Time periodic saturated flow boiling heat transfer of R-134a in a narrow annular duct due to heat flux oscillation

International Journal of Heat and Mass Transfer 106 (2017) 35–46

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Time periodic saturated flow boiling heat transfer of R-134a in a narrow annular duct due to heat flux oscillation C.A. Chen a, T.F. Lin a, Wei-Mon Yan b,⇑ a b

Department of Mechanical Engineering, National Chiao Tung University, Hsinchu 30010, Taiwan Department of Energy and Refrigerating Air-Conditioning Engineering, National Taipei University of Technology, Taipei 10608, Taiwan

a r t i c l e

i n f o

Article history: Received 27 June 2016 Received in revised form 4 October 2016 Accepted 4 October 2016

Keywords: Time periodic saturated flow boiling Heat transfer R-134a Heat flux oscillation Mini-channel

a b s t r a c t The time periodic saturated boiling heat transfer in a horizontal annulus was investigated experimentally where the walls are under an oscillating heat flux. The fluid enters the duct with zero vapor quality (saturated liquid state). The amplitude of the imposed heat flux oscillation Dq varies from 0% to 50% of mean  and four different periods of heat flux oscillation t p , including 20, 30, 60 and 120 s are imposed heat flux q applied to the system. The measured data display that when the applied heat flux is close to that for the onset of stable flow boiling, intermittent flow boiling appears in which nucleate boiling on the heated surface only exists in a partial interval of each periodic cycle and the heat flux oscillation does not noticeably affect the time-average boiling curves and heat transfer coefficients. Besides, the heated wall temperature and evaporating flow pattern are found to oscillate periodically in time as well and at the same frequency as the imposed heat flux oscillation. Furthermore, in the persistent boiling the resulting oscillation amplitudes of the heated surface temperature, heat transfer coefficient gets larger for a longer period and larger amplitude of the imposed heat flux oscillation and for a higher mean imposed heat flux. The substantial time lag in the heated surface temperature oscillation is observed. In the first half of the periodic cycle in which the heat flux reduces with time, after the time lag the heated wall temperature decreases with time. The inverse processes occur in the second half of the cycle in which imposed heat flux increases with time. Finally, flow regime maps are provided to explain the boundaries separating different boiling regimes for the R-134a saturated boiling in the duct. Ó 2016 Elsevier Ltd. All rights reserved.

1. Introduction Global warming reduction is one of the main reasons to improve the energy utilization efficiency of various engineering systems. Recently, it is established that using compressors with variable frequency (instead of ON/OFF) in air-conditioning and refrigeration systems will increase their efficiency dramatically. It is important to note that the systems are subjected to varying thermal loads. How this time dependent heat flux affect the characteristics of boiling processes in the refrigeration cycles employed in these air-conditioning and refrigeration systems remains generally unknown. Besides, in cooling future ultra-high component density electronic devices, methods based on the phase change heat transfer are often considered. Moreover, the power dissipated and cooling load in these devices are time dependent. The associated time dependent heat transfer processes in the specific electronics cooling methods are also poorly understood. ⇑ Corresponding author. E-mail address: [email protected] (W.-M. Yan). http://dx.doi.org/10.1016/j.ijheatmasstransfer.2016.10.014 0017-9310/Ó 2016 Elsevier Ltd. All rights reserved.

Transient single-phase forced convection in a horizontal plane channel with diverse arrangements of time varying imposed heat flux on the channel walls was investigated by Girault and Petit [1] with oscillating wall heat flux. Results illustrate that a small wall temperature oscillation was noted for the power-off situation. Two-phase dynamic instabilities in the flow boiling of various liquids in a long heated channel have been recognized [2,3]. Significant temporal oscillations in pressure, temperature, mass flux and boiling onset occur at a certain operating condition. Specifically in flow boiling of R-11 in a vertical channel, the pressuredrop and thermal oscillations were observed by Kakac et al. [4]. The presence of the density wave oscillation superimposed on the pressure-drop oscillation was further noted by Kakac et al. [5]. In a continuing study for R-11 in a horizontal tube of 106 cm long, the authors [6] examined the dependence of the oscillation amplitude and period on the system parameters and located the boundary of various types of oscillations. A parallel study was carried out by Comakli et al. [7] and the effect of the channel length on the two-phase flow dynamic instabilities was considered.

