Group Title: 7th International Conference on Multiphase Flow - ICMF 2010 Proceedings
Title: 14.7.1 - Performance Evaluation of a Horizontal Helically Coiled Tube in Boiling Heat Transfer Enhancement and its Effect on Pressure Drop
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 Material Information
Title: 14.7.1 - Performance Evaluation of a Horizontal Helically Coiled Tube in Boiling Heat Transfer Enhancement and its Effect on Pressure Drop Boiling
Series Title: 7th International Conference on Multiphase Flow - ICMF 2010 Proceedings
Physical Description: Conference Papers
Creator: Akhavan-Behabadi, M.A.
Shemirani, F.M.
Aria, H.
Publisher: International Conference on Multiphase Flow (ICMF)
Publication Date: June 4, 2010
 Subjects
Subject: boiling
coiled tube
heat transfer
helically
pressure drop
 Notes
Abstract: An experimental study has been carried out to investigate flow boiling heat transfer and pressure drop characteristics of R-134a inside horizontal straight and helically coiled tubes. Both test-evaporators are tube-in-tube heat exchangers, with an inner copper tube of 8.3 mm inside diameter and 9.52 mm outside diameter and an outer tube with 29 mm inner diameter. The length of straight and helically coiled evaporators is 1200 mm and 5870 mm, respectively. Refrigerant R-134a flows in the inner tube and warm water flows in the annulus in opposite direction. The experiments were performed with three refrigerant mass velocities of 112, 132, and 152 kgm-2s-1. It has been found that the use of helically coiled tubes enhance the heat transfer coefficient on relatively higher pressure drop penalty, in comparison to that for the straight tube.
General Note: The International Conference on Multiphase Flow (ICMF) first was held in Tsukuba, Japan in 1991 and the second ICMF took place in Kyoto, Japan in 1995. During this conference, it was decided to establish an International Governing Board which oversees the major aspects of the conference and makes decisions about future conference locations. Due to the great importance of the field, it was furthermore decided to hold the conference every three years successively in Asia including Australia, Europe including Africa, Russia and the Near East and America. Hence, ICMF 1998 was held in Lyon, France, ICMF 2001 in New Orleans, USA, ICMF 2004 in Yokohama, Japan, and ICMF 2007 in Leipzig, Germany. ICMF-2010 is devoted to all aspects of Multiphase Flow. Researchers from all over the world gathered in order to introduce their recent advances in the field and thereby promote the exchange of new ideas, results and techniques. The conference is a key event in Multiphase Flow and supports the advancement of science in this very important field. The major research topics relevant for the conference are as follows: Bio-Fluid Dynamics; Boiling; Bubbly Flows; Cavitation; Colloidal and Suspension Dynamics; Collision, Agglomeration and Breakup; Computational Techniques for Multiphase Flows; Droplet Flows; Environmental and Geophysical Flows; Experimental Methods for Multiphase Flows; Fluidized and Circulating Fluidized Beds; Fluid Structure Interactions; Granular Media; Industrial Applications; Instabilities; Interfacial Flows; Micro and Nano-Scale Multiphase Flows; Microgravity in Two-Phase Flow; Multiphase Flows with Heat and Mass Transfer; Non-Newtonian Multiphase Flows; Particle-Laden Flows; Particle, Bubble and Drop Dynamics; Reactive Multiphase Flows
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Bibliographic ID: UF00102023
Volume ID: VID00363
Source Institution: University of Florida
Holding Location: University of Florida
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Resource Identifier: 1471-AkhavanBehabadi-ICMF2010.pdf

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7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010


Performance Evaluation of a Horizontal Helically Coiled Tube in Boiling Heat Transfer
Enhancement and its Effect on Pressure Drop



M.A. Akhavan-Behabadi, F.M. Shemirani and H. Aria

School of Mechanical Engineering, College of Engineering, University of Tehran, Tehran, IRAN
akharan~ut. ac. ir, fshemirani~ut. ac. ir, razi. pooyan agmail. com


Keywords: Boiling, Coiled Tube, Heat transfer, Helically, Pressure Drop




Abstract

An experimental study has been carried out to investigate flow boiling heat transfer and pressure drop
characteristics of R-134a inside horizontal straight and helically coiled tubes. Both test-evaporators are
tube-in-tube heat exchangers, with an inner copper tube of 8.3 mm inside diameter and 9.52 mm outside
diameter and an outer tube with 29 mm inner diameter. The length of straight and helically coiled
evaporators is 1200 mm and 5870 mm, respectively. Refrigerant R-134a flows in the inner tube and warm
water flows in the annulus in opposite direction. The experiments were performed with three refrigerant
mass velocities of 112, 132, and 152 kgm-2S-1. It has been found that the use of helically coiled tubes
enhance the heat transfer coefficient on relatively higher pressure drop penalty, in comparison to that for
the straight tube.


