Download - Thermo Syphons
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Heat pipes and Thermosyphons
Cold end
Hot end
Inside the system, there is a fluid (usually termed refrigerant)
Heat pipes and Thermosyphons Heat is transferred as latent
heat of evaporation which
means that the fluid insidethe system is continuouslychanging phase from liquidto gas.
The fluid is evaporating atthe hot end, therebyabsorbing heat from thecomponent.
At the cold end, the fluid iscondensed and the heat isdissipated to a heat sink(usually ambient air).
Hot end
Cold end
Heat pipes and Thermosyphons Heat pipes
Heat pipes
In Heat Pipes, capillary forces in the wick
ensures the liquid return from the hot end to
the cold end.
This means that a Heat Pipe can operate
independent of gravity. The heat pipe was
actually developed for zero gravity (i.e.
space) applications.
Heat pipes
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Heat pipes Heat pipes - Applications
Heat pipes - Applications Thermosyphons
Are always gravity driven!
Loop system enables enhancement of heat
transfer and minimization of flow losses
(pressure drop).
Generally have better performance compared to
Heat Pipes working with gravity.
Schematic of a Thermosyphon
PCB
LiquidHot
Component
Liquid-
Vapor
Mixture
Evaporator
Condenser
Air
Example of a
Thermosyphon
cooling three
components in
parallel
1200
988
Falling tube length=1750mm
Rising tube height=1200 mm
27
Liquid head:988+27=1015 mm
Condenser
101510
273
Fallingtube
5 hole with d_f=1.5 mm
Evaporator
Risingtube
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Example of aThermosyphon
cooling three
components in
series
Areas in a thermosyphon
Component, 1 cm2
Evaporator, front, 2.2 cm2
Evaporator, inside, 3.5 cm2
Condenser, inside, 108 cm2
Condenser, facing air,
(heat sink included), 5400 cm2
4 times
Advantages with Thermosyphon cooling:
Large heat fluxes can be dissipated from small
areas with small temperature differences
(150 W/cm2)
Heat can be transferred long distances without
any (or with very small) decrease in
temperature.
Hot side Cold side
Temp
Saturation tempBoiling
Condensation
Temperatures obtained experimentally in aThermosyphon system that has three evaporators thateach cool one component. The total heat dissipation is170 W.
Component Contact
resistance
Evaporation
Saturation
temperature
Condensation
Contact
resistance
Thermosyphon
Fin to
air
Air
Condenser
Evaporator
Temperature difference as a function of the
heat dissipation(Prototype C, Condenser is fan cooled)
Data:P8F2MAX.STA10v * 23c
P (W)
Temp.d
ifference(C)
0
2
4
6
8
10
12
0 40 80 120 160
Filling Ratio = 39% Evaporator2
Condenser
R142b
Evaporator geometries
14.7 mm
d=1.1 mm
10mm
d=1.5 mm
Tc, d=0.8 mm
d=2.5 mm d=3.5 mm
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Cooling of Power Amplifiers in a
Radio Base Station
Thermosyphons - Applications
Thermosyphons - Applications Thermosyphons - Applications
Immersion cooling Two phase flow in a
large diameter tube:
Flow regimes determine heat transfermechanism
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Classification and application of
thermosyphon systems.
Open thermosyphon Closed thermosyphon
Pipe thermosyphon
Single-phase flow
Two-phase flow
Simple loop Thermosyphon
Single-phase flow
Two-phase flow
Closed advanced two-phase flow thermosyphon loop
Thermosyphon is a circulating fluid system whose motion is
caused by density difference in a body force field which result
from heat transfer.
Thermosyphon can be categorized according to:
1. The nature of boundaries (Is the system open or closed to mass flow)
2. The regime of heat transfer (convection, boiling or both)
3. The number of type of phases present (single-or two-phase state)
4. The nature of the body force (is it gravitational or rotational)
All thermosyphon systems removes heat from prescribed source
and transporting heat and mass over a specific path and rejecting the
heat or mass to a prescribed sink.
gas turbin blade cooling
electrical machine rotor cooling
transformer cooling
nuclear reactor cooling
steam tubes for bakers oven cooling for internal combustion engines
electronics cooling.
