Page 1
ORIGINAL
A review on saturated boiling of liquids on tube bundles
Abhilas Swain • Mihir Kumar Das
Received: 25 October 2012 / Accepted: 28 October 2013
� Springer-Verlag Berlin Heidelberg 2013
Abstract A review of recent investigation on boiling of
saturated liquids over plain and enhanced tube bundles has
been carried out taking the earlier review works as refer-
ence point. The experimental observations of various
geometry and performance parameters studied by
researchers are analyzed keeping current demand of
industries in design and development of compact, efficient
heat exchanging devices. The study shows that tube spac-
ing plays an important role in determination of compact-
ness of the heat exchanger.
List of symbols
b Constants in equation
B Scaling factor in equation
CA Factors in the Eq. (2)
Cq Factors in the Eq. (2)
csf Constant depending on type of surface
D Diameter of the tube (mm)
Db Bubble diameter (mm)
Fp Pressure correction factor
Fb Bundle boiling factor
G Mass flux (kg/s m2)
HTC Heat transfer coefficient
h Heat transfer coefficient (W/m2 K)
L Length of the tube (mm)
Lfg Latent heat of vapourisation
m Mass flow rate (kg/s)
M Molecular mass (gm)
N Row number in the bundle
Nu Nusselt number
Pr Reduced pressure (Psat/Pcrit)
Pr Prandlt number
q Heat flux from heater surface (W/m2)
Ra Roughness average value (lm)
Re Reynolds number
s Slip ratio
U Superficial velocity
Vo Voidage number
x Vapor quality
YIB Boiling intensity parameter
Z Convective parameter in Eq. (12)
Greek letters
l Viscosity (P)
e Void fraction
q Density (kg/m3)
Subscripts
nb Nucleate boiling
pb Pool boiling
cv Convection
cb Convective boiling
2u Two phase
f Liquid phase
g/G Vapour or gaseous phase
ext External
Dry Dry out condition
1 Introduction
The advancement in enhancing the performance of heat
transfer equipment used in Chemical, Petroleum and Power
A. Swain (&) � M. K. Das
School of Mechanical Science, Indian Institute of Technology
Bhubaneswar, Samantapuri Campus, Bhubaneswar 751013,
Odisha, India
e-mail: [email protected]
M. K. Das
e-mail: [email protected]
123
Heat Mass Transfer
DOI 10.1007/s00231-013-1257-1
Page 2
sector industries is at pace. Researchers are consistently
thriving for achieving a high rate of heat transfer to
improve the efficiency, reduce the energy consumption,
operating cost, material cost, manufacturing cost, refrig-
erant inventory, operational space and to improve the
durability of these equipments.
The fluids of single phase in natural convection has heat
transfer coefficient values (in W/m2 K) of the order of 5–10
(for gases), 100–200 (for liquids): for forced convection it is of
the order of 30–150 (for gases) and 100–1,000 (for liquids):
for phase change process i.e. boiling and condensation it is of
the order of 1,000–2,000 (low values) and 4,000–5,000 (high
values). The order of magnitude of heat transfer coefficient
achieved in boiling heat transfer seems to offer solutions to the
factors responsible for development of heat transfer equip-
ment to meet modern demand of industries.
The difficulty in getting proper insight into the boiling
over tube bundles in industrial heat exchangers is difficult
because industrial scale experiment is expensive. The
researchers conduct laboratory scaled experiments over
tube & tube bundles to analyze the process. In two phase
shell-tube heat exchangers the shell side boiling is used
because it is easy to remove the vapor produced. Thus in
laboratories the conditions are realized by set up in which
the tubes are supplied with hot fluid inside the tubes and the
boiling liquid flows outside. In some conditions cartridge
heaters are inserted inside the tubes for supply of heat. The
arrangement of tubes in heat exchangers may be triangular
pitched or square pitched. The Fig. 1 highlights the typical
geometry of the tube bundle of the shell and tube heat
exchangers. The tube bundles may be arranged in an inline
or staggered configurations. The staggered configuration
may be with squared pitch or equilateral triangle pitch and
also presence of cleaning lanes is important. A lot of cor-
relations, empirical and semi-empirical on boiling heat
transfer on plain tube and tube bundles are available.
However, most of the correlations do not perform widely in
designing industrial shell and tube heat exchangers with
shell side boiling heat transfer because of the complex
nature of boiling over different boiling mechanisms.
The researchers are also focusing on the enhanced sur-
faces to improve heat transfer coefficient of the tubes for
better heat transfer mechanism. The different structures are
available in engineering data book III of Wolverine tube,
Inc. Lea [1] patented the first externally finned tube for
condensation. By then Jones [2] described the test of an
industrial sized tube bundle evaporator for evaporating
R-11. The tube may be of plain metallic material like steel
or it may be having the enhanced surface. The enhanced
surfaces can be categorized into two types such as coated
surfaces and structured surfaces. The structured surfaces
are mainly of integral fin, re-entrant grooves, re-entrant
cavity machined porous surface. The porous coated sur-
faces may be created by any of the techniques like flame
spraying techniques, metallic deposition and sintering
process. Bukin et al. [3] performed small tube bundles test
taking all these kinds of techniques. The sintered porous
layers performed among all these coatings and also better
than low finned tubes. Many review literatures were pre-
sented by in the past by Thome [4], Webb [5], and Collier
and Thome [6], Browne and Bansal [7] and Ribatski and
Thome [8] concerning boiling on tube bundles. There are
some review in literatures concerned to different aspects of
boiling [9–13]. There is also extensive research going on
nanofluids to enhance heat transfer coefficient by improv-
ing the thermo-physical properties, particularly thermal
conductivity and thermal diffusivity.
The latest review regarding boiling on tube bundle pre-
sented in the Ribatski and Thome [8] include the studies in
which the vapour quality and void fraction measured keeping
the view to study the two phase flow dynamics. The research
works not included in [8] are taken into discussion in the
present article because of having insight to the effect of the
many influential factors like tube spacing, bundle geometry,
surface characteristics, pressure and mass flow.
The research in the area of boiling on tube and tube
bundle is growing at a faster rate. This can be evident from
the huge number of publications available in the literature.
It is therefore required to have a hand on information in the
recent developments in the area of boiling over tube and
tube bundle. This will help in the design and development
of commercial efficient and compact heat transfer equip-
ments. Following is a review in the area of boiling on tube
and tube bundles took place during last 5 years. Boiling
heat transfer on vertical tube & tube bundle and Two-phase
industrial shell and tube heat exchangers are also reviewed.
Fig. 1 Tube bundle geometries
(after Kernn [77])
Heat Mass Transfer
123
Page 3
2 Complex phenomena of boiling
The phenomena of boiling is complex in nature because of
large number of influential parameters like heater surface
characteristics, heater size, shape, material, diameter and
orientation, degree of surface wetting or surface tension,
sub-cooling, inclusion of surfactants, thermodynamic and
transport properties of the fluid affecting the heat transfer
process and bubble dynamics of pool and flow boiling.
Effects of these factors on boiling heat transfer are sig-
nificant individually or in combined manner and were
studied by different researchers extensively.
In pool boiling the vital mechanism involves formation
of vapor bubbles and its departure. Although there is no
bulk fluid movement natural convection and micro-con-
vection play significant role in pool boiling heat transfer. In
case of tube bundles the tube surface and arrangement are
the main factors and other influencing factors are bubble
formation, rising of bubbles and liquid movement. In
convective boiling the mechanism is different form pool
boiling. The influential factors in convective boiling are
the: (1) The convective effects due to the fluid velocity and
the rising bubbles (2) the effects of static head which
causes increased saturation temperature at the lower tubes
than at the upper tubes. The bundle pitch to diameter ratio
and arrangement also affects the heat transfer rate. The heat
transfer rates and mechanism vary as the flow pattern
across the tube bundle changes. In case of tube side boiling
the two phase flow is confined to the tube and is compar-
atively simpler than the shell side boiling.
The bottleneck concerning the thermal design is the
prediction of heat transfer on heated tube in a horizontal
bundle with the complicated two-phase flow. The com-
plexity lies at different scales. On the tube periphery at
different angular positions the heat transfer process varies
differently. The two phase flow inside the shell influences
the heat transfer mechanism to a great extent. The mech-
anism between the conjoint tubes is also affected by the
two-phase flow. Therefore the correlations proposed should
be the accountable for all these phenomena.
3 Boiling on a single tube or cylindrical surface
Plenty of literatures are available on the study of boiling
liquids over single tube. Some important correlations pro-
posed for single tube as summarized by Browne and Bansal
[7] are presented in the Table 1. As the phenomena of
boiling on tube involve many affecting variables, the dif-
ferent mathematical models proposed have their own
limitation.
List of works carried out in the recent years specifically
on single plain and enhanced tube is given in Table 2. The
inspirations behind the single tube studies are generally to
analyze mechanism in the boiling or sometime to study the
performance of an enhanced tube over the plain tube.
4 Studies on boiling heat transfer on across tube
bundles
The two phase flow dynamics or bubble dynamics over
tube bundles differs greatly from that of single tube. The
bubbles or vapour released from the single tube is not
affected by the other tubes. This phenomenon is quite
significant in tube bundles and affects the local HTC and
overall HTC.
