Slug Pump Heat Pipe

Abstract

The slug pump heat pipe allows for passive heat transport of an enclosed two-phase (liquid/vapor) system under multiple orientations with respect to gravity or under an inertial force, such as a centrifugal force. While the fluid flow is driven by a gravitational or inertial force (e.g. centrifugal force) the device enables condensation heat transfer below the evaporator, with respect to such a body force. Since condensation heat transfer can be achieved below the evaporator, the condenser's effective area is nearly uniform under various orientations.

Description

PRIORITY STATEMENT UNDER 35 U.S.C. §119

This application claims priority under 35 U.S.C. §119 based upon prior U.S. Provisional Patent Application Ser. No. 61/984,916, filed Apr. 28, 2014, in the name of Jeremy Rice, entitled “SLUG PUMP HEAT PIPE,” the disclosure of which is incorporated herein in its entirety by this reference.

BACKGROUND OF THE INVENTION

Passive heat transfer devices are of much interest in applications such as electronics cooling. These devices are generically called heat pipes and thermosyphons. What differentiates many of these devices is the method in which the condensate is returned to the evaporator.

A capillary wick heat pipe is a tube with a wick structure bonded to the inner diameter of the tube and a hollow core. Heat can be input at any location along the pipe, and anywhere heat is applied is called an evaporator. In the evaporator, the liquid that saturates the wick is vaporized and flows down the hollow core. The sections along the heat pipe where heat is removed is called the condenser. In the condenser, vapor from the core condenses on the wick. Liquid is pumped by capillary action from the condenser to the evaporator. The most common operating limit of a heat pipe is the capillary dry-out limit. In this limit, the capillary pressure is insufficient to pump the liquid through the wick from the condenser to the evaporator, resulting in a dry wick (no liquid) in the evaporator. Increased transport distance limits the heat load an individual heat pipe can tolerate. Pumping against gravity also limits the transport length.

A looped heat pipe operates similarly to a conventional heat pipe in that it is driven by capillarity. The main difference is that there is only a wick in the evaporator, where the vapor flows to the condenser in one tube while the liquid is returned to the evaporator in another tube. Since the liquid flows in a tube, not a wick, the resistance to the liquid flow is reduced significantly, compared to a capillary wicked heat pipe, thus increasing the transport distance. The major drawback of a looped heat pipe is that vapor plugs in the liquid line that pass through the wick in the evaporator can make the unit fail to function, since the wick needs to be saturated with liquid. Special consideration needs to be taken to ensure these conditions don't arise.

A thermosyphon is another two phase device, where the condensate is returned from the condenser to the evaporator by gravity. Vapor flows upwards, against the pull of gravity, through the center of a relatively large diameter tube while liquid condensate flows downwards along the tube walls. The pool of liquid that accumulates in the bottom of the tube may be boiled to continue the process. The condensation process may only happen above the evaporator, therefore these units have a significant orientation dependence. When orientation is favorable, these units can be used to transfer several watts to several kilowatts or more.

A looped thermosyphon is similar to a thermosyphon in that gravity returns condensate to the evaporator, however, the liquid is returned to the evaporator via a distinct tube. Vapor is supplied to the condenser through a separate tube as well. There is liquid build up in or immediately following the condenser in these devices. The difference in this liquid build up height, to the liquid height in the evaporator drives the fluid flow. For a low impedance system, this liquid build up may be as low as 5 mm, but can also be several meters or more. The major drawback of this system is the same as a thermosyphon, in that condensation can only occur above the evaporator.

