Patent Abstract:
A system including a primary evaporator facilitating heat transfer by evaporating liquid to obtain vapor is disclosed. The primary evaporator receives a liquid from a liquid line and outputs the vapor to a vapor line. The primary evaporator also outputs excess liquid received from the liquid line to an excess fluid line. A condensing system receives the vapor from the vapor line, and outputs the liquid and excess liquid to the liquid line. The excess liquid is obtained at least partially from a reservoir. A primary loop includes the condensing system, the primary evaporator, the liquid line, and the vapor line, and provides a heat transfer path. Similarly, a secondary loop includes the condensing system, the primary evaporator, the liquid line, the vapor line, and the excess fluid line. The secondary loop provides a venting path for removing undesired vapor within the liquid or excess liquid from the primary evaporator.

Full Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims priority to U.S. patent application Ser. No. 60/486,467, filed Jul. 14, 2003, and is a continuation-in-part of U.S. patent application Ser. No. 10/602,022, filed Jun. 24, 2003 now U.S. Pat. No. 7,004,240, issued Feb. 28, 2006, which claims priority to U.S. patent application Ser. No. 60/391,006, filed Jun. 24, 2002, and is a continuation-in-part of U.S. patent application Ser. No. 09/896,561, filed Jun. 29, 2001 now U.S. Pat. No. 6,889,754, issued May 10, 2005, which itself claims priority to U.S. patent application Ser. No. 60/215,588, filed Jun. 30, 2000. These applications are herein incorporated by reference in their entirety. 
    
    
     TECHNICAL FIELD 
     This description relates to a system for heat transfer. 
     BACKGROUND 
     Heat transport systems are used to transport heat from one location (the heat source) to another location (the heat sink). Heat transport systems can be used in terrestrial or extraterrestrial applications. For example, heat transport systems may be integrated by satellite equipment that operates within zero- or low-gravity environments. As another example, heat transport systems can be used in electronic equipment, which often requires cooling during operation. 
     Loop Heat Pipes (LHPs) and Capillary Pumped Loops (CPLs) are passive two-phase heat transport systems. Each includes an evaporator thermally coupled to the heat source, a condenser thermally coupled to the heat sink, fluid that flows between the evaporator and the condenser, and a fluid reservoir for expansion of the fluid. The fluid within the heat transport system can be referred to as the working fluid. The evaporator includes a primary wick and a core that includes a fluid flow passage. Heat acquired by the evaporator is transported to and discharged by the condenser. These systems utilize capillary pressure developed in a fine-pored wick within the evaporator to promote circulation of working fluid from the evaporator to the condenser and back to the evaporator. The primary distinguishing characteristic between a LHP and a CPL is the location of the loop&#39;s reservoir, which is used to store excess fluid displaced from the loop during operation. In general, the reservoir of a CPL is located remotely from the evaporator, while the reservoir of a LHP is co-located with the evaporator. 
     SUMMARY 
     According to one general aspect, a system includes a primary evaporator operable to facilitate heat transfer by evaporating received liquid to obtain vapor, the primary evaporator including a first port for receiving the liquid from a liquid line, a second port for outputting the vapor to a vapor line, and a third port for outputting excess liquid received from the liquid line to an excess fluid line. A condensing system is operable to receive the vapor from the vapor line, to condense at least some of the vapor, and to output the liquid to the liquid line. A reservoir is in fluid communication with the condensing system, and the liquid is obtained at least partially from the reservoir. In the system, a primary loop includes the condensing system, the primary evaporator, the liquid line, and the vapor line, the primary loop being operable to provide a heat transfer path, and a secondary loop includes the condensing system, the primary evaporator, the liquid line, the vapor line, and the excess fluid line. The secondary loop is operable to provide a venting path for removing other vapor that is present within the liquid from the primary evaporator. 
     Implementations may include one or more of the following features. For example, the liquid in the primary evaporator and received from the liquid line may include the excess liquid in excess of a liquid amount necessary to maintain saturation of a primary wick within a core of the primary evaporator. In this case, the primary evaporator may include a secondary wick that is operable to perform phase separation of the other vapor from the liquid for output through the excess fluid line. Further, the primary wick and the secondary wick of the primary evaporator may maintain capillary pumping of the liquid, the excess liquid, and the vapor, so as to maintain flow control to and through the primary evaporator. 
     A mechanical pump may be included that is operable to facilitate the heat transfer by actively pumping the liquid for evaporation by the primary evaporator, and for output as the excess liquid flows through the third port to the excess fluid line. In this case, the reservoir may be positioned between an output of the condensing system and an input of the mechanical pump, or the mechanical pump may be positioned between an input of the condensing system and an output of the primary evaporator. 
     A bypass valve may be included in parallel with the mechanical pump, and may be operable to bypass the mechanical pump during a passive pumping operation of the liquid for evaporation by the primary evaporator. The mechanical pump may include a liquid pump that is oriented in series with the liquid line and positioned between the condensing system and the primary evaporator, or a vapor compressor that is oriented in series with the vapor line and positioned between the primary evaporator and the condensing system. 
     A sensor may be included that is operable to communicate a saturation level of a wick of the primary evaporator to the mechanical pump, wherein a pumping pressure delivered by the mechanical pump is adjusted, based on the saturation level, so as to maintain saturation of the wick with the liquid. A liquid bypass valve may be connected between the liquid line and the vapor line and may be operable to maintain constant pump speed operations of the mechanical pump. The primary evaporator may include a primary wick and a secondary wick, compositions of which may comprise metal. 
     A priming system may be included within the secondary loop, and the priming system may include a secondary evaporator coupled to the vapor line, and a secondary reservoir may be in fluid communication with the secondary evaporator and coupled to the primary evaporator by the excess fluid line, wherein the priming system may be operable to provide the liquid to the primary evaporator at least partially from the secondary reservoir. The condensing system may include a first condenser within the primary loop and coupled to the liquid line and to the vapor line, and a second condenser within the secondary loop and coupled to the excess fluid line and to the secondary reservoir. 
     The third port of the primary evaporator may be primarily used to output the excess liquid to the excess fluid line, and the third port may include a subport for outputting the other vapor to a vapor line, such that the vapor line is included within the secondary loop. 
     The liquid line may be coaxial to and contained within the excess fluid line. A second primary evaporator may be connected in parallel with the primary evaporator within the primary loop. A back pressure regulator may be oriented in series with the vapor line and positioned between the primary evaporator and the condensing system, and may be operable to substantially equalize heat load between the primary evaporator and the secondary primary evaporator. In this case, the back pressure regulator may restrict vapor from reaching the condensing system until a vapor space of the primary evaporator and of the second primary evaporator is substantially devoid of liquid. 
     A second primary evaporator may be oriented in series with the primary evaporator within the primary loop. The condensing system may include a plurality of condensers connected in parallel to one another. In this case, liquid outputs may be associated with each of the plurality of condensers and may be operable to output the liquid to the primary evaporator, and condenser regulators may be coupled to the liquid outputs and operable to regulate liquid flow therefrom. 
     A second primary evaporator may be connected with the primary evaporator within the primary loop, and a thermal storage unit may be coupled to the second primary evaporator. A second primary evaporator may be connected with the primary evaporator within the primary loop, and first and second flow controllers may be connected to the primary evaporator and the second primary evaporator, respectively, and may be operable to regulate liquid flow to the primary evaporator and the second primary evaporator, respectively, so as to ensure a substantially equal heat load distribution between the evaporators. 
     A second primary evaporator may be connected with the primary evaporator within the primary loop, and a condensing heat exchanger may be coupled to the second primary evaporator. A spray-cooled evaporator may be coupled to the condensing heat exchanger by way of a mechanical pump. The condensing system may include a body-mounted radiator, or may include a deployable or steerable radiator. 
     According to another general aspect, liquid is evaporated from a primary wick of a primary evaporator to thereby obtain vapor, the vapor is output through a vapor line coupled to the primary evaporator, and the vapor from the vapor line is condensed within a condensing system. The liquid is returned to the primary evaporator through a liquid line coupled to the primary evaporator, where a saturation amount of the liquid is provided so as to maintain a saturation of the primary wick during the evaporating. Excess liquid beyond the saturation amount is provided to the primary evaporator at least partially from a reservoir, through the liquid line, and the excess liquid and other vapor within the primary evaporator is swept to the condensing system. 
     Implementations may include one or more of the following features. For example, in evaporating liquid from the primary wick of the primary evaporator capillary pumping of the liquid, the excess liquid, and the vapor may be maintained, so as to maintain flow control to and through the primary evaporator. 
     Also, in outputting the vapor, the vapor may be output through a first port of the primary evaporator. In returning the liquid and providing excess liquid, the liquid and excess liquid may be returned through a second port of the primary evaporator. In sweeping the excess liquid and undesired vapor, the excess liquid and undesired vapor may be swept from a third port of the primary evaporator. 
     Outputting the vapor may include outputting the vapor through a first port of the primary evaporator. Returning the liquid and providing excess liquid may include returning the liquid and excess liquid through a second port of the primary evaporator, and sweeping the excess liquid and other vapor may include sweeping the excess liquid from a third port of the primary evaporator, and sweeping the other vapor from a fourth port of the primary evaporator. 
     Sweeping the excess liquid and other vapor may include separating the liquid and excess liquid from the other vapor with a secondary wick of the primary evaporator. Providing the excess liquid may include pumping the excess liquid from the reservoir using a mechanical pump. In this case, the mechanical pump may be bypassed using a bypass valve in parallel with the mechanical pump, during a passive pumping operation of the liquid for evaporation by the primary evaporator. 