The most common industrial thermosyphon
applications include: Open Thermosyphon:Single-phase, natural-
convection open system in the
form of a tube open at the top
and closed at the bottom.
For open thermosyphon
Nua=C1Raam(a/L)C2,
Nua=(ha)/k
a: based on radius
Closed Thermosyphon
(simple pipe)A simple single-phase natural-
convection closed system in the form
of a tube closed at both ends.
It has been found that the closed
single-phase thermosyphon can be
treated as two simple open
thermosyphon appropriately joined at
the midtube exchange region.
The primary problem is that of
modeling the exchange region.
It has been found that the exchange
mechanism is basically convective.
Simple thermosyphon
loop
Advanced thermosyphon
loop
Evaporator
Condenser
Thermosyphon pipe
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Closed loop thermosyphon
Two distinct advantages make the closed-loop
thermosyphon profitable to study:
1. Natural geometric configuration which can be found or
created in many industrial situation.
2. It avoid the entry choking or mixing that occurs in the
pipe thermosyphon
3. For single phase loop:
4. NuL=0.245(GrPr2L/d)0.5 can be used
Two-phase thermosyphon
The advantages of operating two-phase
thermosyphons are:
1. The ability to dissipate high heat fluxes due to
the latent heat of evaporation and condensation
2. The much lower temperature gradients
associated with these process.
3. Reduced weight and volume with smaller heat
transfer area compared to other systems.
Heat pipe and thermosyphon
Thermosyphon and heat pipe cooling both rely on
evaporation and condensation. The difference between
the two types is that in a heat pipe the liquid is
returned from the condenser to the evaporator by
surface tension acting in a wick, but thermosyphon
rely on gravity for the liquid return to the evaporator .
However the cooling capacity of heat pipes are lowerin general compared to the thermosyphon with the
same tube diameter.
Closed advanced two-phase thermosyphonloop
Thermosyphon cooling offers passive circulationand the ability to dissipate high heat fluxes withlow temperature differences between evaporatorwall and coolant when implemented with surfaceenhancement.
An advanced two-phase loop has the possibility of
reducing the total cross section area of connectingtubes and better possibility of close contactbetween the component and the refrigerantchannels than a thermosyphon pipe or a heat pipe.
Thermosyphons
Heat Transfer and Pressure Drop
Rahmatollah Khodabandeh
Heat Transfer Coefficient
At least two different mechanisms behind flow boiling heattransfer: convective and nucleate boiling heat transfer.
General accepted that the convective boiling increasesalong a tube with increasing vapor fraction and mass flux.Increasing convective boiling reduces the wall superheatand suppresses the nucleate boiling. When heat transferincreases with heat flux with almost constant vaporfraction and mass flux, the nucleate boiling dominates theflow boiling process. Due to the fact that the mechanism ofconvective and nucleate boiling can coexist, a goodprocedure for calculating flow boiling must have bothelements.
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all heat transfer correlations can be divided into three basic
models: 1) Superposition model 2) Enhancement model 3)
Asymptotic model
In the superposition model, the two contributions are
simply added to each other, while in the enhancement
model the contribution of nucleate and convective boiling
are multiplied to obtain a single-phase model. In the
asymptotic model the two mechanisms are respectively
dominant in opposite regions.
The local heat transfer coefficient as sum of the two
contributions
Where n is an asymptotic factor equal to 1 for the
superposition model and above 1 for the asymptotic model
( ) ( )nnbnL
nb
ncb
ntp hFhEhhh +=+=
With larger n, the htp is implying more asymptotic behavior
in the respectively dominant region. hL and hnb are the heat
transfer coefficients for one-phase liquid flow and poolboiling respectively. E and F are enhancement and
suppression factors.
