Condensation on Inclined Surfaces

7/26/2019 Condensation on Inclined Surfaces

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Condensation on downward-facing surfaces subjected to upstream flowof airvapor mixture

Fernando F. Czubinski a, Marcia B.H. Mantelli a,, Jlio C. Passos b

a Departamento de Engenharia Mecnica, LEPTEN/Labtucal Centro Tecnolgico, Universidade Federal de Santa Catarina, 88040-900 Florianpolis, SC, Brazilb Departamento de Engenharia Mecnica, LEPTEN/Boiling Centro Tecnolgico, Universidade Federal de Santa Catarina, 88040-900 Florianpolis, SC, Brazil

a r t i c l e i n f o

Article history:Received 26 September 2012

Received in revised form 7 January 2013

Accepted 10 January 2013

Available online 4 February 2013

Keywords:

Film condensation

Condensing surfaces

Noncondensable gases

Inclined smooth surfaces

Grooved surfaces

a b s t r a c t

This paper reports experimental results for film condensation on vertical and downward inclined smooth

andgrooved surfaces subjected to ascendingstreams of a vaporair mixture. An experimental facility was

built in order to evaluate the heat fluxes for different surface inclinations, surface sub-cooling tempera-

tures and air to vapor ratios in the mixture. Two experimental procedures were employed to evaluate the

heat absorption rate during the condensation process. The employment of grooved surfaces resulted in a

10% enhancement of the condensation rate in the case of the pure vapor condensation and a negligible

effect when noncondensable gases (NCGs) were present in the mixture.

2013 Elsevier Inc. All rights reserved.

1. Introduction

The condensation heat transfer phenomenon is encountered in

a variety of engineering and industrial applications. Knowledge of

the physical mechanisms which drive condensation is important

for the design of several types of equipment.

The first film condensation model was proposed by Nusselt in

1916. It assumes negligible effects of the interfacial shear stress

at the vapor/liquid interface, and equates gravity and viscous

forces for a quiescent pure vapor environment in contact with an

isothermal vertical smooth plate. A linear temperature profile

was considered. Subsequently, a number of researchers improved

Nusselts model by removing some of the original restrictive

assumptions and including effects like sub-cooling, a nonlinear

temperature profile, inertial terms and interfacial stress [14].

The wettability of the condenser surface plays an importantrole, as the interaction between the condensate liquid and the con-

densing surface dictates the mode in which condensation occurs.

The dropwise condensation mode is associated with higher heat

transfer coefficients when compared to film condensation. There-

fore, many researchers have directed their efforts to the develop-

ment of suitable condensation promoters such as polymeric

films, a monolayer of organic materials and ion-plating technology.

These techniques provide poor wettability of the substrate and

thus dropwise condensation is maintained for a longer period [5

8]. However, as is well known, dropwise condensation is often dif-

ficult to maintain for all the period.Wavy condenser surfaces, also known as Gregorig surfaces [9],

are used to produce better heat transfer rates for the film conden-

sation mode. The enhancement is caused by the variation in the

interface radius of the liquid film over the substrate, which, to-

gether with the surface tension, produces a gradient pressure in

the condensate, driving the film from the crest to the valley of

the surface grooves, improving the condensation performance.

The tests performed showed that this configuration results in heat

transfer coefficients up to five times larger than those for smooth

surfaces, for the same projected area[1012].

In some practical operations, noncondensable gases (NCGs) may

be found in the condensing vapor, deteriorating the heat transfer

process. This is caused by the formation of a gas boundary layer

over the condensate liquidfilm, reducing the partial vapor pressureat the condensate interface. Several studies have been conducted

on stagnant and forced convection occurring over smooth plates

or inside and outside tubes in order to investigate the effect of

an NCG. For a quiescent condensing mixture on smooth plates,

the decrease in heat transfer rates could be around 50% with an

air fraction of only 0.5% (by mass) in the airvapor mixture. In

cases where the mixture flows, the reduction in the heat transfer

decreases since the stream disperses the gas boundary layer. This

effect is dependent on the Reynolds number of the mixture. An-

other effect that could spread the gas boundary layer is the undu-

lation in the liquid filmcondensation interface, which occurs in the

wavy and turbulent flow regime[1316].

0894-1777/$ - see front matter 2013 Elsevier Inc. All rights reserved.http://dx.doi.org/10.1016/j.expthermflusci.2013.01.004

Corresponding author. Tel./fax: +55 48 3721 9379.

