s1) and saturation temperature (35 and 45 °C) have been examined. In addition, results obtained from dimpled helical coiled tube are compared with those for the straight tube and smooth helical coiled tube.
2. Experimental apparatus and procedures The simplified schematic diagram of the experimental set-up was depicted in Fig. 1. The main component of experimental setup was composed of a pre-heater, test-condenser, postcondenser, three refrigerant magnetic gear pumps, coriolis mass flow meter and data acquisition system. The test section was a counter flow type heat exchanger. The test section was connected to post condenser, where two phase quality of refrigerant coming from test section was sub-cooled to a liquid state by extracting heat to cold water circulated from the condensing unit. The municipal water was employed for cooling purpose which was circulated inside the shell of the test-section. The flow rate of the cooling water was regulated by a manually valve. The temperature of water inside the test-section was controlled by variac of 2 kW capacity. A slight glass after the post-condenser was placed to ensure the continuous liquidity of refrigerant before entering the three magnetic gear pump connected side by side. The flow rate of refrigerant was monitored by operating the speed of the magnetic gear pump through frequency inverter. The coriolis mass flow meter followed by magnetic gear pump was accommodated to measure the refrigerant flow rate. In order to control the vapor quality at the inlet of the test condenser, an appropriate preheater was designed from a 6 m long U-bend stainless steel tube. By supplying high current and low voltage through step down auto transformer, liquid refrigerant was evaporated by heating the tube. A standard filter-drier was incorporated following the pre-heater to remove any foreign particles and moisture in refrigerant loop. A number of hand shut-off valves were facilitated to provide facile installation of a component in the apparatus. Fig. 2 depicts a schematic diagram of the test section. The test section was a counter flow helical coiled tube in shell type heat exchanger in which the refrigerant flowing inside the inner tube extracts the heat to the cooling water flowing the opposite direction to refrigerant flow inside the shell. There were three type of tubes taken as inner tube, one smooth straight tube, one smooth helical coiled tube and another one dimple helical coiled tube. The inner helically coiled tube, which was made by wrapping the straight tube over the wooden pattern. As a next step, the helically coiled tube was processed to make the projections on its surface by using a round tip tool and thus a dimple helically coiled tube was obtained. The shape of dimples in the inner surface of the helical coiled tube were spherical. The Schematic diagram of the inner tubes are shown in Fig. 3. Details of the inner tubes are given in Table 1. Two quartz sight glass were mounted at the inlet and outlet of the test section to visualize the flow patterns. The pho-
538
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
Fig. 1. Schematic diagram of experimental set-ups.
Water outlet
25mm
120 mm 100 mm 140 mm Refrigerant out
Refrigerant in
200 mm Water inlet Fig. 2. Schematic diagram of test-section.
tograph of the visualization section can be seen in Fig. 4. The inside diameter of the sight glass is identical to that of the test tube. The photographs of the flow patterns at the downstream of the test section i.e. sight glass, are captured by using high speed camera at 1000 frame per second. The sight glass with negligible surface roughness and shorter length results minor disruption in the downstream flow pattern. Hence, the redevelopment of a flow pattern in a sight glass is insignificant. The T-type thermocouples were located to measure the refrigerant temperature at the inlet of the pre-heater, inlet and outlet of the test-section. The outer wall temperature of the helical coiled tube was also measured with T-type thermocouple at six positions. In each position, four thermocouple were placed 90° apart over the top, bottom and two
sides. A special glue was used to fix thermocouples on the wall to avoid water resistance. In the present study, Presys T-25 N calibrator having temperature range 25 °C to 125 °C was used to calibrate the T-type thermocouples. The thermocouples were calibrated in the operating temperature range of 10 °C to 70 °C with an accuracy of ±0.1 °C of full scale. The pressure Transducers with accuracy 0.25% were utilized to measure the refrigerant pressure at the inlet and outlet of the test-section and at the entrance of the pre-heater. Also Pressure drop between the test-section was measured by a differential pressure drop transducer. The coriolis flow meter with an accuracy of 0.1% was employed to measure the refrigerant mass flow rate. The water flow rate in annulus tube was measured by a turbine
539
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
(a)
s θ
(b) o
p e
(c) Fig. 3. Sketch diagram of the tubes: (a) Smooth straight tube, (b) Smooth helically coiled tube and (c) Dimpled helically coiled tube.