Introduction

Curved tubes have been introduced as one of the passive
heat transfer enhancement techniques and are widely used
in various industrial applications due to their compact
structure and high heat transfer coefficient. Helical coils
are well known types of curved tubes which have been
used in a wide variety of applications such as heat recover
processes, air-conditioning and refrigeration systems
chemical reactors, food and dairy processes.
Heat transfer and pressure drop characteristics of
refrigerants during condensation have been studied by a
large number of researchers. The study of the boiling heat
transfer and pressure drop inside helically coiled tubes has
received comparatively little attention in the literature.
Garimella et al.[1] studied the forced convection heat
transfer in coiled annular ducts. They found that the heat
transfer coefficients obtained from the coiled annular ducts
were higher than those obtained from a straight annulus-
especially in the laminar region.
Xin et al.[2] investigated the single-phase and two-phase
air-water flow pressure drop in annular helicoidal pipes
with horizontal and vertical orientations. Experiments were
performed for the superficial water Reynolds number from
210 to 23,000 and superficial air Reynolds number from 30
to 30,000. A friction factor correlation for single-phase
flow in laminar, transition and turbulent flow regime was
proposed. The two-phase flow pressure drop multipliers in
annular helicoidal pipe were found to be dependent on the


Lockhart-Martinelli parameter and the flow rate of air or
water. The effect of flow rate tended to decrease as the
pipe diameter decreased.
An experimental study on condensation heat transfer of
R-134a flowing inside a helical pipe while cooling water
flowing in annulus, was reported by Yu et al. [3]. The
experiments were performed for mass velocities of 100-
400 kg/nr's and the Reynolds number of cooling water in
the ranges of 1500-10000. They found that the orientation
of the helical pipe had significant effect on the heat
transfer coefficient.
Wongwises and Polsongkram [4] investigated both
condensation and evaporation of R-134a experimentally to
study the effect of heat flux, mass velocity, and evaporation
temperature on heat transfer coefficients and pressure drop.
They conducted their experiments for high mass velocities
ranging from 400-800 kgm s-', and provided correlations
for refrigerant Nusselt number and pressure drop.
Al Hajeri et al. [5] reported an experimental investigation
of condensation heat transfer and pressure drop of
refrigerant R-134a inside a helical tube. Their study
concerns the condensation of R-134a flowing through
annular helical tubes with different operating refrigerant
saturated temperatures including 360C, 420C, and 480C.
The average pressure drop is measured and compared. The
measurements of R-134a were performed on mass
velocities ranges from 50 to 680 kgm s '. Their
experimental results shows that the refrigerant side heat
transfer coefficient and overall heat transfer coefficients






















































FIGURE 1
(A) STRAIGHT AND HELICALLY COILED TUBES USED AS TEST
SECTION (B) EXPERIMENTAL SET-UP


TABLE 1
DIFFERENT ELEMENTS OF FIGURE 1
1Compressor 6 Expansion valve
2 Condenser 7 Pre evaporator
3 Flow meter 8 Test evaporator
4 IReceiver 9 After evaporator
5 Filter drier 10 Accumulator
Thermocouple P Pressure gauge
DP Differential pressure transducer



Data Reduction

Figure 2 shows the refrigerant flow diagram from the
entrance of expansion valve to the exit of after evaporator.
The following procedure is used to determine the quality
of refrigerant at the inlet and exit of the test evaporator, the
heat transfer coefficient and pressure gradient, from the
data recorded during each test run at steady state