In the recent years many works have been carried out on
pool and convective boiling on tube bundles. Browne and
Bansal [7] reviewed over 80 papers in the area of boiling
on the outside of horizontal tubes stating the advantages
and disadvantages of different approaches of researchers.
Table 1 Correlations for
boiling on single tube [7]No. Authors Correlations
1 Rosenhow [78] q ¼ qnb þ qc h ¼ hnb þ hc
2 Cornwell and Leong [42]Nu ¼ VomNu2u þ Re0:7
g 1� Vom Nu2u
Nu
� �Pr�0:7 D
B
� �0:7
Nuf ¼ CfRemf Prn; Ref ¼ GD
1�xð Þl , Vo ¼ 1
1�ε
3 Singh et al. [79]hcb ¼ 1þ bu
usu
� �0:67
hnb þ hcv,
hcv ¼ k
D Prð Þo:3 0:35þ 0:56 DGlf
� �0:56� �
hnb ¼ 1
Csf
k
Db
GbDb
l
� �Pr�0:7
4 Kutatelazde et al. [60]h
hcv
¼ 1þ hnb
hc
� �nh i1=n
5 Palen et al. [80]h ¼ 0:00417P0:69
c q0:7Fp; Fp ¼ 1:8 pp
c
� �0:17
6 Cooper et al. [32] h ¼ 55:1q 0:67Pr 0:12 �log Prð Þ�0:55M �0:55
Heat Mass Transfer
123
Page 4
Ta
ble
2R
ecen
tex
per
imen
tal
stu
die
so
nb
oil
ing
on
sin
gle
tub
e
Au
tho
rsT
yp
eo
fh
eati
ng
tub
eL
iqu
idR
emar
ks
Rib
atsk
ian
dS
aiz-
jab
ard
o
[33
]
Co
pp
er,
bra
ssan
dst
ain
less
stee
ltu
bes
wit
hsa
nd
pap
erfi
nis
hed
surf
ace
Ra
=0
.02
–3
.3lm
,D
=1
9
R-1
1,
R-1
2,
R-2
2,
R-1
23
a,
R-1
34
h qn w¼
f wP
r�
log
Pr
½��
0:8
Ra
0:2
M0:5
wh
ere
n¼
0:9�
0:3
P0:2
ran
dth
ev
alu
eso
ff w
=1
00
(co
pp
er),
11
0(b
rass
)an
d8
5(s
tain
less
stee
l)
Das
etal
.[8
1]
Sm
oo
thst
ain
less
stee
l,D
=8
,6
.5an
d
4
R-1
23
Wat
erT
he
dia
met
ero
fth
eh
eate
ris
var
ied
bet
wee
nw
ire
and
ind
ust
rial
tub
e.
Inth
era
ng
eth
eh
eat
tran
sfer
rate
incr
ease
sw
ith
dia
met
er
Yan
get
al.
[82
]C
op
per
Tu
rbo
-Etu
bes
D=
19
.05
,H
elix
ang
les
=9
0,
45
and
13
5
e/D
=0
.03
5an
d0
.04
1
HF
C1
34
aE
ffec
to
fh
elix
ang
leo
fT
urb
o-E
tub
eis
stu
die
d
h¼
CK
f
Db
hi
q=
Að
ÞDb
KfT
sat
hi i
qg qfhi j P
k rc �
l
wh
ere
i,j,
kan
dl
are
exp
on
ents
dep
end
ing
on
hea
tfl
ux
and
tub
ety
pe.
c=
cav
ity
dia
met
eran
de
=b
ase
op
enle
ng
tho
fca
vit
y
Sai
zJab
ard
oet
al.
[83]
Co
pp
er,
bra
ssan
dst
ain
less
stee
ltu
bes
Ra
=0
.02
–1
0.5
lm
R-1
23
,R
-13
4a
Ver
yro
ug
hsu
rfac
esp
erfo
rmb
ette
rat
low
hea
tfl
ux
esb
ut
dev
iate
sat
hig
her
hea
tfl
ux
es
Sat
eesh
etal
.[8
4]
Sai
nle
ssst
eel
tub
esD
=2
1,
33
Wat
erH
TC
and
tem
per
atu
rev
aria
tio
nw
asst
ud
ied
wit
hin
clin
atio
nan
gle
fro
mv
erti
cal
toh
ori
zon
tal
Das
and
Kis
ho
re[8
5]
Po
rou
sco
pp
er
Co
ated
tub
e,
D=
31
.85
,co
atin
g
thic
kn
ess
=2
9–
11
8lm
Dis
till
edw
ater
,m
eth
ano
lan
d
iso
-pro
pan
ol
h¼
C1q
mP
rn
wh
ere
the
C1,m
and
nar
eco
nst
ants
dep
end
ing
on
coat
ing
thic
kn
ess
and
typ
eo
fli
qu
id
Sar
ma
etal
.[8
6]
Co
pp
ertu
be
D=
12
.7F
ora
ne
ql�
llL
fg
¼3:3
6�
10�
5l� d thi 1:
18
Pr½��
0:5
8i8
Psa
tl�
ll
Lfg
�� 0:
406
wh
ere
l*=
char
acte
rist
icle
ng
th
Cie
slin
ski
and
Kac
mar
czy
k
[87
]
Sm
oo
thC
op
per
tub
ean
dS
tain
less
Ste
eltu
be,
D=
10
Wat
er-A
l 2O
3
Wat
er-C
u
Fo
ra
par
ticu
lar
sup
erh
eat
the
HT
Cin
ver
sely
var
ies
wit
hco
nce
ntr
atio
n.
Sta
inle
ss
stee
lp
erfo
rmed
bet
ter
than
cop
per
tub
es
Sar
afar
azan
d
Pey
gh
amb
arza
deh
[88]
Sta
inle
ssst
eel
tub
e,B
rass
tub
e,co
pp
er
tub
eo
fsa
me
dim
ensi
on
and
rou
gh
nes
s,D
=2
0
Al 2
O3-w
ater
and
TiO
2–
wat
erE
nh
ance
dra
teo
fh
eat
tran
sfer
isse
eno
nS
San
db
rass
tub
eb
ut
no
tin
cop
per
surf
ace
Tri
sak
sri
and
Wo
ng
wis
es
[88
]
Co
pp
ertu
be
D=
28
.5
TiO
2-R
14
1a
Hea
ttr
ansf
erra
ted
eter
iora
ted
Go
rgy
and
Eck
les
[89]
Tu
rbo
-BII
HP
Tu
rbo
-BII
LP
An
dsm
oo
thtu
be,
D=
19
.05
,
L=
1m
R-1
23
R-1
34
a
h¼
C1q
mw
her
efo
rT
BH
P,
C1
=2
,97
0.2
8&
m=
0.5
49
and
for
TB
LP
,
C1
=3
,82
9.8
4&
m=
0.4
55
inth
en
ucl
eate
bo
ilin
gre
gim
e
Lee
etal
.[9
0]
Pla
inal
um
iniu
mS
urf
ace
Al 2
O3-n
ano
po
rou
ssu
rfac
e(1
.5lm
)
Sat
ura
ted
Wat
erL
ow
wal
lsu
per
hea
tis
req
uir
edfo
rN
ano
-po
rou
ssu
rfac
eto
on
set
nu
clea
teb
oil
ing
Heat Mass Transfer
123
Page 5
Ribatski and Thome [8] have critically reviewed the
research works involving mainly study of the local flow
conditions, vapor qualities and their effects on the heat
transfer behavior. These investigations also unveil about
different aspects of induced convection effects and the
onset of dry out or mist flow highlighted in Table 3.
The above review work carried out by Ribatski and
Thome [8] have not included some of the important
parameters responsible in design and development of
compact and efficient heat transfer equipment. Therefore,
following review comprise works not embraced in [8]
along with the recent works on plain tube bundles.
An overview of the recent studies on boiling outside
tube bundles is represented below and earlier works not
included in [8] are also presented here.
4.1 Plain tube bundles
Although these are so called as plain tubes there will be
some surface roughness or cavities which will be the per-
forming as active nucleation sites. The general manufac-
turing procedure such as extrusion or machining leaves
some cavities on the tubes produced. These are called as
plain tubes because no special surface texture preparation
techniques are applied.
Burnshide and shire [14] studied the boiling of R113 over
a 17 row, 5 column of plain machined tube bundle under
atmospheric condition with uniform heat flux of 10–65 kW/
m2 and maximum Reynolds numbers Remax, between 7,800
and 27 000. Their experimental set up was a thin slice of
kettle reboiler closed by plates to constrain the flow. They
observed that at lower heat fluxes the heat transfer rate
approaches towards higher value with vapor quality. Flow
quality is also seen to be low in nucleate boiling regime
controlled region. The Reynolds number considered here is
related to liquid phase. The author compared the values with
the predicted values of isolated tube by Mostinski [15]
correlation and the convective heat transfer coefficient of
only liquid phase at different Reynolds number, heat flux
range and vapor quality range. The HTC for vapor quality 0
is observed to be greater than the predicted nucleate pool
boiling coefficient and that of the liquid phase convective
HTC at all flow rates. The HTC again decreases with
increase in heat flux over a range of quality and flow rates.
This may be due to suppression of nucleate boiling because
of two phase flow and coalescence. The quality range was
from 0 to 0.3. The vapor phase Reynolds number may have
magnified effect at high quality values.