A bubble pump is a two-phase device, and consists of an evaporator and a condenser with tubes connecting the two devices. During the evaporation/boiling process, vapor and liquid are supplied to the condenser/radiator. Vapor bubbles generated in the evaporator drive liquid slugs upwards, with respect to gravity, towards a condenser/radiator. A bubble pump operates by latent heat transfer as well as sensible heat transfer, since the liquid flow rates induced by the bubble pump are much greater than the liquid flow rates produced by condensing vapor alone. While driven by gravity, these systems can achieve better heat transfer than a thermosyphon below the evaporator, since the sensible heat transfer is more capable of moving heat, with the increased liquid flow rates. The heat transfer is dominated by sensible heat transfer below the evaporator in a bubble pump device. The drawbacks of such a device are that common working fluids, such as hydrocarbons, hydrofluorocarbons, hydrofluoroethers, and fluoroketones, have poor thermal conductivities, which limit their performance in sensible heat transfer applications. Additionally, the liquid flow rates of these devices are also limited, which further limits the sensible heat transfer capabilities.

A pulsating heat pipe consists of a serpentine capillary tube. The tube is structured to have several parallel flow paths between the evaporator and the condenser. The flow is a capillary liquid slug and vapor plug flow. Since the length and position of these slugs and plugs is not the same in each channel, the expansion of the vapor slugs in the evaporator and contraction in the condenser causes the fluid to oscillate in the tubes. The operation of the pulsating heat pipe depends on a perpetually unstable thermodynamic condition. These devices can operate against gravity, and heat can travel a relatively large distance (several meters); however, there is contradictory evidence to the limitations and characteristics of such devices, such as heat limits, and temperature differentials required for operation.

SUMMARY OF THE INVENTION

The slug pump heat pipe addresses limitations in gravity or inertial driven two-phase heat transfer systems, such as looped thermosyphons. In these systems, liquid and vapor are stratified in the bottom of the condenser (or collector) with an interface height above the evaporator. This condition does not allow for condensation heat transfer below the evaporator, thus limiting the application base where the technology can be applied. The slug pump heat pipe allows for condensation heat transfer below the evaporator. This feature is enabled by the use of many parallel capillary channels in the condenser which promote the formation of discrete liquid slugs and vapor plugs along the length of the channel. When these slugs form above the evaporator, it enables the existence of vapor plugs below the evaporator, thus enabling condensation heat transfer. In many space constrained applications, such as electronics cooling, having effective condenser heat transfer both above and below the evaporator will aid in the overall heat removal performance. Additionally, the slug pump heat pipe can enable reversible heat transfer, where the evaporator and condenser may swap functionality.

BRIEF DESCRIPTION OF THE DRAWINGS

For a more complete understanding of the present invention and the advantages thereof, reference is now made to the following description taken in conjunction with the accompanying drawings in which like reference numerals indicate like features and wherein:

FIG. 1 is a schematic drawing of a looped thermosyphon in accordance with prior art;

FIG. 2 is a schematic drawing of one embodiment of a slug pump heat pipe of the present invention;

FIG. 3 is a schematic drawing of the equivalent pressure head of a series of slugs in a capillary tube relative to a large diameter tube;

FIG. 4 is a schematic drawing of a second embodiment of a slug pump heat pipe in which the evaporator and condenser are configured to switch functionality;

FIG. 5 is a schematic drawing of one embodiment of a slug pump heat pipe for use in an electronics system from the front view and top view;

FIG. 6 is a schematic drawing of one embodiment of a condenser with numerous parallel capillary channels;

FIG. 7 is a schematic drawing of one embodiment of an evaporator with numerous parallel capillary channels; and

FIG. 8 is a schematic drawing of one embodiment of a slug pump heat pipe air-to-air heat exchanger.