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Nomenclature Dh G hr k L Lc P  q Q tp t1 t2 t3 T Tw Tsat,

hydraulic diameter, m, Dh = (Do  Di) time-average mass fluxes, kg/m2 s boiling heat transfer coefficient, W/m2 °C thermal conductivity, W/m °C heating length of the annular duct, m characteristics length, m system pressure, kpa time-average imposed heat flux, W/m2 heat transfer rate, W period of mass flux and heat flux oscillation, sec conduction time scale, sec convective time scale, sec time scale for bubble growth or departure, sec temperature,°C wall temperature, °C saturated temperature of refrigerant, °C

The dynamic characteristics of a horizontal boiling duct connected to surge tank was examined using nonlinear analysis by Mawasha and Gross [8]. The dynamic oscillations are described as relaxation oscillation and the qualitative features of the response in their study. The boiling onset in an upward flow of subcooled water in a vertical tube of 7.8-m long connected with a liquid surge tank could cause substantial flow pressure and density-wave oscillations [9]. These boiling onset oscillations were attributed to a sudden growth of pressure-drop across the channel and a large fluctuation in the water flow rate at the onset of nucleate boiling. The pressure-drop oscillations of n-pentane liquid in a vertical small rectangular channel was reported by Brutin et al. [10]. A two-phase flow pressure drop analysis by ranging several mass flow rates in a minichannel has been performed by Brutin and Tadrist [11]. The effects of the inlet flow conditions on the boiling instabilities were found to be relatively significant. A similar study for subcooled flow boiling of deionized water was conducted by Shuai et al. [12] by considering pressure-drop oscillation. Thorncroft et al. [13] experimentally investigated upflow and downflow boiling of FC-87 in a vertical rectangular channel. They showed that both bubble growth and bubble departure rates increase with increases in Jacob number, but decrease with increasing the mass flux. Thorncroft and Klausner [14] found that the sliding of the bubbles on the boiling surface enhances heat transfer rate in forced convection boiling of FC-87 in a vertical upflow and downflow. An experimental investigation of low pressure subcooled flow boiling inside a vertical annular duct was performed by Zeitoun and Shoukri [15]. They confirmed that the bubble departure was not the reason for the net vapor generation (NVG). However, the mean size and lift duration of the bubbles increased at decreasing liquid subcooling. Klausner et al. [16] developed a criterion for the bubble departure diameter from the heated surface in the forced convection boiling. The study was carried out for a saturated two-phase mixture of refrigerant R-113 flowing through a 25
25 mm2 visual boiling section. They found that the average bubble departure diameter decreased with increasing mass flux and with decreasing heat flux. Chien and Webb [17] studied the bubble dynamics for pool boiling on an enhanced tubular surface, which consists of an integral-fin tube having a copper foil wrapped over the fin tips. They disclosed that the mean bubble diameter was smaller for a higher heat flux and the bubble growth period was shorter for smaller bubbles and greater values of heat flux. The predictions of the active nucleation sites from knowing the size and cone angle of the cavities actually present on the surface

z

coordinate (downstream coordinate for annular duct flow), mm

Greek symbols Dq amplitude of heat flux oscillation, W/m2 DTsat wall superheat, (T w  T sat ), °C d gap size, mm l viscosity, N s=m2 q density, kg/m3 Subscripts r refrigerant side s heater surface sat saturated flow boiling tp two-phase w duct wall

were proposed by Yang and Kim [18]. Using an electron microscope and a differential interference contrast microscope, they obtained the cavity probability density function involving cavity size (ranging from 0.65 to 6.2 lm) and b (cone half angle). Gaertner [19] observed that active nucleation sites were randomly located and could be expressed in terms of Poisson distribution function. Sultan and Judd [20] reached the same conclusion from their observations. Zeng and Klausner [21] obtained experimental data for the active nucleation site density Nac during flow boiling of R113 on a horizontal 25
325 mm test section with a nichrome heating strip. Their experiments were performed for varying vapor quality at inlet, system pressure and wall heat flux. They examined the effects of vapor velocity, liquid velocity, liquid film thickness, system pressure, and wall heat flux on Nac. They concluded that even if Nac was dependent on the critical radius of the nucleus rc, their data were not sufficient for correlating Nac. Kocamustafaogullari and Ishii [22] developed a relation for active nucleation site density in pool boiling from the data available in the literature. They also applied the correlation to few available forced convection nucleate boiling data. Basu et al. [23] suggested an experimentally based correlation considering the effects of contact angle on the active nucleation site density during forced convective boiling of water on a vertical surface. In the experiments, they utilized mirror-finished copper surfaces prepared by a well-defined procedure [24,25] examined flow boiling heat transfer and associated bubble characteristics of R-134a in a narrow annular duct. They concluded that the bubbles are suppressed to become smaller and less dense by raising the refrigerant mass flux and inlet subcooling. The mean bubble departure frequency rises with the growing refrigerant mass flux and saturated temperature and with the decreasing duct size. Moreover, the active nucleation site density is much high at low refrigerant mass fluxes, particularly at a high heat fluxes. Chen et al. [26] experimentally studied the effects of the imposed time periodic refrigerant flow rate oscillation on flow boiling heat transfer and associated bubble characteristics of refrigerant R-134a in a horizontal narrow annular duct with the duct gap fixed at 2.0 mm. The measured data revealed that the effects of the mass flux oscillation on the size of the departing bubble and active nucleation site density dominate over the bubble departure frequency. Recently, Wang et al. [27] conducted the transient oscillatory boiling heat transfer and associated bubble characteristics of FC-72 flow over a small circular plate with a time varying heat flux. They indicated that the imposed the temporal oscillations in the heated plate temperature, boiling heat transfer