Table 1 Dimension of inner tubes. Parameter
Smooth straight tube
Smooth helically coiled tube
Dimpled helically coiled tube
Inner diameter of tube, mm Outer diameter of tube, mm Length of tube, mm Inside tube area, mm2 Coil mean diameter, mm Coil pitch, mm Number of turns Dimple pitch (s), mm Helical pitch (p), mm Dimpled depth (e), mm Helix angle (h), ° Dimpled diameter (o), mm
8.92 9.52 1000 26,156 – – – – – – – –
8.92 9.52 2200 57,543 110 25 6 – – – – –
8.92 9.52 2200 61,041 110 25 6 8.5 10 1 49 2
100 mm
A quartz glass tube
in Table 2. The uncertainties of experimental results are evaluated by methodology which is suggested by Klein and McClintock [23] as shown in Table 3. 3. Data reduction The data reduction analysis is employed to evaluate the average heat transfer coefficient and the average frictional gradient during an experiment at steady state conditions. For calculating the thermo-physical properties of R-134a, REFPROF-9 [24] is utilized. The average heat transfer coefficient (havg) is obtained by the following equation:
hav g ¼
Qw Ai ðT sat T wi Þ
ð1Þ
where Qw, Ai and Twi are the heat transfer rate inside the test section from tube to flowing water, inner surface area on tube and average inner wall surface temperature of the tube, respectively. Tsat is the saturation temperature of refrigerant corresponding to the saturation pressure which is calculated by the inlet pressure of the test section and the total pressure drop along the test section. For pressure drop, a linear profile is assumed along the test section. The average inner wall surface temperature of the tube is given by
T wi ¼ T wo þ Q w ln
Do 2pKL Di
ð2Þ
where Two, Do, Di, K and L are the average temperature of outside tube wall, outer diameter of the tube, inner diameter of the tube, thermal conductivity of the copper and length of the tube, respectively. Fig. 4. Visualization section.
flow meter. The turbine flow meter has an accuracy of 1% of the full scale. All thermocouple and pressure transducer signal was transferred to the multichannel data acquisition system with PXI controller. The steady state condition was assumed when reading of temperatures, pressure and mass flow rate remain constant for at least 20 min. The details of the experimental condition are shown
Table 2 Experimental conditions. Parameter
Range
Refrigerant Average saturation temperature (°C) Mass flux (kg m2 s1) Inner tube material Heat flux (kW m2) Vapor quality
R-134a 35, 45 75, 115, 156, 191 Copper 1.5–14.34 0.1–0.84
540
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
For horizontal tube, the gravitational pressure gradient (dP/ dz)tp,g is considered as negligible. The acceleration pressure gradient is defined as following equation:
Table 3 Experimental uncertainties. Measurement
Uncertainty
Temperature (°C) Refrigerant mass flow rate (kg s1) Water mass flow rate (kg s1) Pressure (kPa) Pressure drop (kPa) Average vapor quality Heat transfer rate at pre-heater Heat transfer rate at test-section Heat transfer coefficient
0.1 0.050 103 1.41 103 5.201 0.125 5% 7.2% 11% 15%
( !) dP d x2 ð1 xÞ2 ¼ G2 þ dz tp;a dl qg a ql ð1 aÞ
where a is the void fraction which is calculated from Abdul-Razzak et al. correlation [25]:
a ¼ ð1 þ 0:49v0:3036 Þ tt
ð3Þ
_ w , Cpw, Tw,out and Tw,in are the mass flow rate of water inside where m the test section, specific heat of water, water temperature at outlet of test section and water temperature at inlet of test section, respectively. The average vapor quality of the refrigerant is the arithmetic mean of vapor quality at inlet and outlet of the test section which can be defined as following equation:
xav g;ts ¼
x þ x out;ts in;ts 2
ð4Þ
where inlet vapor quality ðxin Þ of the test section is evaluated by applying energy balance equation at the preheater as follows:
Q ph 1 xin ¼ hin;ph þ hf ;in;ts hfg;ts;in mR
ð5Þ
where Qph is the electrical power supplied to the pre-heater. The preheater is well insulated with ceramic wool to prevent the heat loss from the preheater to environment. The ceramic wool is also covered with the steel strips. It is assumed that, the heat loss from preheater is negligible. The hin,ph is the enthalpy at the inlet of the preheater which is calculated by measuring the temperature and pressure of sub-cooled liquid refrigerant. It is assumed that energy loss between the preheater outlet and test section inlet is neglected. Hence, the enthalpy at the test section inlet is equal to the enthalpy at the outlet of pre-heater. The enthalpy of saturated liquid (hf,in,ts) and enthalpy of vaporization (hfg,ts,in) are determined based on the temperature at the inlet of the test-section. The outlet vapor quality of the test section is given by
xout ¼
hout;ts hf ;out;ts hfg;out;ts
ð6Þ
where hf,out,ts and hfg,out,ts are the enthalpy of the saturated liquid and vaporization at the outlet of the test section, respectively, which are calculated by outlet temperature of the test section. The hout,ts is the enthalpy of the refrigerant at the outlet of the test section which is determined by energy balance equation on the test section as follows:
Q hout;ts ¼ hin;ts w mR
ð7Þ
where hin,ts and mR are the enthalpy of refrigerant at the inlet of the test section and mass flow rate of the refrigerant, respectively. The total pressure gradient (dP/dz)T is directly measured by a differential pressure transducer at inlet and outlet of the test section. The frictional pressure gradient (dP/dz)tp,f is calculated by subtracting the acceleration pressure gradient (dP/dz)tp,a and the gravitational pressure gradient (dP/dz)tp,g from the total pressure drop, as
