Enhancement of flow boiling in meso scale channels ... - Springer Link

J. Micro-Nano Mech. (2009) 5:93–102 DOI 10.1007/s12213-010-0027-0

RESEARCH PAPER

Enhancement of flow boiling in meso scale channels with subsonic vibrations Yasir M. Shariff

Received: 22 October 2009 / Revised: 1 April 2010 / Accepted: 19 April 2010 / Published online: 29 April 2010 # Springer-Verlag 2010

Abstract Heat transfer coefficients were measured for forced convection refrigerant mixture (R-407C) flow in meso-scale heat exchangers enhanced by subsonic vibration. Flow in three horizontal meso-channels (ID 1.0 mm, 2.0 mm, and 4 mm), and a length of 60 mm for subcooled and saturated boiling conditions are reported in this study. Experiments were conducted at heat fluxes of 2 and 11 kW/m2 for subcooled boiling, and 15 and 29 kW/m2 for saturated boiling. The mass flow rate was varied from 0.45 to 1.85 kg/min. The frequency was varied, freq=0–70 Hz. An experimental setup composed of heating elements provided heat flux variations in the meso-channels and excitation elements were attached to the side of the channels to produce the subsonic vibrations. The heat transfer coefficient was found to be dependent on both the heat flux as well as mass flux levels. Results show that subsonic vibration enhanced the heat transfer performance over that of non enhanced meso-channels by 8% as compared to regular flow convective boiling process. Keywords Heat transfer . Subsonic vibration . Enhancement . Meso scale Nomenclature A internal surface area (m2) cp specific heat (J/kgK) D meso-channel diameter (mm) d micro-coil diameter (mm) e diameter ratio (d/D) f friction factor (−)

Y. M. Shariff (*) Mechanical Engineering Department, College of Engineering, Taibah University, Madina, P.O. Box 1020, Saudi Arabia e-mail: [email protected]

Freq h I iout im if,out ig,out k m Nu P p Q Q1 Q2 q q″ Re Ts T# Tin Tout v V WRe Wρ Wμ WD WNu WT x

frequency (Hz) heat transfer coefficient (W/m2K) current (amps) outlet enthalpy (kJ/kg) enthalpy (kJ/kg) saturated liquid outlet enthalpy (kJ/kg) saturated vapor outlet enthalpy (kJ/kg) thermal conductivity (W/mK) mass flowrate (kg/min) Nusselt number (−) power (W) pitch (mm) volumetric flowrate (m3/s) volumetric flowrate at entrance of channel (m3/s) volumetric flowrate at exit of channel (m3/s) heat transfer (W) heat flux (W/m2) Reynolds number (−) average surface temperature (K) temperature of # thermocouple (K) inlet temperature (K) outlet temperature (K) flow velocity (m/min) voltage (volts) uncertainty in Reynolds number uncertainty in density uncertainty in viscosity uncertainty in diameter uncertainty in Nusselt number uncertainty in Temperature vapor quality (−)

Greek ΔP pressure drop (Pa) ρ density (kg/m3) μ viscosity (Pas)

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1 Introduction The use of subsonic vibration in enhancing heat transfer characteristics for refrigerant mixtures at the meso-scale has an important impact on various daily used thermal systems. The use of subsonic vibration was investigated, but there is a need for predicting the heat transfer coefficients enhancement for subcooled and saturation boiling processes in such meso scale. Therefore, combining meso-scale channels along with ultrasonic vibration for subcooled and saturated boiling needs to be investigated in conjunction with refrigerant mixtures. The relationship between the flow behavior induced by vibration and the consequent heat transfer enhancement in natural convection and pool boiling regimes were investigated by Kim et al. Experimental results showed that the effects of vibration on flow behavior are vastly different depending on the heat transfer regime and the amount of dissolved gas. In the natural convection and subcooled boiling regimes, behavior of cavitation bubbles strongly affected the degree of heat transfer enhancement. In saturated boiling, no cavitation occured thus the reduced thermal bubble size at departure and acoustic streaming were major factors in enhancing the heat transfer rate. The highest enhancement ratio was obtained in natural convection regime where the effect of vibration is manifested through violent motion of cavitation bubbles [1]. The effects of vibration on critical heat flux (CHF) has been experimentally investigated by Jeong and Kwon under natural convection condition. Flat bakelite plates were coated with thin copper layer and distilled water were used in their experiment as heated specimens and working fluid, respectively. Measurements of CHF on flat heated surface were made with and without vibration applied to working fluid similarly to the study done but with the addition of size and working fluid impact on the heat transfer coefficient. An inclination angle of the heated surface and water subcooling were varied as well by Jeon and Kwon as they examined water subcoolings at 5°C, 20°C, 40°C and the angles are 0°, 10°, 20°, 45°, 90°, 180°. The measurements showed that wave applied to water enhances CHF and its extent is dependent upon inclination angle as well as water subcooling this particular comparison was not done as in this experimental investigation a flat horizontal setting is investigated. The rate of increase in CHF increases with an increase in water subcooling while it decreases with an increase in inclination angle. Visual observation showed that the cause of CHF augmentation is closely related with the dynamic behaviour of bubble generation and departure in acoustic field [2]. Other investigations were done by Tillery et al on twophase cooling cell based on channel boiling and a vibrationinduced cavitation jet whose collective purpose is to delay

