Heat pipes and Thermosyphons. Heat pipes and Thermosyphons. Heat pipes and Thermosyphons. Heat pipes. Heat pipes. Heat pipes

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 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).

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

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 Liquid Hot Liquid-Vapor Mixture Condenser Air

Example of a

Thermosyphon

cooling three

components in

parallel

12 00 98 8

Falling tube length=1750mm

27 Condenser 10 15 10 27 3 Fa llin g t u b e 5 hole with d_f=1.5 mm Evaporator R is ing tu be

Example of a

Thermosyphon

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/cm

2

₎

• Heat can be transferred long distances without

any (or with very small) decrease in

temperature.

Hot side Cold side

Temp

Saturation temp Boiling

Condensation

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 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.STA 10v * 23c P (W) Te mp .d iffe re n c e (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 10 mm d=1.5 mm Tc, d=0.8 mm d=2.5 mm _{d=3.5 mm}

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

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 baker’s 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=C1·Raam(a/L)C2,

Nua=(h·a)/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

• 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. Nu_L=0.245·(Gr·Pr2_·L/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 lower in general compared to the thermosyphon with the same tube diameter.

• 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.

Thermosyphons –

Heat Transfer and Pressure Drop

Rahmatollah Khodabandeh

• Heat Transfer Coefficient

• At least two different mechanisms behind flow boiling heat transfer: convectiveand nucleate boilingheat 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 boilingdominates 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.

• 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

( ) ( )n nb n L n b n cb n tp h h Eh Fh h = + = · + ·

• With larger n, the htpis implying more asymptotic behavior

in the respectively dominant region. h_Land h_nbare 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.

• 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 ( )

(

)

[

]

( ) ( )( ) ( ) ( ) ( )

[

]

( ) ( ) ( )0.4 l 8 . 0 l l l 1 16 . 0 l 1 . 0 35 . 0 g l l 67 . 0 5 . 0 55 . 0 r 12 . 0 r pool 5 . 0 2 pool 2 l tp Pr Re d k 023 . 0 h Re E 055 . 0 1 s 1 Pr x 1 E q M p 10 log p 55 h h s h E h ⋅ ⋅ ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ ⋅ = ⋅ ⋅ + = ⎥ ⎥ ⎦ ⎤ ⎢ ⎢ ⎣ ⎡ ⎟ ⎟ ⎠ ⎞ ⎜ ⎜ ⎝ ⎛ − ρ ρ ⋅ ⋅ + = ⋅ ⋅ − ⋅ ⋅ = ⋅ + ⋅ = − − −

• 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:

•

• Rtot=Rcr+Rbo+Rco+Rcv (K/W) •

• R_cris the contact resistance between the simulated component and the evaporator front wall. In order to reduce Rcra 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.

• R_cvis 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.

• 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 / 1 1 2 1 · · · · 241 . 0 · · · · 4 4 · · / 4 · · · Re ·Re 158 . 0 · · · fg fg fg h d Q L p d h Q d h Q d m A V w d w f d L w f p ρ μ π ρ π ρ π ρ υ ρ = Δ = ≈ = = = = = Δ −

•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.

•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) Fi g ur e of m e ri t (D p) R32, M=52.02 NH3, M=17.03 R12, M=120.9 R134a, M=102 R22, M=86.47 R600a, M=58.12

• 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.

0 5000 10000 15000 20000 25000 30000 35000 40000 45000 0 5 10 15 20 25 30 35 40 Ps (bar) h-C o o p er ( W/ m ²· K ) 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 with 3-4 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 Reduced pressure CH F ( 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 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

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.

cr r

p p p =

•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/m²) pr=0.02 pr=0.3 Isobutane Smooth tube

•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 •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 20 22 24 0 20 40 60 80 100 120 Q (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 = constant·pr0.317 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/m².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 will be 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 80 120 160 200 240 280 q (kW/m²) h ( k W/m ². K ) R600a h=f (qn₎ h=f (q0.57₎ •Comparison between Cooper’s correlation and experimental results

•The Fig. shows heat transfer coeff. comparison between Cooper’s pool boiling correlation versus experimental results for smooth tube surfaces at different reduced pressure.

•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.

0 10 000 20 000 30 000 40 000 50 000 0 10000 20000 30000 400 00 50000 h-e xp (W/m² ·K) h -C o o p e r (W /m ²· K ) Q=10 W Q=30 W Q=50 W Q=70 W Q=90 W Q=110 W 25% 25%

•Comparison between Liu-Winterton’s correlation and experimental results

•The Fig. shows heat transfer coeff., comparison between Liu-Winterton’s correlation versus experimental results for smooth tube surfaces at different reduced pressure. •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 10000 20000 30000 40000 50000 0 10000 20000 30000 40000 50000 h-exp (W/m²·K) h-L W (W /m ²· K ) 10 W 30 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/p

cr

)

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 ²) h-exp. (kW/m²·K ) d=6 m m d=3.5 m m d=2.5 m m d=1.9 m m 1.5 m m d=1.1m m

Conclusions

• 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. • 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

• 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.