C.A. Chen et al. / International Journal of Heat and Mass Transfer 106 (2017) 35–46

coefficient, bubble departure diameter and frequency, and active nucleation site density were caused the imposed heat flux oscillations. The measured heat transfer data for the oscillatory subcooled flow boiling of R-134a resulting from the refrigerant mass flux oscillation in the narrow annular duct have been examined by Wang et al. [28]. In their study, effects of the mean level and oscillation amplitude and period of the refrigerant mass flux on the oscillatory subcooled R-134a flow boiling are presented in details. Saturated flow boiling characteristics of deionized water in parallel microchannels with 9 parallel square microchannels having hydraulic diameters of 150 lm are investigated experimentally by Markal et al. [29]. Heat transfer and pressure drop are examined for varying values of the governing parameters. Simultaneous high-speed video images have been taken as well as temperature and pressure measurements. The flow visualization results lead to key findings for flow boiling instabilities and underlying physical mechanisms of heat transfer in microchannels. The boiling heat transfer in minichannel at high saturation temperatures in a horizontal 3.00 mm inner diameter stainless steel tube with R-245fa as working fluid was examined experimentally by Charnay et al. [30]. The experimental work is characterized by a saturation temperature ranging from 100 to 120 °C. In their study, four flow patterns were found: intermittent flow, annular flow, dryout flow and mist flow regimes. They concluded that the kind of flow pattern has a major influence on the heat transfer mechanisms. Besides, at high saturation temperatures, the experimental results clearly show the dominance of nucleate boiling over a wide range of vapor quality. Kim and Mudawar [31,32] developed universal predictive tools for saturated flow boiling pressure drop and heat transfer for mini/micro-channel flows. In Ref. [31], they culminated in a technique for determining the dryout incipience quality corresponding to substantial deterioration in the heat transfer coefficient. In Ref. [32], a consolidated database for flow boiling in mini/micro-channels is amassed from 31 sources, of

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which 10,805 data points are designated as pre-dryout. A generalized correlation was proposed by superpositioning the contributions of nucleate boiling and convective boiling. This correlation was indicated to provide very good predictions against the entire pre-dryout database for all working fluids and all ranges of the database parameters. The heat transfer and pressure drop data for R134a at a saturation temperature was studied by Manavela Chiapero et al. [33] for two different mass fluxes and heat fluxes. In their study, the flow patterns have been characterized by means of a fast speed camera and a visualization section between the heat transfer and pressure drop measurements. Recently, experiments have been conducted by Chen el al. [34] to investigate how the imposed time periodic heat flux oscillation affects the bubble characteristics of saturated flow boiling with refrigerant R-134a in a horizontal narrow annular pipe. The results show that the bubble departure diameter, bubble frequency and active nucleation site density are found to oscillate periodically in time as well and at the same frequency as the imposed heat flux oscillation. The saturated flow boiling heat transfer of nitrogen in a vertical upward 11.9 mm inner diameter stainless steel tube was experimentally investigated by Fang et al. [35]. The measured data showed that for upward flow in vertical macro-tubes, there are two occurrences of critical heat flux along a uniformly heated channel. Effects of aspect ratio (AR) on the saturated flow boiling characteristics of deionized water in parallel rectangular microchannels are investigated experimentally by Markal et al. [36]. Detailed flow visualizations are conducted as well as temperature and pressure measurements. They indicated that heat transfer coefficient increases with an increase in the aspect ratio up to AR = 3.54 and then decreases. AR = 1.22 has been appeared as a threshold value for the heat transfer coefficient. Huang and Thome [37] conducted a comprehensive experimental study to measure the local heat transfer coefficients during flow boiling of refrigerants (R245fa and R236fa) in multi-microchannel evaporators. The experimental

Fig. 1. Schematic of experimental setup.