DETAILED DESCRIPTION

Various embodiments of the slug pump heat pipe of the present invention allow for the passive heat transport of a two-phase (liquid/vapor) flow under multiple orientations by utilizing pressure head built up in the condenser by recurring liquid slugs and vapor plugs along the length of capillary channels under gravitational forces. In order to have capillary slug flow in the condenser, the channel diameter or width must be less than the critical dimension defined as follows:

$\begin{array}{cc}{D}_{\mathrm{crit}}<2\sqrt{\frac{\sigma }{g\left({\rho }_{\mathrm{liq}}-{\rho }_{\mathrm{vap}}\right)}}& \mathrm{EQ}.1\end{array}$

The surface tension is denoted by σ, the density by ρ, and gravity by g. When the channel size, D, is less than the critical value, the meniscus at the liquid/vapor interface is stable, and its integrity is maintained, thus separating the liquid and vapor phases into discrete slugs along the length of the channel. The approximate critical channel diameter for water at 30° C. is 5.4 mm, while the critical diameter for R134a at 30° C. is 1.6 mm. The critical diameter may change due to variances in cross-sectional shape of the capillary channels, for example a round channel versus a square channel versus a triangular channel. Also, high fluid velocity may make the interface unstable, so smaller channels may be necessary. Since the critical diameter is small, it is necessary to have many parallel channels along the length of the condenser, which increases the total cross-sectional area for fluid to travel with less resistance, thus reducing the pressure loss. The high number of channels also increases the total surface area for heat transfer to occur.

As shown in FIG. 1, the liquid 102, with some intermittent vapor bubbles 103, flows from the condenser 105 to the evaporator 104 through a first tube connecting the two components. Heat 100 is applied to the evaporator 104 and the liquid 102 is vaporized. Vapor 103 flows through a second tube connecting the evaporator 104 to the condenser 105 where heat is rejected 101, closing the loop. The flow in the tube filled with vapor 103 will have some liquid 102 in it. The flow in this tube may be annular flow, with vapor 103 in the core and liquid 102 on the periphery, or may be stratified with liquid 102 on the bottom and vapor 103 on the top, depending on the mass flux of each phase and the tube diameter.

At the surface, a looped thermosyphon appears to be a similar device. It can also achieve operation under multiple orientations, however, due to the stratification of the liquid and vapor phases in the condenser, much of the condenser length below the evaporator is ineffective. The poor condenser performance below the evaporator is the result of sub-cooling a liquid (below the saturation temperature) under the low liquid flow rates in the system. FIG. 1 demonstrates this phenomena, where the evaporator 104 sits mid-way up the condenser 105, with respect to gravity 106. By comparison, a slug pump heat pipe is presented in FIG. 2, in which the condenser 105 and evaporator 104 sit in the same relative position as with the looped thermosyphon. In the slug pump heat pipe, the liquid slugs 107 start forming near the top of the condenser 105, due to the surface tension effects in the small diameter channels. The formation of the these liquid slugs 107 allow for vapor plugs 108 to exist below the evaporator 104, while still maintaining a net pressure head to drive the condensate back to the evaporator 104. Wherever a vapor plug 108 comes into contact with a capillary channel wall, condensation will occur, enabling good heat transfer attributes. The length of the vapor plugs 108 will decrease along the length of the condenser 105, as the specific volume of the vapor 103 is much greater than that of liquid 102. Towards the bottom of the condenser, the vapor plugs 108 may shrink into bubbles once there isn't enough volume to fill up the cross-section of the channels. These bubbles may condense all the way or may flow back to the evaporator 104. In the latter case, the bubble will aide in the flow, and will act like a bubble pump below the evaporator 104. The existence of vapor bubbles at the bottom of the condenser 105 will depend on many factors, including total heat flow, channel diameter, length of condenser, as well as other factors.

An illustration of the equivalent hydrostatic pressure build up in a capillary tube with liquid slugs 107 compared to a large tube with stratified phases is presented in FIG. 3. The equivalent liquid height, h, 109 of the capillary tube may be obtained by taking an integral along the centerline 110 of the capillary tube as follows:

