Patent Publication Number: US-7708053-B2

Title: Heat transfer system

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This application claims priority to U.S. Provisional Patent Application Ser. No. 60/421,737, filed Oct. 28, 2002, which is incorporated herein by reference. 
     This application also claims priority to U.S. Provisional Patent Application Ser. No. 60/514,670, titled “HEAT TRANSFER SYSTEM FOR A CYCLICAL HEAT EXCHANGE SYSTEM,” filed Oct. 28, 2003, which also is incorporated herein by reference. 
     This application is a continuation-in-part of U.S. patent application Ser. No. 10/676,265, titled “EVAPORATOR FOR A HEAT TRANSFER SYSTEM,” filed Oct. 2, 2003, which claimed priority to U.S. Patent Application Ser. No. 60/415,424, filed Oct. 2, 2002, which are also incorporated herein by reference. 
     This application 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 the benefit of U.S. Provisional 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 claims the benefit of U.S. Provisional Patent Application Ser. No. 60/215,588, filed Jun. 30, 2000. All of these applications are incorporated herein by reference. 
    
    
     TECHNICAL FIELD 
     This description relates to heat transfer systems for use in cyclical heat exchange systems. 
     BACKGROUND 
     Heat transfer systems are used to transport heat from one location (the heat source) to another location (the heat sink). Heat transfer systems can be used in terrestrial or extraterrestrial applications. For example, heat transfer systems may be integrated by satellite equipment that operates within zero- or low-gravity environments. As another example, heat transfer 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 transfer 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 transfer 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 an 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 an LHP is co-located with the evaporator. 
     SUMMARY 
     In one general aspect, a heat transfer system for a cyclical heat exchange system includes an evaporator including a wall configured to be coupled to a portion of the cyclical heat exchange system and a primary wick coupled to the wall and a condenser coupled to the evaporator to form a closed loop that houses a working fluid. 
     Implementations may include one or more of the following aspects. For example, the condenser includes a vapor inlet and a liquid outlet and the heat transfer system includes a vapor line providing fluid communication between the vapor outlet and the vapor inlet and a liquid return line providing fluid communication between the liquid outlet and the liquid inlet. 
     The evaporator includes a liquid barrier wall containing the working fluid on an inner side of the liquid barrier wall, which working fluid flows only along the inner side of the liquid barrier wall, wherein the primary wick is positioned between a heated wall and the inner side of the liquid barrier wall; a vapor removal channel that is located at an interface between the primary wick and the heated wall, the vapor removal channel extending to a vapor outlet; and a liquid flow channel located between the liquid barrier wall and the primary wick, the liquid flow channel receiving liquid from a liquid inlet. 
     The working fluid is moved through the heat transfer system passively. 
     The working fluid is moved through the heat transfer system without the use of external pumping. 
     The working fluid within the heat transfer system changes between a liquid and a vapor as the working fluid passes through or within one or more of the evaporator, the condenser, the vapor line, and the liquid return line. 
     The working fluid is moved through the heat transfer system passively. 
     The working fluid is moved through the heat transfer system with the use of the wick. 
     The heat transfer system further includes fins thermally coupled to the condenser to reject heat to an ambient environment. 
     In another general aspect, a thermodynamic system includes a cyclical heat exchange system and a heat transfer system coupled to the cyclical heat exchange system to cool a portion of the cyclical heat exchange system. The heat transfer system includes an evaporator including a wall configured to be coupled to a portion of the cyclical heat exchange system and a primary wick coupled to the wall and a condenser coupled to the evaporator to form a closed loop that houses a working fluid. 
     Implementations may include one or more of the following features. The evaporator is integral with the cyclical heat exchange system. The evaporator is thermally coupled to the portion of the cyclical heat exchange system. The cyclical heat exchange system includes a Stirling heat exchange system. The cyclical heat exchange system includes a refrigeration system. The heat transfer system is coupled to a hot side of the cyclical heat exchange system. The thermodynamic system heat transfer system is coupled to a cold side of the cyclical heat exchange system. 
     In another general aspect, a method utilizes the systems recited above. 
     The evaporator may be used in any two-phase heat transfer system for use in terrestrial or extraterrestrial applications. For example, the heat transfer systems can be used in electronic equipment, which often requires cooling during operation or in laser diode applications. 
     The planar evaporator may be used in any heat transfer system in which the heat source is formed as a planar surface. The annular evaporator may be used in any heat transfer system in which the heat source is formed as a cylindrical surface. 