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results revealed that the local heat transfer coefficient decreased with the inlet subcooling, increased with the saturation temperature and with the size of the inlet orifice width. A deep study of phase-change processes in the channel flow subject to time varying heat flux is vital in the design of airconditioning and refrigeration systems. In the last years, many researches has been carried out on the two-phase flow and heat transfer of refrigerants under a constant heat flux, but the corresponding research for the time dependent heat flux remains largely unexplored. In the present study an initial attempt is made to unravel how an imposed time periodic heat flux oscillation affects the basic channel flow boiling heat transfer and associated bubble characteristics for refrigerant R-134a. Experiments will be conducted here for the saturated boiling of R-134a in a horizontal narrow annular duct. Particularly, the time dependent flow boiling characteristics affected by the amplitudes and periods of the heat flux oscillations will be examined in details.

2. Experimental method The schematic of the experimental setup was shown in Fig. 1 which was used in the previous studies [24,34]. The experimental apparatus consists of three main loops, namely, the refrigerant, water-glycol and hot-water loops, and a data acquisition system. Refrigerant R-134a is circulated in the refrigerant loop. The temperature and flow rate in the water-glycol loop was controlled to

Table 1 Summary of the uncertainty analysis. Parameter

Uncertainty

Annular duct geometry Length, width and thickness (%) Gap size (%) Area (%)

±1.0% ±5.0% ±2.0%

Parameter measurement Temperature, T (°C) Temperature difference, 4T (°C) System pressure, P (MPa)

±0.2 ±0.28 ±0.002 ±2

Mass flux of refrigerant, G (%)  (%) Amplitude of mass flux oscillation, DG=G Period of mass flux oscillation, tp (sec)  (%) Imposed heat flux, q  (%) Amplitude of mass flux oscillation, Dq=q Period of mass flux oscillation, tp (sec)

±4.8 ±0.25 ±4.5 ±0.2 ±0.25

Saturated flow boiling heat transfer Heat transfer coefficient, hr,sat (%)

±14.5

achieve enough cooling capacity for condensing the refrigerant vapor and for maintaining the refrigerant liquid at a preset temperature. The main components in the refrigerant loop include an oilfree variable-speed refrigerant pump, an accumulator, a mass flow meter, a test section, a condenser, a sub-cooler, a receiver, a filter/ dryer, and four sight glasses. An AC motor is used to control the refrigerant mass flow rate bu using the inverter frequency. The refrigerant at the outlet of the refrigerant pump is kept subcooled to avoid any vapor flow through the mass flow meter. The flow meter (Micro motion RFT9739) has an accuracy of ±1%. The preheater is used to heat the refrigerant to a specified refrigerant state before entering the test section. The vapor generated in the test section is re-liquefied in an oversized condenser/subcooler in the cold water-glycol loop. Leaving the subcooler, the liquid refrigerant flows back to the receiver at the bottom of the system. An accumulator is connected to a high-pressure nitrogen tank to dampen the fluctuations of the flow rate and pressure. The filter/dryer is used to filter the impurities and non-condensable gas possibly existing in the loop. Varying the temperature and flow rate of the waterglycol mixture flowing through the condenser and subcooler allows us to control the pressure of the refrigerant loop. Two absolute pressure transducers are respectively installed at the inlet and exit of the test section with a resolution up to ±2 kPa. All the refrigerant and water temperatures are measured by copper-constantan thermocouples (T-type) with a calibrated accuracy of ±0.2 °C. The test section is thermally insulated with a polyethylene insulation layer of 19.5 mm thick so that heat loss from it can be reduced significantly. The detailed description of the experimental system is available in the previous studies [24,34] and is not repeated here. The test section of the experimental setup is a horizontal annular duct. The outer pipe is 160-mm long and 4-mm thick with inside diameter of 20 mm and was made of Pyrex glass for visualization of boiling processes in the refrigerant flow. Both ends of the pipe are connected with a copper tube of the same size by means of flanges and are sealed by O-rings. The inner pipe with 10.0, 16.0 or 18.0 mm nominal outside diameter (wall thickness is 1.0, 1.5 or 2.5 mm) was used. The corresponding hydraulic diameter of the annular duct Dh is 10.0, 4.0 or 2.0 mm (corresponding to the gap size of 5.0, 2.0 or 1.0 mm for the duct), respectively. It is noted that the flow enters the duct long before the heated section so that the entrance effects on the boiling are small. An electric cartridge heater of 160 mm in length and 7.5 mm or 12.5 mm in diameter with a maximum power output of 800 W is inserted into the inner pipe. Furthermore, the pipe has an inactive heating zone of 10-mm long at each end and is insulated with Teflon blocks and thermally nonconducting epoxy to minimize heat loss from it. Thermal contact

