Heat pipes and Thermosyphons Cold end Hot end Inside the system, there is a fluid (usually termed refrigerant) Heat pipes and Thermosyphons Cold end 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 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 1
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. Liquid- Vapor Mixture PCB Schematic of a Thermosyphon Condenser Example of a Thermosyphon cooling three components in parallel 1151 1 Condenser Rising tube Falling tube Evaporator 988 7 73 Hot Component Liquid Evaporator Air 5 hole with d_f=1.5 mm Falling tube length=175mm Rising tube height=1 mm Liquid head:988+7=115 mm
Areas in a thermosyphon Example of a Thermosyphon cooling three components in series 4 times Component, 1 cm Evaporator, front,. cm Evaporator, inside, 3.5 cm Condenser, inside, 18 cm Condenser, facing air, (heat sink included), 54 cm Advantages with Thermosyphon cooling: Large heat fluxes can be dissipated from small areas with small temperature differences (15 W/cm ) Heat can be transferred long distances without any (or with very small) decrease in temperature. Temperatures obtained experimentally in a Thermosyphon system that has three evaporators that each cool one component. The total heat dissipation is 17 W. Condenser Component Contact resistance Evaporation Thermosyphon Saturation temperature Condensation Contact resistance Temp Boiling Saturation temp Condensation Evaporator Fin to air Air Hot side Cold side Temperature difference as a function of the heat dissipation (Prototype C, Condenser is fan cooled) Evaporator geometries Temp.difference (C) 1 1 8 6 4 Data: P8FMAX.STA 1v * 3c R14b Filling Ratio = 39% Evaporator Condenser 1 mm d=1.1 mm 14.7 mm d=1.5 mm Tc, d=.8 mm d=.5 mm d=3.5 mm 4 8 1 16 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 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). 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. 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. 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 Nu a =C1 Ra am (a/l) C, Nu a =(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. Thermosyphon pipe Simple thermosyphon loop Condenser Evaporator 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 3. For single phase loop: 4. Nu L =.45 (Gr Pr L/d).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. 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: 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 ) 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 n n n n h tp = h ( ) ( ) n cb + hb = E hl + F hnb Where n is an asymptotic factor equal to 1 for the superposition model and above 1 for the asymptotic model With larger n, the h tp is implying more asymptotic behavior in the respectively dominant region. h L and h nb 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. 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.5 h tp = [( E hl ) + ( s h pool ) ].1 ( ( )) (.55 ) (.5 ).67 h = 55 p log1 p M q pool E = 1 + s = ( x).1.16 [ ( ) ] ( 1 1+.55 E Re ) l kl hl =.3 d r.35 l Pr ρ l 1 ρg.8 ( Re ) ( Pr ). 4 l r l 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: R tot =R cr +R bo +R co +R cv (K/W) R cr is the contact resistance between the simulated component and the evaporator front wall. In order to reduce R cr a thermally conductive epoxy can be used. R bo, is the boiling resistance. R co, 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 cv 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. 7
For turbulent singlephase 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. L Δp = f1 ρ w d 1/ 4 f1 =.158 Re w d Re = υ V m Q/ h fg 4 Q w = = = A π d π d h fg ρ π d ρ ρ 4 4 7 / 4 1/ 4 L Q μ Δp =.41 19/ 4 7 / 4 d ρ h fg 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 -6 C. Figure of merit (Dp).5E-8.E-8 1.5E-8 1.E-8 5.E-9.E+ 5 1 15 5 3 35 4 Pressure (bar) R3, M=5. NH3, M=17.3 R1, M=1.9 R134a, M=1 R, M=86.47 R6a, M=58.1 Cooper s pool boiling correlation is plotted versus saturated pressure for different fluids: (for saturated temp. between -6 C) As can been seen heat transfer coefficient generally increases with increasing pressure and decreasing the molecular weights. 45 4 35 NH3, M=17.3 3 R3, M=5. 5 R6a, M=58.1 R134a, M=1 15 R1, M=1.9 1 R, M=86.47 5 R11, M=137.4 5 1 15 5 3 35 4 Ps (bar) h-cooper (W/m² K) 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. CHF (W) 4 1 18 15 1 9 6 3.1..3.4.5.6.7 Reduced pressure R6a, M=58.1 R11, M=137.4 NH3, M=17.3 R134a, M=1 R1, M=1.9 R, M=86.47 R3, M=5. 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 R6a (Isobutane) Tests were done at five reduced pressures ; p p r = p ;.,.5,.1,. and.3. cr Two types of evaporators: smooth and threaded tube surfaces. 8