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.

• Operation condition of an advanced two-phase

thermosyphon loop

• 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 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 the system.

• 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 for the friction factor can used:

• l l d L G p _ρ ⋅ ⋅ ⋅ ⋅ = Δ 2 ² Re 16 l l d L G p ρ ⋅ ⋅ ⋅ ⋅ ⋅ = Δ − 2 ² Re 079 . 0 0.25

• 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 G p ρ ξ ⋅ ⋅ = Δ 2 ² ξ

•

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.

• 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-phase flow is valid for homogeneous density and viscosity. The homogeneous density is given by:

• Several different correlations have been proposed for estimation of two-phase viscosity, such as:

• Cicchitti et al. • • Beattie- Whalley • Mc Adams et al. • Dukler et al. L g h x x ρ ρ ρ − + = 1 1

(

)

L g h

=

x

·

μ

+

1

−

x

·

μ

μ

β μ β β μ μh= L·(1− )·(1+2.5· )+ g· L g h x x μ μ μ − + = 1 1

(

)

L h L g h g h x x ρ ρ μ ρ ρ μ μ = ·· + ·1− · g h x ρ ρ β₌ ·

•

Gravitational pressure drop

• The gravitational or head pressure change at the riser • The momentum equation gives:

• Where αis void fraction • A: total cross-section area (m2₎

• A_g: average cross-section area occupied by the gas phase (m2₎

• Void fraction can be calculated by: 1. Homogeneous model

2. Zivi model [1963]

3. Turner& Wallis two-cylinder model [1965] 4. Lockhart-Martinelli correlation [1949] 5. Thom correlation [1964] 6. Baroczy correlation [1963] r m R G gH p _, =ρ · · Δ L g m αρ α ρ ρ = · +(1− )· A Ag = α

• For the homogeneous flow the phase velocities are equal,

u_L=u_g, , where S is the slip ratio. ⎥ ⎦ ⎤ ⎢ ⎣ ⎡ − + = L g L g x x u u ρ ρ α ) 1 ( 1 1 L g u u S= ⎥ ⎦ ⎤ ⎢ ⎣ ⎡ − + = L g h x x ρ ρ α ) 1 ( 1 1

• 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 x v v G p = 2·( g − L )· Δ Experimental setup Not to scale 93 9 97 4 Condenser 8 95 10 15 10 Evaporator Do w n c o m e r

5 hål med d_f=1.5 mm5 hole with d_f=1.5 mm ID=6.1 mm Abs. pressure transduc er 11 60 18 6 25 5 15 0 glass tube 77 Fig. 1 B C

All dimensions in the figure are in mm

CHF

Testing 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(p

r

, G, x)

Effect of pressure on

CHF:

•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. 0 5 10 15 20 25 30 35 350 400 450 500 550 600 650 700 750 Qto t (W) DT (° C) 0.035 0.1 0.2 pr=0.2 pr=0.1 pr=0.035 smooth channel

•Effect of mass flow 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. •Higher pressure gives higher mass flow on CHF, which facilitates the deposition and replenishment of liquid film. • 0 0.001 0.002 0.003 0.004 0.005 0.006 0 100 200 300 400 500 600 700 Qcr i (W) m _ dot ( k g /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. 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 0 100 200 300 400 500 600 700 Qcri (W) x pr =0.035 pr =0.1 pr =0.2 smooth channel

•Effect of enhanced surface on CHF

•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. •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.

0 100 200 300 400 500 600 700 0 0.05 0.1 0.15 0.2 0.25 pr Qcr i (W ) thr eaded smooth • 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.

• Deviation is less than 15 percent. 0 100 200 300 400 500 600 700 0 100 200 300 400 500 600 700 Q_cri_exp. (W) Q_ c ri _ p b . ( W ) 15% -15%

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 Δp_Riser= 6.21·Δx, where Δp_Riser is in kPa, Δx is the change in vapor quality in the evaporator. The refrigerant is R134a for which the latent heat of vaporization, h_fg= 163 kJ/kg, the liquid density, ρ_L=1146 kg/m³, and dynamic viscosity, μ_L=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 rate, , the change in vapor quality, Δx, and the saturation temperature of the refrigerant (6 credits).