Heat pipes and Thermosyphons
Heat pipes and Thermosyphons
Cold end
Cold end
Hot end
Inside the system, there is a fluid (usually termed refrigerant)
Hot end
• Heat is transferred as latent
heat of evaporation which
means that the fluid inside
the system is continuously
changing phase from liquid
to gas.
• The fluid is evaporating at
the hot end, thereby
absorbing heat from the
component.
• At the cold end, the fluid is
condensed and the heat is
dissipated to a heat sink
(usually ambient air).
Heat pipes and Thermosyphons
Heat pipes
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.
1
Heat pipes
Heat pipes - Applications
Heat pipes - Applications
Thermosyphons
LiquidVapor
Mixture
Liquid
Hot
Component
Air
988
1200
Falling tube
Condenser
101510
PCB
Example of a
Thermosyphon
cooling three
components in
parallel
Rising tube
Condenser
Schematic of a Thermosyphon
27 273
• 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.
Evaporator
5 hole with d_f=1.5 mm
Falling tube length=1750mm
Rising tube height=1200 mm
Evaporator
Liquid head:988+27=1015 mm
2
Areas in a thermosyphon
Example of a
Thermosyphon
cooling three
components in
series
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.
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
Temperatures obtained experimentally in a
Thermosyphon system that has three evaporators that
each cool one component. The total heat dissipation is
170 W.
Component
Contact
resistance
Saturation
temperature
Evaporation
Condensation
Condenser
Thermosyphon
Temp
Fin to
air
Evaporator
Boiling
Contact
resistance
Saturation temp
Condensation
Hot side
Air
Cold side
Temperature difference as a function of the
heat dissipation
Evaporator geometries
(Prototype C, Condenser is fan cooled)
Data: P8F2MAX.STA 10v * 23c
12
R142b
Filling Ratio = 39%
10 mm
Temp.difference (C)
10
Evaporator2
8
6
d=1.1 mm
14.7 mm
d=1.5 mm
d=2.5 mm
d=3.5 mm
Tc, d=0.8 mm
4
Condenser
2
0
0
40
80
120
160
P (W)
3
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 transfer
mechanism
4
Classification and application of
thermosyphon systems.
• Open thermosyphon
• Closed thermosyphon
– Pipe thermosyphon
• Single-phase flow
• Two-phase flow
•
•
•
•
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.
2.
3.
4.
The nature of boundaries (Is the system open or closed to mass flow)
The regime of heat transfer (convection, boiling or both)
The number of type of phases present (single- or two-phase state)
The nature of the body force (is it gravitational or rotational)
– Simple loop Thermosyphon
• Single-phase flow
• Two-phase flow
• Closed advanced two-phase flow thermosyphon loop
The most common industrial thermosyphon
applications include:
•
•
•
•
•
•
•
gas turbin blade cooling
electrical machine rotor cooling
transformer cooling
nuclear reactor cooling
steam tubes for baker’s oven
cooling for internal combustion engines
electronics cooling.
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.
Open Thermosyphon:
•Single-phase, naturalconvection open system in the
form of a tube open at the top
and closed at the bottom.
•For open thermosyphon
•Nua=C1·Raam(a/L)C2,
Nua=(h·a)/k
•a: based on radius
•Closed Thermosyphon
(simple pipe)
•A simple single-phase naturalconvection 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.
Condenser
Evaporator
Thermosyphon pipe
Simple thermosyphon
loop
Advanced thermosyphon
loop
5
•
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.
It avoid the entry choking or mixing that occurs in the
pipe thermosyphon
For single phase loop:
NuL=0.245·(Gr·Pr2·L/d)0.5 can be used
•
2.
3.
4.
• 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 lower
in general compared to the thermosyphon with the
same tube diameter.
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.
• Closed advanced two-phase thermosyphon
loop
• Thermosyphon cooling offers passive circulation
and the ability to dissipate high heat fluxes with
low temperature differences between evaporator
wall and coolant when implemented with surface
enhancement.
• An advanced two-phase loop has the possibility of
reducing the total cross section area of connecting
tubes and better possibility of close contact
between the component and the refrigerant
channels than a thermosyphon pipe or a heat pipe.
• Heat Transfer Coefficient
Thermosyphons –
Heat Transfer and Pressure Drop
Rahmatollah Khodabandeh
• At least two different mechanisms behind flow boiling heat
transfer: convective and nucleate boiling heat transfer.