E-mail address: [email protected](M.B.H. Mantelli).

Experimental Thermal and Fluid Science 47 (2013) 9097

Contents lists available at SciVerse ScienceDirect

Experimental Thermal and Fluid Science

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http://dx.doi.org/10.1016/j.expthermflusci.2013.01.004mailto:[email protected]://dx.doi.org/10.1016/j.expthermflusci.2013.01.004http://www.sciencedirect.com/science/journal/08941777http://www.elsevier.com/locate/etfshttp://www.elsevier.com/locate/etfshttp://www.sciencedirect.com/science/journal/08941777http://dx.doi.org/10.1016/j.expthermflusci.2013.01.004mailto:[email protected]://dx.doi.org/10.1016/j.expthermflusci.2013.01.004


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As mentioned above, several studies have focused on theoretical

assessments, modes of fluxes and flow regimes, and the composi-

tion and type of mixture in different geometrical configurations.

In this study, experimental analysis and models presented in the

literature will be employed to investigate the downward conden-

sation of vapor contained in airvapor mixtures, in vertical ascen-

sion, reaching the cooled surface of a condensing plate from below.

1.1. Literature review

Condensation studies using this particular configuration have

been carried out by Gerstmann and Griffith [17]and Chung et al.

[18,19]. This phenomenon can be observed in condensers designed

for the recovery of water from humid air which is released to theatmosphere in industrial cooling towers[20].

Gerstmann and Griffith [17] investigated the condensation of

pure Freon-113 on the underside of a smooth plate in a closed

chamber. The heat transfer rates were of the same order of magni-

tude as those predicted by Nusselt analysis (which considers a qui-

escent condensing vapor). Chung et al. [18], also used a smooth

plate as the condensing surface and showed that in the case of pure

water vapor the condensation heat transfer agreed with the Nus-

selt theoretical model, even though the experimental tests were

conducted under a slow flow. With NCGs, when this slow flow

reaches the surface, the NCG boundary layer is spread, so that

the effect of the NCG on the heat transfer rate is not so obvious.

Chung et al.[18]used a deflector before the mixture flow hit the

cooled condensing surface within the condensing chamber, toavoid the impact of the direct flow on the surface. They also stud-

ied configurations where the condensing chamber was closed, to

create a quiescent environment.

The main objective of the study reported herein was to investi-

gate experimentally the condensation of the water of an upstream

airvapor mixture flow, which hits directly the downward face of a

cooled surface. It should be noted that the condensation of water

under such conditions is a subject not previously explored in the

literature. The heat transfer rate resulting from the film condensa-

tion process was measured. Models and results available in the lit-

erature were used as benchmarks for the analysis of the physical

phenomenon involved. The effects of smooth and grooved surfaces

of electrolytic copper and aluminumalloys were also studied, vary-

ing the inclination angles, the sub-cooling temperature and theamount of air in the airvapor mixture.

2. Experiment

2.1. Experimental apparatus

Fig. 1shows a scheme of the experimental facility. It consists of

a test section and auxiliary equipment (steam generator, cooling

water system, NCG supplier and data acquisition system) which

were grouped into three main sections: boiler, vapor supply line

Nomenclature

AlphabeticNCG Noncondensable gaseQ Heat transfer rate (kW)_m mass flow rate (kg s1)

Cp Specific heat (kJ kg1 K1)

Tout Outlet cooling water temperature (K)Tin Inlet cooling water temperature (K)hlv Latent heat (kJ kg

1)h Mean heat transfer coefficient (kW m2 K1)

g Gravity (m s2)L Surface length (m)DT Sub-cooling temperature (K)A Area (m2)T Temperature (K)Sc=l/qD Schmidt dimensionless number ()W Air mass fraction ()M Molecular weight (kg Kmol1)D Diffusivity (m2 s1)

P Pressure (atm)

Greek Lettersq Density (kg m3)h Surface inclination ()

j Thermal conductivity (kW m1 K1)l Dynamic viscosity (kg m1 s1)

Subscriptw Cooling watercond Condensateliq Liquidvap Pure water vaporsat SaturationG Gas (liquid vapor mixture)1 Bulk0 Condensate interface

Fig. 1. Schematic view of experimental apparatus.