Many researchers (Hsieh and Hsu [16], Webb and Oais
[17], Chang and You [18], Chen and Webb [19], Memory
et al. [20, 21]) investigated the enhanced heat transfer effect
taking different structured tubes and commercial enhanced
tubes. These enhanced surfaces prove to be better tech-
niques but the cost associated is too high. The structured
tubes can be applied to pure liquids but not for impure water
or liquids with some particles as in case of desalination
devices and solar power absorption type chillers. The pores,
channels and micro-tunnels will be filled with deposition of
impurities. Therefore a substitution for the above situation
is enhancement of heat transfer on plain tubes by making
the bundle compact creating more restricted spaces. The
restricted spaces assist to create a thin superheated liquid
film between the heater surface and the bubbles enhancing
evaporation of the liquid layer. With low heat fluxes the
temperature of the liquid increases rapidly. Fujita et al. [22]
studied the effect of narrow space between two parallel
rectangular surfaces on nucleate boiling heat transfer and
critical heat flux. In their experimental results the heat
transfer rate increases with decreasing the gap size up to a
certain point but beyond that the heat transfer rate decreases
with gap size. This is due to early vapor blanketing on the
surface and obstruction to the flow.
Ishibashi and Liu [23] investigated about first (1993) the
effect of restricted spaces in tube bundles with roll worked
tubes under sub-atmospheric conditions. Liu and Ishibashi
[24] studied boiling of pure water on smooth tubes with
varying tube spacing, tube position and salt concentration.
The authors varied the tube spacing from 4 to 0.1 mm. The
boiling heat transfer rate found to be increasing with
reducing the gap size (Fig. 2). The superheat was 1 K for
heat flux less than 100 kW/m2, which represents that the
fully developed nucleate boiling performance was
observed in the natural convection region.
The heat transfer coefficient can be magnified by 10
times compared with general bundle configurations under
the same heat flux. At high heat fluxes the heat transfer
coefficient is not magnified and approaches to single tube
performance because of reduced liquid movement in the
confined spaces (Fig. 3).
In studies on general tube bundles (not compact or pitch
to diameter ratio varying around 1.2–2) there is a signifi-
cant change of boiling heat transfer performance along the
height. The heat transfer coefficient increases with height:
the upper tubes perform better than the lower ones. This
effect is due to two main contributing factors.
1. The thermal expansion of the liquid leads to density
difference. Due to this density difference the buoyancy
effect will appear creating convection of the liquid.
This convective effect increases along the height from
down to top increasing the heat transfer rate.
2. The large amounts of vapor in the flow create a thin
superheated liquid film on the outside of many tubes
which increases the heat transfer coefficient similar to
film condensation [4].
Heat Mass Transfer
123
Page 6
Ta
ble
3S
tud
ies
rev
iew
edb
yR
ibat
ski
and
Th
om
e[8
]
Au
tho
rF
low
dir
ecti
on
Tu
be
bu
nd
lech
arac
teri
stic
s
(co
lum
ns
9ro
ws)
:s/
D,
D(m
m)
Flu
ids,
tem
per
atu
re(�
C),
and
/
or
pre
ssu
re(k
Pa)
Hea
tin
gm
eth
od
x,G
(kg
/m2s)
,an
dq
(kW
/m2)
Do
wla
tiet
al.
[91]
:5
92
0p
lain
tub
esin
asq
uar
ein
lin
e
arra
y:
s/D
=1
.3,
D=
12
.7
R1
13
(48
-6
1�C
)H
ot
oil
,p
roce
du
reto
calc
ula
te
hw
asn
ot
exp
lain
ed
0\
x\
0.5
,5
0\
G\
79
0,
0\
q\
8
Jen
sen
etal
.[9
2]
:5
91
5tu
bes
ina
stag
ger
edar
ran
gem
ent,
hig
h-fl
ux
and
Tu
rbo
-B:
s/D
=1
.17
and
1.5
0,
D=
19
.05
R1
13
(69
.7,
11
4�C
)E
lect
rica
lly
hea
ted
0\
x\
0.8
0,
50\
G\
50
0,
5\
q\
80
Web
ban
dC
hie
n[9
3]
15
pla
intu
bes
dis
trib
ute
din
six
row
sfo
ra
stag
ger
edar
ran
gem
ent
of
equ
ilat
eral
tria
ng
le:
s/D
=1
.42
,D
=1
6.8
R1
13
,R
12
3(1
8.9
,3
7.8
�CE
lect
rica
lly
hea
ted
0.1
\x\
0.9
0,
0.2
8\
G\
40
Gu
pte
and
Web
b[9
4]
:1
5tu
bes
dis
trib
ute
din
six
row
sfo
ra
stag
ger
edar
ran
gem
ent
of
equ
ilat
eral
tria
ng
le,
inte
gra
lfi
ntu
bes
1,0
24
fin
s/m
:
s/D
=1
.25
,D
=1
8.9
R1
1(4
.4,
26
.7�C
)E
lect
rica
lly
hea
ted
0\
x\
0.9
0,
7\
G\
18
,
15\
q\
45
Gu
pte
and
Web
b[9
5,
96]
:1
5tu
bes
dis
trib
ute
din
six
row
s,st
agg
ered
arra
ng
emen
to
feq
uil
ater
altr
ian
gle
,
Tu
rbo
-Ban
dG
ewa-
SE
:s/
D=
1.2
5,
D=
19
R1
1,
R1
23
,R
13
4a
(4.4
,
26
.7�C
)
Ele
ctri
call
yh
eate
d0\
x\
0.9
0,
10\
G\
30
,
15\
q\
45
Fu
jita
etal
.[9
7]
:5
91
0in
squ
are
inli
ne
arra
ng
emen
t;5
0
tub
esd
istr
ibu
ted
in1
1ro
ws
for
a
stag
ger
edar
ran
gem
ent
of
equ
ilat
eral
tria
ng
le,
pla
intu
bes
:s/
D=
1.3
and
1.5
,
D=
14
R1
13
,4
7.6
�CE
lect
rica
lly
hea
ted
G=
33
,1
06
,3
32
Th
on
on
etal
.[9
8]
:4
5tu
bes
dis
trib
ute
din
18
row
sac
cord
ing
toan
inv
erse
tria
ng
ula
rp
itch
Gew
a-K
tub
es(1
,18
1fi
ns/
m):
s/D
=1
.33
,
D=
19
Pro
pan
e(6
00
,7
00
,8
00
kP
a)
and
n-p
enta
ne
(35
,5
0,
13
0k
Pa)
Mea
nro
wa
calc
ula
ted
by
usi
ng
atu
be
sid
eco
rrel
atio
n
0.2
0\
x\
1.0
0, *
12\
G\
60
,1
0\
q\
50
Ro
ser
etal
.[9
9]
:5
91
8p
lain
tub
es(R
a=
0.4
lm
)in
a
60�
(in
ver
seeq
uil
ater
altr
ian
gle
)la
yo
ut:
s/D
=1
.33
,D
=1
9
N-p
enta
ne
(20
,3
0,
50
kP
a)H
ot
wat
er/g
lyco
l?
Wil
son
plo
tm
eth
od
app
roac
hin
ord
erto
ob
tain
mea
nh
x\
0.6
0,
14\
G\
44
,
10\
q\
60
Tat
ara
and
Pay
var
[10
0]
:3
1tu
bes
dis
trib
ute
din
nin
ero
ws,
tria
ng
ula
rp
itch
,lo
w-fi
nn
edtu
bes
1,0
24
fin
s/m
:s/D
=1
.17
,D
=1
9
R1
23
?m
iner
alo
il
R1
34
a?
po
lyo
lest
ero
ilw
ith
xB
0.1
5an
dR
22
-wit
ho
ut
oil
Ele
ctri
call
yh
eate
d0
.15\
x\
1.0
0,
3.6
\G
\1
6.4
,
8.2
\q\
33
Tat
ara
and
Pay
var
[10
1]
:3
1tu
bes
dis
trib
ute
din
nin
ero
ws,
tria
ng
ula
rp
itch
,st
ruct
ure
den
han
ced
tub
efo
rlo
w-p
ress
ure
-ref
rig
eran
tssu
ch
asR
12
3:s
/D=
1.1
7,
D=
19
R1
23
(4.4
�C)
?m
iner
alo
il
ISO
vis
cosi
tyg
rad
e6
8,x
up
to0
.15
Ele
ctri
call
yh
eate
d0
.15\
x\
1.0
0,
7.9
\G
\2
9.9
,
8.2
\q\
33
Tat
ara
and
Pay
var
[10
2]
:3
1tu
bes
dis
trib
ute
din
nin
ero
ws,
tria
ng
ula
rp
itch
,st
ruct
ure
den
han
ced
tub
ed
esig
ned
for
inte
rmed
iate
-pre
ssu
re
refr
iger
ants
such
asR
13
4a:
s/D
=1
.17
,
D=
19
R1
34
a(4
.4�C
)?