• General accepted that the convective boiling increases
along a tube with increasing vapor fraction and mass flux.
Increasing convective boiling reduces the wall superheat
and suppresses the nucleate boiling. When heat transfer
increases with heat flux with almost constant vapor
fraction and mass flux, the nucleate boiling dominates the
flow boiling process. Due to the fact that the mechanism of
convective and nucleate boiling can coexist, a good
procedure for calculating flow boiling must have both
elements.
6
• 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 htpn = h n + h n = (E·hL )n + (F ·hnb )n
cb
b
• Where n is an asymptotic factor equal to 1 for the
superposition model and above 1 for the asymptotic model
• Lazarek-Black, Tran and Crnwell-Kew have developed heat
transfer correlations for small diameter channel.
• Cooper’s pool boiling correlation or Liu-Winterton’s flow
boiling correlation can be used for heat transfer coefficient in
an advanced closed two-phase flow thermosyphon loop.
• Liu-Winterton correlation
0.5
h tp = [(E ⋅ h l )2 + (s ⋅ h pool )2 ]
0.12
⋅ (− log 10(p r ))(−0.55 ) ⋅ M (−0.5 ) ⋅ q 0.67
h pool = 55 ⋅ p r
⎡
⎛ρ
⎞⎤
E = ⎢1 + (x ) ⋅ Prl ⋅ ⎜ l − 1⎟⎥
⎜ ρg
⎟⎥
⎢⎣
⎝
⎠⎦
[
0.35
s = 1 + 0.055 ⋅ E 0.1 ⋅ (Re l )0.16
](
−1)
⎛k ⎞
h l = 0.023 ⋅ ⎜ l ⎟ ⋅ (Re l )0.8 ⋅ (Prl )0.4
⎝ d ⎠
• 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.
• 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 pool
boiling respectively. E and F are enhancement and
suppression factors.
• Chen, Gungor-Winterton [1986] and Jung’s correlations
are based on superposition model.
• Shah, Kandlikar and Gungor-Winterton’s [1987]
correlations are based on enhancement model.
• Liu-Winterton, Steiner-Taborek and VDI-Wärmeatlas are
based on asymptotic model.
• Total thermal resistance in an advanced closed two-phase
flow thermosyphon loop
• The thermosyphon’s thermal resistance can be considered to the sum
of four major component resistances:
•
(K/W)
• Rtot=Rcr+Rbo+Rco+Rcv
•
• 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 due
to the high heat transfer coefficient in condensation and the large
condensing area.
• Rcv is the convection resistance between the condenser wall and the air.
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.
7
w·d
• For a certain tube
length, diameter and
cooling capacity, the
pressure drop is a
function of viscosity,
density and heat of
vaporization.
υ
·
V
w= =
A
·
·
m
π ·d 2
ρ·
Q/ h fg
≈
π ·d 2
ρ·
4
· 7/4
Δp = 0.241·
L·Q
d
19 / 4
·
·
=
4·Q
h fg ·ρ ·π ·d 2
4
μ1/ 4
h-Cooper (W/m²·K)
NH3, M=17.03
R32, M=52.02
R600a, M=58.12
20000
15000
10000
R134a, M=102
R12, M=120.9
R22, M=86.47
5000
0
R11, M=137.4
10
15
20
25
30
35
Ps (bar)
40
•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 with
3-4 times higher than the
other fluids.
•
•
•
•
•
•
In immersion boiling FC fluids have been used
FC fluids generally have poor heat transfer
properties:
-Low thermal conductivity
-Low specific heat
-Low heat of vaporization
-Low surface tension
-Low critical heat flux
-Large temperature overshoot at boiling
incipience
NH3, M=17.03
R12, M=120.9
R134a, M=102
1.00E-08
R22, M=86.47
5.00E-09
R600a, M=58.12
0.00E+00
0 5 10 15 20 25 30 35 40
2400
2100
R600a, M=58.12
1800
R11, M=137.4
1500
NH3, M=17.03
1200
R134a, M=102
900
R12, M=120.9
600
R22, M=86.47
300
R32, M=52.02
0
0
0.1 0.2 0.3 0.4 0.5 0.6 0.7
Reduced pressure
Influence of system pressure and
threaded surface
FC fluids
•
•
R32, M=52.02
1.50E-08
•For Saturated temperature
between 0-60 °C.
