BHT of Ammonia

8/10/2019 BHT of Ammonia

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Local boiling heat transfer characteristics of ammonia ina vertical plate evaporator

H. Arima *, J.H. Kim, A. Okamoto, Y. IkegamiInstitute of Ocean Energy, Saga University, 1-48, Hirao, Kubara-aza, Yamashiro-machi, Imari, Saga, 849-4256 Japan

a r t i c l e i n f o

Article history:Received 27 December 2008Received in revised form19 September 2009Accepted 26 September 2009Available online 8 October 2009

Keywords:Heat exchangerPlate exchangerHeat transferBoiling Ammonia

MeasurementHeat transfer coefcientVisualization

a b s t r a c t

Ocean thermal energy conversion systems are expected to be the next-generation energyproduction systems. In these systems, a plate heat exchanger is used for improving thepower generation efciency, and ammonia or an ammonia/water mixture is used asa working uid.

In this study, boiling heat transfer coefcients of pure ammonia are measured ona vertical at PHE (a plate heat exchanger), for elucidating and characterizing the behaviorof ammonia on a compact plate evaporator, a type of PHE

The measurement results show that local boiling heat transfer coefcients increase withincreasing vapor quality. Further, the effects of saturation pressure, mass ow rate, andaverage heat ux on the boiling heat transfer coefcient are elucidated. An empiricalcorrelation for the local boiling heat transfer coefcient is derived using the Lockhart-Mar-tinelli parameter. Further, a visualization experiment of boiling phenomena of ammonia isperformed to elucidate the relation between boiling behavior and heat transfer.

2009 Elsevier Ltd and IIR. All rights reserved.

Caracte ristiques de transfert de chaleur lors de le bullitionlocale dammoniac dans un e vaporateur a ` plaque verticale

Mots cles : Echangeur de chaleur ; E changeur a ` plaque ; Transfert de chaleur ; E bullition ; Ammoniac ; Mesure ; Coefcient de transfert dechaleur ; Imagerie

1. Introduction

It is well known that greenhouse gases such as CO 2 contributeto global warming. Further, abnormal weather conditionscontinue to be observed worldwide due to global warming.

Therefore, the reduction of CO 2 emissions has become animportantissue worldwide. The best method forthe reductionof CO2 emissions is to reduce the use of fossil fuels such aspetroleum and coal. Furthermore, it is important to employrenewable energy sources. Recently, ocean thermal energy

* Corresponding author . Institute of Ocean Energy, Saga University, 849-4256 Japan (IOES). Tel.: 81 955 20 2190; fax: 81 955 20 2191.E-mail address: [email protected] (H. Arima).

www. i i i r.o rg

ava i l ab l e a t www.sc i enced i r ec t . com

jou rna l homepage : w w w. e l s e v i e r. c o m/ l o c a t e / i j r e f r i g

0140-7007/$ see front matter 2009 Elsevier Ltd and IIR. All rights reserved.doi:10.1016/j.ijrefrig.2009.09.017

i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 3 ( 2 0 1 0 ) 3 5 9 3 7 0
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conversion (OTEC) systems or hot spring thermal energyconversion (STEC) systems have attracted considerableattention as sources of renewable energy.

OneofthedrawbacksoftheOTECsystemisthatitgeneratesless electricity than conventionalpowerplants suchas nuclearand thermal powergeneration plantsdo. This is because in anOTEC plant, a small temperature difference between heatsources on surface and in deep ocean water is used; therefore,the thermal efciency of the OTEC system is very low. Hence,for improving the power generation efciency, plate heatexchangers (PHEs) such as evaporators and condensers areemployed in OTEC plants. PHEs facilitate temperature controland provide a large heat transfer area per unit volume

Temperatures of both heat sources of the OTEC system arevery small; therefore, the use of a low boiling point refrigerantas a working uid is required in the OTEC system. In thesesystems, pure ammonia or ammonia/water binary mixturesare commonly used as the working uid. Since the ozonedepletion potential (ODP) and global warming potential (GWP)of ammonia are zero, it is a very good refrigerant from theviewpointofpreserving the qualityof the earths environment.

Further, for improving the performance of PHEs, it isimportant to improve the heat transfer coefcient of theworking uid. However, the boiling heat transfer performanceof ammonia has not yet been elucidated. Some studies

(Kushibeetal.,2005 ) determinedforcedconvective boiling heattransfer coefcients of ammonia on the plate evaporator ofanexperimental OTEC plant. However, in these studies, localboiling heat transfers on the working uid side of the plateevaporator were not determined, because the overall heattransfers coefcients which include boiling heat transfers onthe working uid and heat source side, were calculated.