J. Micro-Nano Mech. (2009) 5:93–102

the onset of critical heat flux by forcibly dislodging the small vapor bubbles attached to a solid surface during nucleate boiling and propelling them into the cooler bulk liquid within the cell. The submerged turbulent cavitation jet is generated by a vibrating piezoelectric diaphragm operating at resonance. The piezoelectric driver induced pressure oscillations in the liquid near the surface of the driver, resulting in the time-periodic formation and collapse of cavitation bubbles that entrain surrounding liquid and generate a strong liquid jet. The resultant jet, which was directed at a heated surface, enhances boiling heat transfer by removing attached vapor bubbles that insulate the surface while providing additional forced fluid convection on the surface. By introducing a crossflow within the cell, the heat transfer was increased even further due to the fact that the flow sweeped the bubbles downstream while keeping the temperature of the water within the cell regulated [3]. The effects of vibration on heat transfer were also investigated by Shinfuku et al in a natural convection region and a nucleate boiling region. The Nusselt number by vibration, cavitation intensity using aluminum sheet, and electric input power to oscillator were measured systematically in their study. As the height of heating elements was varied, both the profile of the electric input power and the cavitation intensity were periodically changed due to standing waves formed in the water vessel. These results agreed with the profile of the augmentation ratio estimated from the Nusselt number by vibration. The augmentation ratio for a downward facing surface was found to be greater than that for vertical surface in a natural convection region, however, for nucleate boiling, it was less than the augmentation ratio for the vertical surface. The enhancement effect became negligible in a well developed nucleate boiling region. The mechanism of heat transfer enhancement by vibration was the agitation generated by cavitation in a natural convection region, and the removal effect of the vapor bubbles by acoustic streaming in a nucleate boiling region [4]. Experimental investigations were done by Kim et al on the effects of tube vibration on critical heat flux (CHF) in order to gain an understanding of the relationship between CHF and flow-induced vibration (FIV). The experiment was carried out in the following range of parameters: diameter (D)=0.008 m; heated length (L)=0.2, 0.4 m; pressure (P)=101 kPa; mass flux (G)=403–2,551 kg/m2·s; quality (x)=−0.045–0.289; amplitude (a)=0.0001–0.001 m; frequency (f)=0–70 Hz which is similar to the presented results of this paper but differ with the working fluid. The CHF generally increases with vibration intensity, which was represented by vibrational Reynolds number (Rev); the CHF enhancement was more dependent on amplitude than on frequency. CHF enhancement seems to come from the