∫Hdy=h H dy=h  EQ. 2

Δp=└(ρ_liq−ρ_vap)gh┘_condenser−└(ρ_liq−ρ_vap)gh┘_evaporator  EQ. 3

Where h is the equivalent liquid height, and H is a Heaviside step function that is 0 in the vapor phase and 1 in the liquid phase. In practical terms, the pressure head along the capillary tube is approximately equivalent to the total length of the liquid slugs 107 along the center line. The total pressure differential of the slug pump heat pipe, Δp, is the pressure head difference on the condenser side and the evaporator side, as can be calculated by EQ 3. If the frictional losses in the system are less than this pressure differential, then the system will function, if not, dry-out conditions will occur. The charge of the working fluid is also critical in these systems. An approximate starting point to determine the charge is to calculate the volume inside the tubes, evaporator 104, condenser 105 and any reservoirs or headers. The volume will be occupied by 30-70% liquid and the rest vapor. Experimentation will usually be required to fine tune the charge amount for a particular application.

An alternative embodiment is represented in FIG. 4 where both the evaporator 104 and condenser 105 both contain several parallel capillary tubes. In this embodiment, the vapor plugs 108 will be longer in the evaporator 104 and the liquid slugs 107 will be longer in the condenser 105. This difference creates the pressure head necessary for the circulation of liquid 102 and vapor 103. Since the evaporator 104 and condenser 105 both have many parallel capillary channels, the direction of heat flow may be reversed, and the condenser 105 may have heat applied to it instead of rejected from it, thus making it an evaporator. The reverse is true for the evaporator 104.

The general components of a slug pump heat pipe and how they may be designed for a lower profile electronics system are presented in FIG. 5. This embodiment has two evaporators 104 connected to a common condenser 105, by a tube 111 supplying liquid 102 to the evaporators and a second tube 112 returning vapor 103 to the condenser 105. The evaporators 104 make contact with an electronics component, such as a CPU (not shown), via a thermal interface material and an external force that creates a pressure between the two devices. Since the condenser 105 has many capillary channels, a high degree of functionality may be achieved when the slug pump heat pipe is orientated in many ways with respect to gravity 106.

For instance, gravity 106 may point down or up, as seen in FIG. 5, in which case the evaporators 104 sit mid-way up the condenser 105. In this scenario, the formation of liquid slugs 107 in the capillary channels inside the condenser 105 help promote condensation below the evaporator 104. If the same design were to be operated in an orientation where the condenser 105 was above the evaporators 104 with respect to gravity (gravitational vector pointed to the left), the unit can operate at nearly the same performance. In this orientation, liquid slugs 107 will still form in the condenser 105 as a result of the capillary channels, however, they do not aid in the pressure head build up since they are perpendicular to the gravitational force 106. Finally, the unit can operate with gravity 106 oriented towards the z axis. In this scenario, the condenser 105 must be above the evaporators 104, with respect to gravity 106, in order to drive liquid back to the evaporators 104. Similar to the prior orientation, the liquid slugs 107 do not aid in the pressure head build up since the capillary channels are normal to gravity 106. Even when the capillary channels don't aid in the buildup of liquid pressure head, they do aid in the overall heat transfer characteristics of the condenser 105, since there is a significant increase in overall surface area inside the condenser 105.

Although not intended to be limiting, the size of the condenser 105 is approximately 1.5″ tall by 9″ wide and 1.5″ deep in the embodiment presented in FIG. 5, the evaporators 104 are approximately 1.5″×1.5″×0.25″, and the evaporators 104 are located approximately 4 and 8 inches from the front of the condenser 105. It may be desirable to size the tube diameter for the horizontal implementation (gravity pointed into the z axis), since the pressure head to drive the fluid through the system is approximately 10 mm, which is the lowest pressure head of all the viable orientations.

