Effects of spray axis incident angle on heat transfer ... · Effects of spray axis incident angle...

Journal of Mechanical Science and Technology 26 (3) (2012) 681~688

www.springerlink.com/content/1738-494x DOI 10.1007/s12206-011-1232-z

Effects of spray axis incident angle on heat transfer performance of rhombus-pitch

shell-and-tube interior spray evaporator† Ru-Li Lin, Tong-Bou Chang* and Chih-Chang Liang

Department of Mechanical Engineering, Southern Taiwan University, Tainan, Taiwan

(Manuscript Received January 30, 2011; Revised October 6, 2011; Accepted November 24, 2011)

----------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------

Abstract An interior spray method is proposed for enhancing the heat transfer performance of a compact rhombus-pitch shell-and-tube spray

evaporator. The experimental results show that the shell-side heat transfer coefficient obtained using the proposed spray method is sig-nificantly higher than that achieved in a conventional flooded-type evaporator. Four different spray axis incident angles (0°, 45°, 60° and 75°) are tested in order to investigate the effect of the spray inclination angle on the heat transfer performance of the spray evaporator system. It is shown that the optimal heat transfer performance is obtained using a spray axis incident angle of 60°.

Keywords: Rhombus-pitch tube bundle; Shell-and-tube evaporator; Spray axis incident ---------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------- 1. Introduction

Spray evaporation heat exchangers have two major advan-tages compared to conventional flooded evaporators as re-ported by Nakayama et al. [1]. First, they achieve a better heat transfer performance than pool boiling. Second, they enable the chiller refrigerant inventory to be reduced by 20 ~ 90%. As a result, spray evaporation systems are widely used throughout various industries, including wineries, steel manu-facturing, thermal management of electronics, and so on.

Spray cooling heat transfer was first investigated by Hodg-son and Sutherland [2] in 1968. In their experiments, the tem-perature of the heated surface was higher than the Leidenfrost temperature, and thus film boiling was observed. Since their seminal study, many other researchers have also investigated the complex mechanisms associated with spray heat transfer. For example, Choi and Yao [3] investigated the film boiling heat transfer mechanisms of horizontal impacting sprays and found that the heat transfer performance was determined pri-marily by the liquid mass flux. Pais et al. [4] showed that the impact of the refrigerant droplets in spray cooling caused a break-up of the liquid film on the heated surface, and therefore allowed the vapor bubbles to be more easily released.

In a refrigeration cycle, an expansion process is required to reduce the pressure of the refrigerant to a level at which evaporation can take place. In practice, the resulting pressure

drop can be used to drive a liquid spray. However, spray evaporation is seldom applied in shell-and-tube heat exchang-ers since the coolant can not impact directly on the lower sur-faces of the horizontal tubes, and thus the overall heat transfer performance is poorer than that of pool boiling (flooded evaporator). Moeykens and Pate [5] used four overhead noz-zles to spray coolant on a triangular-pitch low-finned tube bundle containing four rows of tubes. They found that the heat transfer performance of the lower three rows was significantly poorer than that of the upper row since the sprayed liquid droplets were unable to impact the lower tubes directly and therefore the tubes dried out on their lower surfaces. Mo-eykens et al. [6, 7] and Chyu et al. [8-10] investigated the ef-fect of the spray flow rate on the heat transfer performance of overhead shell-and-tube spray evaporator systems with trian-gular-pitch and square-pitch tube bundles, respectively. The results showed that the dry-out effect could be postponed slightly by increasing the spray flow rate. Chang et al. [11, 12] showed that the cooling performance of the tubes in the lower rows of a shell-and-tube spray evaporator could be improved by fitting liquid collectors to the underside of every tube in the bundle. Chang and Chiou [13] examined the boiling heat transfer mechanism within the gap between the liquid collec-tors and the tube surface, and found that a smaller gap size improved the heat transfer coefficient at lower surface heat fluxes, but reduced the heat transfer performance at higher surface heat fluxes. Ribatski and Jacobi [14] reviewed the experimental parameters affecting the heat transfer perform-ance of spray and falling-film evaporation in tube bundles. Chang et al. [15, 16] recently showed that the dry-out problem

† This paper was recommended for publication in revised form by Associate EditorKwang-Hyun Bang

*Corresponding author. Tel.: +886 6 2533131, ext:3533, Fax.: +886 6 2425092 E-mail address: [email protected]

© KSME & Springer 2012

682 R.-L. Lin et al. / Journal of Mechanical Science and Technology 26 (3) (2012) 681~688

in compact shell-and-tube evaporators can be prevented by positioning the liquid sprays within the bundle interior rather than above the upper row of tubes.