reinforced flow turbulent mixing effect by vibration in the vicinity of heat transfer surface. Based on the experimental results, an empirical correlation was proposed by Kim et al for the prediction of CHF enhancement by tube vibration. The correlation predicted the CHF enhancement ratio (En) with reasonable accuracy, with an average error rate of −2.18% and 27.75% for RMS [5]. In order to gain an understanding of the relationship between critical heat flux (CHF) and flow-induced vibration (FIV), an experimental investigation was carried out by Lee et al with vertical round tube at the atmosphere. In the both condition of departure from nucleate boiling (DNB) and the liquid film dryout (LFD), CHF increases up to 12.6% with vibration intensity, represented by vibrational Reynolds number (Rev). CHF enhancement by tube vibration seems to come from the reinforced flow turbulent mixing and the increment of deposition of droplet into the liquid film. Based on the experimental results presented by Lee et al, an empirical correlation was proposed for the prediction of CHF enhancement ratio. The correlation predicted the CHF enhancement ratio (En) with reasonable accuracy, with an average error rate of 4.5 and 26.5% for RMS. Vibration is an effective method for heat transfer enhancement as well as CHF. Nonetheless, the risk of system failure by FIV has made it very difficult to take advantage of vibration in heat transfer facilities. Therefore, it is necessary to find out optimal design enhancing the CHF but preventing damage in an acceptable vibration range [6].

2 Experimental set-up An experimental setup has been built for measuring heat transfer and pressure drop under various flow conditions in micro- and meso-channels. In Fig. 1, the schematic diagram of the experimental setup is shown. It consists of a storage tank, pump, by-pass loop, flowmeter, preheater, test section, VARIAC power controller, a condenser, chiller, and a data acquisition system. The setup has been designed and built such that it is flexible enough to use with different refrigerants. The pump circulates the refrigerant in the loop at a rated capacity of 3 gpm. The by-pass loop diverts most of the flow exiting the pump from entering the test section and redirects it to the condenser. The amount of refrigerant entering the test section is measured using a digital flowmeter with a capacity of 0–2,000 ml/min. Heating tape is wrapped around a pipe segment approximately 1 m long and controls the temperature of the refrigerant entering the test section. The heating tape has a capacity of 750 W at 110 V, and is supplied with a controller that changes the power from 0–100%. The temperature of the refrigerant entering the preheater section is monitored by a T-type thermocouple.

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The test section consists of the temperature/pressure measurement stations at the inlet and outlet of the minichannels, where for the temperature measurements, T-type thermocouples are used at strategic locations on the test section’s surface. The pressure at the inlet and outlet are measured using both pressure gauges and pressure transducers, where the pressure gauges has a range of 0–600 psi. The pressure transducers are made of thin film vapor deposited strain gauges having a range of 0–6.8 bar. The refrigerant enters the inlet header of the test section, which distributes the flow to the channels where the refrigerant was heated with the heating elements (five cartridge heaters were used), and exits through the outlet header of the test section, then flows towards the condenser, and next to the chiller. Four thermocouples located in an array on the test section provide information on the surface temperature of the channels. The heat input to the test section is controlled using a VARIAC with an input of 110 V / 50 Hz, and it gives a variable output of 0–110 V / 50 Hz. The output is controlled in 5 V increments, and the voltage output is measured directly using a voltmeter. For cooling the refrigerant as it exits the test section, two stations are used in the process. The first station is the condenser, which is composed of a steel shell and integrally finned copper tubes, and its function is to reduce the temperature of the refrigerant exiting the test section and the by-pass loop before it enters the chiller. Condenser water was circulated through the tubes while heated refrigerant flows through the shell of the condenser. Copper-constantan thermocouples (type-T) were used for this experiment. Thermocouple output was interfaced to a personal computer through a National Instruments Data Acquisition System. LabView was used to control the system with a software graphical interface that allows the user to select the input sensor type, data acquisition rate, filtering and gain, and analysis and storage of the data. Two test sections (Fig. 2) were micro-machined by precision cutters. The larger diameter test section was made from solid copper, and the smallest diameter test section was machined from aluminum Al-6061; and five heaters were used to heat the test sections. The length of the channels in all test sections is kept constant at 50.8 mm. There are 8 parallel channels for each test section. The advantage of using an n number of arrays in a parallel formation gives a broad spectrum of temperature averages and clearly shows which side of the test section is affected. The length of the channels in all test sections is kept constant at 50.8 mm. Copper was chosen because of its superior heat conduction characteristics, and R-407C is stable in the presence of copper over the normal operating temperature range. The channels are spaced equally across the heaters, to ensure uniform energy distribution to all channels and minimize temperature gradients. The channel surface

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J. Micro-Nano Mech. (2009) 5:93–102 Thermocouple Pressure Gauge Control Valve

N2 Pump Tank

Digital Flow meter

Pressure Relief Valve

Preheater -

Power Controller

Bypass Loop

Test Section

Condenser

Chiller

Water in

Water out

Fig. 1 Schematic diagram of the experimental set-up

temperature is measured by averaging the temperature of the four thermocouples inserted into strategic locations in the test section. Different ultrasonic vibrations were produced by the excitation element installed on the middle side of the meso-channel and was varied between 0 to 70 Hz.