     Pumping the excess liquid may include pumping the liquid and the excess liquid using a liquid pump that is oriented in series with the liquid line and positioned between the condensing system and the primary evaporator, or may include pumping the vapor to the condensing system using a vapor compressor that is oriented in series with the vapor line and positioned between the primary evaporator and the condensing system. 
     Providing excess liquid may include providing the excess liquid from a priming system in which the reservoir is in fluid communication with a secondary evaporator, where the reservoir may be coupled to the primary evaporator. In this case, condensing the vapor may include condensing the vapor within a first condenser of the condensing system, the first condenser being coupled to the liquid line and to the vapor line, and sweeping the excess liquid and undesired vapor may include condensing undesired vapor within a second condenser of the condensing system, where the second condenser may be coupled to a mixed fluid line and to the reservoir. 
     According to another general aspect, a system includes a heat transfer system including a main evaporator having a core, a primary wick, a secondary wick, a first port, a second port, and a third port, as well as a condenser coupled to the main evaporator by a liquid line and a vapor line. A heat transfer system loop is defined by the condenser, the liquid line, the vapor line, the first port, and the second port. A venting system is configured to remove vapor bubbles from the core of the main evaporator. The venting system includes a pumping system operable to provide excess liquid to the main evaporator beyond a saturation amount of liquid needed for saturating the primary wick, and a reservoir in fluid communication with the pumping system and providing the excess liquid. The vapor bubbles are vented from the core of the main evaporator through the third port, and a venting loop is defined by the condenser, the liquid line, the vapor line, the first port of the main evaporator, and the third port of the main evaporator. 
     Implementations may include one or more of the following features. For example, the pumping system may include a mechanical pump. 
     The primary wick and the secondary wick of the main evaporator may maintain capillary pumping of the liquid, the excess liquid, and the vapor, so as to maintain flow control to and through the primary evaporator. In this case, the pumping system may include a secondary evaporator in fluid communication with the reservoir and coupled to the vapor line. Further, the reservoir may be in fluid communication with the secondary wick of the main evaporator through a mixed fluid line coupled to the third port of the main evaporator. The excess liquid may be substantially removed from the core of the main evaporator through a fourth port of the main evaporator. 
     Other features will be apparent from the description, the drawings, and the claims. 
    
    
     
       DESCRIPTION OF DRAWINGS 
         FIG. 1  is a schematic diagram of a heat transport system. 
         FIG. 2  is a diagram of an implementation of the heat transport system schematically shown by  FIG. 1 . 
         FIG. 3  is a flow chart of a procedure for transporting heat using a heat transport system. 
         FIG. 4  is a graph showing temperature profiles of various components of the heat transport system during the process flow of  FIG. 3 . 
         FIG. 5A  is a diagram of a three-port main evaporator shown within the heat transport system of  FIG. 1 . 
         FIG. 5B  is a cross-sectional view of the main evaporator taken along dashed section line  5 B- 5 B of  FIG. 5A . 
         FIG. 6  is a diagram of a four-port main evaporator that can be integrated into a heat transport system illustrated by  FIG. 1 . 
         FIG. 7  is a schematic diagram of an implementation of a heat transport system. 
         FIGS. 8A ,  8 B,  9 A, and  9 B are perspective views of applications using a heat transport system. 
         FIG. 8C  is a cross-sectional view of a fluid line taken along dashed section line  8 C- 8 C of  FIG. 8A . 
         FIGS. 8D and 9C  are schematic diagrams of the implementations of the heat transport systems of  FIGS. 8A and 9A , respectively. 
         FIG. 10  is a schematic diagram of another implementation of a heat transport system. 
         FIG. 11  is a schematic diagram of an implementation of an actively pumped heat transport system. 
         FIGS. 12-16  are schematics of implementations of the system of  FIG. 11  that demonstrate various examples of thermal management components and features. 
         FIGS. 17A-17E  are examples of mechanical pumps that may be used in the systems of  FIGS. 11-16 . 
         FIGS. 18A-18C  illustrate examples of different evaporator and condenser architectures for use with the systems of  FIGS. 11-16   
         FIG. 19  is a diagram of an example of a condenser flow regulator. 
         FIG. 20  is a diagram of an example of a back pressure regulator. 
         FIGS. 21 and 22  are diagrams of evaporator failure isolators. 
         FIGS. 23 and 24  illustrate examples of capillary pressure sensors. 
         FIG. 25  is a pressure drop diagram for a thermal management system. 
     
    
    
     Like reference symbols in the various drawings generally indicate like elements. 
     DETAILED DESCRIPTION 
     As discussed above, in a loop heat pipe (LHP), the reservoir is co-located with the evaporator, the reservoir is thermally and hydraulically connected with the evaporator through a heat-pipe-like conduit. In this way, liquid from the reservoir can be pumped to the evaporator, thus ensuring that the primary wick of the evaporator is sufficiently wetted or “primed” during start-up. Additionally, the design of the LHP reduces depletion of liquid from the primary wick of the evaporator during steady-state or transient operation of the evaporator within a heat transport system. Moreover, vapor and/or bubbles of non-condensable gas (NCG bubbles) vent from a core of the evaporator through the heat-pipe-like conduit into the reservoir. 
     Conventional LHPs require liquid to be present in the reservoir prior to start-up, that is, application of power to the evaporator of the LHP. However, liquid will not be present in the reservoir prior to start-up if, prior to start-up of the LHP, the working fluid in the LHP is in a supercritical state in which a temperature of the LHP is above the critical temperature of the working fluid. The critical temperature of a fluid is the highest temperature at which the fluid can exhibit a liquid-vapor equilibrium. For example, the LHP may be in a supercritical state if the working fluid is a cryogenic fluid, that is, a fluid having a boiling point below −150° C., or if the working fluid is a sub-ambient fluid, that is, a fluid having a boiling point below the temperature of the environment in which the LHP is operating. 
     Conventional LHPs also require liquid returning to the evaporator to be subcooled, that is, cooled to a temperature that is lower than the boiling point of the working fluid. Such a constraint makes it impractical to operate LHPs at a sub-ambient temperature. For example, if the working fluid is a cryogenic fluid, the LHP is likely operating in an environment having a temperature greater than the boiling point of the fluid. 
     Referring to  FIG. 1 , a heat transport system  100  is designed to overcome limitations of conventional LHPs, which may include those noted above. The heat transport system  100  includes a heat transfer system  105  and a priming system  110 . The priming system  110  is configured to convert fluid within the heat transfer system  105  into a liquid, thus priming the heat transfer system  105 . As used in this description, the term “fluid” is a generic term that refers to a substance that may be both a liquid and a vapor in saturated equilibrium. 
     The heat transfer system  105  includes a main evaporator  115 , and a condenser  120  coupled to the main evaporator  115  by a liquid line  125  and a vapor line  130 . The condenser  120  is in thermal communication with a heat sink  165 , and the main evaporator  115  is in thermal communication with a heat source Q in    116 . The heat transfer system  105  also may include a hot reservoir  147  coupled to the vapor line  130  for additional pressure containment, as needed. In particular, the hot reservoir  147  increases the volume of the heat transport system  100 . If the working fluid is at a temperature above its critical temperature, that is, the highest temperature at which the working fluid can exhibit liquid-vapor equilibrium, its pressure is proportional to the mass in the heat transport system  100  (the charge) and inversely proportional to the volume of the system. Increasing the volume with the hot reservoir  147  lowers the fill pressure. 
     The main evaporator  115  includes a container  117  that houses a primary wick  140  within which a core  135  is defined. The main evaporator  115  includes a bayonet tube  142  and a secondary wick  145  within the core  135 . The bayonet tube  142 , the primary wick  140 , and the secondary wick  145  define a liquid passage  143 , a first vapor passage  144 , and a second vapor passage  146 . The secondary wick  145  provides phase control, that is, liquid/vapor separation in the core  135 , as discussed in U.S. application Ser. No. 09/896,561, filed Jun. 29, 2001, now U.S. Pat. No. 6,889,754, issued May 10, 2005, which is incorporated herein by reference in its entirety. As shown, the main evaporator  115  has three ports, a liquid inlet  137  into the liquid passage  143 , a vapor outlet  132  into the vapor line  130  from the second vapor passage  146 , and a fluid outlet  139  from the liquid passage  143  (and possibly the first vapor passage  144 , as discussed below). Further details on the structure of a three-port evaporator are discussed below with respect to  FIGS. 5A and 5B . 
     The priming system  110  includes a secondary or priming evaporator  150  coupled to the vapor line  130  and a reservoir  155  co-located with the secondary evaporator  150 . The reservoir  155  is coupled to the core  135  of the main evaporator  115  by a secondary fluid line  160  and a secondary condenser  122 . The secondary fluid line  160  couples to the fluid outlet  139  of the main evaporator  115 . The priming system  110  also includes a controlled heat source Q sp    151  in thermal communication with the secondary evaporator  150 . 
     The secondary evaporator  150  includes a container  152  that houses a primary wick  190  within which a core  185  is defined. The secondary evaporator  150  includes a bayonet tube  153  and a secondary wick  180  that extends from the core  185 , through a conduit  175 , and into the reservoir  155 . The secondary wick  180  provides a capillary link between the reservoir  155  and the secondary evaporator  150 . The bayonet tube  153 , the primary wick  190 , and the secondary wick  180  define a liquid passage  182  coupled to the secondary fluid line  160 , a first vapor passage  181  coupled to the reservoir  155 , and a second vapor passage  183  coupled to the vapor line  130 . The reservoir  155  is thermally and hydraulically coupled to the core  185  of the secondary evaporator  150  through the liquid passage  182 , the secondary wick  180 , and the first vapor passage  181 . Vapor and/or NCG bubbles from the core  185  of the secondary evaporator  150  are swept through the first vapor passage  181  to the reservoir  155  and condensable liquid is returned to the secondary evaporator  150  through the secondary wick  180  from the reservoir  155 . The primary wick  190  hydraulically links liquid within the core  185  of the secondary evaporator  150  to the controlled heat source Q sp    151 , permitting liquid at an outer surface of the primary wick  190  to evaporate and form vapor within the second vapor passage  183  when heat is applied to the secondary evaporator  150 . 