dP dP dP dP ¼ dz tp;f dz T dz tp;a dz tp;g
1
ð10Þ
The two phase frictional pressure gradient multiplier, /2l is expressed as following equation:
The heat transfer rate (Qw) is determined from:
_ w C pw ðT w;out T W;in Þ QW ¼ m
ð9Þ
ð8Þ
/2l ¼
ðdP=dzÞtp;f ðdP=dzÞl;f
ð11Þ
The single phase liquid pressure gradient (dP/dz)l,f can be determined from:
dP G2 ð1 xÞ2 ¼ fl 2 dz l;f ðdi ql Þ
ð12Þ
The fanning friction factor fl is determined from the correlation proposed by Srinivasan et al. [26] for the coil,
( )15 1=2 2 D di fl ¼ 0:084 Rel di D
ð13Þ
where Rel is the liquid Reynolds number which is given by following equation
Rel ¼
Gð1 xÞdi
ll
ð14Þ
4. Results and discussion The experimental data points of smooth and dimpled helically coiled tube-in-shell type heat exchanger during condensation of refrigerant of R134a at the saturation temperature of 35 °C are drawn on the Taitel and dukler flow map [27] (Martinelli parameter, vtt vs vapor Froud number, Frv), as shown in Fig. 5. The vapor Froud number, Frv and Martinelli parameter vtt are given by following equation,
Gx Frv ¼ pffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi qv di gðql qv Þ
vtt ¼
0:9 0:5 0:1 1x qv ll x ql lv
ð15Þ
ð16Þ
The solid line drawn in Taitel and Dukler flow map [27] are the transition line between stratified-wavy, intermittent flow and annular flow of the air-water mixture in the straight smooth tube. Similarly, the flow pattern transitions for smooth and dimpled helically coiled tube are shown in discontinuous line. It can be found that the smooth and dimpled helical coiled tube have nearly the same transition line from stratified wavy to annular as well as stratified wavy to intermittent, while the transition criteria for the smooth straight tube comes to below that of those both tubes. The intermittent flow region falls into the annular region of the map. This may be because, the flow pattern map suggested by Taitel and Dukler was constructed for the air–water experimental data, which have different thermo-physical properties from R-134a studied in the present work. The Martinelli parameter for the transition criteria in the straight tube from intermittent to annular is 1.6, which was given by Taitel and Dukler [27]. In the current tests,
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
541
Fig. 5. Comparison of smooth and dimpled helically coiled tube experimental data on Taitel–Dukler flow map. (a) Smooth helically coiled tube. (b) Dimpled helically coiled tube.
the transition criteria from intermittent to annular flow takes place at Martinelli parameter around 0.51 for smooth helical coiled tube and 0.61 for the dimpled helical coiled tube. To better understand the flow patterns of R-134a inside the smooth and dimpled helically coiled tube, the axis of the Taitel and Dukler flow pattern map have been converted to mass flux versus vapor quality. It can be clearly seen form Fig. 6(a) and (b) that the stratified-wavy flow occurs at low mass flux (75 kg m2 s1) for both tube. As the mass flux increases from 75 to 115 kg m2 s1, for dimpled helically coiled tube, the intermittent flow takes place at small range of vapor quality, while annular flow occurs at large range of vapor quality. Similarly, for smooth helically coiled tube, the stratified-wavy exists at low vapor quality, then transfers to annular flow at high vapor quality without the presence of any intermittent flow. Yang et al. [28] was also observed similar flow characteristics. Moreover, at mass flux greater than 115 kg m2 s1, the intermittent and annular flow exist for both tube. The flow pattern transition from intermittent to annular flow occurred at a vapor quality in the range of 0.32–0.33 for the smooth helically coiled tube, and 0.27 for the dimpled helically coiled tube. Cui et al. [29] proposed the transition from intermittent to annular flow at vapor quality range 0.2–0.3 for micro-fin helically coiled tube during boiling of R-134a. In Fig. 7, it can be observed that the flow pattern inside the dimpled helically coiled tube at G = 156 kg m2 s1 and x = 0.1310 prevails intermittent flow, whereas in the smooth helical coiled tube at G = 156 kg m2 s1 and x = 0.1028, the stratified flow exists. It should be observed that the difference is that the waves produced in stratified-wavy flow touch the upper portion of the tube wall and wet the upper portion of the tube, intermittently. Furthermore, vapor quality increases, the flow pattern inside the smooth helically coiled tube at x = 0.2158 becomes intermittent. In the dimpled helically coiled tube at x = 0.2728, the waves on the interface of the liquid-vapor push the liquid film and cover the top portion of the tube wall as a thin liquid film experiences in the bottom portion of the tube wall. The thin liquid film and more turbulence in the interface of the liquid-vapor are observed in the dimpled helically coiled tube. It is also observed that the dimpled helical coil quickens the transition from intermittent to annular flow compared to that of smooth helical coiled tube. 4.1. Condensation heat transfer coefficient Firstly, as shown in Fig. 8, the experimental heat transfer coefficient of smooth straight tube were compared with the data pre-
Fig. 6. Mass flux (G) vs average vapor quality (x) flow map for (a) Smooth helically coiled tube (SHLCT) (b) Dimpled helically coiled tube (DHLCT).