35000
30000
25000
5
2.00E-08
Pressure (bar)
ρ ·h7fg/ 4
45000
40000
0
2.50E-08
Figure of merit (Dp)
f1 = 0.158·Re−1 / 4
Re =
• Cooper’s pool boiling
correlation is plotted
versus saturated
pressure for different
fluids: (for saturated
temp. between 0-60
°C)
• As can been seen heat
transfer coefficient
generally increases
with increasing
pressure and
decreasing the
molecular weights.
L
d
Δp = f1·ρ ·w2 ·
CHF (W)
• For turbulent singlephase we can derive
pressure drop as:
•Fig. shows ratio of viscosity
to density and heat of
vaporization vs. Saturated
pressure, we find that the
general trend is decreasing
pressure drop with increasing
pressure and decreasing
molcular weights.
•The Two-phase pressure
drops expected to follow the
same trends.
• R600a (Isobutane)
• Tests were done at five reduced pressures ;
p
• p r = p cr
; 0.02, 0.05, 0.1, 0.2 and 0.3.
• Two types of evaporators: smooth and
•
threaded tube surfaces.
8
350000
pr=0.3
pr=0.02
300000
250000
q (W/m²)
200000
150000
100000
Isobutane
Smooth tube
50000
0
•As the heat transfer coefficient is
the heat flux divided by the temp.
difference, this indicates higher
heat transfer coefficient with
increasing pressure
5
10
15
20
25
DT (°C)
40000
30000
0.317
h = constant·pr
R2 = 0.9957
20000
15000
Q=110 W
5000
0
0
0.05
0.1
0.15
0.2
0.25
0.3
0.35
pr
•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 will
be reduced to less than a third by
increasing the reduced pressure
from 0.02 to 0.3.
DT (C°)
25000
•In the present case, m=0.317,
correlates the experimental data
well for the smooth tube with
Isobutane as refrigerant.
pr=0.3
pr=0.2
pr=0.1
pr=0.05
pr=0.02
0
35000
10000
20
40
60
Q (W)
80
100
120
10 W
10
9
8
7
6
5
4
3
2
1
0
30 W
50 W
70 W
90 W
110 W
0
0.1
0.2
0.3
0.4
pr
•Comparison between
Cooper’s correlation and
experimental results
25
•The Fig. shows heat transfer coeff.
comparison between Cooper’s pool
boiling correlation versus
experimental results for smooth
tube surfaces at different reduced
pressure.
y = 0.8761x0.5755
R2 = 0.9984
R600a
20
h (kW/m².K)
•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)
24
22
20
18
16
14
12
10
8
6
4
2
0
•Effect of threaded surface
at different reduced
pressure on heat transfer
coefficient
45000
h (W/m².K)
•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.
0
•As the heat transfer coefficient is
the heat flux divided by the temp.
difference, this indicates higher
heat transfer coefficient with
increasing pressure
15
n
h=f (q )
10
0.57
h=f (q
)
5
0
0
40
80
120
160
q (kW/m²)
200
240
280
•As can be seen the heat transfer
coeff. calculated by Cooper’s
correlation is in good agreement
with the experimental results
•For the most points the deviation
is less than 25 percent.
Q=10 W
50 000
h- Coope r (W/m ²·K)
•As can be seen, the temperature
difference increases with
increasing heat flux, but with
different slopes, depending on the
saturation pressure in the system
•The Fig. shows temperature
difference between inside wall
temperature and refrigerant vs. heat
input.
•As can be seen, the temperature
difference increases with
increasing heat input, but with
different slopes, depending on the
saturation pressure in the system
DT (°C)
•The picture shows heat flux vs.
temperature difference between
inside wall temperature and
refrigerant.
Q=30 W
25%
Q=50 W
40 000
Q=70 W
Q=90 W
30 000
25%
Q=110 W
20 000
10 000
0
0
10000
20000
30000
400 00
50000
h-e xp (W/m²·K)
9
•Comparison between LiuWinterton’s correlation and
experimental results
10 W
50000
30 W
25%
50 W
40000
h- LW( W/m ²·K)
•The Fig. shows heat transfer coeff.,
comparison between Liu-Winterton’s
correlation versus experimental
results for smooth tube surfaces at
different reduced pressure.
Influence of diameter
Testing condition
70 W
30000
25%
90 W
110 W
20000
10000
•As can be seen the heat transfer
coeff. calculated by Liu-Winterton’s
correlation is in good agreement with
the experimental results
•For the most points the deviation is
less than 25 percent.