In the studies described above, the spray axis was perpen-dicular to the heater surface. However, Shedd and Pautsch [17] developed a full-coverage spray drainage system in which the sprays were inclined in order to achieve an orienta-tion-independent removal of the heat flux. Mudawar and Estes [18] examined the effects of the cone angle and nozzle height in the spray cooling of a square surface, and showed that the critical heat flux (CHF) varied as a function of the volumetric flow distribution on the heater surface. Schwarzkopf et al. [19] studied the effect of the spray inclination angle on the thermal management of electronics. The results showed that the CHF increased slightly as the spray inclination angle was increased from 0° to 40° , but fell significantly as the inclination an-gle was increased beyond 40° . Slik et al. [20] conducted spray cooling heat flux measurements on three enhanced sur-faces at different spray inclination angles. The results indi-cated that an inclined spray axis produces a net momentum flux which sweeps any stagnant pooled regions off the surface and therefore mitigates the effects of such regions in prevent-ing the cooling spray from removing heat flux from the sur-face.

The present study applies the interior spray concept pro-posed in Ref. [15] to enhance the heat transfer performance of a compact rhombus-pitch shell-and-tube spray evaporator. In the proposed approach, the nozzle tubes are positioned in such a way that the surface of each heater tube is sprayed simulta-neously by four cooling sprays. In performing the experiments, four different spray axis incident angles are considered (i.e. 0°, 45°, 60° and 75°) in order to identify the spray inclination angle which optimizes the heat transfer performance of the evaporator system.

2. Experimental apparatus

Fig. 1 presents a schematic illustration of the experimental setup. As shown, the setup comprises a refrigerant flow loop, two cooling water loops, a test section, and a data acquisition system. Note that the setup enables the shell-side heat transfer

coefficients of the tube bundle to be measured in either a spray evaporation mode or a pool boiling mode.

2.1 Refrigerant flow loop

The refrigerant flow loop forms a closed circuit and incor-porates various flow meters, RTD temperature sensors, type T thermocouples and pressure sensors. During the experiments, the flow rate of the spray was regulated by means of a bypass valve installed on the outlet side of the refrigerant pump and was measured using a floating-type flow meter at the inlet of the test section. The temperature and pressure of the inlet liq-uid were also measured using a resistance temperature detec-tor (RTD) and pressure sensor, respectively. In the test section, the liquid refrigerant was sprayed onto the heated tubes by the nozzle tubes and partially evaporated. The evaporated refrig-erant vapor was passed through a water-cooled condenser such that it returned to a liquid state and was then combined with the non-evaporated liquid refrigerant collected from the base of the test section.

The spray tests were conducted using R141-b HCFC refrig-erant. R141-b boils at just 32°C under atmospheric pressure, i.e. only slightly higher than room temperature. As a conse-quence, the power required to achieve evaporation is relatively small and the system can be maintained at a low pressure. In other words, the experimental setup has both a low operational cost and a high operational safety.

2.2 Test section

The test section was fabricated from stainless steel, and had the form of a hollow cylindrical tube with a length of 500 mm, an internal diameter of 300 mm and a thickness of 10 mm (see Fig. 2). The test section was sealed at either end by two circu-lar plates. The nozzle tubes were rigidly attached to the left-hand plate and the heater tubes were rigidly attached to the right-hand plate. Two view-glass ports were installed on the front and rear side surfaces of the test section, respectively, to permit photographic observations. In addition, an outlet pipe was installed in the base of the test section to enable the col-lection of the unevaporated refrigerant. Finally, a thermocou-ple was installed within the test section to measure the refrig-erant vapor temperature and the stability of the vapor pressure was measured using a pressure sensor.

Fig. 1. Layout of experimental system.

Fig. 2. Schematic illustration of test section.

R.-L. Lin et al. / Journal of Mechanical Science and Technology 26 (3) (2012) 681~688 683

2.3 Tube bundle

The left hand side of Fig. 3 shows the diagrammatic illustra-tion of the heater and nozzle tubes arrangement in the rhom-bus-pitch shell-and-tube internal spray evaporator system, which developed in the investigation. Note that in this figure, the circles with the annotation「N」are nozzle tubes, while the circles with oblique lines are heater tubes. Due to budget constraints, the experimental tube bundle comprised just four heater tubes and nine nozzle tubes, as shown in the right-hand schematic in Fig. 3. The nozzle tubes were designed in such a way as to spray four individual sprays. As a result, each heater tube was sprayed simultaneously by four neighboring nozzles. Consequently, the dry-out phenomenon was prevented and the heat transfer performance correspondingly improved.