     The reservoir  155  is cold-biased, and thus, it is cooled by a cooling source that will allow it to operate, if unheated, at a temperature that is lower than the temperature at which the heat transfer system  105  operates. In one implementation, the reservoir  155  and the secondary condenser  122  are in thermal communication with the heat sink  165  that is thermally coupled to the condenser  120 . For example, the reservoir  155  can be mounted to the heat sink  165  using a shunt  170 , which may be made of a heat conductive material, such as aluminum. In this way, the temperature of the reservoir  155  tracks the temperature of the condenser  120 . 
       FIG. 2  shows an example of an implementation of the heat transport system  100 . In this implementation, the condensers  120  and  122  are mounted to a cryocooler  200 , which acts as a refrigerator, transferring heat from the condensers  120 ,  122  to the heat sink  165 . Additionally, in the implementation of  FIG. 2 , the lines  125 ,  130 ,  160  are wound to reduce space requirements for the heat transport system  100 . 
     Though not shown in  FIGS. 1 and 2 , elements such as, for example, the reservoir  155  and the main evaporator  115 , may be equipped with temperature sensors that can be used for diagnostic or testing purposes. 
     Referring also to  FIG. 3 , the heat transport system  100  performs a procedure  300  for transporting heat from the heat source Q in    116  and for ensuring that the main evaporator  115  is wetted with liquid prior to startup. The procedure  300  is particularly useful when the heat transfer system  105  is at a supercritical state. Prior to initiation of the procedure  300 , the heat transport system  100  is filled with a working fluid at a particular pressure, referred to as a “fill pressure.” 
     Initially, the reservoir  155  is cold-biased by, for example, mounting the reservoir  155  to the heat sink  165  (step  305 ). The reservoir  155  may be cold-biased to a temperature below the critical temperature of the working fluid, which, as discussed, is the highest temperature at which the working fluid can exhibit liquid-vapor equilibrium. For example, if the fluid is ethane, which has a critical temperature of 33° C., the reservoir  155  is cooled to below 33° C. As the temperature of the reservoir  155  drops below the critical temperature of the working fluid, the reservoir  155  partially fills with a liquid condensate formed by the working fluid. The formation of liquid within the reservoir  155  wets the secondary wick  180  and the primary wick  190  of the secondary evaporator  150  (step  310 ). 
     Meanwhile, power is applied to the priming system  110  by applying heat from the controlled heat source Q sp    151  to the secondary evaporator  150  (step  315 ) to enhance or initiate circulation of fluid within the heat transfer system  105 . Vapor output by the secondary evaporator  150  is pumped through the vapor line  130  and through the condenser  120  (step  320 ) due to capillary pressure at the interface between the primary wick  190  and the second vapor passage  183 . As vapor passes through the condenser  120 , it is converted to liquid (step  325 ). The liquid formed in the condenser  120  is pumped to the main evaporator  115  of the heat transfer system  105  (step  330 ). When the main evaporator  115  is at a higher temperature than the critical temperature of the fluid, the liquid entering the main evaporator  115  evaporates and cools the main evaporator  115 . This process (steps  315 - 330 ) continues, causing the main evaporator  115  to reach a set point temperature (step  335 ), at which point the main evaporator  115  is able to retain liquid and be wetted and to operate as a capillary pump. In one implementation, the set point temperature is the temperature to which the reservoir  155  has been cooled. In another implementation, the set point temperature is a temperature below the critical temperature of the working fluid. In a further implementation, the set point temperature is a temperature above the temperature to which the reservoir  155  has been cooled. 
     Once the set point temperature has been reached (step  335 ), the system  100  operates in a main mode (step  340 ) in which heat from the heat source Q in    116  that is applied to the main evaporator  115  is transferred by the heat transfer system  105 . Specifically, in the main mode, the main evaporator  115  develops capillary pumping to promote circulation of the working fluid through the heat transfer system  105 . Also, in the main mode, the temperature of the reservoir  155  may be reduced below the set point temperature of the main evaporator  115 . The rate at which the heat transfer system  105  cools down during the main mode depends, in part, on the cold-biasing of the reservoir  155  because the temperature of the main evaporator  115  closely follows the temperature of the reservoir  155 . Additionally, though not necessarily, a heater can be used to further control or regulate the temperature of the reservoir  155  during the main mode (step  340 ). Furthermore, in the main mode, the power applied to the secondary evaporator  150  by the controlled heat source Q sp    151  is reduced, thus bringing the heat transfer system  105  down to a normal operating temperature for the fluid. For example, in the main mode, the heat load from the controlled heat source Q sp    151  to the secondary evaporator  150  is kept at a value equal to or in excess of heat conditions, as defined below. In one implementation, the heat load from the heat source Q sp  is kept to about 5 to 10% of the heat load applied to the main evaporator  115  from the heat source Q in    116 . 
     Thus, in the  FIG. 3  implementation, the main mode is triggered by the determination that the set point temperature has been reached at the main evaporator  115  (step  335 ). In other implementations, the main mode may begin at other times or due to other triggers. For example, the main mode may begin after the priming system  110  is wet (step  310 ) or after the reservoir  155  has been cold-biased (step  305 ). 
     At any time during operation, the heat transfer system  105  can experience heat conditions that cause formation of vapor on the liquid side of the evaporator, such as those resulting from heat conduction across the primary wick  140  and parasitic heat applied to the liquid line  125 . Specifically, heat conduction across the primary wick  140  can cause liquid in the core  135  to form vapor bubbles, which, if left within the core  135 , would grow and block off liquid otherwise supplied to the primary wick  140 , thus causing the main evaporator  115  to fail. One such heat condition is caused by parasitic heat input into the liquid line  125  (referred to as “parasitic heat gains”), which causes liquid within the liquid line  125  to form vapor. 
     To reduce the adverse impact of heat conditions such as those discussed above, the priming system  110  operates at a power level Q sp    450  that is greater than or equal to the sum of the head conduction and the parasitic heat gains. As mentioned above, for example, the priming system  110  can operate at 5 to 10% of the power to the heat transfer system  105 . In particular, fluid that includes a combination of vapor bubbles and liquid is swept out of the core  135  for discharge into the secondary fluid line  160  leading to the secondary condenser  122 . In particular, vapor that forms within the core  135  travels along the bayonet tube  143  and directly into the fluid outlet port  139 . Furthermore, vapor that forms within the first vapor passage  144  travels into the fluid outlet port  139  by either traveling through the secondary wick  145  (if the pore size of the secondary wick  145  is large enough to accommodate vapor bubbles) or through an opening (not shown) at an end of the secondary wick  145  near the outlet port  139  that provides a clear passage from the first vapor passage  144  to the outlet port  139 . The secondary condenser  122  condenses the bubbles in the fluid and pushes the fluid to the reservoir  155  for reintroduction into the heat transfer system  105 . 
     Similarly, to reduce parasitic heat input to the liquid line  125 , the secondary fluid line  160  and the liquid line  125  can form a coaxial configuration such that the secondary fluid line  160  surrounds and insulates the liquid line  125  from surrounding heat. This implementation is discussed further below with reference to  FIGS. 8A and 8B . As a consequence of this configuration, it is possible for the surrounding heat to cause vapor bubbles to form in the secondary fluid line  160 , instead of in the liquid line  125 . As discussed, by virtue of capillary action effected at the secondary wick  145 , fluid flows from the main evaporator  115  to the secondary condenser  122 . This fluid flow, and the relatively low temperature of the secondary condenser  122 , causes a sweeping of the vapor bubbles within the secondary fluid line  160  through the condenser  122 , where they are condensed into liquid and pumped into the reservoir  155 . 
     As shown in  FIG. 4 , data from a test run is shown. In this implementation, prior to startup of the main evaporator  115  at time  410 , a temperature  400  of the main evaporator  115  is significantly higher than a temperature  405  of the reservoir  155 , which has been cold-biased to the set point temperature (step  305 ). As the priming system  110  is wetted (step  310 ), power level Q sp    450  is applied to the secondary evaporator  150  (step  315 ) at a time  452 , causing liquid to be pumped to the main evaporator  115  (step  330 ), the temperature  400  of the main evaporator  115  drops until it reaches the temperature  405  of the reservoir  155  at time  410 . Power Q in    460  is applied to the main evaporator  115  at a time  462 , when the heat transport system  100  is operating in LHP mode (step  340 ). As shown, power input Q in    460  to the main evaporator  115  is held relatively low while the main evaporator  115  is cooling down. Also shown are the temperatures  470  and  475 , respectively, of the secondary fluid line  160  and the liquid line  125 . After time  410 , temperatures  470  and  475  track the temperature  400  of the main evaporator  115 . Moreover, a temperature  415  of the secondary evaporator  150  follows closely with the temperature  405  of the reservoir  155  because of the thermal communication between the secondary evaporator  150  and the reservoir  155 . 