Chen, Gungor-Winterton [1986] and Jungs correlations
are based on superposition model.
Shah, Kandlikar and Gungor-Wintertons [1987]
correlations are based on enhancement model.
Liu-Winterton, Steiner-Taborek and VDI-Wrmeatlas are
based on asymptotic model.
Lazarek-Black, Tran and Crnwell-Kew have developed heat
transfer correlations for small diameter channel.
Coopers pool boiling correlation or Liu-Wintertons flow
boiling correlation can be used for heat transfer coefficient in
an advanced closed two-phase flow thermosyphon loop.
Liu-Winterton correlation( ) ( )[ ]
( )( )( ) ( )
( )
( )[ ]( )
( ) ( ) 4.0l8.0
ll
l
116.0l
1.0
35.0
g
ll
67.05.055.0r
12.0rpool
5.02pool
2ltp
PrRed
k023.0h
ReE055.01s
1Prx1E
qMp10logp55h
hshEh
=
+=
+=
=
+=
Total thermal resistance in an advanced closed two-phase
flow thermosyphon loop
The thermosyphons thermal resistance can be considered to the sum
of four major component resistances:
Rtot=Rcr+Rbo+Rco+Rcv
(K/W)
Rcr
is the contact resistance between the simulated component and the
evaporator front wall. In order to reduce Rcr
a thermally conductive
epoxy can be used.
Rbo
, is the boiling resistance.
Rco, is the condensing resistance. This resistance is in fact very low dueto the high heat transfer coefficient in condensation and the large
condensing area.
Rcv
is the convection resistance between the condenser wall and the air.
Heat transfer depends on pressure level, vapor fraction,flow rate, geometry of evaporator and thermal properties of
refrigerant.
The influence of pressure level, choice of working fluid,
geometry of evaporator, pressure drop, heat transfer
coefficient, critical heat flux and overall thermal resistance
were investigated during the present project.
Considerations when choosing refrigerant
A fluid which needs small diameter of
tubing
A fluid which gives low temp. diff. in
boiling and condensation
A fluid which allows high heat fluxes in the
evaporator.
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For turbulent single-
phase we can derive
pressure drop as:
For a certain tube
length, diameter and
cooling capacity, the
pressure drop is a
function of viscosity,
density and heat of
vaporization.
4/7
4/1
4/19
4/7
2
2
2
4/11
21
241.0
4
4
/
4
Re
Re158.0
fg
fg
fg
hd
QLp
dh
Q
d
hQ
d
m
A
Vw
dw
f
d
Lwfp
=
===
=
=
=
Fig. shows ratio of viscosity
to density and heat of
vaporization vs. Saturated
pressure, we find that the
general trend is decreasingpressure drop with increasing
pressure and decreasing
molcular weights.
The Two-phase pressure
drops expected to follow the
same trends.
For Saturated temperature
between 0-60 C.
0.00E+00
5.00E-09
1.00E-08
1.50E-08
2.00E-08
2.50E-08
0 5 10 15 20 25 30 35 40
Pressure (bar)
Figureofmerit(Dp)
R32, M=52.02
NH3, M=17.03
R12, M=120.9
R134a, M=102
R22, M=86.47
R600a, M=58.12
Coopers pool boilingcorrelation is plottedversus saturatedpressure for differentfluids: (for saturatedtemp. between 0-60C)
As can been seen heattransfer coefficientgenerally increases
with increasingpressure anddecreasing themolecular weights.
0
5000
10000
15000
20000
25000
30000
35000
40000
45000
0 5 10 15 20 25 30 35 40
Ps(bar)
h-Coo
erW/mK
NH3, M=17.03
R32, M=52.02
R600a, M=58.12
R134a, M=102
R12, M=120.9
R22, M=86.47
R11, M=137.4
Another important
parameter when choosing
working fluid is the critical
heat flux.
Figure shows calculation of
Kutateladze CHF correlation
versus reduced pressure for
pool boiling.