po
lyo
lest
er
oil
,v
isco
sity
gra
de
68
,x
up
to0
.12
Ele
ctri
call
yh
eate
d0
.15\
x\
1.0
0,
7.2
\G
\2
9.3
,
8.2
\q\
33
Heat Mass Transfer
123
Page 7
Ta
ble
3co
nti
nu
ed
Au
tho
rF
low
dir
ecti
on
Tu
be
bu
nd
lech
arac
teri
stic
s
(co
lum
ns
9ro
ws)
:s/
D,
D(m
m)
Flu
ids,
tem
per
atu
re(�
C),
and
/
or
pre
ssu
re(k
Pa)
Hea
tin
gm
eth
od
x,G
(kg
/m2s)
,an
dq
(kW
/m2)
Ap
rin
etal
.[1
03]
:4
5tu
bes
dis
trib
ute
din
18
row
sac
cord
ing
toan
inv
erse
tria
ng
ula
rp
itch
,p
lain
tub
es:
s/D
=1
.33
,D
=1
9
N-p
enta
ne
(50
kP
a),
pro
pan
e
(80
0k
Pa)
Ho
tw
ater
-gly
col,
aver
aged
ab
yu
sin
gG
nie
lin
ski
[10
4]
for
the
inte
rnal
a(m
od
ified
Wil
son
plo
tap
pro
ach
)
0\
x\
0.7
0,
G=
15
,
5\
q\
52
Kim
etal
.[1
05]
:D
istr
ibu
ted
infi
ve
row
s,eq
uil
ater
al
tria
ng
lela
yo
ut,
pla
inan
dth
ree
enh
ance
dtu
bes
hav
ing
po
res
wit
h
dif
fere
nt
size
san
dco
nn
ecti
ng
gap
s:s/
D=
1.2
7,
D=
18
.8
R1
34
a(4
.4,
26
.7�C
)E
lect
rica
lly
hea
ted
0.1
\x\
0.9
0,
8\
G\
26
,
10\
q\
40
Th
om
ean
dR
ob
inso
n[1
06]
:2
0T
urb
o-B
IIH
Ptu
bes
dis
trib
ute
din
eig
ht
row
s,st
agg
ered
equ
ilat
eral
tria
ng
lela
yo
ut:
s/D
=1
.17
,D
=1
9
R1
34
a,R
41
0A
,R
50
7A
(4.4
�C)
?sy
nth
etic
oil
,IS
O
vis
cosi
tyg
rad
e3
2,x
up
to
0.1
3
Ho
tw
ater
,m
od
ified
Wil
son
plo
tap
pro
ach
too
bta
inlo
cal
a
0.0
9\
x\
0.7
0,
4.6
\G
\3
9,
4.6
\q\
45
Ro
bin
son
and
Th
om
e[1
07]
:2
0p
lain
tub
esd
istr
ibu
ted
inei
gh
tro
ws,
stag
ger
edeq
uil
ater
altr
ian
gle
lay
ou
t:s/
D=
1.1
7,
D=
19
R1
34
a(4
.45
�C)
Ho
tw
ater
,m
od
ified
Wil
son
plo
tap
pro
ach
too
bta
inlo
cal
a
0.1
0\
x\
0.8
7,
5\
G\
41
,
2\
q\
35
Ro
bin
son
and
Th
om
e[5
2]
:2
0lo
w-fi
nn
ed(1
,02
4fi
ns/
m)
tub
es
dis
trib
ute
din
eig
ht
row
s,st
agg
ered
equ
ilat
eral
tria
ng
lela
yo
ut:
s/D
=1
.17
,
D=
19
R1
34
a(4
.4�C
),R
50
7A
(4.7
�C)
Ho
tw
ater
,m
od
ified
Wil
son
plo
tap
pro
ach
too
bta
inlo
cal
a
0.0
8\
x\
0.8
2,
3\
G\
29
,
2\
q\
50
Ro
bin
son
and
Th
om
e[5
3]
:2
0T
urb
o-B
IIH
Ptu
bes
dis
trib
ute
din
eig
ht
row
s,st
agg
ered
equ
ilat
eral
tria
ng
lela
yo
ut:
s/D
=1
.17
,D
=1
9
R1
34
a(4
.4�C
),R
41
0A
(4.5
6�C
),R
50
7A
(4.7
�C)
Ho
tw
ater
,m
od
ified
Wil
son
plo
tap
pro
ach
too
bta
inlo
cal
a
0.0
8\
x\
0.7
8,
4\
G\
38
,
8\
q\
64
Gu
pta
[54
]:
39
5st
ain
less
stee
lp
lain
tub
esin
squ
are
inli
ne
arra
y:
s/D
=1
.5,
D=
19
mm
Wat
er(1
00
�C)
Ele
ctri
call
yh
eate
d0\
G\
10
,1
0\
q\
40
Zh
eng
etal
.[1
08]
:1
5ca
rbo
nst
eel
tub
esd
istr
ibu
ted
inth
ree
colu
mn
sac
cord
ing
toa
tria
ng
ula
rp
itch
;
on
lyo
ne
tub
ein
the
bo
tto
mro
ww
as
hea
ted
:s/
D=
1.2
5,
D=
19
Am
mo
nia
?p
oly
alk
yle
ne
gly
col
oil
(PA
G)
(-2
3.3
,-
9.4
,7
.2�C
),x
val
ues
of
0,
1,
5,
and
10
Ho
tw
ater
/gly
col,
mo
difi
ed
Wil
son
plo
tm
eth
od
app
roac
h
too
bta
inav
erag
eh
eat
tran
sfer
coef
fici
ent
xo
f0
,0
.2,
0.4
;1
0\
q\
60
Heat Mass Transfer
123
Page 8
However, this effect is seen diminished in the study of
Liu and Ishibashi [24]. The reason behind this effect may
be that, the restricted spaces do not allow the convective
effect to reach the upper tubes properly. For three mea-
sured tubes a little increase in heat transfer is visible but for
large number of tubes the bundle effect may disappear.
The salt–water concentration effect is not so evident
below a value of 10 %. Beyond the salt concentration of
10 % the heat transfer rate decreases with increase in
concentration but is still higher than the single tube per-
formance. Thus due to restricted spaces the temperature of
the liquid rises at low heat fluxes forming a stagnant
superheated thermal boundary layer and leads to incipient
of boiling. If the heat flux increases or the restricted space
decreases the bubble grows and coalesces to form a huge
bubble decreasing the heat transfer rate. Hence there is an
optimum value of tube spacing for which the boiling heat
transfer is greatest.
Liu and Qiu [25, 26] studied the boiling of pure water
and water-salt mixture on smooth tubes and enhanced
tubes, varying the tube gaps under atmospheric conditions.
The enhancement technique adapted is low manufacturing
cost as it is simply rolled tubes by steel rollers with com-
pact needles. The roll worked surface showed better
enhancement at moderate heat fluxes and even better than
High flux tubes at heat fluxes higher than 105 W/m2 k. The
boiling heat transfer rate increase with decreasing tube
spacing. At the lowest value of tube spacing, 0.5 mm the
heat transfer rate was maximum but with increasing heat
flux the performance approaches the single tube. This is
due to restricted movement of the liquid due to vapor
blanketing on the restricted spaces. The authors [27, 28]
observed an optimum spacing of 1 mm for staggered tube
bundles under sub-atmospheric conditions but for atmo-
spheric conditions it is 0.3 mm. They observed the heat
transfer coefficient increases with decreasing tube spacing
up to optimum value. The HTC also increases with pres-
sure for a certain value of tube spacing (Fig. 4). The
authors calculated the bubble departure diameter from the
relation (1) proposed by Fritz [25] and observed that the
optimum tube spacing is of the order of bubble departure
diameter.
Dd ¼ 0:119/ð2r=gðql� qvÞÞ1=2 ð1Þ
It is also observed that for tube spacing of 2 mm
(outside diameter = 18 mm) or pitch to diameter ratio of
1.11 the top tube is performing better than the lower two
tubes but the trend is not followed as in bundle effect. At
this point the lower tube is performing better than the
middle tube which is against the trend in bundle effect.
Other studies (Hahne et al. [29], Ribatski et al. [30]) reveal
that in non-compact tube bundle at a pitch to diameter ratio
of 1.2 the bundle effect is observed but the effect is
deviating as the pitch approaches 1.11. Thus it can be
concluded that decreasing the tube spacing or pitch to
diameter ratio beyond a certain value leads to deviation
from bundle effect. Thus for compact tube bundles (tube
spacing 0.1 mm or pitch to diameter ratio 1.005) the bundle
effect is very much diminished.
Liu and Liao [27] investigated the pool boiling heat
transfer study of a newly compact in-line plain tube bundle
evaporator under reduced pressure conditions using pure
water as the boiling liquid. The heat fluxes in flooded
evaporators in desalination, food processing and refriger-
ation applications are very low to induce the nucleate
boiling process. The authors conducted the experiments
under reduced pressure conditions (20 and 50 kPa) taking a
compact inline bundle, varying the tube spacing or pitch to
diameter ratio (1, 2, 4) and heat flux. The surface tension is
inversely related to the reduced pressure. When the reduced
Fig. 2 Effect of tube spacing on boiling heat transfer (after Liu and
Ishibashi [24])
Fig. 3 Tube position effect on boiling performance (Liu and
Ishibashi [24])
Heat Mass Transfer
123
Page 9
pressure is high the surface tension is low and smaller
cavity can lead to nucleation due to high wetting effect.
The heat transfer enhancement decreases as pressure
decreases. There is an optimum tube spacing (2 mm for
pure water under sub-atmospheric condition) for which the
greatest heat transfer coefficient occurs at a certain pressure
condition. The same authors [28] also investigated for the
optimum tube spacing for compact staggered tube bundles
under sub-atmospheric pressure (Fig. 5).