Furthermore, few studies ( Nishikawa andFujita, 1977; Inoueetal., 2002;Arimaet al., 2003 ) wereconducted formeasuring thepool boiling heat transfer of ammonia. In these studies,data ondifferent saturation pressure, average heat ux, and massfraction wereobtained. In addition, somestudies( Zurcheret al.,2002; Zamrescu and Chiriac, 2002 ) were conducted formeasuring thelocalboilingheat transfer coefcientof ammoniaon a horizontal or vertical tube type evaporator.

However, thus far, no studies have been conducted formeasuring the local forced convective boiling heat transfercoefcient of ammonia on a vertical plate evaporator. There-fore, in the present study, the local convective boiling heattransfer coefcient of ammonia on a plate evaporator (used asa test plate) is measured. Using the results of this study,suitable design criteria for improving the power generationefciency of thermal conversion systems with small temper-ature differences can be established. Further, effects of massux, heat ux, and saturation pressure on the boiling heat

Nomenclature

A constant []b constant []C proportionality factor [(W/m 2)(1-n) /K]C1 w C5 constantCp specic heat [J/(kg K)]Dh hydraulic diameter [m] 2wd /(w d)F constant []F uid-dependent parameter []G mass ux [kg/(m 2 s)]h boiling heat transfer coefcient [W/(m 2 K)]hLZ heat transfer coefcient for two-phase ow and

for ow of only a liquid phase in a channel[W/(m2 K)]

n constant []ifg latent heat of vaporization [J/kg]ipre,in specic enthalpy of preheater inlet [J/kg]ipre,ou t specic enthalpy of preheater outlet [J/kg]isat,l specic enthalpy of saturated liquid [J/kg]itest ,n local specic enthalpy of test plate [J/kg] j supercial velocity [m/s]k thermal conductivity [W/(mK)]li distance between two thermocouples [m]m mass ow rate [kg/s]Psat saturation pressure (absolute) [Pa]q heat ux [W/m 2]Q heat ow [W]T temperature [ C]Twall plate wall temperature [ C]D Tsat wall superheat [K]w width of test plate channel [m]x vapor quality []

D y distance between two neighboring thermocouplewells [m]

Greek symbolsd height of test plate channel [m]m viscosity [Pa s]r density [kg/m 3]

Subscriptsav averageg vapori position of measuring pointin inletl liquidloc localn number of measuring pointout outletpre preheatersat saturationsus SUS304

test test sectionwall wall

Dimensionless number Bo boiling number [] qGHfg Co convection number [] 1 xx

0:8r g r l 0:5

Fr Froude number with all ow [] G2r 2l gDh Prl Prandtl number of the liquid phase [] klmlCplReg Reynolds number of the vapor phase []

(GxDh /mg )Rel Reynolds number of the liquid phase []

(G(1-x)Dh /ml)Xtt Lockhart-Martinelliparameter for turbulent liquid

and vapor phases []



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transfer coefcients of ammonia are examined. In addition,an empirical correlation for the local forced convective boiling heat transfer coefcient of ammonia is derived.

Moreover, in order to elucidate the effect of the localboiling heat transfer performance of ammonia on its boiling phenomenon, an experiment for the visualization of theinteriors of the plate is performed.

2. Experiment

2.1. Experimental apparatus and procedure

Fig. 1 shows a schematic of the experimental apparatus con-sisting of a plate evaporator (test plate), condenser, and threeow circuits d a warm water circuit, cold water circuit, andworking uid circuit. A subcooled working uid (ammonia,approximately 8 K) is pumped up to a preheaterusing working uid pump (Teikoku Electric Mfg. Co., Ltd., reverse circulationtype canned motor pump, head 12 m, power 1.1 kW). Theworking uid is heated by the preheater (brazed plate heatexchanger Tokyo Braze Co., Ltd.) to achieve the recommendedvapor quality at the test plate inlet. Then, the working uid isown into the test plate, and the uidexchanges heat with hotwater. As a result, the state of the working uid changes fromliquid to a two-phase uid. The two-phase uid is transportedto an after-condenser and plate condenser, and then, it iscondensed into liquid using cold water. The condensedworking uid is stored in a working uid tank and transportedto the working uid pump. Further, the hot and cold water aregenerated by a gas boiler and refrigerator and stored in hotand cold water tanks, respectively.

The working uid temperature is measured using resis-tance thermometers (Hayashi Denko Co., Ltd., ER6, JIS A-class,

accuracy of less than 0.15 C); the mass ow rate ismeasured using a Coriolis mass owmeter (Endress Hauser,accuracy of 1% of F.S.); pressure is measured using a gaugepressure transducer (Toshiba, 3051CG, range 0 w 2070 kPa,accuracy of 0.25% of F.S.); and ow rates of hot and coldwater are measured using magnetic owmeters (Toshiba,LF410, accuracy of 0.5% of F.S.).