The heater tubes comprised a copper outer sleeve and an inner resistance-type cartridge heater. Each tube had a length of 300 mm and a diameter of 19.05 mm, and was capable of generating a maximum heat flux of 220 kW/m2. The tempera-ture distribution on each heater tube was measured using four thermocouples installed at intervals of 90° on the heater sur-face. The four nozzles in each nozzle tube had the form of full-cone circular hydraulic nozzles with an orifice diameter of 0.9 mm and a cone angle of 90°. In performing the spray evaporation tests, four different spray axis incident angles were considered, namely 0°, 45°, 60° and 75°. As shown in Fig. 4, four different types of nozzle tube, each with a differ-ent length, were therefore fabricated in order to ensure that the spray jets were incident upon the same region of the heater tube in every case (note that the nozzle height was specified as a constant 13.5 mm). With these design guidelines, the length with the angle of incidence 0°, 45°, 60° and 75° from spray nozzle to nozzle tube bottom are 133.87 mm, 112.44 mm, 96.13 mm and 54.04 mm, respectively.

2.4 Data acquisition system

The output signals from the various sensors (i.e. flow me-ters, RTD temperature sensors, type-T thermocouples, power meter and pressure sensors) were acquired continuously, con-verted into digital form, displayed in real-time on a monitor, and then saved to disk for further processing.

3. Experimental procedure

As described in the previous section, the spray flow rate was controlled by an adjustable bypass valve. To ensure the reliability of the measurement results, each data point was obtained by averaging a minimum of 16 data-acquisition scans, where each scan was performed only after steady-state condi-tions had been attained (as indicated by a variation in the sys-tem saturation temperature of less than 0.1°C /min).

The average shell-side heat transfer coefficient, h, was cal-culated in accordance with Newton's cooling law, i.e.

''

w s

qhT T

=−

(1)

where

q" ＝ wall heat flux = total cartridge heater power / total area of heating surface (W/m2) ,

Tw ＝ average wall temperature of tube bundle (°C), sT ＝ saturation temperature (°C).

The average wall temperature was taken as the arithmetic

mean of the sixteen thermocouple readings, i.e.

,1 ,2 ,3 ,4 ,1 ,2 ,3 ,4

,1 ,2 ,3 ,4 ,1 ,2 ,3 ,4

16A A A A B B B Bw

C C C C D D D D

T T T T T T T TT

T T T T T T T T+ + + + + + +⎛ ⎞

= ⎜ ⎟⎜ ⎟+ + + + + + + +⎝ ⎠

(2)

where subscripts A, B, C and D denote the four heater tubes in the tube bundle, and subscripts 1, 2, 3 and 4 denote the four

Fig. 3. (a) Heater and nozzle tube arrangement in rhombus-pitch tube bundle; (b) Schematic of tube bundle used in current evaporation tests.

Fig. 4. Schematic illustration of spray angle.


thermocouples located on the top, sides and lower surface of each heater tube. Since the liquids are sprayed on these loca-tions, however, the temperatures at these locations will be slightly lower than locations away from them.

3.1 Instrumentation accuracy

To assure the accuracy of the measurement data, each measuring device was calibrated prior to use. The accuracy of each calibrated device is listed in Table 1.

The experimental uncertainty of the average shell-side heat transfer coefficient ( hδ ) was calculated using the propaga-tion-of-error method [21], i.e.

1/ 22 2 2

"" w s

sw

h h hh q T Tq TT∂ ∂ ∂δ δ δ δ∂ ∂∂

⎧ ⎫⎡ ⎤ ⎡ ⎤ ⎡ ⎤⎪ ⎪= + +⎨⎢ ⎥ ⎢ ⎥ ⎢ ⎥ ⎬⎢ ⎥ ⎢ ⎥ ⎢ ⎥⎪ ⎪⎣ ⎦ ⎣ ⎦ ⎣ ⎦⎩ ⎭

(3)

where

"qδ = uncertainty of the heater power, wTδ = uncertainty of the heater surface temperature, sTδ = uncertainty of the saturation temperature.

Based on the calibration results, the uncertainty of the aver-

age shell-side heat transfer coefficient was calculated to be around ± 6.5%.