F.F. Czubinski et al. / Experimental Thermal and Fluid Science 47 (2013) 9097 91


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and test section. All parts of the equipment were made of galva-

nized carbon steel plates, of 2 mm thickness, and were thermally

insulated using 50 mmrockwool sheets in order to reduce the heat

loss to the surroundings.

The boiler, which consists of a 600 mm 500 mm 270 mm

rectangular vessel, has electrical heaters inside to supply the con-

trolled power needed to generate the steam. The NCG employed

was atmospheric air, supplied with the aid of an air blower through

a square duct with 50 mm of internal cross section and 300 mm of

length (see Fig. 1). The air was warmed to the same temperature as

the steam using a controlled electrical power source to provide

power for the air heater, so that the resulting mixture is kept in

the dry saturated condition. The air mass flow rate can be calcu-

lated by the thermal capacity of the air times the temperature dif-

ference before and after the heater measured by thermocouples.

The air-vapor mixture is created by these two streams, which meet

at the vapor supply line. The supply line is a cylindrical vertical

tube with 147 mm of inner-diameter and 1000 mm of length.

The test section, illustrated inFig. 2, is a box with a square base

with 250 mmof edge and 350 mm of height. One wall of the cham-

ber is made of glass to allow the visualization of the condensing

phenomenon. The mixture formed in the vapor supply line enters

the test section from the bottom and directly reaches the condens-

ing surface. This surface is fixed by pins located in the lateral walls,

allowing inclination variations.

Different inclinations of the cooling plate lead to variations in

the cross sectional area of the airvapor stream flow. Therefore,

the flow area beside the plate changes with the inclination, causing

different degrees of confinement of the mixture. To keep approxi-

mately the same confinement (same cross sectional area of the

flow) for all tested cases, plates with different lengths, which vary

according to their inclination, were constructed. Also, one of the

plates was tested with different inclinations to study the confine-

ment effect.

Smooth electrolytic copper plates with lengths of 116.2 mm,

142.8 mm and 200 mm were tested at inclinations of 30, 45,

60, respectively. This latter plate was tested at inclinations of60and 90 (vertical position).

To collect the condensed water, a drain was installed in the low-

er part of each surface, as shown in Fig. 2. Seven thermocouples

were inserted within small holes drilled in parallel and 0.1 mm be-

low the condensing surface, in order to measure the surface tem-

perature distribution, as can be seen inFig. 3a.

The back face of the condensing test plate closes a small heat

exchanger, comprised of a hollow thermal insulated metallic paral-

lelepiped box, where cooling water circulates at controlled rates

and temperatures, allowing different sub-cooling levels of the con-

densing surface (seeFig. 3b).

The influence of the surface material and finishing on the con-

densation was tested by means of surfaces of the same length

(200 mm) and same inclination (60), made of smooth and grooved

aluminum 5052 alloy and grooved electrolytic copper plates. Thegrooves were made in the longitudinal direction of a plate of the

same thickness as the smooth plate being tested, with a distance

of 2 mm between crests and 1.44 mm of depth. Fig. 4 shows the

testing surfaces with their main dimensions.

2.2. Data reduction

The experimental heat transfer rates were measured by two

methods. In the first method, the heat absorbed by the cooling

water was calculated by multiplying the mass flux rate by the dif-

ference between the inlet and outlet water temperatures, using the

expression:

Qw

_mCpwTout

Tin

1

where _mw is the coolant water flow rate verified by a calibrated

rotameter, Cpw is specific heat of the cooling water and Tout, Tinare the outlet and inlet temperatures of the cooling water,

respectively.

In the second method, the condensate heat transfer rate is cal-

culated as:

Qcond _mcondhlv 2

where _mcond is the condensate mass flow rate determined with the

use of a weighing scale and hlv is the latent heat of vaporization.In

order to assist the interpretation of the experimental results for the

Fig. 2. Test section.

Fig. 3. Distribution of thermocouples (a). Surface cooling system (b).

92 F.F. Czubinski et al. / Experimental Thermal and Fluid Science 47 (2013) 9097


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pure vapor, the theoretical heat transfer rate was determined using

the modified Nusselt model[2]:

hNusselt 0:943qliq qliq qvap

gsinhk

3

liqhlv

LlliqDT

24

35

1

4

3

whereq liq, jliq andl liq are the density, thermal conductivity andviscosity of the condensed liquid density respectively, qvap is thepure vapor density, j liq is the thermal conductivity, gis the forceof gravity, L is the surface length, h is the surface inclination from

the horizontal and DTis the difference between the temperatures

of the interface and the wall.