dicted by the renowned correlation given by Dobson and Chato [3] and Shah [30]. It was apparently seen that the data was well predicted by Dobson and Chato [3], followed by Shah [30] correlation within ±30% error band. Additionally, in Fig. 9, experimental frictional pressure drop data of smooth straight tube was compared with well-known correlation proposed by Friedel [31] and Muller-Steinhagen and Heck [32]. It was found that most of experimental data fall into an error band of ±30%. This suitable agreement of experimental heat transfer coefficients and frictional pressure drop with predicted data certifies the experimental setup. Form Fig. 10, the average heat transfer coefficient data of dimple helically coiled tube was compared to smooth helical coiled and straight tube at saturation temperature of 35 °C with the mass flux of 75, 115, 156 and 191 kg m2 s1. It can be seen that dimple helically coiled tube gives higher heat transfer coefficient compared to smooth helical coiled tube and smooth straight tube in the percentage of 18–32% and 51–61%, respectively, while, the smooth helical coiled tube produces 37–48% higher heat transfer coefficient than that of smooth straight tube at all range of vapor quality. This is because, helically coiled tube yields the centrifugal effect on two-phase flow and produces secondary flow which influences the flow of the vapor and liquid film. As vapor quality increases, the vapor velocity rises, consequently, secondary flow becomes stronger. The vapor flowing from center core of the helically tube pushes the liquid film from the outer to inner tube wall, hence, vapor flows at center core of tube again. These produces a
542
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
(a) Smooth helically coiled tube
x = 0.1028
x = 0.2158
x = 0.3318
(b) Dimpled helically coiled tube
x = 0.1310
x = 0.2728
x = 0.3410
Fig. 7. Flow pattern of R-134a at a mass flux of 156 kg m2 s1 in (a) smooth helically coiled tube and (b) dimpled helically coiled tube.
Fig. 8. Comparison of two-phase heat transfer coefficient data for smooth straight tube with existing correlation.
Fig. 9. Condensation frictional pressure drop data of the smooth round tube compared with existing correlations.
chaotic waves on the liquid film and more liquid entrainment, thereby, heat transfer surface area increases. As a results, the heat transfer mechanism of helically coiled was found to be different
from that of straight tube. The similar behaviors for smooth helically coiled tube were reported by Wongwises et al. [12], previously. It can be also seen that dimpled helically coiled tube gives higher heat transfer coefficient. This is because, dimpled helically coiled tube provides a protrusions inside the tube which provides more heat transfer surface area and turbulence on the fluid flow. Therefore, secondary flow becomes more effective and entrainment of liquid droplets also increased which tends to a lower thermal resistance. As a results, higher heat transfer coefficient ensures for dimpled helical coiled tube. The same phenomenon investigated for dimpled helically straight tube were given by Aroonrat et al. [7]. They found that protrusions inside the tube enhance the heat transfer by proper mixing of fluid flow and increase an amount of turbulence on the two phase flow. In Fig. 11, the effect of mass flux on average heat transfer coefficient of dimpled helically coiled tube is revealed. It can be seen that the mass flux has a remarkable influence on the heat transfer coefficient. This is because, mass flux enhances the flow velocity which increases the shearing effect between the layer of vapor and liquid. Moreover, the experimental results also gives, as the mass flux increases from 75 to 191 kg m2 s1, the heat transfer coefficient increases up to 113–187% in the all range of vapor quality. In Fig. 12, saturation temperature has a small effect on the heat transfer coefficient. It was found that the heat transfer coefficient for saturation temperature of 45 °C at same mass flux lowers than that of saturation temperature of 35 °C. This is because, a higher saturation temperature results lower specific volume, leading to the lower vapor velocity of refrigerant. Another reason is that, the thermal conductivity of refrigerant decreases with an increase in saturation temperature, which in turn higher liquid thermal resistance. Hence, the heat transfer coefficient decreased. Fig. 13 illustrates the effect of mass flux (75, 115, 156 and 191 kg m2 s1) on heat transfer enhancement factor for smooth and dimpled helically coiled tubes, which is defined as the ratio of heat transfer coefficient of helically coiled tube (smooth or dimpled helically coiled tube) to that of smooth straight tube. The enhancement factor for smooth helical coiled tube varies from 1.67 to 1.91 at 75 kg m2 s1, 1.61 to 1.90 at 115 kg m2 s1, 1.66 to 1.77 at 156 kg m2 s1 and 1.62 to 1.72 at 191 kg m2 s1, Similarly, for dimpled helical coiled tube, the enhancement factor ranges from 2.10 to 2.43, 2.07 to 2.36, 2.03 to 2.39 and 1.94 to 2.13 for the mass flux of 75, 115, 156 and 191 kg m2 s1, respectively. Moreover, Table 4 reveals that the effect of mass flux and saturation temperature on the enhancement factor. The mass flux and saturation temperature have fewer effect on the enhancement factor. It can
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
543
Fig. 10. Comparison of the heat transfer coefficient for tubes at different mass flux.