Table 2 Time scales for transient R-134a saturated flow boiling (d = 1 mm). Time scales

(Sec)

Conduction time scale t1 ¼ L2c =aw   Convection time scale t2 ¼ L= qG

0.346452 T sat T sat T sat T sat T sat

l

= 15 °C, = 15 °C, = 15 °C, = 15 °C, = 10 °C,

G = 600 kg/m2 s G = 500 kg/m2 s G = 400 kg/m2 s G = 300 kg/m2 s G = 600 kg/m2 s

d

Time scale of saturated flow boiling t3 ¼ 2000ðl p=ðq Dh Þ l

Time constant t4 ¼ t c

0.331528 0.397833 0.497292 0.663056 0.336239 0.0008875

l

T sat = 15 °C, G = 600 kg/m2 s T sat = 15 °C, G = 500 kg/m2 s T sat = 15 °C, G = 400 kg/m2 s T sat = 15 °C, G = 300 kg/m2 s T sat = 10 °C, G = 600 kg/m2 s

60.2 (Single-phase) 43.3 (Saturated flow boiling) 61.6 (Single-phase) 44.4 (Saturated flow boiling) 65.9 (Single-phase) 46.2 (Saturated flow boiling) 67.65 (Single-phase) 48.9 (Saturated flow boiling) 61.6 (Single-phase) 45.2 (Saturated flow boiling)

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C.A. Chen et al. / International Journal of Heat and Mass Transfer 106 (2017) 35–46 Table 3 Time scales for transient R-134a saturated flow boiling (d = 2 mm). Time scales Conduction time scale t1 ¼

(Sec) L2c = w

0.124723

a

  Convection time scale t 2 ¼ L= qG l

T sat T sat T sat T sat

= 15 °C, = 15 °C, = 15 °C, = 10 °C,

G = 500 kg/m2 s G = 400 kg/m2 s G = 300 kg/m2 s G = 500 kg/m2 s

d

Time scale for the bubble growth in saturated flow boiling t 3 ¼ 2000ðl p=ðq Dh Þ l

l

T sat = 15 °C, G = 500 kg/m2 s

Time constants for flow boiling on heated surface t4 ¼ t c

0.397833 0.497292 0.663056 0.403487 0.001775

2

T sat = 15 °C, G = 400 kg/m s T sat = 15 °C, G = 300 kg/m2 s T sat = 10 °C, G = 500 kg/m2 s

27.7 24.0 30.1 25.5 31.1 26.3 29.0 25.1

(Single-phase) (Saturated flow (Single-phase) (Saturated flow (Single-phase) (Saturated flow (Single-phase) (Saturated flow

boiling) boiling) boiling) boiling)

Table 4 Time scales for transient R-134a saturated flow boiling (d = 5 mm). Time scales Conduction time scale t1 ¼

(sec) 0.055432

L2c = w

a

 

T sat T sat T sat T sat

Convection time scale t 2 ¼ L= q l G

= 15 °C, = 15 °C, = 15 °C, = 10 °C,

G = 300 kg/m2 s G = 200 kg/m2 s G = 100 kg/m2 s G = 300 kg/m2 s

d

Time scale of saturated flow boiling t3 ¼ 2000ðl p=ðq Dh Þ l

Time constant t4 ¼ tc

0.663056 0.994584 1.989167 0.672479 0.0044375

l

T sat = 15 °C, G = 300 kg/m2 s 2

T sat = 15 °C, G = 200 kg/m s T sat = 15 °C, G = 100 kg/m2 s T sat = 10 °C, G = 300 kg/m2 s

32.7 26.9 33.6 28.2 37.8 30.1 33.9 28.3

(Single-phase) (Saturated flow (Single-phase) (Saturated flow (Single-phase) (Saturated flow (Single-phase) (Saturated flow

boiling) boiling) boiling) boiling)

 and tp at d = 2.0 mm. Fig. 2. Time-average R-134a saturated flow boiling curves (a) and saturated flow boiling heat transfer coefficients (b) for various Dq=q

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Fig. 3. Time-average R-134a flow boiling curves (a) and flow boiling heat transfer coefficients (b) for various gap sizes.