Thus, the heat transfer rate is obtained by:

QNusselt hNusseltAsurfaceTsat Twall 4

wherehNusseltis determined from Eq. (3),Asurfaceis the surface area

of each condensing plate andTsat, Twall represent the temperatures

of the pure saturated steam at the interface and the surface temper-

ature, respectively, measured by the thermocouples.

To account for the presence of NCG, the Rose model[10,13] wasused to predict the theoretical heat transfer. In this case, in con-

trast to the pure vapor, the saturation temperature at the interface

is not previously known, and thus the following equation is used to

estimate this parameter:

10FSclliqqliqlGqG

W1w0

220

21 Sc

W0W1

8

F2Sc

lGqGlliqqliq

! w0W0

5

28FX

w03

100

21

W1W0

2w0W0

8Sc 5

whereqGandlGare the density and viscosity of the airvapor mix-

ture, respectively, W0 is the air mass fraction at the condensateinterface, W1 is the air mass fraction of the mixture, Sc is the

Schmidt number for the gas mixture, F = jliq (T0 Twall) /lliqhlv,where T0 is the interface temperature, w0 = W0 W1 and X =

(Mair Mvapor)/[Mair W1 (Mair Mvapor)], where Mair and Mvaporare the molecular weight of air and water respectively.

The evaluation of the interface temperature therefore begins by

guessing a value for this parameter, which must be between the

temperatures of the surface and the initial mixture formation.

Therefore,F, in Eq.(5)can be computed and this equation is solved

for W0, the noncondensable air mass fraction at the interface. Since

the vapor mass fraction is equal to 1 W0, the partial pressure of

the vapor at the interface is evaluated by assuming equilibrium

at the interface and using Raoult and Daltons Law. With the partial

vapor pressure, the corresponding saturation temperature can beascertained from steam tables. If this new value for the interface

temperature is in good agreement with the guessed value, the heat

transfer rate can be obtained by using Eq.(3). If there is no agree-

ment, this procedure is repeated until a good agreement is reached.

In the study reported herein, these values were evaluated up to

103.

The thermophysical properties of the condensate liquid bound-

ary layer where evaluated at the temperature suggested by Minko-

wycz and Sparrow [14]. The properties of the mixture were

assessed on a mass fraction basis considering the initial state of

formation. The diffusion coefficient of water vapor in air was calcu-

lated using the following equation given by Marrero and Manson

[21],

DH2O;air 1:87 1010T

2:072

1

P16

whereTandPare the temperature and pressure of the system.

The tests were carried out at atmospheric pressure with a steam

mass flow rate of 2.4 g/s. Air (NCG) was combined with the steam

resulting in mixtures with mass fractions varying from 0 to 50%.

The surface temperature varied (between 35 and 85 C) for each

condition of NCG and each inclination.

The boiler heat loss through the insulation to the environmentwas estimated to be around 1.7% of the power. The heat loss from

the air heating section was estimated to be between 3% and 8%,

depending on the input power.

Uncertainty analysis was performed for the experimental deter-

mination of the heat transfer. The temperatures were measured by

calibrated k-type thermocouples, which provide an accuracy of

0.25 C. 5% of uncertainty was evaluated for the air mass flow

measurement. The uncertainties observed for the heat transfer rate

using Eq.(1)varied from 10 to 50%. With Eq. (2), these uncertain-

ties were less than 5%. Large uncertainties were associated with Eq.

(1) because the difference between the inlet and outlet cooling

water temperature is very small, and virtually the data acquisition

system shows almost the same values as temperature readings. On

the other hand, the uncertainties obtained using Eq.(2)are muchsmaller because the volume of collected condensate, which was

performed in a large time frame, was much larger that the smallest

reading of the weighting scale. These heat transfer uncertainties

are presented by vertical error bars superposing the data in the

plots shown in the figures of the next section. In some of the data

presented, these uncertainties are very small and difficult to ob-

serve in these plots.

One should note that the thermocouples are inserted inside the

cooled surface at 1 mm of distance from the condensation surface,

as already mentioned, causing a maximum wall temperature mea-

surement error of 1 C. The propagated error using Eq. (4) for the

heat transfer calculation is negligible.