Ribatski et al. [30] investigated the performance of
individual horizontal tube in a flooded evaporator tube
bank with refrigerant at different pressure conditions. They
proposed a new correlation for the determination of the
heat transfer coefficient, in which the ratio between the
heat transfer coefficients on any tube and the lowest tube in
the array is related taking into account the reduced pressure
and the heat flux for that particular row. The lowest tube
heat transfer coefficient can be calculated using single tube
correlations by Stephan and Abdelsalam [31], Cooper [32],
and Ribatski and Saiz-jabardo [33].
It was observed that the tube spacing has little effect on
HTC. According to Liu and Qiu [25, 26] the tube spacing
effects on HTC become relevant as the tubes become closer
because the bubbles are confined between consecutive
tubes. The effect of tube spacing on HTC contradicts from
literature to literature. According to Muller [34], Wallner
[35] and Gupta et al. [36] the HTC diminishes with inter
tube spacing under partial nucleate boiling conditions but
Hahne et al. [37] observed an opposite trend. Ribatski et al.
[30] proposed correlation (2) relating the heat transfer coef-
ficient with the heat flux, reduced pressure and Nth row.
hN
h1
¼ 1þ 0:345CAP�1r q�1
� exp �0:37P�0:4r ln q= CqP0:7
r
� � � �2n oð2Þ
Cq ¼ 65þ 1200 expð�0:3NÞ ð3Þ
CA ¼ 160� 85:2 expð�0:3NÞ ð4Þ
They verified their correlation with the correlation
proposed by Kumar et al. [38] for water. They found that
only their correlation is not confirming in the partial
nucleate boiling region. This may be because Kumar et al.
[38] data cover the heat flux range from 19 to 45 kW/m2
but not from 0 to 20 kW/m2 heat flux. However they did
not consider the data of other fluid and also fluid properties
are not considered. Hence may not work for fluids of
different properties like water. They also verified their
model with data of Hseih et al. [39].
They concluded that the ratio between the third row and
the lowest tube HTC gradually increases in the range of
partial nucleate boiling. As shown in the Fig. 6 in fully
developed nucleate boiling region the ratio of HTC
approaches one. The variation with respect to tube spacing
is negligible in the fully developed nucleate boiling region.
The heat transfer also increases with pressure. However the
authors have not investigated the effect of the different heat
fluxes in different tubes.
It is essential to know the two phase flow dynamics
while designing the heat exchangers for industrial purpose.
Without proper data regarding two phase flow parameters
like pressure drop and void fraction designing industrial
shell and tube heat exchangers is carried out with a high
factor of safety and higher dimensions. A lot of studies on
two phase flows are there but with air water adiabatic
flows. The investigations concerning the two phase flows in
case of boiling on tube bundles are limited. The earlier
studies on different two-phase flows, void fraction,
Fig. 4 Effect of pressure on boiling heat transfer at a certain tube
spacing (after Liu and Liao [27])
Fig. 5 Effect of tube spacing on heat transfer coefficient at a
particular pressure (after Liu and Liao [28])
Heat Mass Transfer
123
Page 10
pressure drop and flow visualizations are critically
reviewed by Ribatski and Thome [8].
Aprin et al. [40] investigated the flow pattern by measuring
the local void fraction at different heights in the tube bundle.
To analyse the heat transfer mechanism it is necessary to
measure the local properties like liquid and vapor velocity,
mass quality and void fraction. As the fluids used were n-
pentane, propane and isobutane, a data with different fluid
properties are collected. The liquid mass flow going into test
section and mass of vapor at the exit are measured resulting in
vapor quality at different points of test section. The authors
used an optical probe technique to measure the void fraction
between the tubes. The authors also proposed a correlation for
the calculation of the calculation of void fraction.
e ¼x
qG
1:047 xqGþ 1�x
qL
� �þ 0:23
_m
ð5Þ
The same authors [41] conducted another study to
analyse the local heat transfer analysis integrating the two
phase flow dynamics. The authors have critically reviewed
many investigations and identified that different
mechanisms exist in boiling in different regimes and a
single model will not be able to predict the heat transfer
rate in all the regimes. They correlated the flow regimes
with the appropriate heat transfer regimes.
They adopted a novel control volume approach for the
local heat transfer analysis. The vapor quality at different
heights is calculated by the heat balance method. The cor-
responding heat transfer coefficients are also determined and
plotted against the vapor quality. A transition in the value is
seen which represents the change in flow regime as in Fig. 7.
The bubbly flow is the nucleate boiling regime and the dis-
persed flow is the convective flow regime (Fig. 8).
Bundle effect is also seen along the height of the tube
bundle. At a heat flux of 37 kW/m2 the bundle effect is pre-
sented in the Fig. 9. Due to inlet and exit conditions the per-
formances of the lowest and top rows have been diminished.
The lower tubes are under the nucleate boiling regime or the
bubbly flow regime. With the increase in height the heat
transfer rate increases as with convective boiling effect and
annular dispersed flow. The change in the heat transfer rate
associated with the transition in flow regime is also observable
in the Figs. 10 and 11. Also from these experimental data it is
Fig. 6 Ratio of HTC of third row to first row with heat flux variation
(after Ribatski et al. [30])
Fig. 7 Mean bubble diameter versus mean void fraction (Aprin et al. [40])
Fig. 8 Proposed flow on tube bundles bubbly flow and annular
dispersed flow [40]
Fig. 9 Heat transfer coefficient with the rows (after Aprin et al. [40])
Heat Mass Transfer
123
Page 11
observable that there is significant effect of heat flux leading to
increase in the heat transfer rate and transition points of the
change in regime. The effect of mass flux is quite different
from the heat flux. With increase in mass flux the transition
point is shifting to left or occurring at a lower vapor quality but
the heat transfer coefficient does not vary.
The authors have compared the experimental results with
the different superposition models and found that the experi-
mental data are dispersed widely around the predicted values
from the models. The similar observation is also seen when the
experimental measurements are compared with the asymp-
totic models. Most of the classical correlations, established
considering the heat transfer laws in tubes should not be
implemented while designing horizontal shell side vapour
generating units as the heat transfer laws for outside tube
bundles is altogether a different. The authors applied different
correlations for the nucleate boiling regime and the convective
flow regime. The relation proposed by Cooper [32] predicts
their data very closely. The authors suggested using this for
nucleate boiling regime. They proposed a relation (6) based on
Nusselt number with vapor phase Reynolds number and
reduced pressure for convective boiling regime. The Reynolds
number is based on real gas velocity instead of liquid velocity.
A similar kind of correlation proposed by Cornwell et al. [42]
deviates largely form the experimental data may be due to
liquid Reynolds number. The void fraction is calculated from
the Eq. (5) proposed in their earlier work.
Nu ¼ 387P0:17Re0:37G Pr0:33
G ð6Þ
Nu ¼ h � Dext
kG
ð7Þ
VG ¼_mx
eqG
ð8Þ
The Reynolds and Prandlt number accounts for some physical
properties of the fluid undergoing boiling but here three
hydrocarbons studied are considered for the correlation only.
For intermittent regime the authors suggested to apply
correlation for both the regimes and choose which one is
higher. The transition point mass flow rate are defined by the
authors are 0.15 and 0.35 m/s to distinguish between different
flow regimes and use the respective correlations are to be applied.
Shah [43, 44] proposed various correlations not only for tube
bundles but for different boiling process. Shah [43] proposed
for sub-cooled boiling on a single tube and then modified it
considering the database for cross flow on horizontal inline tube
bundle. The database included water and halocarbon refriger-
ants. Shah [44] proposed a dimensionless correlation for satu-
rated boiling with flow across tube bundles covering a wide
range of parameters for seven fluids with a total of 690 data
points. The author identified three heat transfer regimes and
proposed separate heat transfer equations for each regime. The
author differentiated the regimes by the quantity called the
Boiling Intensity Parameter defined as follows (9).
YIB ¼ FpbBoFr0:3 ð9Þ
where Fpb ¼ hpb;actual=hcooper: The hcooper is the pool boil-
ing heat transfer coefficient calculated by the Cooper
equation. The hpb;actual is same as hcooper unless pool boiling
data is available for the tubes.
Heat transfer regimes according to boiling intensity parameter
Regime Range of
parameter
Mod of heat
transfer
Dominating
effect
Intense
boiling
regime
YIB [ 0.0008 Nucleate
boiling
Heat flux
Convective
boiling
regime
0.00021 B YIB B
0.0008
Nucleate
boiling and
convection
Heat flux and
mass
velocity
Convective
regime
0.00021 B YIB Convection
Process
Mass velocity
and vapor
qualityFig. 10 Heat transfer coefficient with vapor quality at different heat
fluxes (after Aprin et al. [41])
Fig. 11 Heat transfer coefficient with vapor quality at different liquid
mass fluxes (after Aprin et al. [41])
Heat Mass Transfer
123
Page 12
The author employed dimensionless parameters like
Boiling number Bo (ratio between the heat flux and mass
flux), Froude Number and parameter Z used by the same
author in the film condensation correlation representing the
convective effect. The Froude number is applied to repre-
sent the mass flux effect. The boundaries among the
regimes are established by plotting the Bo Vs. Fr. The heat
transfer correlations are as follows.