4. Results and discussion

To confirm the accuracy of the measurement system, the measurement results acquired in a series of pool boiling ex-periments were compared with those derived using the pool boiling correlation presented by Cooper [22], i.e.

0.67 0.5 0.55

1090( ) ( log )nnb cr crh q M P P− −′′= − (4)

where 100.12 0.2log pn R= − , in which Rp is the micro-roughness of the tube surface and has units of μm.

The surface roughness of the heater tubes in the experimen-tal test section was found to vary in the range Rp= 0.2 ~ 0.8 μm (as measured by an SJ-401 coordinate measurement ma-chine). The other parameters used in Cooper’s correlation are summarized in Table 2.

Fig. 5 illustrates the experimental results obtained for the variation of the pool boiling heat transfer coefficient with the heat flux on the heater surface. The figure also shows the pre-dicted heat transfer coefficients obtained from Cooper’s correla-tion for surface roughness values of Rp = 0.2 and 0.8 μm, re-

spectively. It can be seen that the experimental data fall within the predicted range at all values of the heat flux. Therefore, the accuracy of the experimental measurement system is confirmed.

Fig. 6 compares the heat transfer performance of the pro-posed interior spray method with the equivalent pool boiling

Table 1. Calibration results.

Instrument Accuracy after calibration

RTD, TC (Thermocouple) ± 0.2°C

Power meter ± 0.1%

Flow meter ± 0.1 l/min

Table 2. Parameters used in Cooper’s correlation.

M Molecular weight 117

Pc Critical pressure 4.12 MPa

P Saturation pressure (at 20 °C) 0.065 MPa

Pcr P/Pc 0.01578

Fig. 5. Comparison of measured single tube pool boiling data with predicted values obtained using Cooper’s correlation [22].

Fig. 6. Comparison of interior spray heat transfer performance with equivalent pool boiling results.


results (note that the spray axis incident angle is equal to 0°). It is evident that the spray cooling method achieves a far better heat transfer performance than the pool boiling mode at all values of the surface heat flux. Fig. 7 shows the variation of the mean shell-side heat transfer coefficient with the wall heat flux for two different spray axis incident angles (0° and 45°) and three different spray flow rates (1.4 l/min, 2.0 l/min and 2.6 l/min). Note that the saturation temperature is equal to 16°C in every case. It is seen that for all values of the heat flux and spray flow rate, the heat transfer coefficient obtained us-ing a spray axis incident angle of 45° is higher than that ob-tained using a spray axis incident angle of 0°. At a high sur-face heat flux (1.4x105 W/m2), the heat transfer coefficients for a spray axis incident angle of 45° are found to be 3%, 6% and 8% higher than those obtained for a spray axis incident angle of 0° for spray flow rates of 1.4 l/min, 2.0 l/min and 2.6 l/min, respectively. The difference in the heat transfer per-formance when using different incident angles can be attrib-uted to the size of the stagnation region on the heater surface in each case. Specifically, when the droplets are sprayed in a normal direction, any droplets which bounce off the heater surface also travel in a normal direction. As a result, some of the droplets which are sprayed later collide with these rebound droplets, causing a stagnation zone to form on the heater sur-face. Consequently, the heat transfer coefficient is reduced. In the case of a spray axis incident angle of 45°, the stagnation zone is significantly smaller (and the heat transfer coefficient correspondingly higher) since the rebound droplets travel in a different direction from the inbound droplets, and thus colli-sions do not occur. For a spray axis incident angle of 45° and a low heat flux (q"= 3× 410 2 W/m ), the heat transfer coefficient obtained at a spray flow rate of 2.6 l/min is 10% higher than

that obtained at a spray flow rate of 2.0 l/min, which in turn is 25% higher than that obtained at a spray flow rate of 1.4 l/min. However, at a higher heat flux (q"= 1.4× 510 2 W/m ), the heat transfer coefficient obtained at a spray flow rate of 2.6 l/min is 35% higher than that obtained at a spray flow rate of 2.0 l/min, which in turn is 50% higher than that obtained at a spray flow rate of 1.4 l/min. In the other words, the efficacy of a higher spray flow rate in enhancing the heat transfer performance increases with an increasing surface heat flux.