3 Experimental procedure The data acquisition software, LabView, loads on the pc with a direct connection to the T-type thermocouples on different strategic locations with in the experimental loop. The circulation pump was switched on followed by the preheater and the test section heaters controlled by the VARIAC controller, this is done till it reaches 25°C. The power settings are adjusted according to the levels of heat fluxes desired. The flowrate was adjusted to the desired flowrate using the needle valves next to the digital flowmeter. The preheater is adjusted so that the inlet temperature of the refrigerant to the test section was about 10°C below saturation temperature of the refrigerant (∼32°C @ atm P). The process is carefully monitored to reach a steady state at which time data can be logged into the file. The data

recorded into the spreadsheet file generates a list for the current surface, inlet, outlet temperatures which calculates all relevant parameters of the refrigerant such as heat input, heat removed, Reynolds number, heat transfer coefficient, and Nusselt number. The flow instability somewhat increased at higher flow rates when data are taken by keeping the flow rate constant and varying the power supply to the test section. This is due to pressure fluctuations and as were different from regular flow in the same test section. The power input is increased until the critical heat flux condition is encountered, as indicated by a sudden and large increase in surface temperature.

4 Error and uncertainty analysis A major emphasis of experimental error is based on critical measurements of flow-rate and temperature. The flowmeter is calibrated based on actual flow rates, where the calibration curve has a confidence level of 95% with tolerance limits of +/−0.0027 kg/s. It should also be noted that there is a component of human error, which can be considered insignificant and that it is completely unknown. The thermocouples used in the experiment have been calibrated at the ice point (0°C). However, channel surface


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Fig. 2 a Schematic of the test section. b Detailed schematic of the meso/micro channels

(a)

(b)

Heat Sink Side View

Thermocouples

25.4 mm

25.4 mm

D 5 mm

10 mm

Flow Direction

Heaters L = 50.8 mm Excitation element at midway location

temperatures at each station as measured by four different thermocouples were found to vary within a range of +/−1°C. The best estimate of the true value is taken as the arithmetic mean of these four values. A typical set of data is analyzed for precision of measurement. Precision of the temperature measurement is found to be of the order of 0.42°C. In addition, the errors considered are for convective heat transfer calculations, therefore, conduction and radiation effects are not considered, where conduction was not involved in the measured part of the test section as the variations in measured temperatures between the mixed refrigerant flow and the surface of the channels, similarly, radiation was not involved in the measurements and the experiment was done indoors in a laboratory environment and ignoring such effects does not have a large impact on the final results. The calculation of uncertainty in the Reynolds number is found to be WRe =14 at Re=541. Using the following uncertain values: density, Wρ =7.8 kg/m3, velocity, Wv = 3.18×10-3 m/s, viscosity, Wμ =11×10−6 Pa.s, diameter, WD = 7.95×10−6 m. The calculation of uncertainty in the heat transfer coefficient was found to be Wh =19.6 W/m2K, with uncertainty values as follows: q, Wq = 0.098 W, and

2 mm

temperature, WT =0.01°C. The calculation of uncertainty in Nusselt number was found to be WNu =0.54. Finally, the uncertainty in pressure drop is the same as the uncertainty of the pressure transducers, that Was, 68.9 Pa (0.01 psi).