All measured data were recorded using a programmablelogic controller (PLC; Mitsubishi Electric, MELSEC Q series)connected to a personal computer (PC).

2.2. Test plate

Fig.2showsaschematicdiagramofthetestplateevaporator.Thetest plate consists of a main plate, two ames, and two spacers.The test plate has dimensions of 380 nm (width) 850 nm(length) 40 nm (thickness). Further, the two ames havedimensions of 380 mm (width) 850 mm (height) 30 nm(thickness). Spacers on the working uid side and heat sourceside have thicknessesof2 mmand10 mm,respectively.The areaofheatexchangerabove themainplate is250 mm 650 mm;themain plate is polished using #2000 sandpaper.

Themain plate, ames, andspacer on theworking uid sideare made of SUS304, and the spacer on the heat source side ismade of rubber. The rubber spacer is also used for thermalinsulator. The channels of the working uid and hot sourceconsist of the main plate, one spacer, and one ame. Then, thecross-sectional areas of the ow channels are 2 mm 250 mm(working uid side) and 10 mm 250 mm (hot source side).

Inside themain plate, there aresix thermocouple (TC) wellsin which thermocouple sheaths ( Fig. 3) are inserted formeasuring local temperatures. Each TC well is 3 mm in diam-eterand38 mmin length.The thermocouplesheathsconsistof a urethane tube and two xed K-type thermocouples (0.1 mm

Fig. 1 Schematic diagram of experimental apparatus.

i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 3 ( 2 0 1 0 ) 3 5 9 3 7 0 361


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in diameter). The thermocouples are mounted on the surfacesof the sheaths. In order to increase the temperature measure-ment accuracy, the distance l1 between the thermocouples ismaintained to be sufciently large (average distance of 33.5 mm). All data measured using the thermocouples arerecorded using a multimeter (Keithley, model 2701) connectedto the PC. In addition, temperature proofreading is carried outin a constant temperature bath by using thermocouples.Therefore, the temperature measurement is highly accurate(less than 0.1 C accuracy).

In order to visualize the boiling phenomena occurring inside the ow channels, three sight glasses were placed onthe working uid side.

2.3. Visualization experiment

Fig. 4 show a cross section of the sight glass used for visuali-zation. Thesightglass is45 mm indiameter.Visualized imagesof thedottedsquare (having an area of2500 mm 2)areshowninsubsequent gures. Boiling phenomena occurring inside theow channels are observed using images captured using a digital still camera (Pentax *istD); these images are capturedfrom outside the owchannels, as shown in Fig. 5. Thecamerashutterspeedis 1/4000 s.The lightsourceis a 250 W coldlamp.

2.4. Local heat ux

The six TC wells, locatedalong thecenter line of the test plate,are used to measure the local heat ux.

Assuming that local heat uxes ( q) can be estimated fromone-dimensional, steady-state heat conduction, q can beexpressed by Eq. (1):

q ksusT1 T2

l1(1)

where ksus is the thermal conductivityof SUS304 and T1 and T2are local temperatures.

The wall temperature of the working uid side ( Twall ) iscalculated by Eq. (2):

Twall T2q$l2ksus

(2)

where l2 is the distance between the thermocouple and thewall surface.

Fig. 2 Schematic diagram of test plate.

HotwaterWorking

fluid

Urethan tube

Platel2 l1

T 1T 2

T wall

A:

Fig. 3 Position of local thermocouple inside the test plate(enlarged view of area A shown in Fig. 2 ).



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The local heat transfer coefcient ( h) is calculatedby Eq. (3):

h q

Twall Tsat

qD Tsat

(3)

where D Tsat is the wall superheat and Tsat is the bulktemperature of the working uid side, derived using thesaturation pressure at the plate inlet. By the way, since theinside of a plate has pressure drop, it along the ow should beconsidered in deriving the local saturation pressure and localsaturation temperature. However, since change of the satu-ration temperature by pressure drop was about 0.04 C, it wasassumed that a saturation temperature was xed.

The local specic enthalpy ( itest ,n ) on the test plate isderived using the following method. First, the subcooledworking uid is own into the preheater inlet. Then, thepreheater inlet specic enthalpy ( ipre,in ) is calculated from theuid temperature and pressure by using the P-Propathcomputer program package ( PROPATH Group, 2006).

The outlet specic enthalpy ( ipre,out ) at the preheater isobtained from the heattransport rate ( Q pre ) and mass ow rate(m G d w) for the preheater:

ipre ;out ipre ;in Q prem

(4)

The specic enthalpy of the plate inlet ( itest,in ) is dened asbeing equal to that of the preheater outlet ( ipre,out ), because the

pipe connecting the preheater and evaporator is sufcientlyinsulated.