7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010


Location of thennoccuple


increase with increase of the mass flux of flowing R-134a
and decrease of vapor saturated temperatures of R-134a.
The pressure drops in annular helical tubes increase as the
mass velocity of refrigerant increases.
Yang et al. [6] carried out a numerical-experimental study
on flow boiling of R-141B in a horizontal coiled tube.
They conducted their tests in two flow rates of 10 and 15
1/hr and three different heat fluxes, and observed that flow
velocity predicted in the simulations showed a strongly
dependence on the phase distribution. Moreover, they
depicted the evolution of flow pattern inside the tube.
Existing Ozone in the Stratosphere absorbs deadly
radiation of ultraviolet from sun. This is why the depletion
of Ozone layer is becoming a great concern to most
societies. A major cause of this phenomenon is releasing
large amounts of CFCs which converts Ozone to Oxygen.
Since the thermo physical properties of R-134a are very
similar to those of the conventional CFCs, it is the best
choice to replace them. Therefore, the properties of R-134a
should be studied in detail before it is applied.
In the present study, an experimental investigation on flow
boiling heat transfer and pressure drop of R-134a inside
horizontal straight and helically coiled tube evaporators
have been carried out.

Experimental Facility

The experimental set-up which was used in this
investigation is a well instrumented vapor compression
refrigeration system. There are three different evaporators;
namely pre-evaporator, test evaporator, and after
evaporator. Both pre-evaporator and after-evaporator were
electrically heated. The purpose of using pre-evaporator
was to achieve any desired vapor quality at the inlet of test
evaporator. After-evaporator was used to ensure that all the
fluid in the test tube is superheated before entering the
compressor.
Both the test-evaporators were tube-in-tube heat
exchangers, with an inner copper tube and an outer
stainless steel tube. The length of straight and helically
coiled test evaporators was 1200 mm and 5870 mm,
respectively. Refrigerant R-134a flowed in the inner tube
and warm water flowed in the annulus in opposite
direction.
The straight and helically coiled tubes used as test
evaporators and also the schematic diagram of
experimental set-up have been shown in Figure 1.
Different elements of this set-up are introduced in Table 1.
The inside diameter of the inner copper tube is D,=9.52
mm and its thickness is 0.62 mm. The outside diameter of
the outer tube is De=34 mm. The pitch of the coil is b=45
mm and the outer diameter of the coil is D=305 mm. To
measure the outside surface temperatures of inner tube,
T-type thermocouples were used at six different locations
on both straight and helically coiled tubes (see Figure 1). At
each section, two thermocouples were employed on
opposite sides of outer surface of the inner tube. The
pressure drop across the test section was also measured by
an accurate differential pressure transducer apparatus. The
provisions were also made to measure all the other
necessary parameters.


n


~V~---~Vb


































































where, vilwr cywr Twa aand Twru are themass
flow rate, specific heat, inlet temperature and outlet
temperature of hot water flowing in test evaporator,
respectively.

The average vapor quality in test evaporator, x:

The average vapor quality in test evaporator is defined as:


7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010

X, + 3
x= (6)

The average heat transfer coefficient, h:

h = 27
A w~-T)
Where, Tand Al are the average inside wall
temperature, and inside surface area of test evaporator
inner tube, respectively. T, is the average temperature of the
refrigerant at the inlet and outlet of test evaporator.

Al = yr DI L (8)


conditions. The thermodynamic and transport properties of
R-134a are evaluated using ASHRAE Handbook of
Fundamentals [7].


T1 PI Qa P2 p Pa

(I 1 2
use .vawm cnawam mwat" t
FIGURE 2
FLOW DIAGRAM IN EXPANSION VALVE AND EVAPORATORS

The inlet vapor quality of the test evaporator, x2




72

where a2 iS the refrigerant enthalpy at the test evapo
inlet, if is the enthalpy of saturated liquid and ifg i
enthalpy of vaporization.