0
0
10000
20000
30000
40000
50000
h-exp (W/m²·K)
Conclusions
Influence of diameter
• Heat transfer coefficient vs.
heat flux at different diameters.
30
h-exp. (kW/m²·K)
• 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.
• 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)
25
d=6 m m
20
d=3.5 m m
d=2.5 m m
15
d=1.9 m m
10
1.5 m m
5
• Heat transfer coefficients and CHF can be expected to
Increase with increasing reduced pressure and with
decreasing molecular weight
• The effects of pressure, and threaded surface on heat
transfer coefficient have been investigated.
d=1.1m m
0
0
50 100 150 200 250 300 350
Heat flux (kW/m ²)
Conclusion
• Heat transfer coefficient can be improved by using
threaded surfaces.
• Heat transfer coefficient at a given heat flux
is more than three times larger at the reduced
pressure 0.3 than 0.02 on threaded surfaces.
• The experimental heat transfer coefficients are
in relatively good agreement with Cooper’s
Pool boiling and Liu-Winterton’s correlations.
• The pressure level has a significant effect on heat
transfer coefficient.
h=f (prm) m=0.317
h=f (qn) where n=0.57
Conclusion
The effects of pressure, mass flow, vapor quality, and
enhanced 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 coefficients
for these small diameter channels was found to be small
and no clear trends could be seen.
10
• Operation condition of an advanced two-phase
thermosyphon loop
• In design of a compact two-phase thermosyphon system,
the dimensions of connecting tubing and evaporator,
affects the packaging and thermal performance of the
system.
• The net driving head caused by the difference in density
between 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 drop is a limiting factor for small tubing
diameter and compact evaporator design.
• The pressure changes along the thermosyphon loop due to
gravitation, friction, acceleration, bends, enlargements and
contractions.
• 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:
Δpl =
16 2 ⋅ G ² ⋅ L
⋅
Re d ⋅ ρ
l
•
• For the turbulent flow regime, the Blasius correlation for
the friction factor can used:
2 ⋅ G² ⋅ L
Δpl = 0.079 ⋅ Re −0.25 ⋅
•
d⋅ρ
• 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.
• The pressure loss around bends can be calculated by:
Δplb = ξ ⋅
•
G²
2 ⋅ ρl
ξ is an empirical constant which is a function of
where
curvature and inner diameter.
• In the downcomer section, the pressure drop due to friction
is much larger than the pressure loss around bends.
l
•
•
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.
2.
3.
4.
5.
6.
7.
Acceleration pressure drop
Friction pressure drop
Gravitational pressure drop
Contraction pressure drop
Enlargement pressure drop
Pressure drop due to the bends
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.
11
• Homogeneous flow model
• 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.
• One of the simplest predictions of pressure drop in twophase 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.
• 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
• In the homogeneous model, the analysis for single-phase
flow is valid for homogeneous density and viscosity. The
homogeneous density is given by:
x 1− x
1
•
The total two-phase flow pressure drop in the riser is
mainly the sum of two contributions: the gravitationaland the frictional pressure drop.
The most used correlations for calculation of frictional
pressure drop are:
Lockhart-Martinelli correlation
CESNEF-2 correlation
Friedel correlation
Homogeneous flow model correlation
ρh ρ g ρL
• Several different correlations have been proposed for
estimation of two-phase viscosity, such as:
•
Cicchitti et al.
μ h = x·μ g + 1 − x ·μ L
•
x 1− x
1
•
Beattie- Whalley
=
+
•
1.
2.
3.
4.
•
Δp G , R = ρ m · g · H r
The gravitational or head pressure change at the riser
•
•
•
•
•
1.
2.
3.
4.
5.
6.
The momentum equation gives:
ρ m = α ·ρ g + (1 − α )·ρ L
Where α is void fraction
A
α= g
A: total cross-section area (m2)
A
Ag: average cross-section area occupied by the gas phase (m2)
Void fraction can be calculated by:
Homogeneous model
Zivi model [1963]
Turner& Wallis two-cylinder model [1965]
Lockhart-Martinelli correlation [1949]
Thom correlation [1964]
Baroczy correlation [1963]
•
• Drift flux model
=
(
μh
μg
+
)
μL
•
μ h = μ L ·(1 − β )·(1 + 2.5·β ) + μ g ·β
•
μh =
μ g · x·ρ h μ L ·(1 − x )·ρ h
+
ρg
ρL
Mc Adams et al.
β=
x·ρ h
ρg
Dukler et al.