As can been seen ammonia
once again shows
outstanding properties with3-4 times higher than the
other fluids.
0
300
600
900
1200
1500
1800
2100
2400
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7
Reducedpressure
CHF(W)
R600a, M=58.12
R11, M=137.4
NH3, M=17.03
R134a, M=102
R12, M=120.9
R22, M=86.47
R32, M=52.02
FC fluids In immersion boiling FC fluids have been used
FC fluids generally have poor heat transferproperties:
-Low thermal conductivity
-Low specific heat
-Low heat of vaporization
-Low surface tension
-Low critical heat flux
-Large temperature overshoot at boilingincipience
Influence of system pressure and
threaded surface
R600a (Isobutane)
Tests were done at five reduced pressures ;
; 0.02, 0.05, 0.1, 0.2 and 0.3.
Two types of evaporators: smooth and
threaded tube surfaces.
crr
p
pp =
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The picture shows heat flux vs.
temperature difference between
inside wall temperature and
refrigerant.
As can be seen, the temperature
difference increases with
increasing heat flux, but with
different slopes, depending on the
saturation pressure in the system
As the heat transfer coefficient is
the heat flux divided by the temp.
difference, this indicates higher
heat transfer coefficient with
increasing pressure
0
50000
100000
150000
200000
250000
300000
350000
0 5 10 15 20 25
DT (C)
q(W/m)
pr=0.02pr=0.3
Isobutane
Smooth tube
The Fig. shows temperature
difference between inside wall
temperature and refrigerant vs. heat
input.
As can be seen, the temperaturedifference increases with
increasing heat input, but with
different slopes, depending on the
saturation pressure in the system
As the heat transfer coefficient is
the heat flux divided by the temp.
difference, this indicates higher
heat transfer coefficient with
increasing pressure
0
2
4
6
8
10
12
14
16
18
2022
24
0 20 40 60 80 100 120Q (W)
DT(C)
pr=0.3
pr=0.2
pr=0.1
pr=0.05
pr=0.02
The Fig. shows, heat transfer
coeff. vs. reduced pressure for 110
W heat input to each one of the
evaporators.
The dependence of heat transfer
coefficient on reduced pressure are
often expressed in the form of h=f
(prm), in which m is generally
between 0.2-0.35.
In the present case, m=0.317,
correlates the experimental data
well for the smooth tube with
Isobutane as refrigerant.
h = constantpr0.317
R2 = 0.9957
0
5000
10000
15000
20000
25000
30000
35000
40000
45000
0 0.05 0.1 0.15 0.2 0.25 0.3 0.35
pr
h(W/m.K)
Q=110 W
Effect of threaded surface
at different reduced
pressure on heat transfer
coefficient
The fig. shows temp. diff. vs.
reduced pressure from 10 to 110 W
heat input for each one of
evaporators on threaded surface.
Relatively low temp. diff can be
achieved.
Temp. diff. In the most points willbe reduced to less than a third by
increasing the reduced pressure
from 0.02 to 0.3.
0
1
2
3
4
5
6
7
8
9
10
0 0.1 0.2 0.3 0.4
pr
DT(C)
10 W
30 W
50 W
70 W
90 W
110 W
Effect of heat flux on heat
transfer coefficient
Figur shows the relation
between heat transfer
coefficient and heat flux for
Pr=0.1, with smooth tube.
The dependence of heat
transfer coefficient on heat
flux can be expressed as h=f
(qn), n, in most cases varies
between 0.6-0.8
Presented data follows h=f
(q0.57)
y = 0.8761x0.5755
R2 = 0.9984
0
5
10
15
20
25
0 40 80 120 1 60 2 00 2 40 2 80
q (kW/m)
h(kW/m.K)
R600a
h=f (qn)
h=f (q0.57
)
Comparison between
Coopers correlation andexperimental results
The Fig. shows heat transfer coeff.
comparison between Coopers pool
boiling correlation versus
experimental results for smooth
tube surfaces at different reduced
pressure.