5 Reduction of data Test section surface temperatures were measured at four locations. Three of the thermocouples were located on the top and between of the channels and the fourth was located above the middle heater on the side of the test section. The average surface temperature was obtained by taking the arithmetic average of the thermocouple readings as follows Ts ¼

T1 þ T 2 þ T3 þ T4 4

ð1Þ

The power input (electrical power) to the test section was calculated based on the voltage and current measurements. For a purely resistive load the power factor for AC voltage is 1 P ¼ VI

ð2Þ

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The mass flow rate of the refrigerant was calculated using the flow rates as measured by the flow meters as follows m ¼ rðQ1 þ Q2 Þ

ð3Þ

The energy removed by the refrigerant was calculated using the simple expression q ¼ m cp ðTout Tin Þ

ð4Þ

And the heat transfer coefficient, based on the channel surface area A, and the average refrigerant temperature, can be calculated as q

h¼ A Ts ðTout2þTin Þ

The outlet enthalpy was compared with the saturated enthalpy at the outlet temperature to compute the vapor quality from the equation, x¼

iout if ;out ig;out if ;out

ð9Þ

where if,out was the saturated liquid enthalpy at the test section outlet temperature, and ig,out was the saturated vapor enthalpy at the same temperature. Reynolds number was calculated for illustrations of the pressure drop and friction in the flow from the equation, Re ¼ v D=m

ð10Þ

ð5Þ 6 Results and discussion

The average Nusselt number can then be calculated from the heat transfer coefficient, the channel diameter and thermal conductivity as Nu ¼

hD k

ð6Þ

The heat flux was calculated based on the internal surface area of the meso channel as follows q00 ¼

q A

ð7Þ

The outlet enthalpy was calculated using the electric power input as follows iout ¼ im þ

P m

ð8Þ

Fig. 3 Effect of vibration levels on heat transfer in mesochannels as compared to correlations for non-enhanced channels at q″=2 kW/m2

2000 1800

Experiments were carried out as described earlier and the results were reduced to provide heat transfer coefficient and heat flux characteristics for subcooled and saturated boiling processes. These results are discussed in the following two sections. 6.1 Subcooled boiling Experiments were performed over a range of mass flow rates inside horizontal channels. Subcooled tests were performed on three test sections for three types vibration levels (15, 45, and 70 Hz) per channel size. The heat fluxes for the subcooled boiling experiments were maintained at 2 and 6 kW/m2. Figures 3 and 4 show the data trends for different diameter ratios for each channel tested and the two

70HZ 45Hz 15Hz

1600

h (W/m2K)

1400 1200 1000 800 600 400 200 0 0.000

0.200

0.400

0.600

0.800

1.000

1.200

m (kg/min)

1.400

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1.800

2.000


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Fig. 4 Effect of vibration levels on heat transfer in mesochannels with a constant heat flux of q″=6 kW/m2

900

70Hz 45Hz

800

15Hz

700

h (W/m2K)

600 500 400 300 200 100 0 0.000

0.200

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m (kg/min)

values of heat flux. An increase in the value of the heat transfer coefficient with the mass flow increase is observed at both values of heat flux (q″=2, and 6 kW/m2), with the heat transfer coefficient increasing dramatically at a mass flux of 0.8 kg/min. Τhe data is plotted for a constant value of heat flux and a mass flow ranging from 0.48 to 1.85 kg/min. It is seen that the heat transfer coefficient increases at higher mass fluxes. In Fig. 4, the experimental heat transfer coefficient is plotted at different values of vibration levels (15, 45, and 70 Hz) at a heat flux of 6 kW/m2. The heat transfer coefficient for the constant heat flux case of 2 kW/m2 (Fig. 3) ranged from 210 to 440 W/m2K at lower mass fluxes ranging from 0.45 to 0.85 kg/min and 480 to 1,200 W/m2K for higher mass fluxes ranging from 0.85 to 1.85 kg/min. For a heat Fig. 5 Effect of vibration levels on heat transfer in mesochannels with a constant heat flux of q″=15 kW/m2

1800

flux of 6 kW/m2 the heat transfer coefficient ranged between 230 and 520 W/m2K at lower mass fluxes ranging from 0.5 to 0.9 kg/min and 520 to 1,200 W/m2K for higher mass fluxes ranging from 0.9 to 1.9 kg/min. The figures also illustrate the relative influence of the various parameters on the thermal hydraulic characteristics, including the impact due to the channel size D. Comparing Figs. 3 and 4, we can notice that increasing the heat flux from 2 to 6 kW/m2 affects the heat transfer coefficient. The heat transfer coefficient for large heat flux increases up to 11% the value for smaller heat flux values. In Figs. 3 and 4, the vibration levels (15, 45, and 70 Hz) effects on heat transfer coefficient is illustrated. The heat flux applied was kept constant at 2 kW/m2 for Fig. 3, and 4 kW/m2 for Fig. 4. It is observed that increasing the