     The heat transfer system that uses the annular evaporator may take advantage of gravity when used in terrestrial applications, thus making an LHP suitable for mass production. Terrestrial applications often dictate the orientation of the heat acquisition surfaces and the heat sink; the annular evaporator utilizes the advantages of the operation in gravity. 
     The heat transfer system provides a thermally efficient and space efficient system for cooling a cyclical heat exchange system because the evaporator of the heat transfer system is thermally and spatially coupled to a portion of the cyclical heat exchange system that is being cooled by the heat transfer system. For example, if the portion to be cooled (also known as a heat source) has a cylindrical geometry, the heat transfer system may include an annular evaporator. Use of the heat transfer system enables exploitation of cylindrical cyclical heat exchange systems, which are capable of being used in a commercially practical application for cabinet cooling. 
     Integral incorporation of the evaporator or condenser with the heat source of the cyclical heat exchange system can minimize packaging size. On the other hand, if the evaporator or condenser is clamped onto the heat source, the deployment and replacement of parts is facilitated. 
     The heat transfer system may be used to cool a cyclical heat exchange system having a cylindrical geometry, such as, for example, a free-piston Stirling cycle. A heat transfer system provides efficient fluid line connection (one vapor phase and on sub-cooled liquid return line connector) to and from an equally efficiently packaged annular condenser assembly. 
     The heat transfer system incorporates a condenser that is efficiently packaged as a flat plate condenser that is formed into annular sections to which are attached extended air heat exchange surface elements such as corrugated fin stock. 
     The heat transfer system combines efficient heat transfer mechanisms (evaporation and condensation) to couple the fluid of the Stirling cycle (helium) to the ultimate heat sink (ambient air). Consequently, a significant improvement in Stirling cycle efficiency (for example, up to 50%) is provided. 
     The evaporator and the condenser of the heat transfer system can be independently designed and optimized. This allows any number of attachment options to the cyclical heat exchange system. Moreover, the heat transfer system is insensitive to gravity orientation because a wick is incorporated into the evaporator. 
     The heat transfer system provides efficient cooling to a cabinet, such as a refrigerator or vending machine, in a small package at a commercially acceptable cost. 
     According to one implementation, an annular evaporator is clamped onto a cyclical heat exchange system and thermally coupled with thermal grease compound to provide easy assembly and servicing. According to another implementation, an annular evaporator is interference fit onto a cyclical heat exchange system to provide easy assembly with improved thermal efficiency. According to a further implementation, an annular evaporator is integrally formed with a cyclical heat exchange system to provide further improved thermal efficiency. 
     The heat transfer system includes a condenser having finned inner and outer annular portions to provide efficient heat transfer to the air in a reduced packaging space. The condenser may be roll bonded or formed by extrusion. 
     A loop heat pipe of the present invention provides for efficient packaging with a cylindrical refrigerator by adapting the traditional cylindrical geometry of a LHP evaporator to a planar “flat-plate” geometry that can be wrapped in an annular shape. 
     The packaging of the heat transfer system is described with respect to a few exemplary implementations, but is not meant to be limited to those exemplary implementations. Although described with respect to use for cooling a cabinet, such as a domestic refrigerator, vending machine, or point-of-sale refrigeration unit, one of skill in the art will recognize the numerous other useful applications of a compact, energy efficient and environmentally friendly refrigeration unit utilizing the heat transfer system as described herein. 
     Other features and advantages 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  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  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 cross-sectional view of a planar evaporator. 
         FIG. 11  is an axial cross-sectional view of an annular evaporator. 
         FIG. 12  is a radial cross-sectional view of the annular evaporator of  FIG. 11 . 
         FIG. 13  is an enlarged view of a portion of the radial cross-sectional view of the annular evaporator of  FIG. 12 . 
         FIG. 14A  is a perspective view of the annular evaporator of  FIG. 11 . 
         FIG. 14B  is a top and partial cutaway view of the annular evaporator of  FIG. 14A . 
         FIG. 14C  is an enlarged cross-sectional view of a portion of the annular evaporator of  FIG. 14B . 
         FIG. 14D  is a cross-sectional view of the annular evaporator of  FIG. 14B  taken along line  14 D- 14 D. 
         FIGS. 14E and 14F  are enlarged views of portions of the annular evaporator of  FIG. 14D . 
         FIG. 14G  is a perspective cut-away view of the annular evaporator of  FIG. 14A . 
         FIG. 14H  is a detail perspective cut-away view of the annular evaporator of  FIG. 14G . 
         FIG. 15A  is a flat detail view of the heated wall formed into a shell ring component of the annular evaporator of  FIG. 14A . 
         FIG. 15B  is a cross-sectional view of the heated wall of  FIG. 15A  taken along line  15 B- 15 B. 