be seen that enhancement factor declines with the rise of both the mass flux and saturation temperature. Fig. 14 presents the comparison of the experimental Nusselt number for smooth helically coiled tube with renowned correlations given by the Wongwise and Polsongkram [12] and Gupta et al. [20]. The correlations for smooth helical coiled tube are listed in the Table 5. As shown in Fig. 14, these correlations well predict the experimental data for smooth helically coiled tube. It is revealed that Wongwise and Polsongkram [12] overestimated the experimental results, however, the data exists within an error band of ±30%. At present, there is no correlation for prediction of the two phase flow Nusselt number in a dimpled helically coiled tube. In order to develop the correlation, the effect of equivalent Dean number, DeEq, Prandtl number, Prl, Martinelli parameter,vtt and reduced pressure, pr are considered. The new developed correlation to predict the two phase flow Nusselt number in a dimpled helically coiled tube is:
ðNutp ÞDT ¼ 0:0208ðDeEq Þ1:05 ðPr l Þ0:65 ðvtt Þ0:113 ðpr Þ0:169 Fig. 11. Effect of mass flux on condensation heat transfer coefficients in the dimpled helically tube.
where an equivalent Dean number, DeEq is given by equation:
ð17Þ
544
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
Prl ¼
Cpl ll kl
ð19Þ
Reduced pressure, pr is evaluated from
pr ¼
ps pcrit
ð20Þ
The proposed correlation gives prediction of Nusselt number in the range of non-dimensional numbers as follows: 1564 DeEq 112,666; 3.19 Pr l 3.29; 0.062 vtt 2.01; 0.21 pr 0.28. In Fig. 15, the experimental two phase Nusselt number were compared with proposed correlation. As shown in Fig. 14, the correlation is well predicted the experimental data within the error bar ±20%. 4.2. Frictional pressure drop
Fig. 12. Effect of saturation temperature on condensation heat transfer coefficient for tubes.
Fig. 13. Enhancement factor as a function of mass flux at Tsat = 35 °C.
(
l DeEq ¼ Rel þ Rev v ll
ql qv
0:5 )
di 2R
0:5
Prandtl number, Prl, is determined from
ð18Þ
From Fig. 16, the frictional pressure drop of dimpled helically coiled tube was compared to smooth helically coiled and straight tube at saturation temperature of 35 °C with the mass flux of 75 and 115 kg m2 s1. It can be seen that dimpled helically coiled tube offers higher frictional pressure drop compared to smooth helically coiled and smooth straight tube in the percentage of 47–61% and 61–75%, respectively, while, the smooth helically coiled tube gives 25–43% higher frictional pressure drop than that of smooth straight tube at all range of vapor quality. This is because, vapor velocity increases with rise of vapor quality and mass flux and enhances the shear stress between vapor-liquid interfaces, hence, secondary flow inside the helically coiled tube turn out to be more effective which cause the higher pressure drop inside helically coiled tube compared to that of straight tube. From Fig. 16, it is also seen that dimpled helically coiled tube produces a higher frictional pressure drop compared to that of both, smooth helically coiled tube and straight tube. The pressure drop growth was yielded by the protrusions provided inside the dimpled helically coiled tube which rises the turbulence level on the twophase flow, and thereby, secondary flow becomes stronger compared to that of smooth helically coiled tube. Consequently, higher frictional pressure drop occurs in the dimpled helically coiled tube. In Fig. 17, frictional pressure drop data of dimpled helically coiled tube data are revealed. It can be seen that the frictional pressure drop increases when mass flux increases. This is because, mass flux of refrigerant rises the flow velocity, which causes higher interfacial shear stress on the liquid film, hence, the frictional pressure drop raised. The experimental results gives, as the mass flux increases from 75 to 191 kg m2 s1, the frictional pressure drop increases up to 261–310% in the all range of vapor quality. Fig. 18 presents the effect of saturation temperature on frictional pressure gradient inside the tubes at mass flux of 156 kg m2 s1. It can be seen that the frictional pressure drop decreases with the rise of saturation temperature of refrigerant. This is because, at high saturation temperature, the specific volume of vapor decreases, hence the vapor velocity falls, resulting lower interfacial shear stress between the vapor and liquid film. Consequently, lower value of frictional pressure drop takes place at a lower saturation temperature. Fig. 19 illustrates the effect of mass flux (75, 115, 156 and 191 kg m2 s1) on pressure drop penalty factor for smooth and dimpled helically coiled tubes, which is defined as the ratio of frictional pressure drop of helical coiled tube (smooth or dimpled helical coiled tube) to that of smooth straight tube. The penalty factor for smooth helically coiled tube varies from 1.35 to 1.75 at 75 kg m2 s1, 1.38 to 1.70 at 115 kg m2 s1, 1.44 to 1.75 at 156 kg m2 s1 and 1.39 to 1.74 at 191 kg m2 s1, Similarly, for the dimpled helically coiled tube, the penalty factor ranges from 2.57 to 3.92, 2.76 to 4.03, 2.99 to 4.60 and 2.92 to 4.01 for the mass flux