For commercial electronics applications (0-85° C. operating range), the condenser 105 may be made out of aluminum or other suitable materials, and may be brazed together through a controlled atmosphere brazing or vacuum brazing process. The evaporators 104 may be made out of copper, aluminum or other suitable materials. The transport tubes 111, 112 may be made from the same material as the evaporators 104. The tubes 111, 112 connecting the evaporators 104 to the condenser 105 may be flame brazed, when the materials are different (e.g., aluminum or copper) or the same (e.g., aluminum). In the case where the evaporator 104 and the condenser 105 are made from the same base material (e.g., aluminum), the entire assembly may be brazed as a single unit. Working fluids that are suitable for this design are hydrofluorocarbons, hydrofluoroethers, hydrofluoroolefins, hydrocarbons, water or ammonia among others. The tube sizing depends on the operating temperatures, fluid selection and maximum power that is desired to be supported, since these parameters impact the fluid velocities and thus the hydrodynamic losses within the system.

Inside the condenser 105 there are many capillary channels which, in some embodiments, are in a parallel configuration. This is necessary since the hydrodynamic loss inside a single channel will be relatively large compared to a larger channel, which will likely lead to a severe limit in the maximum supported channels. An extruded tube or other suitable alternative may be used to construct these capillary channels. In the case where there needs to be several rows of these channels, limitations of extrusion technology may prevent this from happening. In this case, the channels can be constructed by alternate methods. One example of an alternate method for producing these channels is depicted in FIG. 6. The capillary channels in this embodiment are formed from a formed sheet of metal 113 that has a wavy pattern, and a flat piece of sheet metal 114. In this instance, there are multiple rows of the wavy 113 and flat 114 pieces which can be brazed together. A hermetic seal is created around the stacked pieces by a top piece 115 and a bottom piece 116. The seal is created by a lap joint 117 that may be brazed together, as well as to the internal pieces. The collection of the wavy 113, flat 114, top 115 and bottom 116 may be referred to as the condenser core, since these parts represent where the condensation actually occurs. The bottom piece 116 of the condenser core is brazed to a fin stack 118 in which air may pass through, allowed the heat released by the condensation process to be absorbed by the air passing through. Alternative embodiments may have a series of condenser cores and fin stacks 118 to limit the reliance on thermal conduction within the fins 118 when the condenser size increases.

In the evaporator, it may also be necessary to form capillary channels, to ensure liquid covers much of the fins. One embodiment is presented in FIG. 7 where the same basic components as are used in the condenser are used, such as a wavy fin 113, a flat fin 114, a top cover 115, and a bottom cover 116. In this design, the bottom cover 116 will come into contact with a heat generating component, such as a CPU, via thermal interface material. The fins in the evaporator protrude perpendicularly to the bottom 116 and top 115 covers, which differs from the condenser, where the formed material is generally parallel to these pieces. The designer may choose one of these configurations or an alternate configuration to form the capillary channels. A single row of rectangular or square capillary channels may also be used with a stacked or folded fin.

The present invention may be utilized in air to air heat transfer applications, as presented in FIG. 8. In the evaporator 104, hot air may be cooled as is passes through, as the heat vaporizes the liquid inside of it. In the condenser 105, cool air may pass through and increase in temperature as it passes through, as the vapor condenses. The condenser core 119 consists of many parallel capillary channels and conducts to a stack of fins 118. The evaporator may be of identical construction as the condenser. In this case, the warm and cool air streams may be reversed, thus converting the functionality of the evaporator into the condenser and vice versa. In order for this process to be reversible, gravity 106 must be pointed down or up as seen on the page. When gravity 106 is pointed down, the capillary channels help promote two-phase heat transfer in both the condenser 105 and evaporator 104 along the entire vertical length of each device, thus allowing for a high percentage of the overall cross-sectional area to be utilized for heat transfer. If gravity is pointed towards the left, the evaporator and condenser will still function, but the system will lose the ability to reverse the direction of the heat flow.

Gravity 106 has been described as the driving force for the slug pump heat pipe to operate. A centrifugal force may replace gravity as the driving force as well, and the slug pump heat pipe may also operate, as long as many capillary channels are utilized in the condenser.