70Hz 45Hz

1600 15Hz 1400

h (W/m2K)

1200

1000

800

600

400

200

0 0.000

0.200

0.400

0.600

0.800

1.000 m (kg/min)

1.200

1.400

1.600

1.800

2.000

100


Fig. 6 Effect of vibration levels on heat transfer in mesochannels with a constant heat flux of q″=29 kW/m2

2000 1800

70Hz 45Hz 15Hz

1600

h (W/m2K)

1400 1200 1000 800 600 400 200 0 0.000

0.200

0.400

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m (kg/min)

vibration levels, increases the heat transfer coefficient at certain mass flow values. Increasing freq., increases the heat transfer coefficient up to 11%. It can also be found comparing Figs. 3 and 4 that increasing the heat flux from 2 to 6 kW/m2 increased the heat transfer coefficient by an average of 5%. 6.2 Saturated boiling Convective saturated flow boiling experiments were conducted on the three horizontal test sections for three vibration levels (15, 45, and 70 Hz) per channel size over the same mass flow range (0.45 to 1.85 kg/min) as in subcooled boiling tests. The heat fluxes for the saturated boiling experiments were maintained at 15 and 29 kW/m2. Figures 5 and 6 show the data trends at different vibration Fig. 7 Vibration levels effect on heat transfer in meso-channels at q″=15 kW/m2

2500

levels for each channel tested and two different values of heat flux. An increase in the value of the heat transfer coefficient with a mass flow increase is observed at both values of heat flux (q″=15, and 29 kW/m2), with the heat transfer coefficient increasing dramatically at mass flux of 0.6 kg/min due to the increase in flow causing a mixing effect with increase freq. Τhe data is plotted for a constant value of heat flux and mass flow ranging from 0.48 to 1.85 kg/min. It is seen that the heat transfer coefficient increases at higher mass fluxes, and decreasing the diameter of a meso-scale heat exchanger generally could have an increases in the heat transfer coefficient of the refrigerant mixture. In Fig. 5, the experimental heat transfer coefficient is plotted at different values of freq. at a heat flux of 15 kW/m2. The heat transfer coefficient for a constant heat flux of 15 kW/m2 ranged from 480 to 530 W/m2K at lower mass

70Hz 45Hz 15Hz

2000

h (W/m2K)

1500

1000

500

0 0.000

0.200

0.400

0.600

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1.000 m (kg/min)

1.200

1.400

1.600

1.800

2.000


101

Fig. 8 Vibration levels effect on heat transfer in meso-channels at q″=29 kW/m2

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70Hz 45Hz 15Hz

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h (W/m2K)

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1000

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0 0.000

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m (kg/min)

fluxes ranging from 0.45 to 0.85 kg/min and 530 to 1,630 W/m2K for higher mass fluxes ranging from 0.85 to 1.85 kg/min. For a heat flux at 29 kW/m2 the heat transfer coefficient ranged between 520 and 810 W/m2K at lower mass fluxes ranging from 0.45 to 0.85 kg/min and 830 to 1,610 W/m2K for higher mass fluxes ranging from 0.85 to 1.85 kg/min. The figures also illustrate the relative influence of the various parameters on the thermal hydraulic characteristics, including the impact due to the channel size. Comparing Figs. 5 and 6, we can notice the influence of increasing the heat flux from 15 to 29 kW/m2 on the heat transfer coefficient. The heat transfer coefficient increases in value for the lower heat flux ranging from 0.45 to 0.85 kg/min. Fig. 9 Comparison of Reynolds number effect on pressure drop

Comparing both figures for freq. values and the influence of the heat transfer coefficient increase to channel size indicates that larger channel diameter (3.0 mm) and increasing heat flux from 15 to 29 kW/m2 increased the heat transfer coefficient by an average of 8%, where for smaller diameters (1.0 mm); the heat transfer coefficient increased by an average of 25%. Figures 7 and 8 shows the influence of the vibration levels (frequency) on the heat transfer coefficient. The heat flux applied was kept constant at 15 kW/m2 for Fig. 7, and 29 kW/m2 for Fig. 8. It is observed that increasing freq. also increases the heat transfer coefficient. Increasing freq. from 15 to 45 increased the heat transfer coefficient up to an average of 16%. It can also be found comparing Figs. 7