Fig. 8 illustrates the variation of the mean shell-side heat transfer coefficient with the surface heat flux for spray axis incident angles of 0° and 60° and spray flow rates of 1.4 l/min, 2.0 l/min and 2.6 l/min, respectively (note that the saturation temperature is 16°C in every case). It is seen that the heat transfer coefficient obtained using a spray axis incident angle of 60° is higher than that obtained using a spray axis incident angle of 0° for all values of the surface heat flux and spray flow rate. At a high surface heat flux (1.4x105 W/m2), the heat transfer coefficients obtained for a spray axis incident angle of 60° are 8%, 10% and 11% higher than those obtained for a spray axis incident angle of 0° at spray flow rates of 1.4 l/min, 2.0 l/min and 2.6 l/min, respectively. These percentage values are higher than those in Fig. 7. In other words, a spray axis incident angle of 60° yields a greater improvement in the heat transfer performance than a spray axis incident angle of 45°.

Fig. 9 shows the variation of the mean shell-side heat trans-fer coefficient with the surface heat flux for spray axis inci-dent angles of 0° and 75° and spray flow rates of 1.4 l/min, 2.0 l/min and 2.6 l/min, respectively (note that the saturation tem-perature is again equal to 16°C). In contrast to the results pre-sented in Figs. 7 and 8, which show that the heat transfer per-formance improves as the spray axis incident angle is in-

4.0x104 8.0x104 1.2x105 1.6x1052.0x104 6.0x104 1.0x105 1.4x105

0.0x100

1.0x104

2.0x104

3.0x104

4.0x104

:incident angle=45o

:incident angle=0o

Ts=16oC

m=2.6(l/min)

m=2.0(l/min)

m=1.4(l/min)

q"(w/m2)

h(w

/m2 -k

)

Fig. 7. Variation of heat transfer coefficient with surface heat flux asfunction of spray incident angle (0° and 45°) for spray flow rates of 1.4l/min, 2.0 l/min and 2.6 l/min and a constant saturation temperature of Ts=16 °C.

4.0x104 8.0x104 1.2x105 1.6x1052.0x104 6.0x104 1.0x105 1.4x105

0.0x100

1.0x104

2.0x104

3.0x104

4.0x104

:incident angle=60o

:incident angle=0o

Ts=16oC

m=2.6(l/min)m=2.0(l/min)

m=1.4(l/min)

q"(w/m2)

h(w

/m2 -k

)

Fig. 8. Variation of heat transfer coefficient with surface heat flux as function of spray incident angle (0° and 60°) for spray flow rates of 1.4 l/min, 2.0 l/min and 2.6 l/min and a constant saturation temperature of Ts=16 °C.


creased to 45° and 60°, respectively, the results in Fig. 9 show that for all values of the surface heat flux and spray flow rate, a spray axis incident angle of 75° reduces the heat transfer coeffi-cient compared to that achieved using a spray axis incident an-gle of 0°. For example, at a high surface heat flux (1.4x105 W/m2), the heat transfer coefficients for a spray axis incident angle of 75° are 3%, 6% and 8% lower than those for a spray axis incident angle of 0° given spray flow rates of 1.4 l/min, 2.0 l/min and 2.6 l/min, respectively. The loss in heat transfer per-formance at a higher spray axis incident angle can be attributed to the droplet flight distance. Fig. 4 shows that the distance from the nozzle to the heat surface for a spray axis incident angle of 75° (i.e. tan75°）is 3.73 times longer than that for a spray axis incident angle of 0°. A longer flight distance implies a larger cumulative drag force, and thus a lower impingement velocity on the heater surface. As a result, the droplets have a reduced ability to break up the liquid film on the heater surface, and thus the heat transfer performance is reduced.

Fig. 10 compares the heat transfer performance obtained when using spray axis incident angles of 0°, 45°, 60° and 75°, respectively, for a constant spray flow rate of 2.6 l/min and a saturation temperature of Ts=16°C. The results confirm that for all values of the surface heat flux, a spray axis incident angle of 60° achieves the maximum heat transfer coefficient, while a spray axis incident angle of 75° yields the minimum heat transfer coefficient. As the angle at which the sprayed droplets are incident on the heater surface increases, the size of the stagnation zone reduces, and hence the heat transfer coefficient increases. However, a larger incident angle also reduces the droplet impingement effect. Consequently, the sprayed droplets lose their ability to break up the liquid film on the heated surface, and thus the vapor bubbles remain

trapped. In other words, an increasing spray axis incident an-gle has opposing effects on the heat transfer performance. Fig. 10 shows that the reduction in the droplet impingement effect at a spray axis incident angle of 75° outweighs the enhanced heat transfer performance caused by a reduction in the size of the stagnation zone. Thus, the optimal heat transfer perform-ance is obtained at a slightly lower incident angle of 60°.