Regime I: Intense boiling regime hcb ¼ Fpbhcooper ð10Þ
Regime II: Convective boiling regime u ¼ uo;u ¼ hcb=hf
ð11Þ
Regime III: Convective regime u ¼ 2:3=Z0:8Fr0:22 ð12Þ
Z ¼ 1� x
x
�0:8
Pr0:4 ð13Þ
The parameter Z is defined as in Eq. (13) by Shah [45].
The predicted values are compared with the data points
with a mean deviation of 15.2 %. As the correlation
predicts closely for halocarbon refrigerants, pentane and
also water (properties are very much different from
refrigerants and organics), its applicability for wide range
of liquids is expected with confidence but requires a
validation with experimental data. These correlations are
still to be revised to perform better in regime 2 and 3. Not
much deviation is found in accordance with tube pitch and
bundle orientation. This effect may be significant for
compact tube bundle as no bundle geometry parameters are
involved in the correlation.
Rooyen et al. [46] recently studied different aspects of
the boiling over a tube bundle of refrigerant R-134A and
R-236FA. By using dummy tubes in the bundle, high speed
camera and a laser source/photodiode the flow patterns are
studied. The second row from the top of the bundle is
equipped with transparent middle portion and having a
mirror to capture the image by a camera at the end of the
tube. The two tubes of third row from the top are instru-
mented with laser source and photodiode. The PDF method
applied to the lighter portion along the major axis of the
elliptical image is studied to visualize the flow. When
liquid is present the light passes through it and a bright
image is formed. When vapor is there then a dark image is
formed. The PDF was plotted as image in abscissa and
ordinate time step. The author did not proposed any flow
pattern map as there is a gradual change in patterns. The
authors have rightly stated that the superposing the in tube
flow patterns onto the tube bundle is not evident as the later
is altogether a different process. However, the different
regimes similar to Aprin et al. [40] are expected.
The dry out of the surface occurs when the vapor quality
approaches 1. The authors proposed a model to predict the
dry out points. The vapor quality at which the dry out
occurs is called Xdry. The value of dry out qualities is 0.92
for R-134A and 0.89 for R-236. This is used in the fol-
lowing expression to calculate the superficial liquid
velocity to predict the dry out point.
GL ¼ Gg
qLxdry
qg 1� xdry
� !
ð14Þ
The criticality in prediction of heat transfer in the tube
bundle boiling is relating to void fraction which is essential
to formulate the pressure drop and the heat transfer.
Rooyen et al. [47] as a second part of their study
investigated and presented a new method of prediction of
heat transfer and pressure drop. The pressure drop in a two
phase flow is comprised of three contributing effects like
gravitation, momentum and friction. The void fraction is
very much necessary for the momentum and gravitational
effect. The void fraction is determined from the vapor
quality calculated by heat energy balance by dividing the
test section to different control volumes as applied by
Aprin et al. [37]. The pressure drop due to gravitation and
momentum are determined by the Eqs. (15) and (16).
DPg ¼X
i
qL 1� εiþ1 þ εi
2
� �þ qg
εiþ1 þ εi
2
� �h ig DZi
ð15Þ
DPm¼G2X
i
1�xð Þ2
qL 1�εð Þþx2
qge
" #
iþ1
� 1�xð Þ2
qL 1�εð Þþx2
qgε
" #
i
" #( )
ð16Þ
The total pressure is calculated by using the Eq. (17).
The f2; is the two phase friction factor calculated by
multiplying a factor with homogenous friction factor of
Zukauskas and Ulinskas [48]. The properties are two phase
mixture properties calculated by using the vapor quality.
The frictional pressure drop can be found out by
subtracting the gravitational and momentum from the
total pressure drop.
DP2; ¼ 4NRf2;G
2
2qð17Þ
The temperature profile of the heating fluid water is
expressed as a function of curvilinear coordinate and is
used to calculate local heat flux ultimately the local heat
transfer coefficient [46]. The overall outside heat transfer
coefficient is modeled in terms of thermal resistance of
external surface, wall and internal surface. The constant Ci
is determined by Wilson Plot method and Rwall is the
cylindrical wall resistance.
Twall � Tsat
q¼ 1
Uo
¼ 1
ho
þ 1
Cihl;i
Do
Di
þ Rwall ð18Þ
Heat Mass Transfer
123
Page 13
The author had validated the data with the recent
empirical model of Shah [44] which predicts 51.4 % of
the data within 30 % deviation. The authors suspect that
the model of Shah [44] is not coupled with the super-
ficial velocities as in Noghrehkar et al. [49] and Ulbrich
and Mewes [50]. The flow boiling heat transfer coeffi-
cient as a power function sum of nucleate pool boiling
and the convective component with exponent less than
one. The convective component is established as Eq.
(19).
Nucb ¼hcb d
KL
¼ 0:0082Re1dBo0:42 ð19Þ
where Bo ¼ q
hfg Gdð20Þ
Gd ¼ ulqL; ð21Þ
Red ¼4qLuld
ll
ð22Þ
The d is the liquid thickness above the tube surface
calculated by assuming a hexangular area around the tube
[46]. The whole work [46, 47] includes prediction
methods for pressure drop, onset of dry out and heat
transfer.
The above studies on plain tubes show that the use of
compact tube bundles can be beneficial because of early
onset of nucleate boiling. It can be effective in low and
moderate heat fluxes. At higher heat fluxes the enhance-
ment decreases as observed i.e. around 100 kW/m2 for
staggered compact tube bundle. So a combination of tube
spacing, heat flux and pressure can give best results.
However, for wide application the combined effect of
compactness and mass flow on pressure drop and heat
transfer in flow boiling needs to be investigated. In inline
tube bundles of normal pitch (Pitch to diameter
ratio = 1.2–2) the bundle effect disappears at an early
stage which is around 40 kW/m2 [27] which can be
increased by introducing compactness.
The correlations developed by Shah [44] considered
wide experimental data but the pitch to diameter ratio
range is 1.17–1.5. The correlation developed does not
contain any term for pitch to diameter ratio. Hence its
application can deviate when applied to compact tube
bundles. The studies [40, 41, 44] also show that the heat
transfer rates are dependent on the different flow
regimes.
5 Enhanced tube bundle
The enhanced tubes include the different types of surface
textures prepared for the improved boiling performance.
Research in the area started many years ago. A detailed
analysis and knowledge can be found out in Thome [4] and
Webb [5]. Some of the recent works are reviewed here.
Kim et al. [51] studied the flow boiling of R-123 and R-
134a on an enhanced surface having pores and connecting
gaps. They have presented a comparison of the perfor-
mance of commercial tubes with bundle factor (Ratio of
bundle HTC to nucleate boiling HTC) from the previous
studies. The connecting gaps help in allowing the two
phase mixture to pass through it enhancing the boiling
process. They varied the flow quality, heat flux and mass
flux with different pore sizes (0.2, 0.23 and 0.27 mm).
They observed not so dominant effect of mass flux and
vapour quality. The convective boiling effect is magnified
at higher saturation temperatures. Better enhancement is
found at pore size 0.27 mm for R-134a and 0.23 for R-123.
They empirically fitted their data with asymptotic relation
with exponent value 1.
hbundle ¼ h1nb þ h1
cb
� 1 ð23Þ
The optimum pore sizes are different for R-134a and
R-123. This can happen because departure bubble sizes
would be different for these liquids as the vapour density is
higher for R-134a than that of R-123. The low density
vapour will have large size bubbles than high vapour
density which may obstruct the connecting gaps.
Thome and Robinson [52–54] studied the boiling per-
formance of refrigerants R-134A, R-507A, R-410A on
smooth, low finned and Turbo-BII HP tube bundle. The
enthalpy profile method is applied to measure the local
boiling heat transfer coefficient using water heating and
measuring the temperature in the axial direction. Thome
and Robinson [55] developed different correlations for
each of the tube. The cross sectional void fraction model
(24) proposed by Fenestra et al. [56] is used for the pro-
posed correlation. The slip ratio is given by the relation Eq.
(25), the capillary number is given by (26) and the Rich-
ardson Number by (27).
e ¼ 1
1þ qG
qLS
1�xð ÞqL
� � ð24Þ
S ¼ 1þ 25:7ðRi CapÞ0:5ðP=DÞ�1 ð25Þ
Cap ¼ lG:lL
rð26Þ
Ri ¼ g a qG � qLð Þ2
G2total
ð27Þ
where g = acceleration due to gravity and ‘a’ is tube
spacing
The authors iteratively modified the guess value through
secant method to reach to the correct value by checking the
difference between the successive values. This method is
also validated by the data of Scharge et al. [57]. The
Heat Mass Transfer
123
Page 14
convective boiling HTC is calculated from an asymptotic
relation assumed as follows by taking the nucleate boiling
HTC from Cooper relation.
hbundle ¼ h2nb þ h2
cb
� 1=2 ð28Þ
These values calculated from the above are used to
empirically fit the below (29) correlation as the
convective heat transfer coefficient to the liquid film.
Then this is used for plain tube bundle HTC calculation.
h ¼ 4:032Re0:236d Pr0:4
L
K
d
�ð29Þ
Similarly for finned tubes again the asymptotic relation
(28) is assumed. The nucleate boiling components are
calculated from the best fitted correlations (30 and 31) for
the data obtained in the research work in [54].
hnb ¼ 93:35q0:55 for R� 507Að Þ ð30Þ
hnb ¼ 90:11q0:436 ðfor R� 134AÞ ð31Þ
Then using the experimental bundle HTC and the nucleate
boiling HTC the convective boiling HTC is determined.