, Dq=q  and tp at d = 2.0 mm. Fig. 4. Time variations of imposed heat flux and wall temperature at z = 80 mm for various q

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, Dq=q  and tp at d = 2.0 mm. Fig. 5. Time variations of imposed heat flux and heat transfer coefficient at z = 80 mm in persistent boiling for various q

between the heater and the inner pipe is improved by coating a thin layer of heat-sink compound on the heater surface before the installation of the heater. Then, 8 T-type calibrated thermocouples are electrically insulated by covering their beads with an electrically non-conducting thermal bond before they are fixed on the inside surface of the inner pipe so that the voltage signals from the thermocouples are not interfered by the DC current passing through the cartridge heater. The thermocouples are positioned at three axial stations along the smooth pipe. At each axial station, two to four thermocouples are placed at top, bottom, or two sides of the pipe circumference with 180° or 90° apart. The outside surface temperature Tw of the inner pipe is then derived from the measured inside surface temperature by taking the radial heat conduction through the pipe wall into account. Uncertainties of the heat transfer coefficients are estimated according to the procedures proposed by Kline and McClintock for the propagation of errors in physical measurement [38]. The results from this uncertainty analysis are summarized in Table 1. 3. Results and discussion In this work, the time periodic R-134a saturated flow boiling experiment was performed for the refrigerant mass flux G varying  from 0 to from 100 to 600 kg/m2 s, mean imposed heat flux q 45 kW/m2, and system pressure P set at 414.6 kPa and 488.6 kPa (corresponding to the R-134a saturation temperature Tsat = 10 °C and 15 °C) for the gap of the annular duct d = 1.0, 2.0 and 5.0 mm. The amplitude of the imposed heat flux oscillation Dq is

. Moreover, the perset at 0%, 10%, 30% and 50% of mean heat flux q iod of the heat flux oscillation tp is selected to be 20, 30, 60 and 120 s. Several time scales associated with the saturated flow boiling over the heated inner pipe surface of the annular duct are given in Table 2. Specifically, there are four relevant time scales in this work. The conduction time scale t 1 is estimated by considering a copper slab of the characteristic length Lc (thickness) subject to a step heat input at a certain time instant and t1 is the duration for the other surface to begin to feel the heat input. Accordingly, conduction time scale t1 can be expressed as t1  L2c =aw . The convection time scale t2 can be estimated by t2 ¼ L=ðG=ql Þ. The time scale for the bubble growth or departure t3 is approximated by the empirical expression t 3 ¼ dp =2000ðll =ðql Dh Þ [39]. Finally, the time constant of the present flow boiling on the heated copper duct surface t c is obtained by measuring wall temperature Tw subject to a step heat flux for both single-phase and boiling conditions. The results in Tables 2–4 clearly indicate that the time constants in association with the single-phase liquid flow and saturated flow boiling over the heat surface are much larger than the other time constants and hence dominate the time response of the duct wall to the heat flux oscillation. Besides, the time lag of the heated surface temperature behind the imposed heat flux oscillation is expected to be significant. Therefore, attention will be mainly paid to examining the effects of the amplitude and period of the heat flux oscillation on the time periodic R-134a saturated flow boiling heat transfer performance. Note that for the limiting case of  = 0% we have saturated flow boiling of R-134a at a constant Dq=q refrigerant mass flow rate in the test section, which is designated as ‘‘stable flow boiling’’and has been investigated by Lie and Lin

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[24]. The measured boiling heat transfer data are expressed in terms of the boiling curve and boiling heat transfer coefficient. The thermal characteristics of the time periodic flow boiling is illustrated by the time variations of the instantaneous heated pipe wall temperature and boiling heat transfer coefficient. Moreover, selected flow photos and data deduced from the images of the time periodic boiling processes are presented to show the temporal bubble characteristics in the boiling flow. 3.1. Time-average boiling curve and heat transfer coefficient The time-average boiling curves and heat transfer coefficients measured at the middle axial location (z = 80 mm) of the narrow annular duct for various amplitudes and periods of the heat flux oscillation are shown in Fig. 2 for Tsat = 15 °C and G = 500 kg/m2 s. It is clearly found that in Fig. 2 that the effects of the amplitude and period of the heat flux oscillation on the time-average boiling curves and heat transfer coefficients are not significant. In fact, they are almost the same as those of the stable boiling. The effects of the gap size of annular duct d on the time-average boiling curves are shown in Fig. 3(a). The boiling curve shifts clearly to the left as the gap size d is reduced, indicating that the boiling heat transfer in the smaller gap size is substantially significant. Besides, It is evident that lower imposed heat flux and wall superheat are needed to initiate boiling on the heated surface for the smaller gap size. This mainly results from the fact that for given  q  and T sat the mass flow rate through the duct is lower for a G, smaller gap size d. For the lower refrigerant mass flow rate the axial temperature rise of the refrigerant flow is faster, which, in turn, causes earlier bubble nucleation for a smaller gap size. Then,