     As mentioned, in one implementation, ethane may be used as the fluid in the heat transfer system  105 . Although the critical temperature of ethane is 33° C., for the reasons generally described above, the heat transport system  100  can start up from a supercritical state in which the heat transport system  100  is at a temperature of 70° C. As power Q sp  is applied to the secondary evaporator  150 , the temperatures of the condenser  120  and the reservoir  155  drop rapidly (between times  452  and  410 ). A trim heater can be used to control the temperature of the reservoir  155  and thus the condenser  120  to −10° C. To startup the main evaporator  115  from the supercritical temperature of 70° C., a heat load or power input Q sp  of 10 W is applied to the secondary evaporator  150 . Once the main evaporator  115  is primed, the power input from the controlled heat source Q sp    151  to the secondary evaporator  150  and the power applied to and through the trim heater both may be reduced to bring the temperature of the heat transport system  100  down to a nominal operating temperature of about −50° C. For instance, during the main mode, if a power input Q in  of 40 W is applied to the main evaporator  115 , the power input Q sp  to the secondary evaporator  150  can be reduced to approximately 3 W while operating at −45° C. to mitigate the 3 W lost through heat conditions (as discussed above). As another example, the main evaporator  115  can operate with power input Q in  from about 10 W to about 40 W with 5 W applied to the secondary evaporator  150  and with the temperature  405  of the reservoir  155  at approximately −45° C. 
     Referring to  FIGS. 5A and 5B , in one implementation, the main evaporator  115  is designed as a three-port evaporator  500  (which is the design shown in  FIG. 1 ). Generally, in the three-port evaporator  500 , liquid flows into a liquid inlet  505  and into a core  510 , defined by a primary wick  540 , and fluid from the core  510  flows from a fluid outlet  512  to a cold-biased reservoir (such as reservoir  155 ). The fluid and the core  510  are housed within a container  515  made of, for example, aluminum. In particular, fluid flowing from the liquid inlet  505  into the core  510  flows through a bayonet tube  520 , into a liquid passage  521  that flows through and around the bayonet tube  520 . Fluid can flow through a secondary wick  525  (such as secondary wick  145  of main evaporator  115 ) made of a wick material  530  and an annular artery  535 . The wick material  530  separates the annular artery  535  from a first vapor passage  560 . As power from the heat source Q in    116  is applied to the evaporator  500 , liquid from the core  510  enters a primary wick  540  and evaporates, forming vapor that is free to flow along a second vapor passage  565  that includes one or more vapor grooves  545  and out a vapor outlet  550  into the vapor line  130 . Vapor bubbles that form within first vapor passage  560  of the core  510  are swept out of the core  510  through the first vapor passage  560  and into the fluid outlet  512 . As discussed above, vapor bubbles within the first vapor passage  560  may pass through the secondary wick  525  if the pore size of the secondary wick  525  is large enough to accommodate the vapor bubbles. Alternatively, or additionally, vapor bubbles within the first vapor passage  560  may pass through an opening of the secondary wick  525  formed at any suitable location along the secondary wick  525  to enter the liquid passage  521  or the fluid outlet  512 . 
     Referring to  FIG. 6 , in another implementation, the main evaporator  115  is designed as a four-port evaporator  600 , which is a design described in U.S. application Ser. No. 09/896,561, filed Jun. 29, 2001. Briefly, and with emphasis on aspects that differ from the three-port evaporator configuration, liquid flows into the evaporator  600  through a fluid inlet  605 , through a bayonet  610 , and into a core  615 . The liquid within the core  615  enters a primary wick  620  and evaporates, forming vapor that is free to flow along vapor grooves  625  and out a vapor outlet  630  into the vapor line  130 . A secondary wick  633  within the core  615  separates liquid within the core from vapor or bubbles in the core (that are produced when liquid in the core  615  heats). The liquid carrying bubbles formed within a first fluid passage  635  inside the secondary wick  633  flows out of a fluid outlet  640  and the vapor or bubbles formed within a vapor passage  642  positioned between the secondary wick  633  and the primary wick  620  flow out of a vapor outlet  645 . 
     Referring to  FIG. 7 , a heat transport system  700  is shown in which the main evaporator is a four-port evaporator, such as that illustrated  600  in  FIG. 6 . The system  700  includes one or more heat transfer systems  705  and a priming system  710  configured to convert fluid within the heat transfer systems  705  into a liquid to prime the heat transfer systems  705 . The four-port evaporators  600  are coupled to one or more condensers  715  by a vapor line  720  and a fluid line  725 . The priming system  710  includes a cold-biased reservoir  730  hydraulically and thermally connected to a priming evaporator  735 . 
     Whether using a three-port or four-port evaporator design, design considerations of heat transport systems such as the heat transport systems  100  and  700  may include various advantageous features. For example, with specific reference to elements of the heat transport system  100  (although similar comments may generally apply to the heat transport system  700  of  FIG. 7 , with reference to the corresponding elements as shown therein), such advantages may include startup of the main evaporator  115  from a supercritical state, management of parasitic heat leaks, heat conduction across the primary wick  140 , cold biasing of the cold reservoir  155 , and pressure containment at ambient temperatures that are greater than the critical temperature of the working fluid within the heat transfer system  105 . To accommodate these design considerations, the body or container (such as container  515 ) of the main evaporator  115  or secondary evaporator  150  can be made of extruded 6063 aluminum and the primary wicks  140  and/or  190  can be made of a fine-pored wick. In one implementation, the outer diameter of the main evaporator  115  or secondary evaporator  150  is approximately 0.625 inch and the length of the container is approximately 6 inches. The reservoir  155  may be cold-biased to an end panel of the heat sink  165  using the aluminum shunt  170 . Furthermore, a heater (such as a KAPTON™ heater) can be attached at a side of the reservoir  155 . 
     In one implementation, the vapor line  130  is made with smooth-walled stainless steel tubing having an outer diameter (OD) of 3/16″ and the liquid line  125  and the secondary fluid line  160  are made of smooth-walled stainless steel tubing having an OD of ⅛″. The lines  125 ,  130 ,  160  may be bent in a serpentine route and plated with gold to minimize parasitic heat gains. Additionally, the lines  125 ,  130 ,  160  may be enclosed in a stainless steel box with heaters to simulate a particular environment during testing. The stainless steel box can be insulated with multi-layer insulation (MLI) to minimize heat leaks through panels of the heat sink  165 . 
     In one implementation, the second condenser  122  and the secondary fluid line  160  are made of tubing having an OD of 0.25 inch. The tubing is bonded to the panels of the heat sink  165  using, for example, epoxy. Each panel of the heat sink  165  is an 8×19 inch direct condensation, aluminum radiator that uses a 1/16-inch thick face sheet. KAPTON™ heaters can be attached to the panels of the heat sink  165 , near the secondary condenser  120  to prevent inadvertent freezing of the working fluid. During operation, temperature sensors such as thermocouples can be used to monitor temperatures throughout the heat transport system  100 . 
     The heat transport system  100  may be implemented in any circumstances where the critical temperature of the working fluid of the heat transfer system  105  is below the ambient temperature at which the heat transport system  100  is operating. The heat transport system  100  can be used to cool down components that require cryogenic cooling. Referring to  FIGS. 8A-8D , the heat transport system  100  may be implemented in a miniaturized cryogenic system  800 . In the miniaturized system  800 , the lines  125 ,  130 ,  160  are made of flexible material to permit coil configurations  805 , which save space. The miniaturized system  800  can operate at −238° C. using neon fluid. Power input Q in    116  is approximately 0.3 to 2.5 W. The miniaturized system  800  thermally couples a cryogenic component Q in  (or heat source that requires cryogenic cooling)  816  to a cryogenic cooling source such as a cryocooler  810  coupled to cool the condensers  120 ,  122 . 
     The miniaturized system  800  reduces mass, increases flexibility, and provides thermal switching capability when compared with traditional thermally switchable vibration-isolated systems. Traditional thermally switchable, vibration-isolated systems require two flexible conductive links (FCLs), a cryogenic thermal switch (CTSW), and a conduction bar (CB) that form a loop to transfer heat from the cryogenic component to the cryogenic cooling source. In the miniaturized system  800 , thermal performance is enhanced because the number of mechanical interfaces is reduced. Heat conditions at mechanical interfaces account for a large percentage of heat gains within traditional thermally switchable, vibration-isolated systems. The CB and two FCLs are replaced with the low-mass, flexible, thin-walled tubing used for the coil configurations  805  of the miniaturized system  800 . 
     Moreover, the miniaturized system  800  can function in a wide range of heat transport distances, which permits a configuration in which the cooling source (such as the cryocooler  810 ) is located remotely from the cryogenic component Q in    816 . The coil configurations  805  have a low mass and low surface area, thus reducing parasitic heat gains through the lines  125  and  160 . The configuration of the cooling source  810  within the miniaturized system  800  facilitates integration and packaging of the miniaturized system  800  and reduces vibrations on the cooling source  810 , which becomes particularly important in infrared sensor applications. In one implementation, the miniaturized system  800  was tested using neon, operating at 25 to 40K. 