The determined convective boiling values are used to
empirically fit the following liquid convection type
equation for finned tube where the leading constant and
the exponent of the Reynolds number were found from the
data.
Nucb ¼hcb D
KL
¼ 13:92ReL PrL
L=D
�0:0013
ð32Þ
For Turbo-BII HP tubes the following relation (33) has
been proposed.The nucleate boiling curve is used
particularly for the liquid on Turbo-BIIHP tube to get
hnb:The authors [52–54] presented the nucleate boiling
curves for three refrigerants (R-134A, R-410A, R-507A).
hbundle ¼ hnbFpFe ð33Þ
The Fpand Fe represents the bundle boiling factor with
reduced pressure (pr) and void fraction effect respectively
and are calculated as follows.
Fp ¼ 1:41� 2:66pr ð34Þ
Fe ¼ 1:15� 2ð0:4� eÞ2 ð35Þ
Another important type, Coated Porous tubes also serve
as good enhanced surface for boiling purpose. The porous
surface supports for nucleation of bubbles because of
cavities and connecting lanes between the pores. Thus
vapour trapping in the nucleation sites is easier. This
ultimately increases the heat transfer rate and an early
incipience of boiling at low wall superheat.
Hsieh et al. [39] studied the nucleate pool boiling of
R-134A over plasma coated copper tubes with varying heat
flux conditions. Their experimental data includes a variety
of combinations of heated tubes and instrumented tubes
like only lower middle tube, middle column lower three
tubes and lower two rows heated conditions. The heat
transfer rate at a wall superheat is observed to be magnified
for plasma coated tubes than for plain smooth tubes. The
inline bundle with same tube spacing has a higher heat
transfer rate than the staggered one. The bundle factor is
defined as the ratio of the area averaged bundle HTC to the
isolated tube HTC. The configuration factor is defined as
the ratio of the HTC of a bundle configuration to that of the
bundle in which a single tube is heated (middle one of the
lower row) for both smooth and coated tube bundle.
Schafer et al. [58] also investigated the performance of
plasma coated tube bundle using the R-134a refrigerant.
They observed a good enhancement over smooth tube
bundles. The heat transfer coefficient increased with satu-
ration pressure in their observation.
Lakhera et al. [59] investigated the boiling performance
of water at atmospheric pressure on SS 316 flame sprayed
coated tubes regarding the effect of surface roughness
(0.3296–4.7321 lm), the mass flow rate and heat flux. The
flow boiling experimental data best fitted to the Ku-
tateladze [60] asymptotic relation for cross flow boiling
over tube bundle. The HTC values increases with the sur-
face roughness and heat flux at all mass flow rates. The
pool boiling experimental data are also validated with the
correlation proposed by Gorenflo [61]. The similar kind of
observation is seen here in which the HTC increase with
heat flux but after certain value the due to high bubble
coalescence the HTC decreases. With the increase in mass
flow the nucleation on the tube surface is suppressed except
in the wake region the tube and of the flow. The pool
boiling data are governed best by the (36).
hnb ¼ 0:931q0:686 Ra=Dð Þ0:123 ð36Þ
where the Ra is the roughness average of surface, D = tube
diameter. The flow boiling data are then empirically fitted
to following asymptotic relation.
hnb ¼ hl½1þ hnb=hcvð Þ2:258�1=2:258 ð37Þ
The Fig. 12 shows the variation of heat transfer for
different surface roughness of the coated tubes and it is
observable that for coated tubes also the HTC increases
with roughness. The Fig. 13 is the variation of heat transfer
rate (ratio between the local HTC and the average HTC)
around the tube at different angular positions
Lakhera et al. [62] studied the boiling performance of
plain stainless steel (Ra = 0.3296 lm) and copper coated
(Ra = 8.279 lm, porosity \2 %) 8 9 3 tube bundle with
electrically heated. The central column tubes are taken as
measuring tubes. The different pitches to diameter ratios
Heat Mass Transfer
123
Page 15
studied are 1.4, 1.7 and 2.0. The authors compared the
lower tube performance with Gorenflo [55] correlation for
single tube as also observed in other studies. The bundle
effect is also seen in these studies and the coated tubes
perform better than the plain tubes. The coated tube bundle
with lowest tube pitch performed best.
In refrigeration systems as the refrigerant passes through
the compressor the lubricant oil comes mixes with the
refrigerant. This has a detrimental effect when it enters into
again evaporator. Then the performance of the evaporator
tubes degrades. Some of the researches involving different
types of enhanced surfaces are presented here (Table 4).
The operation of flooded evaporators of large tonnage
refrigeration systems, compressor oil inevitably enters to the
refrigerant path. The oil has a detrimental effect on the per-
formance of the shell side boiling in the evaporators. So it is
essential to investigate the effect of oil for proper designing
and parameters required for performance of the flooded
evaporators. The oil concentration in the refrigeration system
is generally 0.5–3 %. Through the studies, it is concluded that
addition of oil to boiling refrigerants cause degradation of heat
transfer coefficient. The oil layer on the heater surface adds to
heat transfer resistance on the boiling surface and it is mag-
nified for enhanced surfaces compared to smooth tubes.
The above studies on the enhanced surfaces reveal that no
general correlation can be established for all types of enhanced
surfaces. The different types of enhanced surfaces are inves-
tigated for application mostly involving refrigerants. Very
limited studies are there regarding application of enhanced
surfaces to boiling of hydrocarbons and other organic liquids.
The enhanced surfaces perform better than the plain tubes but
bundle effect is still there. Hence the proper heat flux range
should be chosen to make the enhanced tube bundle cost
effective and efficient.
6 Boiling on external surface of vertical tube and tube
bundle:
The boiling on outside of vertical tubes has its application
in the nuclear reactors, vertical long tube evaporators and
thermosyphon reboilers. The modeling of boiling two
phase flow on vertical tube bundles is of great importance
to the safety analysis and design of nuclear power plants
and specifically BWR (Boiling Water Nuclear Reactors).
Yao and Chang [63] studied pool boiling inside the confined
annular space taking three different liquids Freon-113, acetone
and water at atmospheric pressure and different annular gaps.
They associated the Bond number to identify the boiling
regimes. Bond number is a non-dimensional quantity repre-
senting the ratio between the gravitational force and the buoyant
force. When the Bond number is less than one, the isolated
bubble regime at low heat fluxes and coalescence bubble pattern
at high fluxes are observed. When the Bond number is slightly
greater than one the nucleate boiling pattern is observed at high
heat fluxes. This is because high gravitational force than
buoyancy does not allow much bubble coalescence.
Kang [64–66] studied the various aspects of boiling on
vertical tube confined in an annular space. It is observed from
different literatures that confinement increases the heat transfer
rate than unrestricted boiling. Hence the author emphasized on
study of heat transfer on vertical tube in confined annulus.
Fig. 12 Variation of heat transfer with surface roughness [59]Fig. 13 HTC/HTC average value around the periphery of the tube at
different angles [59]
Heat Mass Transfer
123
Page 16
Kang [64] investigated that for the same surface rough-
ness, boiling heat transfer get magnified when the tube is
vertical than horizontal. The reason behind the heat transfer
rate increase is the bubble movement and the agitation of
the liquid due to it. The bubbles here move along the surface
and get released at the upper portion of the tube.
Kang [65] observed that as the outer tube length is
increased the HTC decrease at a constant heat flux. For a
certain outer tube length the ratio of the confined tube HTC
to the unrestricted tube HTC approaches one at a higher
heat flux. This deterioration point decreases as the outer
tube length increases. This is due to increase in bubble
coalescence at higher heat fluxes.
Kang [66] also studied the effect of tube inclination on
pool boiling heat transfer by changing the tube position
from horizontal to vertical. The author observed that the
heat transfer rate increases as the angle increases. Two
factors are influential for this process: (1) The effect of
liquid agitation caused by bubble movement which
enhances the heat transfer (2) The coalescence of bubbles
forming large vapor slugs which causes reduction in heat
transfer. Therefore the increase in slope of heat flux and
degree of super heat after 60� is not substantial. The
increase in open bottom is more than in case of closed
bottom of the annulus.
Gupta et al. [67] studied the nucleate pool boiling of
water in vertical tube bundle under sub-atmospheric con-
dition. They used a test vessel of 100 mm diameter, tubes
of diameter 19.00 mm, a pitch to diameter ratio of 1.66,
850 mm long and heating length of 800 mm. They also
correlated their data and expressed the local heat transfer
coefficient along the tube length as (38).
h ¼ 0:0865 q0:66 H=Dð Þ0:51 ð38Þ
The heat transfer coefficient along the height of the tube
bundle increases in case of a vertical tube bundle because
of the turbulence generated due to onset of boiling in the
lower half of the bundle. It was also observed that at lower
rate of heat flux the effect along the height is strong and the
effect vanishes as the heat flux increases. This effect is
similar to the horizontal tube bundles in which the vapor
bubbles formed at the lower portion gets accumulated at
the upper portion and decreases the heat transfer coefficient
over the heater surface at higher heat fluxes.
Many works has been done for investigation of two
phase flow dynamics and boiling taking real data from
nuclear power plants and also on laboratory scaled exper-
imental set up. These studies are available in literatures
concerning nuclear science Technology. This section is
presented in a limited manner in this review article.