(1) The fricionalpressue graienPF)

The two-phase flow pressure gradient is the sum of three
rator components: the frictional pressure gradient, the
s the momentum pressure gradient and the hydrostatic pressure
gradient as follows:


12 = 11 1 (2)

where i, is the enthalpy of the liquid refrigerant at the inlet
of pre-heater, rizref is the mass flow rate of the refrigerant,

Qz is the heat transfer rate from electrical coil to the
refrigerant in the pre-evaporator.
The outlet vapor quality of the test section, x3 :


i3 f
x3 = (3)


where i3 iS the refrigerant enthalpy at the test evaporator
outlet. As a consequence, the outlet enthalpy of the
refrigerant flow is calculated from


13 = 12 + (4)

where the heat transfer rate from the hot water to the
refrigerant in the test evaporator, Q2, iS obtained from


The hydrostatic pressure gradient is defined as:


(g = gsini9 ap +(1-a)pl,


where p, and pf are the density of vapor and liquid
TOSpectively and a is the void fraction.
The momentum pressure gradient can be calculated by the
results of one dimensional two phase separated flow model
analySIS aS follows:


dP 2d (1 -x)2 2
dz dz p P(1-a) pga


in which, Gt is mass velocity of refrigerant. The void
fraction values in the above equations can be determined
from [8] as follows:


a=Pg [(~~(iPg

1.18(1- x) [go-(pf pg )]o 5 P
G2 05


e2 = iwat p,wat (Twat~,,n Twat,out


Finally, the frictional pressure gradient is computed by
subtracting the calculated momentum and hydrostatic
pressure gradients from measured two phase total pressure
gradient by Equation (13).


dP)= dPF dP I dP
dz dz dz dz


dP dP dP dP
-F --a -I-g
dzdz dz dz






7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010

It is found from Figure 4 that by increasing the vapor
quality, the pressure drop increases. At a same average
quality, it can be seen that the pressure drop increases with
the increasing of mass velocity. This can be explained by
the fact that, any increasing in mass velocity accelerates
the two-phase flow velocity, which leads to a greater tube
wall shear stress.

Helically coiled tube

Heat transfer coefficients of horizontal helically coiled
tube versus average vapor quality for different mass
velocities are depicted in Figure 5. It can be seen that same
as straight tube, the heat transfer coefficient increases by
increasing of vapor quality and mass velocity.
Besides the reasons stated about this phenomenon in
straight tube section, the effect of strengthening the
Secondary flow should be noted. By increasing the strength
Of secondary flow, the redopisition and entrainment rates
increase. This enhancement induces the increased number
and larger size of waves on the liquid film surface, which
leads to increasing of heat transfer area. In addition, the
higher velocity, due to the increasing mass velocity,
increases the degree of turbulence of the fluid, which
results in a higher heat transfer coefficient.


s000


Results and Discussion

Straight Tube

The comparison of heat transfer coefficient of evaporative
R-134a flow versus average vapor quality for three
different mass velocities is illustrated in Figure 3.


4500 ,,


2 4000
S3500
3000


S2000
S1500
S1000
1500


O "


a .

.


*G=112 Kg/m2-s
o G=132 Kg/m2-s
A G=152 Kg/m2-s

0 8 1


0 0 2 0 4 0 6
Vapor Quality


FIGURE 3
HEAT TRANSFER Ci)EFFICIENT <)F Hi)RIZi)NTAL STRAIGHT TUBE

It is obvious from Figure 3 that the refrigerant side heat
transfer coefficient continuously increases by increasing
vapor quality. This trend has two major reasons. Firstly
because of the boiling, the mass of liquid refrigerant which
has a low specific volume decrease, which leads to
increase the velocity of fluid and therefore a higher heat
transfer coefficient is achieved. Secondly, as the average
vapor quality increases, the liquid film thickness decreases,
which reduces the thermal resistance in the liquid film and
a higher heat transfer coefficient is obtained. It is also
observed that by increasing the mass velocity the heat
transfer coefficient increases. However, the mass velocity
is found to have less effect on heat transfer coefficient at
low average vapor qualities.
The frictional pressure gradient of R-134a flowing in
straight horizontal tube versus average vapor quality for
different mass velocities is depicted in Figure 4.


--8000
S7000
S6000
~5000
S4000
'3000
S2000
1000


a =
a


As a


I

4 *


* G=112 Kg/m2-s
SG=132 Kg/m2-s
A G=152 Kg/m2-s


O 0 2 0 4 0 6 0 8 1
Vapor Q~uality

FIGURE 5
HEAT TRANSFER CO EFFICIENT <)F HO RIZO NTAL HELICALLY CO ILED TUBE

The variation of frictional pressure gradient of horizontal
helically coiled heat exchanger versus average vapor
quality is depicted in Figure 6 for different mass velocities.
The trend of changes of pressure drop is just same as that
of straight tube. Except the reasons stated for these trends
in straight tube section, there are two more reasons: the
longer distances that fluid particles should move and the
turbulence causes by secondary flows.
Comparison of heat transfer coefficient and frictional
preSsure gradient of helically coiled tube and straight tube
veTSus average vapor quality for different mass velocities
iS depicted in Figures 7-8. From these figures it is found
that at a same mass velocity and vapor quality, both heat
transfer coefficient and frictional pressure gradient of
horizontal helically coiled tubes are higher than those of
Straight tubes due to the effect of secondary flows.