As can be seen the heat transfer
coeff. calculated by Coopers
correlation is in good agreement
with the experimental results
For the most points the deviation
is less than 25 percent.
0
10000
20000
30000
40000
50000
0 10000 20000 30000 400 00 50000
h-exp (W/mK)
h-Cooper(W/mK)
Q=10 W
Q=30 W
Q=50 W
Q=70 W
Q=90 W
Q=110 W
25%
25%
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Comparison between Liu-
Wintertons correlation and
experimental results
The Fig. shows heat transfer coeff.,
comparison between Liu-Wintertons
correlation versus experimental
results for smooth tube surfaces at
different reduced pressure.
As can be seen the heat transfer
coeff. calculated by Liu-Wintertons
correlation is in good agreement with
the experimental results
For the most points the deviation is
less than 25 percent.
0
10000
20000
30000
40000
50000
0 10 00 0 2 000 0 300 00 40 00 0 50 000
h-exp (W/mK)
h-LW(W/mK)
10 W30 W
50 W
70 W
90 W
110 W
25%
25%
Influence of diameter
Testing condition
R600a as refrigerant
Tests were done with 7, 5,4, 3, 2 and 1 vertical
channels with diameter of 1.1, 1.5,1.9, 2.5 3.5 and
6 mm.
Smooth surface
At reduced pressure 0.1 (p/pcr)
Influence of diameter
Heat transfer coefficient vs.
heat flux at different diameters.
The influence of diameter on
the heat transfer coefficients for
these small diameter channels
was found to be small and no
clear trends could be seen.
0
5
10
15
20
25
30
0 50 100 150 200 250 300 350
Heat flux (kW/m)
h-exp.
(kW/mK) d=6 mm
d=3.5 mm
d=2.5 mm
d=1.9 mm
1.5 m m
d=1.1mm
Conclusions
Heat transfer coefficients and CHF can be expected toIncrease with increasing reduced pressure and withdecreasing molecular weight
The effects of pressure, and threaded surface on heattransfer coefficient have been investigated.
The pressure level has a significant effect on heattransfer coefficient.
h=f (prm) m=0.317
h=f (qn) where n=0.57
Conclusion
Heat transfer coefficient can be improved by usingthreaded surfaces.
Heat transfer coefficient at a given heat fluxis more than three times larger at the reducedpressure 0.3 than 0.02 on threaded surfaces.
The experimental heat transfer coefficients arein relatively good agreement with CoopersPool boiling and Liu-Wintertons correlations.
Conclusion
The effects of pressure, mass flow, vapor quality, andenhanced surface on CHF have been investigated.
Threaded surface has a minor effect on CHF.
The pressure level has a significant effect on CHF.
The CHF can be increased by using a higher pressure.
The influence of diameter on the heat transfer coefficientsfor these small diameter channels was found to be smalland no clear trends could be seen.
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Operation condition of an advanced two-phase
thermosyphon loop
The net driving head caused by the difference in densitybetween the liquid in the downcomer and the vapor/liquid
mixture in the riser must be able to overcome the pressure
drop caused by mass flow, for maintaining fluid circulation.
The pressure changes along the thermosyphon loop due to
gravitation, friction, acceleration, bends, enlargements and
contractions.
In design of a compact two-phase thermosyphon system,
the dimensions of connecting tubing and evaporator,
affects the packaging and thermal performance of thesystem.
The pressure drop is a limiting factor for small tubing
diameter and compact evaporator design.
By determining the magnitude of pressure drops at
different parts of a thermosyphon, it may be possible to
reduce the most critical one, therby optimizing the
performance of the thermosyphon system.
Single-phase flow pressure drop in downcomer
The total pressure drop in the downcomer consists of two
components: frictional pressure drop and pressure drop due
to bends respectively.