7 Studies on shell and tube heat exchanger
The investigations reviewed above substantiated their
studies by experimenting with laboratory scaled models.
Whereas the real application involving industrial scale
models is altogether different. The industrial shell and tube
heat exchanger has many components which differently
govern the boiling process. The baffles, impingement plate,
opening nozzles and many others affect the phenomena of
boiling heat transfer.
Doo et al. [68] investigated about the shell side boiling
in a TEMA E-type shell and tube heat exchanger. The test
unit is a horizontal single tube pass with R-134A on shell
side and low pressure steam condenses on the tube side. Six
segmental baffles are present resulting in up-down and side
to side horizontal flow. The results show a drop in heat
transfer in low mass flow and high vapor quality conditions
which may be due to the transition in the flow pattern. The
temperature, pressure and flow rates are measured at dif-
ferent points in the heating fluid line and the shell side
Table 4 Studies on boiling of
refrigerant and oil mixtures over
tube bundles
Authors Tubes Pool/convective Refrigerants
Marvillet et al. [109] Porous aluminum tubes Nucleate Pool
boiling
R12/R22
Gan et al. [110] Flame sprayed surface tube bundle Nucleate Pool
boiling
R-113/R-11
Webb et al. [111] Plain and enhanced tubes Pool boiling R-11/R-123 oil
Memory et al. [20] Porous or gapped GEWA-K tube Porous
TURBO-B
Nucleate Pool
boiling
R-114/oil
Memory et al. [21] Smooth and Enhanced tubes Pool boiling HCFC-124
with oil
Chyu et al. [112] Enhanced tubes Convective
boiling
Ammonia and
oil
Tatara and Payvar [101,
102]
Turbo-BII tubes Flow boiling R-123 and
R-134a
Kim et al. [113, 114] Tubes with pores and connecting gaps Pool and
convective
R-123 and oil
Heat Mass Transfer
123
Page 17
boiling liquid line. The mean temperature difference was
calculated by using the steam temperature and the satura-
tion temperature of R-134A. The heat supplied is calcu-
lated from condensation rate. The shell side heat transfer
rate is calculated by deducting the inner side and the wall
resistance (from the HTFS simulation software) (Fig. 14).
At low mass flux values the boiling HTC decreases with
decrease in mass flux and is relatively independent of heat
flux and the effect vanishes at high values of mass fluxes.
At higher mass flux the heat transfer rate increases with
heat flux.
The tubes are surrounded by vapors at high vapor
quality and low mass flux resulting in reduction in boiling
HTC. Another hypothetical effect may be high gravita-
tional effect than the inertia force due to which the liquid
remains in the lower tubes and the vapor is accumulated at
the top tubes. The flow maps are presented by same
method as in Grant and Murray [69] by plotting the mass
flux values with the vapor quality (Fig. 15). The flow
transitions can also be marked in the boiling HTC data.
The study also includes the two phase pressure multiplier
from the data of the pressure transducer between the first
and last baffle.
Doo et al. [70] proposed a prediction method using shell
and tube heat exchanger data which suggests the extent of
tube wetting for upper tube bundle and also takes into
account the baffle orientation. General correlations assume
that the two phases are well mixed and heater surface is
fully wetted but in actual industrial heat exchangers the
flow regime changes may not lead to ideal conditions. The
shell side geometry is divided into number of segments.
The different types of flow are identified as the cross flow,
bypass flows and baffle and tube leakage flows are showed
in Fig. 16. The homogenous models assume that the liquid
and vapor are well and uniformly distributed in each path
as discussed in Doo et al. [68]. There is substantial drop in
heat transfer at the low mass flux region below 300 kg/
m2 s. The results also suggested that due to stratified flow
vapor at the top and liquid at the bottom the heat transfer is
affected. The stratified homogenous local shell side heat
transfer rate is proposed in terms of tube side heat transfer
rate, fouling n shell side and wall resistance and shell side
convective factors.
1
h¼ 1
hshellðstrtÞþ rs þ
yd
Kw
Do
Dw
þ 1
htube
þ rl
�Do
Di
ð39Þ
hshellðstrtÞ ¼ 1� eð Þhcb þ ehvapor ð40Þ
The stratified flow shell side coefficient is calculated as
the void fraction weighted sum of the single phase vapor
and the boiling coefficient Eq. (40). When the vapor
velocity on stratified liquid layer is increased it blocks the
open path leading to intermittent flow. This transition is
called critical velocity (ugðcritÞ) suggested by Taitel and
Dukler [71] whose flow map model is found best to be best
suited to the experimental data by Doo et al. [68]. Based on
this the shell side coefficient is expressed as weighted sum
of stratified and homogenous heat transfer coefficient as in
Eq. (41).
hshell ¼ 1�Wð Þhstratified þW hhomo ð41Þ
The value of W is chosen as follows.
If usg\b1ugðcritÞ then W ¼ 0 and
if usg [ b2ugðcritÞ then W ¼ 1:ð42Þ
If b1ugðcritÞ\usg\b2ugðcritÞ then W
¼ ðusg=ug critð Þ � b1Þ=ðb2 � b1Þ ð43Þ
where b1 and b2 are the factors to be considered for upper
and lower critical velocity boundary limits. The model
predicts the value close to the experimental data in a better
manner than the homogenous model with a range of 30 %.
Fig. 14 boiling HTC with mass flux for vertical cross flow and
horizontal cut baffles (after Doo et al. [68])Fig. 15 Flow map for shell and tube heat exchanger with vertical
cross flow and horizontal cut baffles (after Doo et al. [68])
Heat Mass Transfer
123
Page 18
Nasr and Tahmasbi [72] studied on improving the per-
formance of thermosyphon reboilers by using different tube
inserts like twisted tape inserts, Mesh inserts and helical
coil inserts which increases the tube side heat transfer
coefficients.
Al-anizi and Al-otaibi [73] investigated the effect of a
double perforated type impingement plate (DPIP) to erad-
icate the problem of fouling at the inlet nozzle in the first
tube row due to low velocity and vortices production. The
perforation in the plate reduces the turbulence and vortices
due to the DPIP there by increasing the utilization of heater
surface. The simulation of the shell and tube heat
exchanger with modified impingement plate shows
enhancement of heat transfer and restricts the velocity to be
within specified range to improve the life of the tubes.
McNeil et al. [74, 75] modeled the two phase flow inside
the kettle reboiler with one fluid model and two fluid
models. The one dimensional model assumes only one
column of the shell and tube heat exchanger for analysis in
which liquid enters from the bottom and vapor exits from
the top. The one fluid model treats that the both the phases
move in same direction and with different velocities. The
one fluid flow model requires correlations for void fraction,
flow resistance and pressure distribution around them.
The two fluid models use the conservation equations for
mass, momentum and energy for each phase. The tube
bundle is presented as a porous media and the boundary
condition to the free surface of the liquid on shell side
affects greatly to flow patterns. From the numerical simu-
lation the superficial velocity and pressure drop are
analyzed.
Recently Cielinski and Fluk [76] investigated about the
two phase thermosyphon heat exchanger taking water,
methanol and R-141b as boiling liquid. The set up is an
evaporator in combination with a condenser which is
generally used for removal of heat. The investigation
includes the study of effect of liquid head above the top
tube in the evaporator. The heat transfer coefficient
increases with the head up to a certain limit and then
decreases. The effect is more magnified as the heat flux
increases.
Different kinds of researches are going on to improve
the performance of the two phase shell and tube heat
exchanger.
8 Conclusion
The boiling over tube bundle is very much essential from
industrial shell and tube heat exchanger point of view. The
mechanism of boiling over tube bundle is very much
complicated and different than in-tube boiling. The
important outcomes of the present review work are con-
cluded as follows:
1. The performance of the enhanced tube bundle is
definitely higher than the plain tube bundle except in
certain cases where there is chance of mixing of oil in
refrigerant. However, there is no general prediction
method that can be applied to all kinds of surfaces and
liquids.
2. The tube bundle performance is substantially higher
than a single tube and the bundle effect is seen in all
tube bundles. The bundle effect gradually vanishes at
higher heat fluxes due to bubble coalescence or vapour
blanketing on top tubes.
3. The flow regimes identified are bubble flow regime
and convective flow regime but it needs more insight.
4. The compact tube bundles with low tube spacing assist
in incipience of nucleate boiling and high heat transfer
coefficient at a lower heat flux. The application of
compact tube bundles in all industrial two phase heat
exchangers is yet to be investigated in a detailed
Fig. 16 The different flows in a
shell and tube heat exchanger
(after Doo et al. [70])
Heat Mass Transfer
123
Page 19
manner. The compact bundle may limit the velocity of
the liquid which is a dominant factor. The lower limit
of velocity is determined by Fouling and upper limit
corresponds to erosion.
5. The real phenomenon involved in industrial heat
exchangers is quite a different than that in laboratory
scaled tube bundle set up. The prediction methods
developed considering the observations in laboratory
models may not be successful in designing the
industrial shell and tubes heat exchangers very accu-
rately. The studies concerned with actual industrial
heat exchangers are very much limited.
The review suggests that further research works are
needed to get a clear picture of the boiling mechanism in
shell side in case of tube bundles. More research works also
are needed to extend the results of studies on tube bundles
to implement on shell side boiling of industrial shell and
tube heat exchangers for better performance and design.
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