545
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548 Table 4 Heat transfer enhancement factor (EF) and pressure drop penalty factor (PF) of the smooth and dimpled helically coiled tube. Tube
T (°C)
G (kg m2 s1)
x
EF
PF
Smooth helical coiled tube
35
75 115 156 191
0.23 0.25 0.21 0.23
1.78 1.75 1.74 1.70
1.47 1.48 1.54 1.51
45
75 115 156 191
0.24 0.22 0.20 0.25
1.77 1.71 1.72 1.66
1.50 1.49 1.66 1.62
35
75 115 156 191
0.20 0.24 0.19 0.22
2.26 2.16 2.14 2.08
3.16 3.23 3.35 3.53
45
75 115 156 191
0.23 0.26 0.23 0.20
2.18 2.12 2.06 1.92
3.20 3.25 3.43 3.41
Dimpled helical coiled tube
little effect of the penalty factor. It can be seen that penalty factor increases with the rise of both the mass flux and saturation temperature. In Fig. 20, the experimental liquid two- phase frictional pressure drop multiplier value is compared to well-known correlations proposed by the Wongwises and Polsongkram [12] and Gupta et al. [20]. The correlations for smooth helically coiled tube are listed in the Table 5. As shown in Fig. 20, these correlation well predicted the experimental data for the smooth helically coiled tube. At present, there is no correlation for predicting the two phase flow frictional pressure drop in a dimpled helical coiled tube. In order to develop the correlation, a two-phase frictional multiplier is used to correlate the frictional pressure drop of two-phase flow. A newly correlation is developed which is given by following Eq. (21),
DPtp;f 4:8671 1 ¼ ð/2l ÞDT ¼ 4:5033 1 þ 1:4919 þ 2 p0:321 DPl;f DT vtt vtt r Fig. 14. Comparison of experimental Nusselt number for smooth helically tube with existing correlation.
of 75, 115, 156 and 191 kg m2 s1, respectively. Moreover, Table 4 reveals that the effect of mass flux and saturation temperature on the penalty factor. The mass flux and saturation temperature have
ð21Þ
where pr and vtt are the reduced pressure of the refrigerant, defined in Eq. (19) and Martinelli parameter, respectively. The proposed correlation gives prediction of two-phase frictional multiplier in the range of non-dimensional numbers as follows: 0.062 vtt 2.01; 0.21 pr 0.28. In Fig. 21, the experimental two phase frictional pressure drop were compared with proposed correlation. As shown in Fig. 21, the experimental frictional pressure drop for dim-
Table 5 Wongwise and Polsongkram [12] and Gupta et al. [20] correlations for smooth helically coiled tube. Correlation Wongwise and Polsongkram [12]
Two phase flow Nusselt number (Nutp) ðNutp Þ ¼ 0:1352ðDeEq Þ0:7654 ðPr l Þ0:8144 ðvtt Þ0:0432 ðpr Þ0:3356 ðBo 104 Þ 0:5 0:5 C l di DeEq ¼ Rel þ Rev llv qql , Pr l ¼ plk l , pr ¼ pps 2R l
v
l
0:112
crit
q i i Rel ¼ Gð1xÞd , Rev ¼ Gxd l l , Bo ¼ Gilv ; v
l
The two-phase frictional pressure gradient multiplier ð/2l Þ 1 ð/2l Þ ¼ 1 þ v5:569 1:496 þ v2 tt
Gupta et al. [20]
tt
Two phase flow Nusselt number (Nutp) 0:63 ¼ 0:4 1 þ 3:024 v0:8 pr
Nutp Nul
tt
Nul is the single phase Nusselt number which is proposed by Mori and Nakayama [33] (
16 ) 1=12 2:5 5=6 di 1 Rel Pr0:4 1 þ 0:061 Rel dDi Nul ¼ 41 l D The two-phase frictional pressure gradient multiplier ð/2l Þ 1 ð/2l Þ ¼ 2:76 1 þ v7:094 p0:7 1:378 þ v2 r tt
tt
546
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
Fig. 15. Experimental Nusselt number of dimpled helically coiled tube compared with proposed correlation.
Fig. 17. Effect of mass flux on frictional pressure drop in the dimpled helically tube.
Fig. 18. Effect of saturation temperature on frictional pressure drop for tubes.
Fig. 16. Comparison of the frictional pressure drop for tubes at mass flux of 75 & 115 kg m2 s1.
Fig. 19. Penalty factor as function of mass flux at Tsat of 35 °C.
pled helical coiled tube is well predicted by new proposed correlation, almost, within an error band of ±20%. Furthermore, the maximum heat transfer rate is concern, it is always preferable to provide the more surface area and the dim-
pled helically coiled tube absolutely fulfill the same. The design of a heat exchanger comprising dimpled helically coiled tube in context to condensation need a prediction tool. Therefore, a non-dimensional correlations predicting the Nusselt number and
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
2.
3.
4.