3. Results and discussion

Fig. 6shows the test results for the heat flux transferred during

the condensation of pure vapor on a vertical smooth plate, as a

function of the surface temperature. The solid line represents the

predictions provided by the Nusselt model, Eq. (4), while the solid

and open symbols show the experimental values obtained from

Eqs. (1) and (2), respectively. The heat fluxes obtained from Eqs.

(1) and (2)showed a good agreement between them but not with

the Nusselt theory. As the condensing vapor is not stagnant (the in-

let and outlet of the test section are fully opened), a large volume of

the vapor stream does not have contact with the condensing sur-

face, flowing freely inside the condensing chamber. In order to ver-

ify this situation, two modifications were made to create a

stagnant environment. Firstly, as performed by Chung et al. [18],a deflector was used to avoid the direct impact of the upstream

Fig. 4. Surfaces tested.



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flow on the downward face of the condensing plate and the test

section outlet was almost completely closed, as can be seen inFig. 5.Table 1shows the results obtained for one testing case with

this modification, showing a good agreement with the Nusselt the-

ory This means that the experimental apparatus is well designed

and that its testing parameters are well controlled, since the heat

flux results approach the Nusselt predictions as the flow ap-

proaches the stagnant conditions, as assumed in the Nusselt mod-

el. One should note it was not possible to obtain more data with

this set up because, as the test chamber was full of warm vapor,

more cooling power in the condensate plate would be necessary.

The thermal bath used was not able to cool down the plate to tem-

perature levels tested in the other experiments (from 40 to 80 C),

so that only the plate temperature of around 83 C was tested.

Fig. 7illustrates the heat transfer results observed in the con-

densation of vapor from a mixture of airvapor, condensing on avertical smooth copper plate. The line represents the theoretical

results obtained by Rose, Eq. (5), the symbols the experimental

measurements and the vertical bars the degree of uncertainty.

As can be seen in Fig. 7, the heat transfer rate decreases system-

atically as the surface temperature and the presence of NCG in-

crease, both for experimental and theoretical results. Actually the

NCG tends to accumulate at the liquid interface, resulting in a

reduction in the partial pressure of the steam, decreasing the satu-

ration temperature at which the condensation takes place, and,

therefore, reducing the heat flux. In other words, an additional

thermal resistance is built up, reducing the heat transfer rate. On

the other hand, the heat fluxes predicted by the Rose model for

mixtures in quiescent environment are smaller than those ob-

served in streams, where the flow hitting the condensation surface

spreads the NCG boundary layer. In the latter case, the saturation

temperature is not so strongly affected by the accumulation of

the NCG. In other words, Rose considered a stagnant environmen-

tal while in this work the mixture was in an upstreamflow. For the

case with a 20% air mass fraction, the heat flux decrease, compared

with the pure vapor case, was between 15% and 22%, depending on

the surface temperature. Rose [13] reported that for a quiescent

mixture with a 0.5% air mass fraction the NCG caused a large ther-

mal resistance, with a decrease of more than 50% in the heat flux.

Similar trends were observed for plate inclinations of 60, 45

and 30, as shown in Figs. 810, respectively, for smooth copper

plate

The Nusselt theory shows that the heat transfer for vertical

plates (in a quiescent environment) is larger than for tilted sur-

faces, as the gravity pulls the liquid film downward. However, on

comparingFig. 6(vertical case) withFigs. 810(tilted cases), it is

clear that this effect is not observed. It was noted that, as the sur-

face approximates the horizontal position the ascending stream

tends to be more trapped by the condensing surface and the vapor

of the airvapor mixture exchanges more heat with the cooled sur-

face, increasing the production of condensate and, therefore, the

heat transfer coefficient. It is well known that the mean liquid film

thickness, which acts as thermal resistance, increases as the sur-

face length increases (larger condensing area). As already ex-

plained, the length of the tested plates decreases as their

inclinations tend to the horizontal position, and therefore, the film

thickness and the thermal resistance are expected to decrease.

Actually this is not observed. Taking Fig. 7 (vertical case) and 8

(60 case), although the tested surface is the same (and so the same

mean liquid film thickness), the heat transfer rate for the inclinedsurface is larger than for the vertical case. Therefore, the tests show

that the effect of the liquidfilm thickness is less important than the

influence of inclination. Consequently, the condensation heat

fluxes were higher for an inclination of 30 (Fig. 10) than for 45

(Fig. 9) and 60 (Fig. 8).