For fully developed laminar flow in circular tubes, the
frictional pressure drop can be calculated by:
For the turbulent flow regime, the Blasius correlation forthe friction factor can used:
ll
d
LGp
=2
Re
16
l
ld
LGp
= 2
Re079.0 25.0
The pressure loss around bends can be calculated by:
where is an empirical constant which is a function of
curvature and inner diameter.
In the downcomer section, the pressure drop due to friction
is much larger than the pressure loss around bends.
l
lb
Gp
=
2
Two-phase flow pressure drop Two-phase flow in the riser and evaporator:
The total two-phase flow pressure drop consists of six
components:
1. Acceleration pressure drop
2. Friction pressure drop
3. Gravitational pressure drop
4. Contraction pressure drop
5. Enlargement pressure drop
6. Pressure drop due to the bends
7. Frictional and gravitational pressure drop are most important
pressure drops in the riser
Method of analysis two-phase flow pressure drop
The methods used to analyse a two-phase flow are often
based on extensions of single-phase flows.
The procedure is based on writing conservation of mass,
momentum and energy equations.
To solve these equations, often needs simplifying
assumptions, which give rise different models.
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Homogeneous flow model
One of the simplest predictions of pressure drop in two-
phase flow is a homogeneous flow approximation.
Homogeneous predictions treat the two-phase mixture as a
single fluid with mixture properties.
In the homogeneous flow model it is assumed that the two
phases are well mixed and therefore have equal actual
vapor and liquid velocities.
In other words in this model, the frictional pressure drop is
evaluated as if the flow were a single-phase flow, by
introducing modified properties in the single-phase friction
coefficient.
Separated flow model
The separated flow model is based on assumption that two
phases are segregated into two separated flows that have
constant but not necessarily equal velocities.
Drift flux model
This model is a type of separated flow model, which looks
particularly at the relative motion of the phases. The model
is most applicable when there is a well-defined velocity in
the gas phase
Pressure drop in the riser
The total two-phase flow pressure drop in the riser is
mainly the sum of two contributions: the gravitational-
and the frictional pressure drop.
The most used correlations for calculation of frictional
pressure drop are:
1. Lockhart-Martinelli correlation
2. CESNEF-2 correlation
3. Friedel correlation
4. Homogeneous flow model correlation
In the homogeneous model, the analysis for single-phaseflow is valid for homogeneous density and viscosity. Thehomogeneous density is given by:
Several different correlations have been proposed forestimation of two-phase viscosity, such as:
Cicchitti et al.
Beattie- Whalley
Mc Adams et al.
Dukler et al.
Lgh
xx
+=
11
( ) Lgh x1x +=
)5.21)(1( gLh ++=
Lgh
xx
+=
11
( )
L
hL
g
hg
h
xx
1 +=
g
hx
=
Gravitational pressure drop
The gravitational or head pressure change at the riser
The momentum equation gives:
Where is void fraction A: total cross-section area (m2)
Ag: average cross-section area occupied by the gas phase (m2)
Void fraction can be calculated by:
1. Homogeneous model
2. Zivimodel [1963]
3. Turner& Wallis two-cylinder model [1965]
4. Lockhart-Martinelli correlation [1949]
5. Thom correlation [1964]
6. Baroczy correlation [1963]
rmRG Hgp , =
Lgm )1( +=
A
Ag= For the homogeneous flow the phase velocities are equal,
uL=ug, , where Sis the slip ratio.
+
=
L
g
L
g
x
x
u
u
)1(1
1
L
g
u
uS=
+
=
L
g
h
x
x
)1(1
1
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Acceleration pressure drop
Acceleration pressure drop in the evaporator, resulting
from the expansion due to the heat input during the
evaporation process can be calculated:
(homogeneous model)
v specific volume
xvvGp Lg )(2 =
Experimental setup
Not to scale
939
974
Condenser
8
95
10
15
10
Evaporator
Downcomer
5 hl med d_ f=1.5 mm5 hole with d_f=1.5 mm
ID=6.1 mm
Abs.pressuretransduc
er
1160
186
255
150
glass tube
77
Fig. 1
B
C
All dimensions in the figureare in mm
CHFTesting condition
R600a (Isobutane)
Tests were done at three reduced pressures;
0.035, 0.1, and 0.2.