6000 1.0mm 2.0mm 3.0mm 5000

ΔP (Pa)

4000

3000

2000

1000

0 0

100

200

300

400

500 Re

600

700

800

900

1000

102

and 8 that increasing the heat flux constant from 15 to 29 kW/m2 increased the heat transfer coefficient by 9% at lower mass flow rates and 15% at higher mass fluxes. This increase in the heat transfer coefficient is due to the ultrasonic vibration effects on the meso/micro channels, which causes a mixing effect especially for lower mass flow rates. Therefore, size of the channel clearly affects heat transfer coefficient as well as the heat flux level. It can be seen also that the increase in heat transfer coefficient is maintained over the entire mass flow range. Some of the increase in the heat transfer coefficients was due to pressure drop in the channels and also due to flow variations. In Figs. 4, 5, 6, 7, and 8 it is shown that heat transfer coefficient decreases over the range of mass flow rate of 0.4 to 1.4 kg/min, and this is due to the transformation of the flow from laminar to pre turbulent. The pressure drop versus the Reynolds number is shown in Fig. 9 for all test sections. There is a transition region from Re= 250 to 500 with pressure drop initially decreasing and then increasing with increasing Reynolds number. Another inference is that the pressure drop is decreasing with decreasing diameter size for a given Reynolds number.

7 Conclusions This paper outlines heat transfer characteristics of subcooled boiling and saturated boiling of the refrigerant mixture R407C in meso-channels. The experimental data trends indicate that the heat transfer coefficients are larger for R-407C in smaller diameter test sections. The ‘h’ values are larger by 8 to 12% for the 1.0 mm channel as compared to these of the 3.0 mm channel. The mass flow rate has a positive influence on the ‘h’ values, and more so near the onset of nucleate boiling. The heat fluxes for the three channels are also noticeably higher. Furthermore, it also can be concluded that increasing heat flux by 60% increases the heat transfer


coefficient by approximately 5% for the subcooled boiling process, but increasing heat flux by 50% increases the heat transfer coefficient by 12% for the saturated boiling process. Therefore, using smaller size channels and the heat flux level affects the heat transfer coefficient in the saturated boiling process more than in the subcooled boiling process. Enhancement of heat transfer was also studied by implementing vibration levels (15, 45, and 70 Hz) as the diameter of the meso-channel was also varied. Increasing freq. from 15 to 45 Hz increased the heat transfer coefficient up to 8% for the subcooled process, where increasing freq. from 45 to 70 Hz increased the heat transfer coefficient up to 12% for the saturated boiling process. Using larger gradients in the freq. values with smaller diameters will increase the heat transfer coefficient, as we can find an optimum value of freq. but to a limit that freq. gradient versus channel sizes.

References 1. Kim HY, Kim YG, Kang BH (2004) Enhancement of natural convection and pool boiling heat transfer via ultrasonic vibration. Int J Heat Mass Transfer, Elsevier 2. Jeong JH, Kwon YC (2006) Effects of ultrasonic vibration on subcooled pool boiling critical heat flux. Heat Mass Transf 42:1155–1161, Springer-Verlag, 2006 3. Tillery SW, Heffington S, Smith MK, Glezer A (2004) Boiling heat transfer enhancement by submerged vibration induced jets, Thermal and Thermomechanical Phenomena in Electronic Systems, 2004. ITHERM apos;04. The Ninth Intersociety Conference on 2 (1–4), June 2004 4. Shinfuku N, Koichi M, Hiroaki S (1999) The mechanism for ultrasonic enhancement of natural convection and nucleate boiling heat transfer. Memoirs of the Faculty of Engineering, Ehime University Journal 18:37–45, Code: F0035A, ISSN: 0285-6107 5. Kim DH, Lee YH, Chang SH (2006) The effect of vibration on critical heat flux in a vertical round tube. Nucl Eng Des, Elsevier 6. Lee YH, Kim DH, Chang SH (2004) An experimental investigation on the critical heat flux enhancement by mechanical vibration in vertical round tube. Nuclear Engineering and Design, Elsevier