Next, the local specic enthalpy ( itest, n) in each TC well atpoint n is calculated by adding the increase in stock of thespecic enthalpy from a plate entrance.

Here, the increase in stock of specic enthalpy betweenneighboring TC wells is calculated using the heat ow ( Q n)

between two neighboring TC wells and the mass ow rate ( m),similar to the calculation shown in Eq. (1). Here, Q n isobtainedfrom the local heat ux ( qn) and the heat transfer area.However, since the heat ux of each TC well is different, theheat ow of each area Q 0n is calculated using the area of eachplate part An yn w, as shown in Fig. 6. Finally, Q n is calcu-lated from the average value of the heat ow Q 0n and Q 0n -1

Further, itest,n is expressed as follows:

(a) in the case that n 1

itest ;1 itest ;in Q 1m ;

Q 1 q1A1

(b) in the case that n 2w 6

itest ;n itest ;n 1 Q n

m; Q n Q

0n Q 0n 12 ; Q

0n qnAn n 2w 6 (5)

The local vapor quality ( xn) is dened as follows:

Fig. 5 Setup for visualization.

45

View area

Fig. 4 Cross section of sight glass.

Fig. 6 Schematic diagram of effective heat transfer surfacefor calculating local enthalpies.



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xn itest ;n isat ;liq

ifg n 1w 6 (6)

where isat,liq is the specic enthalpy of the saturated liquid atthe saturation pressure of the plate inlet ( Psat ) and ifg is thelatent heat, which can be calculated using P-Propath.

The experimental conditions are shown in Table 1 .

3. Experimental results

3.1. Boiling curve

Fig. 7 shows theboilingcurve (at Psat 0.7 MPa)with changes inthe mass ux and average heat ux. The local heat uxincreases linearly with increasing D Tsat . This tendency wasobserved in the pool boiling curve of ammonia ( Arima et al.,2003). However, change of the wall superheat temperature isonly 2 K against changeof heat ux being 10 kW/m 2. Therefore,theeffect of thewallsuperheat temperature isvery smallon theheat ux. Furthermore, the local heat ux was not affected by

the mass ux in present study. Because each mass ux differ-ences is small in itself and the result is close to the pool boiling condition due to the very small mass ux for forced convectionexperiment.Thesolid lineand diamondplots in Fig.7 show thattheexperimental data of theboiling curve forthe pool boilingof ammonia by Arima et al. (2003). And the short and long dashedlineshowsa pool boilingprediction,which isobtained byEqs. (1)and (3). Eq. (7) has been proposed by Stephan and Abdelsalam(1980) and Nishikawa and Fujita (1977) .

h Cqn (7)

where C 1.05, n 0.745 (Stephan and Abdelsalam, 1980 ) andC 4.41, n 2/3 (Nishikawa, 2000 ) at Psat 0.7 MPa forammonia. It is found that D T

sat for forced convective boiling

and the result of pool boiling by Arima et al. (2003) are mostlyin agreement. On the other hand, D Tsat for forced convectiveboiling is 8 K or 3 K lower than each prediction. According toArima et al. (2003) or Inoue et al. (2002) , the boiling pool boiling correlation of Nishikawa and Fujita (1977) is more agreementcompared with that of Stephan and Abdelsalam (1980) .Therefore, even if it compares with the prediction of Nishi-kawa andFujita (1977) , it was found out that thepresent resultindicates near pool boiling condition. The reason why thelocal heat ux as not affected by the mass uxes that theresult is close to the pool boiling condition.

3.2. Local boiling heat transfer

3.2.1. Inuence of mass uxFig. 8(a) and (b) show that plots of the measured local boiling heat transfer coefcient hloc versus vapor quality x at various

mass uxes; the average heat uxes in these cases are 15 kW/m 2 and 20 kW/m 2, respectively, and the saturation pressure inboth these cases remains constant at 0.70 MPa. In case of x < 0.3, the local heat transfer coefcients remain almostconstant with increasing x. However, in case of 0.3 < x < 0.7,that tend to increase with increasing x. In general, at theforced convective boiling in a vertical tube, it is known that incase of nucleate boiling region, wall superheat is constantwith increasing x and in case of forced convective heattransfer through liquid lm region (forced convective region),the wall superheat is little decrease with increasing x (Tong and Tang, 1997 ). Since change of the gradient of heat transferwas observed bordering on x 0.3 as shown in Fig. 8.