Two types of evaporators: smooth and
threaded tube surfaces.
CHF=f(pr, G, x)Effect of pressure onCHF:The Fig shows temperaturedifference between inside walltemperature and refrigerant for threeevaporators, vs CHF.
For pr =0.2 the CHF is 690 W whichcorrespond to 230 W/cm front area ofthe component which correspond to650 kW/m heat flux for smoothchannels.
As can be seen, the saturationpressure strongly affected the temp.diff. With increased pressure thetemp. diff. decreases in the total rangeof heat load up to CHF.
0
5
10
15
20
25
30
35
350 400 450 500 550 600 650 700 750
Qtot (W)
DT(C)
0.035
0.1
0.2
pr=0.2
pr=0.1pr=0.035
smooth channel
Effect of mass flow on
CHF
The mass flow is a function ofboth heat flux and system pressure.
As can be seen simulations atCHF shows that mass flowincreases with increasing reducedpressure.
This is believed to be theexplanation for the higher CHF.
Higher pressure gives higher massflow on CHF, which facilitates thedeposition and replenishment ofliquid film.
0
0.001
0.002
0.003
0.004
0.005
0.006
0 100 200 300 400 500 600 700
Qcri
(W)
m_
dot(kg/s)
pr=0.035
pr=0.1
pr=0.2
smooth channel
Effect of vapor
quality on CHF
The Fig. shows, vapor
quality vs. CHF for three
evaporators.
According to the
simulations the vapor
quality at different
pressure on CHF is almost
constant.
00.1
0.20.30.40.50.60.7
0.80.9
1
0 100 200 300 400 500 600 700
Qcri (W)
x
pr=0.035
pr=0.1
pr=0.2
smooth channel
-
7/31/2019 Thermo Syphons
14/14
14
Effect of enhanced surfaceon CHF
Generally at enhanced surfaces
increases the heat transfer.In this study threaded surfaceshave been used to investigate theeffect of surface structure on CHF.
The picture shows the CHF versusreduced pressure for both surfaces.
However the CHF is independenton surface condition.
The fact that the surface conditionis unimportant for CHF werereported by other researcher.
0
100
200
300
400
500
600
700
0 0.05 0. 1 0. 15 0. 2 0.25
pr
Qcri(W)
threaded
smooth
Comparison between
Kutateladzes
correlation and
experimental results
The Fig. shows CHF,
comparison between
Kutateladzes pool boiling
correlation versus
experimental results for
smooth tube surfaces.
Deviation is less than 15
percent.
0
100
200
300
400
500600
700
0 100 200 300 400 500 600 700
Q_cri_exp. (W)
Q_
cri_
pb.
(W)
15%
-15%
Old Exam Problem 2003-03-07A thermosyphon can be quite complex to model. In this assignment
we will investigate the behavior of a simplified thermosyphon. The
difference in height between the condenser and the evaporator is 15
cm. The tube diameter is 5 mm and the downcomer tube length is
16 cm. The heat exchanger area in the condenser and the evaporator
is 40 cm and 4 cm respectively. The total pressure drop in the
rising tube can be calculated using pRiser = 6.21x, where pRiseris in kPa, x is the change in vapor quality in the evaporator. Therefrigerant is R134a for which the latent heat of vaporization,
hfg = 163 kJ/kg, the liquid density,L=1146 kg/m, and dynamicviscosity, L=1.7810-4 Pas. The temperature of the evaporatorwalls is 50 C, the boiling heat transfer coefficient is 20.000
W/(mK), and the heat dissipation is 60 W. Calculate the mass flow
rate, , the change in vapor quality, x, and the saturationtemperature of the refrigerant (6 credits).
m&