     Referring to  FIGS. 9A-9C , the heat transport system  100  may be implemented in an adjustable mounted or gimbaled system  1005  in which the main evaporator  115  and a portion of the lines  125 ,  160 , and  130  are mounted to rotate about an elevation axis within a range of ±45° and a portion of the lines  125 ,  160 , and  130  are mounted to rotate about an azimuth axis within a range of ±220°. The lines  125 ,  160 ,  130  are formed from thin-walled tubing and are coiled around each axis of rotation. The system  1005  thermally couples a cryogenic component (or heat source that requires cryogenic cooling)  1016  such as a sensor of a cryogenic telescope to a cryogenic cooling source  1010  such as a cryocooler coupled to cool the condensers  120 ,  122 . The cooling source  1010  is located at a stationary spacecraft  1060 , thus reducing mass at the cryogenic telescope. Motor torque for controlling rotation of the lines  125 ,  160 ,  130 , power requirements of the system  1005 , control requirements for the spacecraft  1060 , and pointing accuracy for the sensor  1016  are improved. The cooling source  1010  and the radiator or heat sink  165  can be moved from the sensor  1016 , reducing vibration within the sensor  1016 . In one implementation, the system  1005  was tested to operate within the range of 70 to 115K when the working fluid is nitrogen. 
     The heat transfer system  105  may be used in medical applications, or in applications where equipment must be cooled to below-ambient temperatures. As another example, the heat transfer system  105  may be used to cool an infrared (IR) sensor that operates at cryogenic temperatures to reduce ambient noise. The heat transfer system  105  may be used to cool a vending machine, which often houses items that preferably are chilled to sub-ambient temperatures. The heat transfer system  105  may be used to cool components such as a display or a hard drive of a computer, such as a laptop computer, handheld computer, or a desktop computer. The heat transfer system  105  can be used to cool one or more components in a transportation device such as an automobile or an airplane. 
     Other implementations are within the scope of the following claims. For example, the secondary condenser  120  and heat sink  165  can be designed as an integral system, such as, a radiator. Similarly, the secondary condenser  122  and heat sink  165  can be formed from a radiator. The heat sink  165  can be a passive heat sink (such as a radiator) or a cryocooler that actively cools the condensers  120 ,  122 . 
     In another implementation, the temperature of the reservoir  155  is controlled using a heater. In a further implementation, the reservoir  155  is heated using parasitic heat. In another implementation, a coaxial ring of insulation is formed and placed between the liquid line  125  and the secondary fluid line  160 , which surrounds the insulation ring. 
       FIG. 10  is a schematic diagram of an implementation of a heat transport system  1000 . In  FIG. 10 , four-port evaporators  600  are arranged in a serial orientation. 
     More particularly, the heat transport system  1000  includes multiple heat transfer systems  1005  and a priming system  1011  configured to convert fluid from within the heat transfer systems  1005  into a liquid capable of priming the heat transfer systems  1005 . The heat transfer systems  1005  each include four-port evaporators  600  that are coupled to one or more condensers  1015  by a vapor line  1020  and a fluid line  1025 . The priming system  1011  includes a cold-biased reservoir  1030  hydraulically and thermally connected to a priming evaporator  1035 . 
     Similarly to the four-port, parallel arrangement shown in  FIG. 7 , and in accordance with the general principles associated with an operation of the heat transport system  100  described above with respect to  FIG. 1 , the heat transport system  1000  is capable of starting the main evaporators  600  from a super critical state, managing parasitic heat leaks, sweeping excess vapor and non-condensable gas bubbles (NCG) from the cores of the main evaporators  600 , and various other features and advantages described herein. 
     Moreover, as illustrated by  FIGS. 7 and 10 , various implementations of heat transport systems may be used in many different operating environments, providing flexibility and a wide scope of use to designers of heat transport systems. For example, arrangements may be optimized to account for, for example, locations and types of heat sources, heat load sharing between the evaporators  600 , a type of fluid used in the system(s), and various other operating parameters. Of course, it should be understood that the parallel and serial evaporator configurations of  FIGS. 7 and 10  also may be implemented using three-port evaporators, such as, for example, the three-port evaporator  500  of  FIGS. 5A and 5B . 
       FIG. 11  is a schematic diagram of an implementation of an actively pumped heat transport system  1100 . In  FIG. 11 , active loop pumping is enabled for the purpose of, for example, supporting improved waste heat rejection and heat transport capability when compared to heat transport systems that rely solely on passive (e.g., capillary) pumping. 
     More particularly, the actively pumped heat transport system  1100  includes multiple heat transfer systems  1105 , having evaporators  600 , and a mechanical pump  1110  that is arranged in series between a condenser  1115  (and a vapor line  1120  feeding the condenser  1115 ) and the evaporators  600 , along a liquid line  1125 . A reservoir  1130  is disposed between the mechanical pump  1110  and the condenser  1115 , where the reservoir  1130  may be used for, for example, managing excess fluid flow, fine temperature control through cold-biasing, and other features and uses as described herein and as are known. 
     The actively pumped heat transport system  1100  including the mechanical pump  1110  shares certain features and advantages with the passive heat transport systems described above with respect to  FIGS. 1-10 . For example, the heat transport system  1100  includes a primary loop including the vapor line  1120  and the liquid line  1125 , as well as secondary loop(s) defined by the secondary fluid outlets  640  and the secondary vapor outlet  645  (where it should be understood that the outlets  640  and  645  may be replaced with the secondary fluid line  160  of  FIG. 1  in a system using the three-port evaporator  500 ). 
     The mechanical pump  1110  thus provides a source of pumping power for moving fluid through the primary loop and/or the secondary loop of the heat transport system  1100 . This pumping power may be used during various operations of the heat transport system  1100 , and may be in addition to, or in the alternative to, other sources of pumping power. 
     For example, the pumping power provided by the mechanical pump  1110  may be used to provide liquid to the evaporators  600  during a start-up operation of the evaporators  600 , perhaps in conjunction with a separate priming system. Such a priming system may include, for example, the priming system  110  of  FIG. 1 , or some other, conventional priming system (not shown). 
     The mechanical pump  1110  also may be used during steady-state operation of the actively pumped heat transport system  1100 , either continuously or intermittently, as needed to maintain a desired operational state of the heat transport system  1100 . For example, the mechanical pump  1110  may be activated during start-up of the heat transport system  1100 , and then may be bypassed or otherwise de-activated during steady-state operation of the heat transport system  1100 , unless and until a secondary pumping source (e.g., passive pumping supplied by capillary pressure) is insufficient to provide adequate heat transfer. In this sense, the heat transport system  1100  may be considered a dual-pumping system, in which mechanical pumping, capillary pumping, or some combination of both, is available on an as-needed basis to an operator or designer of the heat transport system  1100 . In particular, for instance, when the heat transport system  1100  is used to provide heat transfer over relatively large distances (e.g., 10 meters or more), the mechanical pump  1110  may be required to be used continuously to ensure adequate pumping power. 
     As a final example, and as discussed in more detail below, pumping power of the mechanical pump  1110  also may be used to ensure sweeping or venting of vapor bubbles from the cores of the evaporators  600 . As such, a use or extent of the pumping power of the mechanical pump  1110  may be dependent on the extent to which such vapor bubbles exist (or are thought to exist) within the evaporator cores or, similarly, may be dependent on the extent to which conditions for creating such vapor bubbles within the evaporator cores exist within and around the heat transport system  1100 . 
     As just referenced, and as described above in detail, the construction of three- and/or four-port evaporators permit control and management of liquid and vapor phases within the evaporator core(s). Specifically, for example, fluid within the cores  615  of evaporators  600  that includes a combination of liquid and vapor bubbles may be swept out of the cores  615  for discharge into the secondary fluid outlets  640  and vapor outlets  645  (or into the mixed secondary fluid line  160  in a three-port evaporator configuration). 
     As also described above, such mixed-phase fluid within the core  615  may result from various causes. For example, the mixed-phase fluid may result from heat conduction across the primary wick  620  and/or parasitic heat gains through the liquid line  1125  (e.g., when routing the liquid line through a “hot” environment). Whatever the cause of the mixed-phase flow, the heat transport system  1100  (using the mechanical pump  1110 ), and the systems described above (using the priming or secondary evaporators  150 / 710 / 1011  and associated reservoirs), are operable to provide excess liquid to the evaporators  600 , above and beyond the minimum needed to maintain operation of the heat transport system (e.g., an amount needed to maintain saturation of the wicks and associated capillary pumping). 
     As a result, the heat transport system  1100 , and the systems described above, are able to use this excess liquid to vent or sweep the gaseous portion of the mixed-phase flow from the evaporators  600 , using the secondary flow loops that include the secondary fluid/liquid outlets  640 / 645  or the secondary fluid line  160 . In this way, excess vapor enters the secondary loop either through the secondary wick  635  (if feasible for a given pore size of the secondary wick  635 ), or through an opening at an end of the secondary wick near an outlet port for the secondary loop(s), and is returned to the condenser  1115  for condensation and subsequent return through the liquid line  1125  and/or to the reservoir  1130 . 
     In one implementation, an amount of excess liquid provided to the cores of the evaporators  600  is optimized. In this implementation, the amount of excess liquid is sufficient to sweep all of the evaporator cores present in the system, but not substantially more than this amount, since excess fluid in the heat transport system  1100  may affect performance and reliability of the heat transport system  1100 . However, sweeping all of the evaporators  600  may be problematic, particularly, for example, when the evaporators  600  are not powered equally or, in the limiting case, where one of the evaporators  600  receives no heat (or actually acts as a condenser). 
     One technique for optimizing an amount of excess fluid flow to the evaporators  600  includes an appropriate selection of line diameters of the evaporator wicks, and/or for the liquid line  1125  or the vapor line  1120 . By selecting these line diameters appropriately, an amount of excess fluid beyond that required for operation of the evaporators  600  may be reduced or minimized, while still ensuring that the amount of excess fluid is sufficient to completely sweep or vent all of the evaporators  600 . 