Gravitational pressure drop
α =
•
1
⎡ u g (1 − x) ρ g ⎤
1+ ⎢
⎥
⎣uL x ρL ⎦
For the homogeneous flow the phase velocities are equal,
uL=ug, S = u g , where S is the slip ratio.
uL
αh =
1
⎡ (1 − x) ρ g ⎤
1+ ⎢
⎥
⎣ x ρL ⎦
12
Fig. 1
Experimental setup
glass tube
77
• Acceleration pressure drop
186
255
150
Condenser
• Acceleration pressure drop in the evaporator, resulting
from the expansion due to the heat input during the
evaporation process can be calculated:
• (homogeneous model)
Δ p = G 2 ·( v g − v L )· x
• v specific volume
1160
ID=6.1 mm
974
Downcomer
939
Abs.
pressure
transduc
er
8
Evaporator
10 15 10
Not to scale
C
B
95
5 hål
holemed
with d_f=1.5
d_f=1.5 mm
mm
All dimensions in the figure
are in mm
CHF
•CHF=f(pr, G, x)
Effect of pressure on
CHF:
• 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.
•Effect of mass flow on
CHF
0.006
m _dot (k g/s )
• The Fig. shows, vapor
quality vs. CHF for three
• evaporators.
pr =0.2
0.003
0.002
smooth channel
0.001
0
0
100 200 300 400 500 600 700
Qcri (W)
• According to the
simulations the vapor
quality at different
pressure on CHF is almost
constant.
pr=0.1
20
pr=0.2
15
10
5
0
smooth channel
0.035
0.1
0.2
350 400 450 500 550 600 650 700 750
Qto t (W)
x
pr =0.1
0.004
pr=0.035
25
pr =0.035
0.005
•Higher pressure gives higher mass
flow on CHF, which facilitates the
deposition and replenishment of
liquid film.
•
30
• Effect of vapor
quality on CHF
•The mass flow is a function of
both heat flux and system pressure.
•As can be seen simulations at
CHF shows that mass flow
increases with increasing reduced
pressure.
•This is believed to be the
explanation for the higher CHF.
35
•The Fig shows temperature
difference between inside wall
temperature and refrigerant for three
evaporators, vs CHF.
•For pr =0.2 the CHF is 690 W which
correspond to 230 W/cm² front area of
the component which correspond to
650 kW/m² heat flux for smooth
channels.
•As can be seen, the saturation
pressure strongly affected the temp.
diff. With increased pressure the
temp. diff. decreases in the total range
of heat load up to CHF.
DT(°C)
Testing condition
1
pr =0.035
0.9
0.8 pr =0.1
0.7 pr =0.2
0.6
0.5
0.4
0.3
0.2 smooth channel
0.1
0
0 100 200 300 400 500 600 700
Qcri (W)
13
700
600
Qcr i (W)
•Generally at enhanced surfaces
increases the heat transfer.
•In this study threaded surfaces
have been used to investigate the
effect of surface structure on CHF.
500
400
300
200
•The picture shows the CHF versus
reduced pressure for both surfaces.
•However the CHF is independent
on surface condition.
•The fact that the surface condition
is unimportant for CHF were
reported by other researcher.
threaded
100
smooth
0
0
0.05
0.1
0.15
0.2
0.25
• Comparison between
Kutateladze’s
correlation and
experimental results
• The Fig. shows CHF,
comparison between
Kutateladze’s pool boiling
correlation versus
experimental results for
smooth tube surfaces.
pr
700
15%
600
Q_cri_pb. (W)
•Effect of enhanced surface
on CHF
500
-15%
400
300
200
100
0
0
100 200 300 400 500 600 700
Q_cri_exp. (W)
• Deviation is less than 15
percent.
Old Exam Problem 2003-03-07
A 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 cm² and 4 cm² respectively. The total pressure drop in the
rising tube can be calculated using ΔpRiser = 6.21·Δx, where ΔpRiser
is in kPa, Δx is the change in vapor quality in the evaporator. The
refrigerant is R134a for which the latent heat of vaporization,
hfg = 163 kJ/kg, the liquid density, ρL=1146 kg/m³, and dynamic
viscosity, μL=1.78·10-4 Pa·s. The temperature of the evaporator
walls is 50 °C, the boiling heat transfer coefficient is 20.000
W/(m²·K), and the heat dissipation is 60 W. Calculate the mass flow
& , the change in vapor quality, Δx, and the saturation
rate, m
temperature of the refrigerant (6 credits).
14