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 q (W/m²) 35 3 pr=.3 pr=. 5 15 1 Isobutane 5 Smooth tube 5 1 15 5 DT ( C) 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 DT ( C) 4 pr=.3 pr=. 18 pr=.1 16 pr=.5 14 pr=. 1 1 8 6 4 4 6 8 1 1 Q (W) The Fig. shows, heat transfer coeff. vs. reduced pressure for 11 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 (pr m ), in which m is generally between.-.35. In the present case, m=.317, correlates the experimental data well for the smooth tube with Isobutane as refrigerant. h (W/m².K) 45 4 35 3 5 15 1 5 h = constant pr.317 R =.9957 Q=11 W.5.1.15..5.3.35 pr Effect of threaded surface at different reduced pressure on heat transfer coefficient The fig. shows temp. diff. vs. reduced pressure from 1 to 11 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. to.3. DT (C ) 1 1 W 9 3 W 8 5 W 7 7 W 6 9 W 5 4 11 W 3 1.1..3.4 pr Effect of heat flux on heat transfer coefficient Figur shows the relation between heat transfer coefficient and heat flux for Pr=.1, with smooth tube. The dependence of heat transfer coefficient on heat flux can be expressed as h=f (q n ), n, in most cases varies between.6-.8 Presented data follows h=f (q.57 ) h (kw/m².k) y =.8761x.5755 5 R6a R =.9984 15 1 5 h=f (q n ) h=f (q.57 ) 4 8 1 16 4 8 q (kw/m²) 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 5 percent. h-cooper (W/m ² K) 5 4 3 1 Q=1 W Q=3 W Q=5 W Q=7 W Q=9 W Q=11 W 5% 5% 1 3 4 5 h-exp (W/m² K) 9
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 5 percent. h- LW( W/m ² K) 5 4 3 1 1 W 3 W 5 W 7 W 9 W 11 W 5% 5% 1 3 4 5 h-exp (W/m² K) Influence of diameter Testing condition R6a as refrigerant Tests were done with 7, 5,4, 3, and 1 vertical channels with diameter of 1.1, 1.5,1.9,.5 3.5 and 6 mm. Smooth surface At reduced pressure.1 (p/p cr ) Influence of diameter Conclusions 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. h-exp. (kw/m² K) 3 5 d=6 mm d=3.5 mm d=.5 mm 15 d=1.9 mm 1 1.5 m m 5 d =1.1m m 5 1 15 5 3 35 Heat flux (kw/m ²) 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 (pr m ) m=.317 h=f (q n ) where n=.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.3 than. 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. 1
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: 16 G² L Δpl = Re d ρl For the turbulent flow regime, the Blasius correlation for the friction factor can used:.5 G² L Δp =.79 Re l d ρ l The pressure loss around bends can be calculated by: G² Δp = ξ lb ρ where ξ 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 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. 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. 11
Homogeneous flow model 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. 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 gravitationaland the frictional pressure drop. The most used correlations for calculation of frictional pressure drop are: 1. Lockhart-Martinelli correlation. CESNEF- 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: 1 x 1 x = + ρh ρ g ρ L Several different correlations have been proposed for estimation of two-phase viscosity, such as: μ Cicchitti et al. h = x μg + ( 1 x) μl 1 x 1 x = + Beattie- Whalley μ μ μ h g L μ h = μ L ( 1 β ) (1 +.5 β ) + μ β x ρh g McAdams et al. β = ρ g μ g x ρ h μ L ( 1 x) ρh μ Dukler et al. h = + ρ ρ g L Gravitational pressure drop Δp G, R = ρ m g H r The gravitational or head pressure change at the riser 1 α = The momentum equation gives: ug (1 x) ρ g ρ m = α ρ g + (1 α) ρ L 1+ Where α is void fraction ul x ρ L A: total cross-section area (m A g ) α = A For the homogeneous flow the phase velocities are equal, A g : average cross-section area occupied by the gas phase (m ) u L =u g, ug S =, where S is the slip ratio. Void fraction can be calculated by: ul 1. Homogeneous model 1. Zivi model [1963] α h = (1 x) ρ g 1+ 3. Turner& Wallis two-cylinder model [1965] x ρ L 4. Lockhart-Martinelli correlation [1949] 5. Thom correlation [1964] 6. Baroczy correlation [1963] 1
Fig. 1 Acceleration pressure drop Experimental setup glass tube 77 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 ( v g v L ) v specific volume x 15 55 939 Downcomer 8 Abs. pressure transduc er 186 ID=6.1 mm 974 116 Not to scale 1 15 1 C Evaporator B 95 5 hål hole med with d_f=1.5 mm All dimensions in the figure are in mm CHF Testing condition R6a (Isobutane) Tests were done at three reduced pressures;.35,.1, and.. 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 =. the CHF is 69 W which correspond to 3 W/cm² front area of the component which correspond to 65 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) 35 3 pr=.35 pr=.1 5 15 pr=. 1 5.35 smooth channel.1. 35 4 45 5 55 6 65 7 75 Q to t (W) 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. m_dot (kg/s).6 pr=.35.5 pr=.1 pr=..4.3. smooth channel.1 1 3 4 5 6 7 Q cr i (W) 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. x 1 pr=.35.9.8 pr=.1.7 pr=..6.5.4.3. smooth channel.1 1 3 4 5 6 7 Qcri (W) 13
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. Qcri (W) 7 6 5 4 3 threaded 1 smooth.5.1.15..5 pr 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. Q_cri_pb. (W) 7 15% 6 5-15% 4 3 1 1 3 4 5 6 7 Q_cri_exp. (W) Old Exam Problem 3-3-7 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 4 cm² and 4 cm² respectively. The total pressure drop in the rising tube can be calculated using Δp Riser = 6.1 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 1-4 Pa s. The temperature of the evaporator walls is 5 C, the boiling heat transfer coefficient is. W/(m² K), and the heat dissipation is 6 W. Calculate the mass flow rate, m&, the change in vapor quality, Δx, and the saturation temperature of the refrigerant (6 credits). 14