     More particularly, in an implementation such as the one just described, such line sizing may be a factor in determining an efficiency of the sweeping of the evaporators  600 . In the case of  FIG. 11 , this sweeping efficiency may determine how much more liquid must be supplied to the evaporators  600  through the liquid line  1125  than what is required to satisfy the heat load(s) of the evaporators  600 . Similarly, in the case of  FIG. 1  or  FIG. 7 , the sweeping efficiency may determine how much power must be applied to the secondary evaporator in excess of what is required to satisfy the heat load of the main evaporators  115  or  600 , respectively. 
     One parameter for describing the appropriate sizing criteria includes a ratio of the flow resistance of the secondary fluid/vaport outlets  640 / 645  (or, in  FIG. 1 , the mixed secondary fluid line  160 ) to a sum of the resistances of the liquid line  1125  ( 125  in  FIG. 1 ) outside of the evaporator  600  and the liquid flow passage in the evaporator core  615  ( 135  in  FIG. 1 ). In general, a relatively large value of this ratio is preferred, and serves to decrease a sweepage power required to completely sweep all evaporator cores. 
     With such complete sweepage being provided, the heat transport system  1100  may use a narrow-diameter, small-pore, metal wick (e.g., 1 micron pore metal wick), which provides high thermal conductivity and increased pumping capability, relative to the polyethylene wicks that often are used in conventional heat transport systems. Such polyethylene wicks may be used despite their reduced pumping capacity, in part due to their relatively wide diameter and large pore size, which tends to reduce their thermal conductivity and, therefore, tends to reduce a presence of vapor within the liquid line  1125  and liquid  615 . 
     In other words, since the structure and function of the heat transport system  1100  enable venting or sweeping of such undesirable vapor from the core  615 , the heat transport system  1100  may not be required to resort to disadvantageous measures to avoid the presence of this vapor in the first place. As a result, the system  110  may enjoy the advantages of narrow-diameter, small-pore, metal wicks, and, in particular, increased pumping against gravity by a factor of ten, relative to polyethylene wicks, for example. Similarly, the heat transport system  1100  may not require subcooled liquid to be returned to the core  615 , such that the liquid line  1125  may be routed through hotter environments than are feasible with conventional systems that do not offer vapor sweepage, as it is described herein. 
     Accordingly, the heat transport system  1100  may provide many advantageous features for the transport and disposal of heat. For example, in addition or as an alternative to one or more of the features just described, the mechanical pump  1110  of the heat transport system  1100  may provide increased flow, increased flow controllability, and increased waste heat transportation and rejection, relative to passive systems (for example, heat transport may occur on the order of 50 kW or more, over a distance of 10 meters or more). As another example, the mechanically pumped heat transport system  1100  may greatly reduce temperature gradients across phased array antennas that may include thousands of elements arranged in complex arrays, thereby reducing an overall size of such arrays and reducing or eliminating the need for separate heat pipes to maintain acceptable element temperatures within the arrays. 
     The heat transport system  1100  offers one or more of the following or other advantages over conventional actively pumped systems as well, including those that have been deployed, for example, in geosynchronous communication satellites. For instance, the two-phase nature of the heat transport system  1100  is beneficial to heat transfer at the thermal interfaces, and reduces required pumping power. Additionally, the sweepage of excess vapor and its complete condensation within the condenser  1115  may reduce an amount of mixed fluid (i.e., two-phase) flow experience by the mechanical pump  1110 . As a result, a lifetime and reliability of the mechanical pump  1110  may be improved, since vapor within a liquid mechanical pump such as the mechanical pump  1110  tends to provide excessive stress within the pump. 
     In addition to some or all of these and other advantages, the heat transport system  1100  is compatible with a wide variety of thermal management components and features. Accordingly,  FIGS. 12-16  are schematics of implementations of the heat transport system  1100  of  FIG. 11  that demonstrate examples of such thermal management components and features. 
     In  FIG. 12 , a system  1200  operates essentially as described above with respect to the heat transport system  1100 . The mechanical pump  1110  is illustrated as a liquid pump  1202  that is in series with a liquid line  1204  that is connected to evaporators  1206 . The evaporators  1206  vent or sweep two-phase fluid flow from their respective liquid cores through a mixed fluid line  1208 , as already described. The evaporators  1206  also output vapor through a vapor line  1210  to a condenser  1212 , which, in  FIG. 12 , includes a body-mounted radiator (discussed in more detail below). 
     The mixed fluid line  1208  is shown as a dashed line in  FIG. 12  to indicate the variety of forms it may take within the system  1200 . For example, the mixed fluid line  1208  may be implemented in a coaxial fashion with respect to the liquid flow line  1204 , as described above with respect to, for example,  FIG. 8C . Such an implementation assists in protecting the liquid line  1204  from parasitic heat effects that may cause vapor and/or NCG bubbles within the liquid line  1204 , and allows the liquid line  1204  to be routed through relatively hot environments without experiencing parasitic heat gain. 
     Further, the mixed fluid line  1208  may be used in conjunction with a secondary evaporator  1214 , which, when used with a (cold-biased) two-phase reservoir  1216  in one of the various manners described above, provides for advantages such as, for example, operation of the system  1200  (or the heat transport system  1100 ) in a passive mode, in which the mechanical pump  1202  (or  1110 ) is bypassed, perhaps using a pump bypass valve  1218 , and the system  1200  (or  1100 ) relies solely on capillary pumping for fluid flow. 
     To the extent that the system  1200  uses fine-pore metal wicks, as described above with respect to  FIG. 11 , its passive pumping capacity in this mode may be improved relative to other passive, capillary-pumped loops. Although the secondary evaporator is shown only conceptually in  FIGS. 12-15 , its use should be apparent based on the above descriptions of secondary evaporators or priming systems  150 ,  710 , and  1011 . Moreover, a particular implementation for using such a secondary evaporator in the context of a mechanically pumped heat transfer system is discussed in detail with respect to  FIG. 16 . 
     As referred to above with respect to  FIG. 11 , the secondary evaporator  1214  is not required for the system  1200  to operate in passive mode. For example, in such a passive mode, a conventional priming system may be used for starting the system  1200  (e.g., for wetting the primary wicks of the evaporators  1206 ). Alternatively, the liquid pump  1202  may be used to prime the evaporator(s)  1206  initially for starting, and/or may be used to maintain saturation of the primary wicks of the evaporators  1206  intermittently thereafter. The choice of which startup method(s) to use, or whether or when to use the system  1200  in a passive mode at all, is, of course, dependent on various operational and environmental factors of the system  1200 , such as, for example, one or more of the type of working fluid, a critical temperature of the working fluid, an ambient operating temperature of the system  1200 , the amount of heat to be dissipated, and various other factors. 
     The above discussion of a general operation of the system  1200  included reference to the evaporators  1206 , similar in structure and function to one or more of the various evaporators discussed herein, and using a cold plate as a heat transfer surface. However, it is a strength of the system  1200  that multiple types and arrangements of evaporators and heat transfer surfaces may be used. 
     For example, in  FIG. 12  the system  1200  includes an evaporator  1220  that is interfaced with a thermal storage unit  1222 . In one implementation, the thermal storage unit  1222  may be used as a heat load transformer for pulsed power applications, such as, for example, space-based laser applications. The thermal storage unit may include, for example, 250 W-hr graphite hardware and a paraffin-based, lightweight composite design. 
     Further in  FIG. 12 , the system  1200  may include an evaporator  1224  that is interfaced with a condensing heat exchanger  1226 , which is used to couple a spray-cooled evaporator  1228  into the system  1200 . The heat exchanger  1226  may be, for example, a high efficiency, two-phase/two-phase heat exchanger. A liquid pump  1230  is used to pump liquid from the condensing heat exchanger  1226  through the spray-cooled evaporator  1228 , to thereby form a separate loop coupled to the loop(s) of a primary thermal bus of the system  1200 . 
     In particular, such a separate loop may be used to connect the spray-cooled evaporator  1228  to the system  1200 , due to the fact that a nozzle pressure drop (e.g., 0.7 bar) of the spray-cooled evaporator  1228  relative to a capillary pressure rise (e.g., 0.4 bar) in the system  1200  may make parallel arrangement of the spray-cooled evaporator  1228  difficult in some use environments. In other implementations, however, the spray-cooled evaporator  1228  may be integral to the system  1200 , instead of being coupled through the condensing heat exchanger  1226 . 
     The spray-cooled evaporator  1228  may be used for efficient thermal control of high heat flux sources. For example, 500 W/cm 2  has been demonstrated with a heat transport system using ammonia as the working fluid. A loop using the spray-cooled evaporator  1228  may be operated near saturation in order to maximize heat transfer. 
     Such a spray-cooled evaporator  1228  may be particularly useful, for example, in spacecraft thermal management. For instance, in spacecraft electronics, heat fluxes at transistor gates are approaching 1 MW/in 2 . As component size continues to shrink and heat fluxes rise further, state-of-the-art systems may be used to offset the associated increases in local temperature drops. The significantly higher heat-transfer coefficient afforded by spray cooling, using the spray-cooled evaporator  1228 , may be advantageous in this respect. 
     Factors to consider in using the spray-cooled evaporator  1228  include, for example, nozzle optimization and scalability of the spray-cooled evaporator  1228  to extended surface areas. In one implementation, the spray-cooled evaporator  1228  may be used for cooling laser diode applications. 
     In  FIGS. 11 and 12 , and in light of the above discussion, it should be understood that the capillary pumping developed by the evaporator wicks, as described above, may generally maintain phase separation at each heat source interface of the evaporators, and thereby assure excellent heat transfer characteristics and automatic flow control among the evaporators, even when no flow controllers are used. A pressure diagram illustrating this phenomenon is described in more detail below with respect to  FIG. 25 . 