D *
*


*G=112k/2
aG= 132k/2
G=152k/2


0 0 1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 1
Vapor Qualty

FIGURE 4
FRICTI<)NAL PRESSURE GRADIENT <)F Hi)RIZi)NTAL STRAIGHT TUBE
















E` r .






2 *G=112 Kg/m2-s
1-I a G=132 Kg/m2-s
rG=152 Kg/m2-s

0 0 1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 1
Vapor Q~ualitr

FIGURE 6
FRICTIONAL PRESSURE GRADIENT OF HORIZONTAL HELICALLY COILED
TUBE


9000
8000
S7000
6000


O 4000
"3000
2000
loo


helical G=112Kg/m2-s
a helical G=132Kg/m2-s
helical G=152Kg/m2-s
xstraight G=112Kg/m2-s
straight G=132Kg/m2-s
+straight G=152Kg/m2-s


.


A


r a.

xxx x x x x
x


0 01 02 03 04 05 06 07 08 09 1
Vapor Quality

FIGURE 7
COMPARISON OF HORIZONTAL STRAIGHT AND HELICALLY COILED TUBE
HEAT TRANSFER COEFFICIENT


7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010

them and can produce the conditions for the comparison of
these two parameters. This third parameter can be the
compressor power consumption for test-evaporator length
to heat transfer coefficient ratio, W/h, or compressor power
consumption to heat transfer in it, W/Q. The reason is that
the pressure drop increasing rate directly depends upon the
COmpressor power consumption rate.
IH perfOrmance evaluation of twisted tape insert on boiling
heat transfer of R-12 inside horizontal tubes, Agrawal and
Varma [9] also used the ratio of pumping power to heat
transfer enhancement rate as the evaluation criterion. They
calculated the increased pumping power due to the
preSence of twisted tape in test-evaporator by multiplying
the volumetric flow rate and produced pressure drop in test
evaporator:


W = v AP,, (14)

Finally, based on this criterion, they prepared a figure for
different tubes. In the present work, firstly, the ratio of the
pOWeT COHSumption of compressor to heat transfer
COefficient of straight tube (Winl is calculated; then
the same ratio for helically coiled tubes (Winl is
COmputed. The compressor power is calculated from (14).
Eventually, the ratio for straight tube is divided by that for
the flow in helically coiled tube.



MP, Tes it hflelical 1 R

h Stagt hStraight R
(15)
(Test Helical

~M~s(MrTesragt1 Stah


If the ratio of Rh p, iS greater than one, to apply the coiled
tube is useful.
The variation of Rh/R, with mass velocity has been shown
in Figure 9. From this figure it is concluded that among the
different vapor qualities and different mass velocities, at
G=112 kgm-2S-1 and x=0.82, the horizontal helically coiled
tube has the best performance with Rh/R,=1.15. At a
constant vapor quality, by reducing mass velocity, Rh/R,
decreases.


14


*helical G=112Kg/m2-s

xstrlaigh G=112Kg/m2-s
straight G=132Kg/m2-s a. --.
+straight G=152Kg/m2-s A a **



t a


e
a a


0 02 04 06 08 1
Vapor Quality

FIGURE 8
COMPARISON OF HORIZONTAL STRAIGHT AND HELICALLY COILED TUBE
FRICTIONAL PRESSURE GRADIENT


Performance evaluation of helically coiled tubes

The present experimental data analysis shows that the use
of horizontal helically coiled tube increases the heat
transfer coefficient with a pressure drop penalty. The
pressure drop and heat transfer coefficient are two
independent parameters which do not relate by any
equation, so comparison is difficult. Therefore, a third
parameter should be considered which relates to both of