Fig. 20. Comparison of experimental liquid two-phase frictional multiplier for smooth helically tube with Wongwises and Polsongkram [12] and Abhinav Gupta et al. [20] corrrelations.
5.
547
effect on the heat transfer coefficient. On the other hand side, the frictional pressure drop for dimpled helically coiled tube also increases with the rise of mass flux and vapor quality, but, it decreases with the increase of saturation temperature. In a dimpled helically coiled tube, the transition from intermittent to annular flow occurred at lower vapor quality (x = 0.27) than the smooth helical coiled tube (x = 0.31–0.33). Dimple helically coiled tube gives a higher heat transfer coefficient compared to smooth helically coiled tube and smooth straight tube in the percentage of 18–32% and 51–61%, respectively, while, the smooth helically coiled tube produces 37–48% higher heat transfer coefficient than that of smooth straight tube at all range of vapor quality. The frictional pressure drops of dimpled helically coiled tube were about 2.57–4.01 times higher than those of smooth straight tube, while, for smooth helically coiled tube, the frictional pressure drops were almost 1.35–1.75 times higher than those of smooth straight tube. Experimental Nusselt number and frictional pressure drop of dimpled helically coiled tube are not adequately predicted by correlation, hence, a newly correlation is developed which predicts the experimental results of dimpled helical coiled tube in the mass flux range of 75–191 kg m2 s1.
Appendix A. Supplementary material Supplementary data associated with this article can be found, in the online version, at https://doi.org/10.1016/j.applthermaleng. 2017.10.026. References
Fig. 21. Comparison of the two-phase frictional pressure drop data for dimpled helically tube by proposed correlation.
frictional pressure drop, were developed and best suitable to small scale heat exchanger. Since the effect of dimples was not considered while developing the proposed correlation, the applicability is limited to specified geometry i.e., pitch and depth of the dimple. However, in order to generalize the correlation in context to effect of dimples, further research would be done i.e. multiple tubes of varying pitch and depth of dimple. 5. Conclusion In this experiment, the effect of mass flux (75, 115, 156 and 191 kg m2 s1), vapor quality in the range of 0.1–0.8 and saturation temperature (35 & 45 °C) on the average heat transfer coefficient and average frictional pressure drop inside a smooth and dimpled helically coiled tube have been examined. The results obtained from smooth and dimpled helically coiled tube were also compared with results from those smooth straight tube. The following results are obtained. 1. The average heat transfer coefficient for dimpled helically coiled tube increases, as mass flux or vapor quality of refrigerant increases, while, the saturation temperature has a small
[1] D.A. Reay, Heat transfer enhancement-a review of techniques and their possible impact on energy efficiency in the U.K., Heat Recover. Syst. CHP. 11 (1991) 1–40. [2] J.G. Collier, J.R. Thome, in: Convective Boiling and Condensation, third ed., Oxford Univ. Press, Berlin, 1994, pp. 99–280. [3] M.K. Dobson, J.C. Chato, Condensation in smooth horizontal tubes, J. Heat Transf. 120 (1998) 193. [4] A. Cavallini, G. Censi, D. Del Col, L. Doretti, G.A. Longo, L. Rossetto, Experimental investigation on condensation heat transfer and pressure drop of new HFC refrigerants (R134a, R125, R32, R410A, R236ea) in a horizontal smooth tube, Int. J. Refrig. 24 (2001) 73–87. [5] D. Jung, K.H. Song, Y. Cho, S.J. Kim, Flow condensation heat transfer coefficients of pure refrigerants, Int. J. Refrig. 26 (2003) 4–11. [6] M. Anowar Hossain, Y. Onaka, H.M.M. Afroz, A. Miyara, Heat transfer during evaporation of R1234ze(E), R32, R410A and a mixture of R1234ze(E) and R32 inside a horizontal smooth tube, Int. J. Refrig. 36 (2013) 465–477. [7] K. Aroonrat, S. Wongwises, Experimental study on two-phase condensation heat transfer and pressure drop of R-134a flowing in a dimpled tube, Int. J. Heat Mass Transf. 106 (2017) 437–448. [8] A. Sarmadian, M. Shafaee, H. Mashouf, S.G. Mohseni, Condensation heat transfer and pressure drop characteristics of R-600a in horizontal smooth and helically dimpled tubes, Exp. Therm. Fluid Sci. 86 (2017) 54–62. [9] H.J. Kang, C.X. Lin, M.A. Ebadian, Condensation of R134a flowing inside helicoidal pipe, Int. J. Heat Mass Transf. 43 (2000) 2553–2564. [10] B. Yu, J.T. Han, H.J. Kang, C.X. Lin, A. Awwad, M.A. Ebadian, Condensation heat transfer of R-134A flow inside helical pipes at different orientations, Int. Commun. Heat Mass Transf. 30 (2003) 745–754. [11] J.T. Han, C.X. Lin, M.A. Ebadian, Condensation heat transfer and pressure drop characteristics of R-134a in an annular helical pipe, Int. Commun. Heat Mass Transf. 32 (2005) 1307–1316. [12] S. Wongwises, M. Polsongkram, Condensation heat transfer and pressure drop of HFC-134a in a helically coiled concentric tube-in-tube heat exchanger, Int. J. Heat Mass Transf. 49 (2006) 4386–4398. [13] C.X. Lin, M.A. Ebadian, Condensation heat transfer and pressure drop of R134a in annular helicoidal pipe at different orientations, Int. J. Heat Mass Transf. 50 (2007) 4256–4264. [14] L. Shao, J.Tian Han, G.Ping Su, J.Hong Pan, Condensation heat transfer of R134A in horizontal straight and helically coiled tube-in-tube heat exchangers, J. Hydrodyn. 19 (2007) 677–682. [15] M.H. Al-Hajeri, A.M. Koluib, M. Mosaad, S. Al-Kulaib, Heat transfer performance during condensation of R-134a inside helicoidal tubes, Energy Convers. Manage. 48 (2007) 2309–2315.