     Also, it should be apparent from  FIG. 12  and the above discussion that many variations exist with respect to a number, type, and arrangement of evaporators that may be used in the system  1200 . Further examples of evaporator configurations are discussed below with respect to  FIGS. 18A-18C . 
     Similarly, many types of condenser configurations may be used. For example, the condenser  1212  referred to above may include a body-mounted radiator, while a condenser  1232  may include a multi-fold, deployable or steerable radiator. Particularly in high-power spacecrafts, these radiators may be direct condensation or may use discrete heat pipes, depending on, for example, system reliability factors and/or a mass of micro-meteoroid shielding. As just mentioned, the condenser  1232  also may be made steerable for non-geostationary applications, in order, for example, to minimize radiator backloading. Gimbaled heat transport systems used in conventional telecom satellite systems may be used to enable such steerable radiators. Further, passive two-phase loops (e.g., LHPs) also may be incorporated into two-axis gimbaled systems. Other examples of condenser configurations are discussed below with respect to  FIGS. 18A-18C . 
     Finally with respect to  FIG. 12 , a liquid bypass valve  1234  is illustrated that may be used, for example, to maintain constant pump speed operations with the liquid pump  1202 , and which may improve a pump lifetime of the pump  1202 . Further, flexible elements  1236  are illustrated in order to indicate that the various elements of the system  1200  may be routed over and through a wide variety of use environments. 
       FIG. 13  is a schematic illustrating a heat transport system  1300  that shares many elements with the system  1200  of  FIG. 12  (indicated in  FIG. 13  by like-numbered elements). In  FIG. 13 , however, the mechanical pump  1102  of  FIG. 11  is represented by a vapor compressor  1302 , which may be a variable-speed vapor compressor. A liquid/vapor separator  1304  (or a vapor superheater, not shown) may be used to prevent liquid from entering the compressor and, similarly to the pump bypass valve  1218  of  FIG. 12 , a compressor bypass valve  1306  may be used to operate the system  1300  in a passive (capillary) pumping mode. 
     The choice of whether to use the liquid pump  1202  or the vapor compressor  1302  is typically a design consideration. Generally, the liquid pump  1202  offers lighter weight and increased pumping power relative to the vapor compressor  1302  (due to, for example, the lower volumetric flow rate of the former). On the other hand, the vapor compressor  1302  offers heat pumping (i.e., an increased condensation temperature), which may reduce radiator heat and overall system mass and, additionally, may offer a longer operational lifetime. 
     The liquid pump  1202  may include, for example, a hermetically sealed, magnetically driven, centrifugal design. Other liquid pumps for space station applications, e.g., waste water and carbon dioxide, also may be used. 
     The vapor compressor  1302  may be a variable-speed compressor, and may include, for example, a hermetically sealed, oil-less centrifugal compressor with gas or magnetic bearings. A low-lift heat pump, which includes a similar compressor, also may be used. Further examples of specific types of pumps are provided below, and, in particular, with respect to  FIGS. 17A-17E . 
     As also illustrated in  FIG. 13 , a vapor compressor  1308  may be used in the loop formed by the spray-cooled evaporator  1228  and the condensing heat exchanger  1226 , instead of the liquid pump  1230 . The choice between the liquid pump  1230  and the vapor compressor  1308  may be driven by, for example, design choices similar to those just described. 
     Further in  FIG. 13 , flow controllers  1310  may be used to ensure a desired heat load distribution between the evaporators  1206 ,  1220 , and  1224 . For example, the flow controllers  1310  may be used to route more or less liquid to a particular evaporator, depending on, for example, an amount of heat present at that evaporator or, in the case of the evaporator  1220 , an amount of heat to be stored in the thermal storage unit  1222 . In order to provide equal heat load distribution, for example, feedback may be provided from an output of each of the evaporators  1206 ,  1220 , and  1224  to the flow controllers  1310 . An example of this implementation is illustrated in more detail below, with respect to  FIG. 15 . The flow controllers  1310  are shown in  FIG. 13  as liquid flow controllers, but also may include other types of flow controllers, such as, for example, vapor flow controllers. 
     Referring to  FIG. 14 , an implementation of a system  1400  is shown that includes condenser capillary flow regulators  1402 . The capillary flow regulators  1402  are included to increase or maximize condenser efficiency, reduce or minimize condenser size, and ensure subcooled liquid return to the liquid pump  1202 . The capillary flow regulators  1402  are discussed in more detail below with respect to  FIG. 19 . 
     Also in  FIG. 14 , a vapor bypass line  1404  is shown in conjunction with a low temperature heat source  1406  (and/or the spray-cooled evaporator  1228 ). Specifically, the vapor bypass line  1404  bypasses the vapor compressor  1308  and facilitates operation of the condensing heat exchanger  1226 . 
     Referring to  FIG. 15 , an implementation  1500  is shown that includes superheat feedback flow controllers  1502  for regulating evaporator flow control. A regenerator  1504  is connected to the vapor compressor  1302 , and generally is operable to reuse the latent heat in the steam that leaves the compressor  1302  to assist in operation of the compressor  1302 . An expansion valve  1506  is included to meter the liquid flow that enters the evaporators from the liquid line  1204 , such that the liquid flow enters the evaporators at a desired rate, e.g., a rate that matches the amount of liquid being evaporated in the evaporators. 
     Referring to  FIG. 16 , an implementation of a system  1600  is shown that includes a secondary evaporator  1602 , which is used similarly to the secondary evaporator  150  of  FIG. 1 , the secondary evaporator  710  of  FIG. 7 , and the secondary evaporator  1011  of  FIG. 10 . That is, the secondary evaporator  1602  is used as a priming evaporator for ensuring successful start-up of the system  1600 , and for ensuring sufficient excess flow through the primary evaporator cores to enable venting of excess vapor and NCG bubbles therefrom, particularly during a passive (capillary) operation of the system  1600 . 
     More specifically, as should be apparent from the above discussion, the secondary evaporator  1602  is thermally and hydraulically connected to a cold-biased reservoir  1604 . As described with respect to  FIG. 3 , application of power (heat) to the secondary evaporator  1602  causes evaporation therefrom, which travels through a back pressure regulator (BPR)  1606  (discussed in more detail below) and is condensed within one or more condensers  1608 . Flow regulators  1610  (similar to the regulators  1402  discussed above, and co-located with one another or with their respective condensers) regulate the condensed liquid flow from the condensers  1608  through a mechanical pump  1612 . From there, the condensed liquid flows through an inner liquid flow line of a coaxial flow line  1614 . In this way, the liquid reaches cold plate evaporator(s)  1616 , as well as a thermal mass (storage unit)  1618  and a remote evaporator  1620 . 
     Further, an isothermalized plate or structure  1622  may be included. Such a structure may be useful, for example, in settings where a constant temperature surface is desired or required, such as, for example, some laser systems. To the extent that such systems require a constant temperature surface, it may be efficient to use the (waste) heat being transported by the system  1600  to keep the structure  1622  at a constant temperature. When the structure  1622  is used, a flow regulator  1624  (perhaps similar to the regulators  1402  of  FIG. 14 ) may be used to ensure that a proper amount of vapor from a vapor return line  1626  is provided to the structure  1622 . 
     A liquid line heat exchanger  1628  is used to provide subcooling of the liquid before it is routed to the evaporators. Also, as just referred to, the vapor return line  1626  returns vapor to the secondary evaporator  1602  and to the BPR  1606 . The BPR  1606 , generally speaking, ensures that no vapor reaches the condensers unless a vapor space for all evaporators in the system is devoid of liquid. As such, heat load sharing among the many parallel (or series) evaporators may be increased. An example of the BPR  1606  is discussed in detail below with respect to  FIG. 20 . 
       FIGS. 11-16  illustrate various implementations of actively pumped thermal management systems, which include different combinations and arrangements of thermal management components. In order to further illustrate the flexibility of design and use of such thermal management systems, additional examples of such thermal components and their uses are provided below with respect to  FIGS. 17-25 . It should be understood that such thermal components, and others, may be used in conjunction with some or all of the implementations of  FIGS. 11-16 , or in other implementations. 
       FIGS. 17A-17E  are examples of mechanical pumps that may be used in the systems of  FIGS. 11-16 . Specifically,  FIG. 17A  illustrates a bellows pump, while  FIG. 17B  illustrates a centrifugal pump.  FIG. 17C  illustrates a diaphragm pump, and  FIG. 17D  illustrates a gear pump. Finally,  FIG. 17D  illustrates a peristaltic pump. It should be understood that the illustrated pumps are merely examples of known pumps that may be used in an actively pumped thermal management system, and other types of pumps also may be used. 
       FIGS. 18A-18C  illustrate examples of different evaporator and condenser architectures for use with the systems of  FIGS. 11-16 . As already discussed, such architectures may be characterized by virtually any parallel or series arrangement of evaporators and condensers. In  FIG. 18A , a heat flow arrangement involving a centralized thermal bus  1802  is used for defense space applications requiring on-orbit servicing. In this concept, multiple parallel evaporators  1804  are used to cool internal electronics  1806 , thermal storage units  1808 , on-gimbal evaporator  1810  on a gimbaled payload  1812  that is connected to the bus  1802  by a coil  1814 , and on-orbit replaceable electronics modules  1816 . Spot coolers  1818  may be used as needed, and the bus  1802  is connected to a deployable or steerable direct condensation radiator  1820  by a coil  1822 . The deployable radiator  1820  may include a secondary loop heat pipe evaporator/reservoir mounted on the radiator  1820  to ensure that the radiator  1820  is cold-biased. 