the data shown in Fig. 3(b) indicate that the saturated boiling heat transfer coefficient hr increases noticeably with a decreasing gap size Since the shear stress of the flow acting on the heated surface in a smaller gap of the annular duct is higher, the bubbles on the heating surface can be more easily swept away from the heated surface. Moreover, the flow pattern change from the bubbly flow to the slug flow for smaller gap size d = 1.0 mm is noited for a lower imposed heat flux. These effects are thought to be the main reasons for the enhancement of flow boiling heat transfer when the gap size is decreased. 3.2. Time dependent flow boiling heat transfer characteristics The time periodic flow boiling heat transfer characteristics for R-134a flow in the annular duct resulting from the imposed periodic heat flux oscillation are illustrated in Figs. 4–7 by presenting the time variations of the heated wall temperature Tw and boiling heat transfer coefficients hr at the middle axial location in the sta, tp and q . Overall inspection of tistical state for various G, Tsat, Dq=q Figs. 4–7 indicates when the wall heat flux oscillates periodically in time nearly like a triangular wave, significant temporal oscillations in heated wall temperature Tw and boiling heat transfer coefficient hr nearly in the form of sinusoidal waves can occur. Note that the oscillations in both Tw and hr are also periodic in time and are at the same frequency as the imposed wall heat flux. In addition, the stronger oscillations in the Tw and hr are found for a higher amplitude and a longer period of the imposed heat flux oscillation. Moreover, the Tw oscillation is more significant for a higher mean imposed heat flux. Note that single-phase forced convection pre-

Fig. 6. Time variations of imposed mass flux, inlet pressure, and wall temperature at z = 80 mm for various gap sizes.

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Fig. 7. Time variations of imposed mass flux, inlet pressure, and heat transfer coefficient at z = 80 mm in persistent boiling for various gap sizes.

Fig. 8. Flow regime map for time periodic R-134a saturated flow boiling in annular duct at d = 2.0 mm.

Fig. 9. Flow regime map of time periodic R-134a flow boiling in annular duct for various gap sizes.

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Fig. 10. Time variations of imposed heat flux and wall temperature at z = 80 mm for various tp at d = 2.0 mm.

vails in the duct at a relatively low imposed heat flux. The data given in Figs. 4–7 further reveals that the Tw oscillation significantly lags the heat flux oscillation for both single-phase and flow boiling flows. This time lag apparently results from the thermal  , the inertia of the heated pipe wall. For given G, Tsat , and Dq=q duration of the time lag relative to the oscillation period tl =tp is slightly longer at a lower time-average imposed heat flux and a shorter period of the imposed heat oscillation. Furthermore, for  and in the first half of the the single-phase convection at low q periodic cycle in which the heat flux decreases with time, after the time lag the wall temperature is found to decrease with time. The opposite process occurs in the second half of the cycle. It is fur the temperature of the ther noted in the experiment that at a low q heating surface is lower than that required for the onset of nucleate boiling and no bubble nucleates from the heating surface in the whole periodic cycle. Hence heat transfer in the flow results completely from the single-phase forced convection. As the mean imposed wall heat flux is raised gradually, the heated surface temperature increases correspondingly. The corresponding Tw oscillation amplitude can be significant even when the single-phase flow prevails in the duct (Fig. 4(c)). At a certain higher wall temper bubbles start to nucleate from the ature for a small increase in q heated surface at certain instant of time but bubble nucleation is not seen in the whole cycle. Thus, we have intermittent boiling in the flow in which nucleate boiling only exists in a partial duration of the periodic cycle. The time instants for the start and termination of the nucleate boiling are marked on the curves for Tw for the cases with the presence of the intermittent boiling. For some increase in the mean imposed heat flux nucleate boiling persists over the entire period of the cycle and we have persistent flow boiling in the duct. The results in Figs. 4–7 also show that after

the time lag Tw and hr both decrease with time in the first half of the periodic cycle in which the imposed heat flux decreases with time. While in the second half of the cycle the opposite process occurs. Then, the data shown in Fig. 6 indicate that the Tw oscillation amplitude increases noticeably with an increase in the channel gap in the persistent flow boiling. 3.3. Intermittent flow boiling It is of interest to note that over a certain intermediate range of the mean imposed heat flux the intermittent boiling appears. More specifically, in a typical periodic cycle of the imposed heat flux oscillation at a certain time instant in the first half of the cycle as the imposed heat flux decreases to a certain low level and the heated pipe wall temperature decreases to the level below that required for the onset of nucleate boiling, bubble nucleation on the heated inner pipe disappears and boiling stops. Single-phase flow prevails in the duct. At a certain later time instant in the second half of the cycle in which the imposed heat flux increases and Tw rises to exceed the wall superheat required for the onset of nucleate boiling, bubble nucleation is seen on the heated surface and boiling appears. The boiling process continues for some time interval. The above processes of the intermittent flow boiling are repeatedly seen on the heated surface. To be clearer, we mark the time instants at which boiling starts and stops in Fig. 4 on the curves for Tw for the cases with the presence of the intermittent boiling. Note that at a higher imposed heat flux the onset of boiling is earlier and the termination of boiling is later. Besides, the instants for the boiling inception and termination are affected by the amplitude and period of the heat flux oscillation to a noticeable degree (Fig. 4). Flow regime maps to delineate the boundaries sep-