While the present system and method has been disclosed according to the preferred embodiment of the invention, those of ordinary skill in the art will understand that other embodiments have also been enabled. Even though the foregoing discussion has focused on particular embodiments, it is understood that other configurations are contemplated. In particular, even though the expressions “in one embodiment” or “in another embodiment” are used herein, these phrases are meant to generally reference embodiment possibilities and are not intended to limit the invention to those particular embodiment configurations. These terms may reference the same or different embodiments, and unless indicated otherwise, are combinable into aggregate embodiments. The terms “a”, “an” and “the” mean “one or more” unless expressly specified otherwise. The term “connected” means “communicatively connected” unless otherwise defined.

When a single embodiment is described herein, it will be readily apparent that more than one embodiment may be used in place of a single embodiment. Similarly, where more than one embodiment is described herein, it will be readily apparent that a single embodiment may be substituted for that one device.

In light of the wide variety of passive heat transfer devices known in the art, the detailed embodiments are intended to be illustrative only and should not be taken as limiting the scope of the invention. Rather, what is claimed as the invention is all such modifications as may come within the spirit and scope of the following claims and equivalents thereto.

None of the description in this specification should be read as implying that any particular element, step or function is an essential element which must be included in the claim scope. The scope of the patented subject matter is defined only by the allowed claims and their equivalents. Unless explicitly recited, other aspects of the present invention as described in this specification do not limit the scope of the claims.

Claims

1. A closed loop heat transfer apparatus comprising:

an evaporator;

a condenser consisting of a plurality of parallel capillary channels;

a first tube connecting the condenser to the evaporator primarily transporting liquid therebetween through the use of inertial force; and

a second tube connecting the evaporator to the condenser primarily transporting vapor therebetween.

2. The apparatus of claim 1, wherein the plurality of capillary channels are configured from an extruded piece of metal.

3. The apparatus of claim 1, wherein the plurality of capillary channels are configured from alternating wavy formed and flat sheet metal.

4. The apparatus of claim 1, wherein the condenser core comprises multiple rows of capillary channels.

5. The apparatus of claim 1, wherein the evaporator comprises a plurality of capillary channels.

6. The apparatus of claim 1, wherein the evaporator has the same configuration as the condenser.

7. The apparatus of claim 1, wherein the plurality of capillary channels in the evaporator are configured from alternating wavy formed and flat sheet metal.

8. The apparatus of claim 1, wherein the inertial force is gravity.

9. The apparatus of claim 1, wherein the inertial force is a centrifugal force.

10. The apparatus of claim 1, wherein the evaporator is located between 25% and 75% of the bottom to top height of the condenser with respect to the direction of the inertial force.

11. A method for the removal of heat from a device comprising

receiving heat into a liquid in an evaporator resulting in at least a partial vaporization of the liquid;

transporting the vapor through a first tube to a condenser having a plurality of capillary channels, wherein heat is released as the vapor passes through the plurality of capillary channels resulting in at least a partial condensation of the vapor;

transporting the condensate through a second tube through the use of an inertial force to the evaporator, wherein the evaporator, the first tube, the condenser and the second tube form a closed loop heat transfer system.

12. The method of claim 11, wherein the plurality of capillary channels are configured from an extruded piece of metal.

13. The method of claim 11, wherein the plurality of capillary channels are configured from alternating wavy formed and flat sheet metal.

14. The method of claim 11, wherein the condenser core comprises multiple rows of capillary channels.

15. The method of claim 11, wherein the evaporator comprises a plurality of capillary channels.

16. The method of claim 11, wherein the evaporator has the same configuration as the condenser.

17. The method of claim 11, wherein the plurality of capillary channels in the evaporator are configured from alternating wavy formed and flat sheet metal.

18. The method of claim 11, wherein the inertial force is gravity.

19. The method of claim 11, wherein the inertial force is a centrifugal force.

20. The method of claim 11, wherein the evaporator is located between 25% and 75% of the bottom to top height of the condenser with respect to the direction of the inertial force.