It is interesting to note that during the experiments and for all

slopes tested, no condensate drops were observed to fall from the

cooling surface. Instead, all of the condensate liquid film ran down

the downward condensing surface and was collected by the drain.

This means that all the condensate formed over the condensing

surface was collected in the drain, improving the quality of the

experimental data.

Also, one should note that the heat transfer calculated using Eq.

(2) is always larger than that calculated using Eq. (1) for all inclinedtesting cases, although most of the data lie within the experimental

uncertainty range. This is not observed for the vertical case. Even

though it is not possible to explain exactly why this happens, it

is believed that this difference is due to the different stream config-

urations over the condensing plates in both positions. More inves-

tigation would be need to a deep understanding of this

phenomenon. In spite of this difference the comparison between

the heat transfer determined by Eqs.(1) and (2)is better than with

the results of the literature theoretical models.

The influence of other surface materials was also investigated in

this study. Samples of aluminum and copper, in the smooth and

grooved surface configurations, were tested for an inclination of

60. In the following plots the theoretical results and degrees of

uncertainty are not shown, as no new information other than thatpreviously discussed herein can be obtained.

Fig. 5. Modification in the test section.

40 60 80

0

100

200

300

400Eq. (1)

Eq. (2)

Eq. (4) Nusselt

Surface Temperature [C]

q"[kWm

-2]

Fig. 6. Smooth plate in the vertical position using pure vapor.

Table 1

Heat transfer results.

Surface temperature (C) Heat flux (kW/m2)

Eq.(4)Nusselt Eq.(1) Eq.(2)

82.6 138.9 125.4 128.3



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InFigs. 1113, the heat flux is plotted against surface tempera-

ture for the condensing surfaces at 60 degrees of inclination, for

grooved copper, grooved aluminum and smooth aluminum sur-

faces, respectively. In the case of the grooved surfaces, the heat

fluxes were evaluated based on the apparent projected surface

area.

The results for the heat fluxes show that the behaviors of thesmooth and grooved surfaces follow the same trend, that is, an in-

crease in the surface temperature and in the amount of NCG leads

to a decrease in the heat flux.

On comparingFigs. 1113withFig. 8(all experimental data for

surfaces with the same inclination) it can be observed that the

highest heat fluxes were obtained for the pure vapor with the

grooved copper surface. The heat transfer was enhanced by around

10% compared with the smooth copper surface. This result is muchlower than that observed by Markowitz et al. [11], where an

40 60 80

0

10

20

30

Eq.1PureVaporEq.2PureVaporEq.1 - 20% NCGEq.2 - 20% NCGRose - 20% NCGEq. 1 - 30% NCGEq. 2 - 30% NCGRose - 30% NCGEq. 1 - 40% NCGEq. 2 - 40% NCGRose - 40% NCGEq. 1 - 50% NCGEq. 2 - 50% NCGRose - 50% NCG


q"[kWm

-2]

Fig. 7. Smooth plate in the vertical position.

40 60 80

0

10

20

30

40

Eq.1PureVapor

Eq.2PureVapor

Eq.1 - 20% NCGEq.2 - 20% NCG

Rose - 20% NCG

Eq. 1 - 30% NCG

Eq. 2 - 30% NCG

Rose - 30% NCG

Eq. 1 - 40% NCG

Eq. 2 - 40% NCG

Rose - 40% NCG

Eq. 1 - 50% NCG

Eq. 2 - 50% NCG

Rose - 50% NCG


q"[kWm

-2]

Fig. 8. Smooth plate with an inclination of 60.

40 60 80

0

10

20

30

40

50Eq.1PureVaporEq.2PureVaporEq.1 - 20% NCGEq.2 - 20% NCGRose - 20% NCGEq. 1 - 30% NCGEq. 2 - 30% NCGRose - 30% NCGEq. 1 - 40% NCGEq. 2 - 40% NCGRose - 40% NCG

Eq. 1 - 50% NCGEq. 2 - 50% NCGRose - 50% NCG


q"[kWm

-2]




7/8

enhancement of around 150% was achieved for ascending Freon-

113 (comparing grooved with smooth surfaces). In the study re-

ported herein, the vapor flows upwards and the grooves are not

able to hold the vapor for long and, therefore, the improvement ob-

served is due to an increase in the condensing area, while Marko-

witz et al.[11]attributed the improvement to the surface tension

and the augmentation of the liquid vapor film curvature.