O 6


*G= 112 Kg/m2-s
a G= 132 Kg/m2-s
AG= 152 Kg/m2-s

0 4 0 6 08 8
Vapor Quality

FIGURE 9
VERSUS AVERAGE VAPOR QUALITY AT DIFFERENT
MASS VELOCITIES


0 2


EVALUATION OF Ro/Rn






7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30-June 4, 2010

References

[1] Garimella, S., Richards, D.E. and Christensen, R.N.,
Experimental investigation of heat transfer in coiled
annular duct, Transactions of ASME, Vol. 110, 329-
336 (1998).
[2] Xin, R.C., Awwad, A., Dong, Z.F. and Ebadian, M.A.,
An experimental study of single-phase and two-phase
flow pressure drop in annular helicoidal pipes,
International Joumnal of Heat and Mass Transfer, Vol.
18, 482-488 (1997).

[3] Yu, B., Han, J.R., Kang, H.J., Lin, C.X., Awwad, A.
and Ebadian, M.A., Condensation heat transfer of R-
134a flow inside helical pipes at different orientations,
Int. Communications in Heat and Mass Transfer, Vol.
30, 745-754 (2003).

[4] Wongwises, S. and Polsongkram, M., Evaporation
heat transfer and pressure drop of HFC-134a in a
helically coiled concentric tube-in-tube heat
exchanger, Intemnational Joumnal of Heat and Mass
Transfer, Vol. 49, 658-670 (2006).

[5] Al-Hajeri, M.H., Koluib, A.M., Mosaad, M. and Al-
Kulaib, S., Heat transfer performance during
condensation of R-134a inside helically coiled tubes,
Energy Conversion and Management, Vol. 48, 2309-
2315 (2007).

[6] Yang, Z., Peng, X.F. and Ye, P., Numerical and
experimental investigation of two phase flow during
boiling in a coiled tube, Intemnational Joumnal of Heat
and Mass Transfer, Vol. 51, 1003-1016 (2008).

[7] ASHRAE, Handbook of Fundamentals, American
Society of Heating, Refrigerating, and Air
Conditioning Engineers, Atlanta, GA, 1997.

[8] Ould Didi, M.B., Kattan, N. and Thome, J.R.,
Prediction of Two-Phase Pressure Gradients of
Refrigerants in Horizontal Tubes, Intemnational
Journal of Refrigeration, Vol. 25, 935-944 (2002).

[9] Agrawal, K.N. and Varma, H.K., Heat transfer during
forced convection boiling of R-12 under swirl flow,
Journal of Heat Transfer, Vol. 108, 567-572 (1986).


Also from Figure 9 It can be seen that in some points
specially at high vapor quality region, this ratio is greater
than one; in these points the use of helically coiled tube is
recommended.
If the ratio of Rh p, iS less than one, only under the certain
conditions and for specific applications considering the
appropriate consistence between the heat transfer
performance and the rate of energy penalty paid for
increased pressure drop, the horizontal helically coiled
tube can be used. For example, in the places where due to
the limitation of spaces, more compact and smaller heat
exchangers are needed and low cost energy resources are
available to supply the excess power consumed for the
increase in pressure drop.
Finally, it is concluded that, the penalty which should be
paid indirectly for compensation of pressure drop
increasing due to the use of horizontal coiled tube instead
of straight tube, generally, is more than the related heat
transfer enhancement obtained. In the other word, the main
purpose of using the coiled tube instead of straight tube is
to have a compact heat exchanger with a lower occupied
space, not saving in consumption of energy.

Conclusions

1- The heat transfer coefficient of both straight and
helically coiled tube continuously increased by
increasing of vapor quality and mass velocity.
2- The pressure drop for both straight tube and
helically coiled tube increased with the increase
of mass velocity and vapor quality.
3- At a constant mass velocity and vapor quality,
both the heat transfer coefficient and frictional
pressure gradient of horizontal helically coiled
tubes were higher than those of straight tubes.
4- It was found that at mass velocity of 112 kgm-2S-1
and vapor quality of 0.82, the horizontal helically
coiled tube has the best performance with
Rh p,=1.15.
5- Finally, it was concluded that, the main purpose of
using the coiled tube instead of straight tube is to
have a compact heat exchanger with a lower
occupied space, not saving in consumption of
energy.




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