     In  FIG. 18B , an evaporator section  1824  includes multiple cold plates  1826  connected in parallel to a starter pump  1828  and thermal storage units (TSUs)  1830 . A two-axis gimbaled cold plate  1832  is also connected to the evaporator section  1824 , by way of a coil  1834 . The cold plate  1826  may feature equipment mounting locations  1836  having an advanced interface design, as well as additional spot cooler loops  1838 . In this example, a two-axis gimbaled condenser  1840  is connected to the evaporator section  1824  by a coil  1842 , and is connected to a pump  1844  and reservoir  1846 . Additional cooling may be supplied by a chiller  1848  that is connected to the condenser  1840 . 
     In  FIG. 18C , a possible design for use in a space shuttle bay is illustrated, in which an evaporator section  1850  includes a deployable evaporator section  1852  with a coil or hinge  1854 , modular electronic boxes  1856 , and thermal storage units  1858 . A deployable radiator  1860  includes a pump  1862  and reservoir  1864 , as well as a coil or hinge  1866 . 
       FIG. 19  is a diagram of an example of the condenser flow regulator  1402  of  FIGS. 14-16 . In  FIG. 19 , a capillary structure  1902  receives a combined liquid/vapor flow  1904  from an associated condenser, and ensures liquid return to an associated liquid line. As discussed above, the regulator  1402  may thus increase a performance, and reduce a size of, associated parallel condensers. 
       FIG. 20  is a diagram of an example of the back pressure regulator (BPR)  1606  of  FIG. 16 . As discussed above, the BPR  1606  typically is added to a condenser inlet in order to enable heat load sharing in either an active or passive (capillary) pumping mode of a thermal management system, such as the systems of  FIGS. 11-16 . 
     In  FIG. 20 , the BPR  1606  is attached at a vapor transport line  2002  on one end and at a radiator or condenser inlet header  2004  at the other end. The BPR  1606  includes a tubular shell external structure  2006  that has an internal annular wick  2008 . The wick  2008  has a first, sealed end  2010  and a second, unsealed (open) end  2012 . The sealed end  2010  of the wick  2008  is surrounded by an annular gap  2014  filled with vapor. The unsealed end  2012  of the wick  2008  is surrounded by an annular gap  2016  filled with liquid. As shown, the annular gaps  2014 / 2016  extend only a portion of the length of the BPR  1606 . In a central (low conductance) portion  2018  of the BPR  1606 , the tubular shell  2006  makes contact with the wick outer surface, and thereby seals the annular gap  2014  from the annular gap  2016 . 
     Thus, the BPR  1606  typically is positioned at the inlet to the condenser, where the vapor transport line  2002  meets the condenser inlet header  2004 . As such, the unsealed open end  2012  of the internal wick  2008  is thermally linked to a cooling source  2020  (e.g., radiator or other heat sink), and is connected to the condenser inlet header  2004  end of the BPR  1606 . The other end  2010  (sealed end of the internal wick  2008 ) is connected in series to the vapor transport line  2002 . 
     The BPR  1606  ensures that no vapor reaches the condenser unless the vapor space for all evaporators in the system is devoid of liquid. As such, heat load sharing among the many parallel or series evaporators in the system may be increased. The BPR  1606  typically uses pores  2022  selected such that the pore size is larger than the pore size(s) of any of the system evaporators. Thus, as vapor is produced, it is contained within all the evaporator vapor side space, and is thereby given an opportunity to condense. The vapor clears all evaporator vapor side space of liquid and, once that condition is achieved, pushes through the BPR wick  2008  and allows flow to reach the connected condenser. 
       FIGS. 21 and 22  are diagrams of evaporator failure isolators  2100  and  2200 , respectively, which may be used in any multi-evaporator implementations of the systems of  FIGS. 11-16 . The isolators  2100  and  2200  generally are operable to prevent evaporator pump failures at any particular evaporator from propagating throughout an associated thermal management system. 
     In  FIG. 21 , the isolator  2100  includes a first port  2102  for receiving liquid flow from a liquid line  2104  supplying liquid to a plurality of evaporators. A liquid return port  2106  outputs liquid to other isolators, and a liquid outlet port  2108  outputs liquid to an associated capillary pump (evaporator). 
     A tube  2110  defines a body of the isolator  2100  that includes a wick  2112  and a flow annulus  2114 . Along with a swage seal  2116 , the wick  2112  and flow annulus  2114  enable isolation of the liquid flow to an associated evaporator, through the liquid outlet port  2108 . If the associated evaporator experiences pump failure, it may be bypassed by the isolator  2100  until repair may be effected. 
     Similarly, in  FIG. 22 , an evaporator failure isolator  2200  includes a liquid flow annulus  2202  through which subcooled liquid flows from an associated reservoir to remaining pumps. Isolation seals  2204  ensure that liquid flow to associated pumps is maintained through ports  2206 , such that only currently functioning pumps receive liquid flow. 
       FIGS. 23 and 24  illustrate examples of capillary pressure sensors  2300  and  2400 , respectively. Such capillary pressure sensors, generally speaking, provide feedback control for a mechanical pump (e.g., the mechanical pump  1102  of  FIG. 11 ), and enable heat load sharing among multiple evaporators. 
     In  FIGS. 23 and 24 , a liquid line  2302  and vapor line  2304  are coupled hydraulically to the capillary pressure sensors  2300  and  2400 . Particularly, in  FIG. 23 , the liquid and vapor lines are adjacent to one or more evaporators, and the capillary pressure sensor  2300  includes a hermetic envelope  2306 , an internal wicking structure  2308 , and multiple temperature sensors  2310 . 
     The internal wicking structure  2308  includes a continuous wick element  2312  with the same capillary pumping radius  2314  (r pevap ) as an evaporator wick that hydraulically links the liquid line  2302  to one or more wick segments  2316 ,  2318 , and  2320  with larger capillary pumping radii (r p1 , r p2 , and r p3 ). The capillary sensor  2300  is thermally coupled to one or more heat sources  2322 . 
     In operation, the temperature sensors  2310  measure envelope temperature above each wick segment  2316 ,  2318 ,  2320 , and/or temperature differences between the envelopes above each wick segment  2316 ,  2318 ,  2320 . Temperature increases on the envelope indicate that the wick segment below the envelope may no longer be saturated with liquid, due to inability of the wick segment to support the pressure difference between the vapor line  2304  and the liquid line  2302 . Thus, temperature feedback may be used to adjust a pumping pressure delivered by the mechanical pump  1102  by, for example, adjusting pump speed or adjusting a position of an associated pump bypass valve, in order to maintain saturation of the appropriate wick segment(s). 
     In  FIG. 24 , a heat sink  2402  provides cold bias between the wick segments  2316 ,  2318 , and  2320 , and multiple temperature sensors  2310  are used to measure temperature in the cold-biased zone(s). The wick segments  2316 ,  2318 , and  2320  may be arranged in sequence, with the wick segment with the largest capillary radius nearest as associated vapor manifold. 
     In operation, temperature increases on the envelope indicate that the wick segment between the sensor and the vapor manifold may no longer be saturated with liquid due to, for example, an inability of the wick segment to support a pressure difference between the vapor line  2304  and the liquid line  2302 . Then, temperature feedback may be used to adjust the pumping pressure delivered by the mechanical pump, by either adjusting pump speed or the position of a pump bypass valve, to maintain saturation of the appropriate wick segment(s). 
       FIG. 25  is a pressure drop diagram  2500  for a thermal management system, such as the various implementations of thermal management systems discussed above. In  FIG. 25 , the mechanical pump  1110  provides a pressure difference ΔP pump    2502  that is slightly higher than the low pressure point  2504  of the system at the reservoir. Pressure difference ΔP Flow Reg    2506 , the pressure differences provided by the flow regulators  1402 , are lower than the pressure difference ΔP LHP    2508  of the Loop Heat Pipe. Other than the pressure differences ΔP visc 5,6    2510 ,  2512 , where a viscous pressure drop may dominate in effect, pressure differentials ΔP cap 1, 2, 3    2514 ,  2516 ,  2518  demonstrate the positive pressure differentials that enable capillary back pressure(s) the evaporators of the thermal management system, using the evaporator wicks, that allow excellent heat transfer and flow control, in conjunction with the mechanical pump  1110 . Finally, a pressure difference ΔP cap 4    2520  illustrates a pressure difference maintained for regulating flow through the condenser(s)  1115 . 
     As shown in  FIGS. 11-25 , many different implementations exist for actively pumped thermal management systems. Such systems include capillary and/or mechanically pumped two-phase thermal management systems that combine the low input power, passive system advantages (e.g., heat load sharing, no moving parts) of small pore wick (capillary) pumped two-phase loop systems with the operational flexibility advantages (e.g., fluid flow-heat flow decoupling and flow controllability) of mechanically pumped two-phase loop systems. 
     As described, such thermal management systems absorb waste heat from a wide range of sources, including, for example, waste heat of electronics and power conditioning equipment, high-powered spacecraft, antennas, batteries, and laser systems. Military applications, such as space-based radar and lasers, offer a wide suite of potential heat sources and the elements required for their thermal management. Accordingly, such military applications such as those requiring counterspace detection and offensive force projection capabilities, may benefit from such thermal management systems, which provide high heat transport capability and high heat rejection, as well as high flux heat acquisition and efficient thermal storage, all the while minimizing system mass and maintaining operational reliability over the mission life. Commercial applications, such as, for example, soda-dispensing machines and notebook computers, also may benefit from the implementations of heat transport systems discussed herein, or variations thereof.

Technology Classification (CPC): 5