However, regardless of the amount of mass ux, the localboiling heat transfer coefcient remains almost constant fora given vapor quality. Therefore, this tendency shows that anincrease in mass ux has almost no effect on the boiling heattransfer coefcient. At the nucleate boiling region, it isconsidered that the bubble which is generated in the heating surface tends to stagnate into the narrow channel, althoughheat transfer by the forced convection is performed. Becausethese mass uxes are small different and mass uxes of present study are lower than that of previous study. There-fore, it is considered that it depends for heat transfer on theamount of bubbles emergence and heat transfer is not inu-enced by the mass ux. In addition, the boiling heat transfercoefcient decreases rapidly with increasing x for x > 0.7 onFig. 8 (b). The range x > 0.7 implies the occurrence of a dry-out.The same tendency is observed in Fig. 9(a).

3.2.2. Inuence of heat uxFig. 9(a) and (b) show the plot of measured local boiling heattransfer coefcient hloc versus vapor quality x at variousaverage heat uxes; the mass uxes in these cases are7.5 kg/(m 2 s) and 10 kg/(m 2 s), respectively, and the saturationpressure in both these cases remains constant at 0.70 MPa.

Table 1 Experimental conditions.

Working uid Ammonia

Mass ux G [kg/(m2 s)] 7.5, 10, 15Average heat ux qav [kW/m 2] 15, 20, 25Saturation pressure Psat [MPa] 0.7, 0.8, 0.9Saturation temperature Tsat [ C] 13.9, 17.9, 21.6

Vapor quality of test plate inlet xtest,in [] 0.1w

0.4

Fig. 7 Boiling curve at Psat [ 0.7 MPa.



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All local boiling heat transfer coefcients tend to increasewith decreasing average heat ux and increasing quality.However, as shown in Fig. 9(a), at qav 20.0 and 24.5 kW/m 2,boiling heat transfer decreases for x > 0.85, thereby causing a dry-out.

Under the same saturation pressure and mass ux, theboiling heat transfer coefcient decreases with increasing average heat ux. This tendency is different from thatobserved by Kido et al. (1992) and Hsieh et al. (2002): theyreported that the boiling heat transfer coefcient increaseswith increasing average heat ux. The reason for thistendency is considered to be the fact that in their experi-ments, heating was carried out using a uid and not a heater.In the former case, heat ux is determined by the balancebetween heat transfers of the working uid and warm water,whereas in the latter case, even when the operating uid hasan irregular ow with air bubbles, xed heating is carried out.Therefore, it is considered that uid heating is difcult to heatuniformly. The investigation of this phenomenon is thesubject of a future study.

3.2.3. Inuence of saturation pressureFig. 10(a) and (b) show variations in the local boiling heattransfer coefcient h loc with vapor quality x at a given satu-ration pressure. The local boiling heat transfer coefcienttends to increase with increasing quality, as is the case whoseplot is shown in Fig. 9. In addition, the boiling heat transfercoefcient decreases with increasing saturation pressure.This is the cause of the increase in the wall superheat withincreasing saturation pressure. However, the local boiling heat transfer coefcient decreases with increasing saturationpressure at Psat 0.8 and 0.9 MPa at G 10.0 kg/m 2 s, despitethe wall superheat being almost constant.

This result is the same as that obtained by Ishibashi and

Nishikawa (1969) : under a slug and annular ow, the boiling heat transfer coefcient decreases with increasing saturationpressure at a constant heat ux.

3.2.4. Comparisons of previous correlationComparisons between the existing correlations and presentdata which are shown in Fig. 8 are performed. Nishikawa and

Quality x [-]

0.0 0.2 0.4 0.6 0.8 1.0

L o c a

l h e a

t t r a n s f e r c o e f

f i c i e n t

h l

o c

k W / ( m

2 .

K )

5

6

7

8

9

10

G = 7.4 kg/(m 2 . s)

G = 10 kg/(m 2 . s)

qav = 15 kW/m2

P abs = 0.70 MPa

Quality x [-]

0.0 0.2 0.4 0.6 0.8 1.0

L o c a

l h e a


f i c i e n t

h l o

c k W / ( m

2 .

K )

5

6

7

8

9

10

G = 7.5 kg/(m 2 . s)

G = 10 kg/(m 2 . s)

qav = 20 kW/m2

P abs = 0.70 MPa

qav = 15 kW/m 2 qav = 20 kW/m 2

a b

Fig. 8 Plot of local boiling heat transfer coefcient versus vapor quality at various mass uxes.

0.0 0.2 0.4 0.6 0.8 1.0 0.0 0.2 0.4 0.6 0.8 1.05

6

7

8

9

10

qav = 15.4 kW/m2

qav = 20.0 kW/m2

qav = 24.5 kW/m2

G = 7.5 kg/(m 2 . s)P abs = 0.70 MPa

Quality x [-]Quality x [-]

L o c a

l h e a

t t r a n s

f e r c o e f

f i c i e n t

h l o c

k W / ( m

2 .

K )

L o c a

l h e a

t t r a n s

f e r c o e f

f i c i e n t

h l o c

k W / ( m 2

.