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On the other hand, when the oscillatory period is relatively long at 120 and 600 s, the data given in Fig. 10(b) and (c) manifest that the amplitude of the Tw oscillation is very large. Note that the Tw oscillation is slightly stronger for tp raised from 120 to 600 s. The stronger Tw oscillation at a longer tp can be attributed to the longer period of time for the thermal energy storing in and releasing from the wall of the heated inner pipe. This in turn causes the heated pipe wall to accumulate thermal energy over a longer period of time and hence results in a larger amplitude of the Tw oscillation. 3.5. Effect of heat flux oscillation amplitude The effect of the amplitude of the heat flux oscillation on the time periodic flow boiling is examined. The data presented in Fig. 11 for five different amplitudes of the imposed heat flux oscil = 10, 30, 50, 80 and 100% at tp = 120 s show that lation with Dq=q the Tw oscillation is significantly stronger for a larger amplitude of the imposed heat flux oscillation. Note that the time lag in Tw oscillation increases substantially when the amplitude of the imposed heat flux oscillation is raised from 10% to 30%. But a further increase in the amplitude of the heat flux oscillation does not increase the time lag noticeably. 4. Concluding remarks An experimental study was done to unveil the effects of applying a oscillating heat flux on an annulus where the working fluid is R-134a refrigerant. The effects of key parameters such as mean level and oscillation amplitude and period of the heat flux oscillation on the time periodic R-134a flow boiling have been examined in details. The main findings can be summarized as follow:

Fig. 11. Time variations of imposed heat flux and wall temperature at z = 80 mm for  at d = 2.0 mm. various Dq=q

arating different flow boiling regimes in terms of the Boiling number versus the relative amplitude of the imposed heat flux oscillation are given in Figs. 8 and 9. The results show that the intermittent boiling prevails over a significantly wider range of the Boiling number for a higher amplitude and a longer period of the imposed heat flux oscillation and for a lower refrigerant mass flux. Moreover, the intermittent boiling appears over a wider range of the Boiling number for a larger duct gap (Fig. 9).

3.4. Effect of heat flux oscillation at extremely short and long periods How the period of the heat flux oscillation affects the time periodic flow boiling is further investigated by comparing the data for relatively short and long tp in Fig. 10. Now when the oscillatory period is set to be 2.0 s, the heated surface temperature stays at a nearly constant level, as evident from the data shown in Fig. 10 (a) for the oscillation amplitude of q being 50% of the mean imposed heat flux. In fact, for this case the amplitude of Tw oscillation is smaller than 0.2 °C which is below the random thermal disturbances from the background and is close to that in the stable boiling. Due to its heat capacity, the heated copper duct is not able to quickly respond to the very rapid change in the imposed heat flux. This situation appears when the period of the imposed heat flux oscillation is significantly shorter than the dominated time constant of the flow boiling system.

(1) The time-average boiling curves and heat transfer coefficients for the time periodic flow boiling of R-134a are not affected considerably by the amplitude and period of the imposed heat flux oscillation. (2) The heated pipe wall temperature, bubble departure diameter and frequency, and active nucleation site density also oscillate periodically in time and at the same frequency as the heat flux oscillation. Results display that the resulting oscillation amplitudes of the heated surface temperature, heat transfer coefficient get larger for a longer period and larger amplitude of the imposed heat flux oscillation and for a higher mean imposed heat flux. Moreover, a significant time lag in the heated surface temperature oscillation is also noted, which apparently results from the thermal inertia of the copper inner pipe. (3) The dimensionless time lag, defined as the ratio of the time lag to the period of the imposed heat flux oscillation, is greater at a lower mean imposed heat flux and a shorter period and for a larger amplitude of the imposed heat flux oscillation when the amplitude is below 30%. (4) Due to the existence of the thermal inertia of the heated copper duct, the resulting heated surface temperature does not oscillate with time at an extremely short period of the imposed heat flux oscillation. But the oscillation amplitude of the heated surface temperature is reinforced for an extremely long period of the imposed heat flux oscillation.

Acknowledgment The financial support of this study by the Ministry of Science and Technology, R.O.C., through the contract NSC 96-2221-E-009133-MY3 is greatly appreciated.

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