For different types of surfaces, it is observed that with an in-

crease in the NCG in the airvapor mixture the heat fluxes for

the four different surfaces tend to assume the same value. This

can be clearly observed, for instance, in the case of 50% of NCG.

In other words, copper and aluminum with a grooved or smooth

surface showed the same behavior for high amounts of NCG. As ex-

plained by Markowitz et al.[11], the NCG tends to accumulate in

the grooves, blocking the access of the condensate liquid to the val-

ley regions, reducing the efficiency of the surfaces.

In summary, when airvapor mixture ascending flows reach

condensing surfaces, the enhancement provided by the channels

40 60 80

0

10

20

30

40

50Eq.1PureVaporEq.2PureVaporEq.1 - 20% NCGEq.2 - 20% NCGRose - 20% NCGEq. 1 - 30% NCGEq. 2 - 30% NCGRose - 30% NCGEq. 1 - 40% NCGEq. 2 - 40% NCGRose - 40% NCGEq. 1 - 50% NCGEq. 2 - 50% NCGRose - 50% NCG


q"[kWm

-2]


40 60 80

0

10

20

30

40

50


Eq. 1PureVapor

Eq. 2PureVapor

Eq.1 - 20% NCG

Eq. 2 - 20% NCG

Eq.1 - 30% NCG

Eq. 2 - 30% NCG

Eq.1 - 40% NCG

Eq. 2 - 40% NCG

Eq.1 - 50% NCG

Eq. 2 - 50% NCG

q"[kWm

-2]

Fig. 11. Grooved copper surface at an inclination of 60

.

40 60 80

0

10

20

30

40


Eq. 1PureVapor

Eq. 2PureVapor

Eq.1 - 20% NCG

Eq. 2 - 20% NCG

Eq.1 - 30% NCG

Eq. 2 - 30% NCG

Eq.1 - 40% NCG

Eq. 2 - 40% NCG

Eq.1 - 50% NCG

Eq. 2 - 50% NCG

q"[kWm

-2]

Fig. 12. Grooved aluminum surface at an inclination of 60.



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was observed only in the case of the pure vapor. With NCG, the val-

ues for the heat transfer rate are similar for different condensing

surfaces.

4. Conclusions

Film condensation of ascending flows of airvapor mixtures

was experimentally investigated on vertical and downward in-

clined surfaces, in relation to the effects of different inclination an-

gles, sub-cooling levels and the presence of NCG in the upstream

flow, for smooth and grooved surfaces of copper and aluminum.

The test results were evaluated based on the Nusselt and Rose

theories as benchmarks. Although the presence of NCG had an ef-

fect on the decrease in the heat transfer rate, the configuration of

the condensation system plays the main role since the mixture

flows upstream freely inside the condensing chamber.

Reverse trends were observed in the heat transfer rates for the

test results and the theory predictions, considering the effect of

inclination. The experimental results show that the heat fluxes in-

crease as the plate inclination decreases from the vertical to an

inclination of 30. As the amount of NCG increases, a systematic

reduction of the heat transfer rate is observed. Due to the mixture

flow, the formation of the NCG boundary layer over the condensate

liquid is disturbed and the decrease in the heat transfer rate was

not so strongly affected by the NCG, as observed in the literature

for quiescent mixtures. The heat transfer variation with the con-

denser surface sub-cooling level follows the same trend as pre-

dicted by the theoretical models.

Furthermore, the enhancement of the heat transfer rates for

grooved surfaces was small and observed only in the case of the

pure vapor, especially for the copper samples. The effect of grooves

was negligible for streams with NCGs.

Acknowledgement

The authors would like to acknowledge the financial support

provided by Petrobras and CNPq for this research.

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40 60 80

0

10

20

30

40

q"[kWm

-2]


Eq. 1PureVapor

Eq. 2PureVapor

Eq.1 - 20% NCG

Eq. 2 - 20% NCG

Eq.1 - 30% NCG

Eq. 2 - 30% NCG

Eq.1 - 40% NCG

Eq. 2 - 40% NCG

Eq.1 - 50% NCG

Eq. 2 - 50% NCG

Fig. 13. Smooth aluminum surface with an inclination of 60.


Condensation on Inclined Surfaces

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