K )

5

6

7

8

9

10

qav = 15.4 kW/m2

qav = 20.0 kW/m2

qav = 24.5 kW/m2

G = 7.5 kg/(m 2 . s)P abs = 0.70 MPa

G = 7.5 kg/m 2s G = 10.0 kg/m 2s

a b

Fig. 9 Plot of local boiling heat transfer coefcient versus quality at various heat uxes.



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Fujita (1977) and Stephan and Abdelsalam (1980) are proposedcorrelation Eq. (7) for pure ammonia on pool boiling. Thecorrelation of Arimaet al. (2003) is derived by their pool boiling data. Kandlikar (1990) proposed correlation Eq. (8) for localboiling heat transfer on vertical tube using convection numberCo.hlochLZ

ClCoC2 C3BoC4 Ff l (8)

where, hLZ is the heat transfer coefcient for a two-phase owand for the ow of just the liquid phase in the channel. hLZ isexpressed using the Dittus-Boelter equation as follows:

hLZ 0:023 klDh

G1 xDhml

0:8

Pr0:4l (9)

where Prl is the Prandtl number of the liquid phase.The parameters C1w C4 are as follows on different Co

numbers;

(a) Case convective region (at Co< 0.65);

C1 1:1360; C2 0:9; C3 667:2; C4 0:7 (10)

(b) Case nucleate boiling region (at Co> 0.65);

C1 0:6683; C2 0:2; C3 1058:0; C4 0:7 (11)

Incidentally, uid-dependent parameter F for ammoniawas not given by Kandlikar. Therefore, Zamrescu and Chir-

iac, 2002 proposed F 0.7 for ammonia.On the other hand, Shah (1982) proposed correlation Eq.

(12) for local boiling heat transfer on vertical tube.

hlochLZ

j (12)

The value of j is given by following Eqs. (13) to (17) with thevalue of convection number Co.

j cb 1:8=Co0:8 (13)

(a) Case 0.1 < Co 1.0

j bs FBo0:5exp 2:74Co 0:1 (14)

(b) Case Co 0.1

j bs FBo0:5exp 2:47Co 0:15 (15)

when j > j bs and j cb , Thus if j bs > j cb , j j bs . If j cb > j bs ,j j cb

where, constant F in Eqs. (14) and (15) is dened by Eqs. (16)and (17).

F 14:7; Bo > 11 10 4 (16)

F 15:43; Bo < 11 10 4 (17)

The values of both correlations are plotted into Fig. 11.Fig. 11 shows that the predicted heat transfer coefcients byShah (1982) and Kandlikar (1990) correlations are very smallerthan present study data. It is found that both correlationscannot predict present data on vertical plate. It is consideredthat the reason for disagreement is the magnitude of thepresent mass uxes is very smaller than that of assumed inthe tube experiment. Generally the correlation in tubeexperiment is made on the conditions of a high mass ux.Since the effect of heat transfer by forced convection is quitelarge, the effect of nucleate boiling will be underestimated. Onthe other hand, in present study, since a mass ux is verysmall and the effect by nucleate boiling is large, the large heattransfer is shown compared with the correlation. Further-more, the correlation of pool boiling heat transfer by Nishi-kawa and Fujita (1977) and Stephan and Abdelsalam (1980) arealso very smaller than present study data. On the other hand,Arimas correlation is more close to present study data.Therefore, it was found that the present data was able to bewell expressed with the correlation of the pool boiling whichderived from Arimas et al.(2003) experiment.

3.3. Nondimensional correlation

The nondimensional correlations for forced convective heattransfer on different refrigerants were proposed by manyresearchers ( Mandrusiak and Carey, 1989; Wen and Ho, 2005;Kushibe et al., 2005 ). Eq. (18) is the general correlation forboiling heat transfer. The correlation is expressed using theLockhart-Martinelli parameter X. In their studies, X wasdened as Xtt where the liquid was turbulent and vapor wasturbulent ow, which can be expressed as Eq. (19).

5

6

7

8

9

10

P abs = 0.70 MPa

P abs = 0.80 MPa

P abs = 0.90 MPa

G = 7.5 kg/(m 2 . s)qav = 20 kW/m

2

0.0 0.2 0.4 0.6 0.8 1.05

6

7

8

9

10

P abs = 0.7 MPa

P abs = 0.8 MPa

P abs = 0.9 MPa

G = 10.0 kg/(m 2 . s)qav = 20 kW/m

2

0.0 0.2 0.4 0.6 0.8 1.0

L o c a

l h e a

t t r a n s

f e r c o e

f f i c i e n t

h l o c k

W / ( m

2 .

K )

L o c a

l h e a


f i c i e n t

h l o c k

W / ( m

2 .