5. Conclusion

This study has proposed a new nozzle / heater arrangement method to enhance the heat transfer coefficient of a rhombus-pitch shell-and-tube spray evaporator. The experimental results have shown that the shell-side heat transfer coefficient obtained using the proposed interior spray method is significantly higher than that achieved in a conventional flooded-type evaporator. The heat transfer performance of the interior spray evaporator has been examined for spray axis incident angles of 0°, 45°, 60° and 75°, respectively. In general, the results have shown that an inclined spray produces a net momentum flux which suppresses the formation of stagnation zones on the heated surface and therefore increases the heat transfer coefficient. The ability of the droplet spray to suppress the stagnation zones increases with an increasing incident angle. However, as the incident angle is in-creased, the distance from the nozzle to the heated surface must also be increased if the droplets are to impact upon the same region of the surface. Consequently, the impingement velocity of the droplets reduces with an increasing incident angle, and thus a reduction in the heat transfer performance occurs. In other words, an increasing incident angle exerts two simultaneous but oppos-ing effects on the heat transfer coefficient. The present results have suggested that the optimal heat transfer performance in a

4.0x104 8.0x104 1.2x105 1.6x1052.0x104 6.0x104 1.0x105 1.4x105

0.0x100

1.0x104

2.0x104

3.0x104

4.0x104

:incident angle=60o

:incident angle=45o

:incident angle= 0o

:incident angle=75o

Ts=16oCm=2.6(l/min)

q"(w/m2)

h(w

/m2 -k

)

Fig. 10. Variation of heat transfer coefficient with surface heat flux as function of spray incident angle (0°, 45°, 60° and 75°) for spray flow rate of 2.6 l/min and saturation temperature of Ts=16 °C.

4.0x104 8.0x104 1.2x105 1.6x1052.0x104 6.0x104 1.0x105 1.4x105

0.0x100

1.0x104

2.0x104

3.0x104

4.0x104

:incident angle=0o

:incident angle=75o

Ts=16oC

m=2.6(l/min)

m=2.0(l/min)

m=1.4(l/min)

q"(w/m2)

h(w

/m2 -k

)

Fig. 9. Variation of heat transfer coefficient with surface heat flux asfunction of spray incident angle (0° and 75°) for spary flow rates of 1.4l/min, 2.0 l/min and 2.6 l/min and a constant saturation temperature of Ts=16 °C.


rhombus-pitch shell-and-tube interior spray evaporator can be obtained by using a spray axis incident angle of 60°.

Acknowledgments

This study was supported by the National Science Council of Taiwan (NSC 96-2221-E-218-034 and NSC 100-2221-E-218-046) and by the Ministry of Education of Taiwan (Re-search and Development of Intelligent Electric Vehicles with Green Energy and Personnel Training Project).

Nomenclature------------------------------------------------------------------------

h : Heat transfer coefficient ( 2W/m C− ° ) m•

: Spray flow rate (l/min) q′′ : Wall heat flux ( 2W/m ) Rp : The micro-roughness of the tube surface (μm) T : Temperature (°C) Tsat : Saturation temperature (°C)

wT : Average wall temperature (°C)

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Tong-Bou Chang received the Ph.D degree in Mechanical Engineering from National Cheng Kung University, Tainan, Taiwan, in 1997. From 1997 to 2001, he was a researcher at Yuloon-Motor Group (Taiwan), whose job function includes design and characterization of the thermal and fluid

flow systems for vehicle. Since 2002, he has been as a Profes-sor at the Department of Mechanical Engineering, Southern Taiwan University. His current research interests include heat transfer with phase change, energy-system optimization, heat and mass transfer in porous medium, enhancement heat trans-fer and high performance heat exchangers.


Ru-Li Lin received his Ph.D degree in Mechanical Engineering from the National Taiwan University, Taipei, Taiwan in 1999. He was a Post-doctorate researcher in the National Taiwan Uni-versity from 1999 to 2003. He joined the Department of Mechanical Engineering, Southern Taiwan University in 2003 and

is currently an associate professor in the Department. His re-search interests focus on heat conduction in the solid, the me-chanics of solid, anisotropic elasticity and piezoelectric elasticity.

Chih-Chang Liang was a graduate student at Department of Mechanical Engineering, Southern Taiwan University, Taiwan. His research interests include spray heat transfer, enhancement heat transfer and high performance heat exchangers. He is currently working at the Ministry of

National Defense as an engineer.

Effects of spray axis incident angle on heat transfer ... · Effects of spray axis incident angle...

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