548
A.K. Solanki, R. Kumar / Applied Thermal Engineering 129 (2018) 535–548
[16] M. El-Sayed Mosaad, M. Al-Hajeri, R. Al-Ajmi, A.M. Koliub, Heat transfer and pressure drop of R-134a condensation in a coiled, double tube, Heat Mass Transf. 45 (2009) 1107–1115. [17] Y. Murai, S. Yoshikawa, S.I. Toda, M.A. Ishikawa, F. Yamamoto, Structure of airwater two-phase flow in helically coiled tubes, Nucl. Eng. Des. 236 (2006) 94– 106. [18] S.B. Genic´, B.M. Jac´imovic´, M.S. Jaric´, N.J. Budimir, M.M. Dobrnjac, Research on the shell-side thermal performances of heat exchangers with helical tube coils, Int. J. Heat Mass Transf. 55 (2012) 4295–4300. [19] S.B. Genic´, B.M. Jac´imovic´, M.S. Jaric´, N.J. Budimir, Analysis of fouling factor in district heating heat exchangers with parallel helical tube coils, Int. J. Heat Mass Transf. 57 (2013) 9–15. [20] A. Gupta, R. Kumar, A. Gupta, Condensation of R-134a inside a helically coiled tube-in-shell heat exchanger, Exp. Therm. Fluid Sci. 54 (2014) 279–289. [21] M. Mozafari, M.A. Akhavan-Behabadi, H. Qobadi-Arfaee, M. Fakoor-Pakdaman, Condensation and pressure drop characteristics of R600a in a helical tube-intube heat exchanger at different inclination angles, Appl. Therm. Eng. 90 (2015) 571–578. [22] M.R. Salimpour, A. Shahmoradi, D. Khoeini, Experimental study of condensation heat transfer of R-404A in helically coiled tubes, Int. J. Refrig. 74 (2017) 584–591. [23] S.J. Kline, F.A. McClintock, Describing uncertainties in single-sample experiments, Mech. Eng. 75 (1953) 3–8. [24] E.W.W. Lemmon, M.L.L. Huber, M.O.O. McLinden, NIST Standard Reference Database 23: Reference Fluid Thermodynamic and Transport Properties (REFPROP), Version 9.0, Phys.
[25] A. Abdul-Razzak, M. Shoukri, J.S. Chang, Characteristics of refrigerant R-134A liquid-vapor two-phase flow in a horizontal pipe, ASHRAE Trans. (1995). [26] P.S. Srinivasan, S.S. Nandapurkar, F.A. Holland, Friction factor for coils, Trans. Inst. Chem. Eng. 48 (1970) T156–T161. [27] Y. Taitel, A.E. Dukler, A model for predicting flow regime transitions in horizontal and near horizontal gas-liquid flow, AIChE J. 22 (1976) 47–55. [28] Z. Yang, M. Gong, G. Chen, X. Zou, J. Shen, Two-phase flow patterns, heat transfer and pressure drop characteristics of R600a during flow boiling inside a horizontal tube, Appl. Therm. Eng. 120 (2017) 654–671. [29] W. Cui, L. Li, M. Xin, T.C. Jen, Q. Liao, Q. Chen, An experimental study of flow pattern and pressure drop for flow boiling inside microfinned helically coiled tube, Int. J. Heat Mass Transf. 51 (2008) 169–175. [30] M.M. Shah, An Improved and Extended General Correlation for Heat Transfer During Condensation, HVAC&R Res. 15 (2009) 37–41. [31] L. Friedel, Improved friction pressure drop correlations for horizontal and vertical two phase pipe flow, in: Eur. Two-Phase Flow Gr. Meet., 1979, pp. 485–491. [32] H. Müller-Steinhagen, K. Heck, H. Müller-Steinhagen, K. Heck, A simple pressure drop correlation for two-phase flow in pipes, Chem. Eng. Prog. 20 (1986) 297–308. [33] Y. Mori, W. Nakayawa, Study on forced convective heat transfer in curved pipes: 3rd report, theoretical analysis of uniform wall temperature and practical formulae, Trans. Jpn. Soc. Mech. Eng. 33 (1967) 104–112.