K )

Quality x [-]

G = 7.5 kg/m 2sQuality x [-]

G = 10.0 kg/m 2s

a b

Fig. 10 Variations in local boiling heat transfer coefcient with quality at various saturation pressures.



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However, in the present study, both of the liquid and vaporphases are laminar conditions, because Rel 40 300 (Liquid

phase Renumber)and Reg 7803600(Vaporphase Renumber).Then, X is dened as Xvv which is expressed as Eq. (20).hlochLZ

A1X

b

(18)

where A and b are constants.

Xtt 1 x

x

0:9 r g r l

0:5 mlmg !

0:1

turbulent-turbulent (19)

Xvv 1 x

x

0:5 r g r l

0:5 mlmg !

0:5

laminar laminar (20)

The relationshipbetween the ratio hloc /hLZ and Xvv1 is shown

in Fig. 12. It is found that h loc /hLZ increases with increasing Xvv1. However, the gradient for Xvv1 < 2 differs from that forXvv1 > 2. This difference in gradients is considered to be due tothe difference in ow patterns. We considered that in the caseof low gradient region, the two-phase ow became slug ow,whereas in the case of high gradient region, the ow becameannular ow when a thin liquid lm covered the entire owchannel. The ow pattern in this study is examined in thenext Section 3.4.

The solid line in Fig. 12 shows the correlation obtained inthe present study. This correlation is obtained by the least-

squares method using all data obtained here.hlochLZ

16:4 1Xvv

1:08

(21)

The empirical correlation expressed in Eq. (21) can predictexperimental results within 25% accuracy for Xvv1 > 2.

On the other hand, the measured local heat transfer coef-cients are predicted using the correlation Eq. (21). Fig. 13shows that the comparison of predicted by Eq. (21) againstmeasured the ratio hloc /hLZ. It is found that almost data can bepredictedby Eq. (21)within 25%accuracy.However,in case of hloc /hLZ < 30of G 7.5,the value is largerthan 25%.Therangeof hloc /hLZ < 30 is indicated Xvv1 < 1.7 which is low vapor quality.Then, the correlation cannot predict the measured data.

3.4. Visualization

Visualization of the boiling phenomena of ammonia is carriedout under various mass uxes, heat uxes, saturation pres-sures, and vapor qualities.

0.0 0.2 0.4 0.6 0.8 1.00

2

4

6

8

10G = 7.4G = 10G = 7.4 (Kandlikar)G = 10 (Kandlikar)G = 7.4 (Shah)G = 10 (Shah)

Pool boiling (Stephan) Pool boiling (Nishikawa) Pool boiling (Arima)

qav = 15 kW/m2

P abs = 0.70 MPa

Kandlikar

Shah

Stephan

Arima

0.0 0.2 0.4 0.6 0.8 1.00

2

4

6

8

10G = 7.5G = 10G = 7.5 (Kandlikar)G = 10 (Kandlikar)G = 7.5 (Shah)G = 10 (Shah)

Pool boiling (Stephan) Pool boiling (Nishikawa) Pool boiling (Arima)

qav = 20 kW/m2

P abs = 0.70 MPa

Kandlikar

Shah

Stephan

Nishikawa

Arima

qav = 15 kW/m 2Quality x [-]

qav = 20 kW/m 2Quality x [-]

L o c a

l h e a t

t r a n s

f e r c o e

f f i c i e n t

h l o c

k W / ( m

2 .

K )

L o c a

l h e a t

t r a n s

f e r c o e

f f i c i e n t

h l o c

k W / ( m

2 .

K )

a b

Fig. 11 Comparisons of local boiling heat transfer coefcient between present and predicted data.

Fig. 12 h loc / h LZ as a function of 1/ X vv .

100

100

60

60

80

80

00

20

20

40

40

Predicted hloc /h LZ [-]

E x p e r i m e n

t a l h

l o c

/ h L Z

[ - ]

+25%

-25%

G=7.5G=10

Fig. 13 Comparison of predicted against experimental

h loc / h LZ .



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3.4.1. Effect of vapor qualityFig. 14(a) and (b) show visualization results of boiling of ammonia at various vapor qualities and constant mass ux,heat ux, and saturation pressure.

The vapor quality x considered for visualization whoseresults are shown in Fig. 14(a) and (b) is determined by a poly-nomial interpolation of the vapor quality values measured inthe ow direction. Fig. 14(a) shows that some small bubblesappear on the plate surface at x 0.28. It is found that the owpattern consists of bubble ow. On the other hand, Fig. 14(b)shows that instead of bubbles, a liquid lm covers the entirevisualized area. It is considered that this ow pattern isannular ow corresponding to ow inside the tube.

In section 3.3, we stated that in the case of Xvv1 < 2 andXvv1 > 2, the ow pattern becomes slug ow and annular ow,respectively. Since the visualization results in shown Fig. 14(a)(x 0.28 and Xvv 1.68) demonstrate that the ow is a bubbleow and those shown in Fig. 14(b) (x 0.63 and Xvv 3.51)demonstrate that the ow is liquid ow, we can conclude thattheclassicationoftheowpatternaccordingto Xvv isaccurate

3.4.2. Effect of mass uxFig.15(a)and(b) showvisualizationresults atvariousmassuxesat a constant vapor quality, heat ux, and saturation pressure.

In the case of a low mass ux ( Fig. 15(a)), some interme-diate-size bubbles are observed over the entire area. On theother hand, in the case of a high mass ux ( Fig. 15(b)), theliquid lm that covers the entire visualized area is observed,similar to the results shown in Fig. 15(b). The ow patternsshown in Figs. 15(a) and (b) are different; however, the boiling heat transfer coefcient for these patterns is almost the same.Therefore, we conclude that the ow pattern does notcontribute to boiling heat transfer. As mentioned in Section3.2, forced convection is dominant in the boiling heat transferunder these conditions. This behavior is also conrmed in thevisualization experiment.

3.5. Flow pattern map

In order to consider the owpattern obtained by visualization,comparison with the existing ow map was performed. Theow pattern maps for horizontal ow were proposed in somepapers. However, the map for vertical upow in tube has atleast the diagram of HewittRoberts map ( Hewitt and Roberts,1969) and there is no map for vertical upow in plate. There-fore, the all data were plotted into the HewittRoberts map asshown in Fig. 16. The mapshows that plots of supercial liquidmomentum ux r l jl2 versus supercial vapor momentum ux

Fig. 14 Boiling ow patterns of ammonia at two different vapor qualities ( G [ 10 kg/m 2 s, Psat [ 0.7 MPa, q av [ 20 kW/m 2 )

Fig. 15 Boiling ow patterns at different mass uxes ( Psat [ 0.8 MPa, q av [ 20 kW/m 2 , x [ 0.4)



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r g jg 2. Where,supercial liquid velocity jl andvapor velocity jg aredened by as follows;

jl G1 x=r l (22)

jg Gx=r g (23)

Fig. 14(a) and (b) show that the ow patterns were bubbleand annularow, respectively. However, Fig. 16 shows that alldata are in churn ow, because both momentum uxes of present data are very lower than that of general study in tube.It is found that the vertical ow on plate could not beexpressed on this map for vertical tube. In the future, it isnecessary to examine the map of the vertical plate ow.

4. Conclusion

The experimental results of the boiling heat transfer coef-cients of pure ammonia under forced convective boiling ina vertical at plate are summarized as follows.

(1) The boiling curve in the case of a forced convective boiling heat transfer shows that in such a heat transfer, thesurface wall superheat is 8 K less than that for a poolboiling heat transfer. Therefore, it is concluded that in theboiling heat transfer under the present experimentalconditions, forced convection is dominant.

(2) The forced convective boiling heat transfer coefcient of ammonia increases with increasing vapor quality x ata constant mass ux, saturation pressure,and average heatux. However, in the case of x > 0.7, a dry-out occurs occa-sionally, and therefore, boiling heat transfer decreases.

(3) An increase in massuxhasalmostno effect on boiling heattransfer. On theother hand,an increase in theheat uxandsaturationpressurecauseadecreaseinboilingheattransfer.

(4) Anempiricalcorrelationfortheforcedconvectiveboilingheattransfer coefcient is derived using the Lockhart-Martinelliparameter. Theboiling heat transfer coefcient estimatedbythis correlation in the range of Xvv > 2 is in good agreementwith the measured boiling heat transfer coefcient.

(5) The results of visualization conrm that the relationbetween the ow pattern and Xvv is correct. Moreover, therelation between boiling heat transferand the ow patternclearly indicates that boiling heat transfer is dominated byforced convection.

Acknowledgement

We thank the Ministry of Education, Culture, Sports, Scienceand Technology, Japan, for nancial support in the form of a grant under their 21st Century COE Program AdvancedScience and Technology for Utilization of Ocean Energy.

r e f e r e n c e s

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10 -3 10 -2 10 -1 100 101 102 103 104 105 10610 -1

100

101

102

103

104

105

Annular Whispy annular

Churn

Bubbly-slug

Bubbly

Superficial liquid momentum fluxl j l 2 [kg/ms2]

S u p e r f i c i a l v a p o r m o m e n t u m

f l u x

g j g

2 [

k g / m s 2 ]

G =7.5 G =10

Fig. 16 HewittRoberts ow pattern map ( Hewitt andRoberts, 1969 )

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