         FIG. 16A  is a perspective view of a primary wick of the annular evaporator of  FIG. 14A . 
         FIG. 16B  is a top view of the primary wick of  FIG. 16A . 
         FIG. 16C  is a cross-sectional view of the primary wick of  FIG. 16B  taken along line  16 C- 16 C. 
         FIG. 16D  is an enlarged view of a portion of the primary wick of  FIG. 16C . 
         FIG. 17A  is a perspective view of a liquid barrier wall formed into an annular ring of the annular evaporator of  FIG. 14A . 
         FIG. 17B  is a top view of the liquid barrier wall of  FIG. 17A . 
         FIG. 17C  is a cross-sectional view of the liquid barrier wall of  FIG. 17B  taken along line  17 C- 17 C. 
         FIG. 17D  is an enlarged view of a portion of the liquid barrier wall of  FIG. 17C . 
         FIG. 18A  is a perspective view of a ring separating the liquid barrier wall of  FIG. 17A  from the heated wall of  FIG. 15A . 
         FIG. 18B  is a top view of the ring of  FIG. 18A . 
         FIG. 18C  is a cross-sectional view of the ring of  FIG. 18B  taken along line  18 C- 18 C. 
         FIG. 18D  is an enlarged view of a portion of the ring of  FIG. 18C . 
         FIG. 19A  is a perspective view of a ring of the annular evaporator of  FIG. 14A . 
         FIG. 19B  is a top view of the ring of  FIG. 19A . 
         FIG. 19C  is a cross-sectional view of the ring of  FIG. 19B  taken along  19 C- 19 C. 
         FIG. 19D  is an enlarged view of a portion of the ring of  FIG. 19C . 
         FIG. 20  is a perspective view of a cyclical heat exchange system that can be cooled using a heat transfer system. 
         FIG. 21  is a cross-sectional view of a cyclical heat exchange system such as the cyclical heat exchange system of  FIG. 20 . 
         FIG. 22  is a side view of a cyclical heat exchange system such as the cyclical heat exchange system of  FIG. 20 . 
         FIG. 23  is a schematic diagram of a first implementation of a cyclical heat exchange system including a cyclical heat exchange system and a heat transfer system. 
         FIG. 24  is a schematic diagram of a second implementation of a cyclical heat exchange system including a cyclical heat exchange system and a heat transfer system. 
         FIG. 25  is a schematic diagram of a heat transfer system using an evaporator designed in accordance with the principles of  FIGS. 11-13 . 
         FIG. 26  is a functional exploded view of the heat transfer system of  FIG. 25 . 
         FIG. 27  is a partial cross-sectional detail view of an evaporator used in the heat transfer system of  FIG. 25 . 
         FIG. 28  is a perspective view of a heat exchanger used in the heat transfer system of  FIG. 25 . 
         FIG. 29  is a graph of temperature of a heat source of a cyclical heat exchange system versus a surface area of an interface between the heat transfer system and the heat source of the cyclical heat exchange system. 
         FIG. 30  is a top plan view of a heat transfer system packaged around a portion of a cyclical heat exchange system. 
         FIG. 31  is a partial cross-sectional elevation view (taken along line  31 - 31 ) of the heat transfer system packaged around the cyclical heat exchange system portion of  FIG. 30 . 
         FIG. 32  is a partial cross-sectional elevation view (taken at detail  3200 ) of the interface between the heat transfer system and the cyclical heat exchange system of  FIG. 30 . 
         FIG. 33  is an upper perspective view of a heat transfer system mounted to a cyclical heat exchange system. 
         FIG. 34  is a lower perspective view of the heat transfer system mounted to the cyclical heat exchange system of  FIG. 33 . 
         FIG. 35  is a partial cross-sectional view of an interface between an evaporator of a heat transfer system and a cyclical heat exchange system in which the evaporator is clamped onto the cyclical heat exchange system. 
         FIG. 36  is a side view of a clamp used to clamp the evaporator onto the cyclical heat exchange system of  FIG. 35 . 
         FIG. 37  is a partial cross-sectional view of an interface between an evaporator of a heat transfer system and a cyclical heat exchange system in which the interface is formed by an interference fit between the evaporator and the cyclical heat exchange system. 
         FIG. 38  is a partial cross-sectional view of an interface between an evaporator of a heat transfer system and a cyclical heat exchange system in which the interface is formed by forming the evaporator integrally with the cyclical heat exchange system. 
         FIG. 39  is a top plan view of a condenser of a heat transfer system. 
         FIG. 40  is a partial cross-sectional view taken along line  40 - 40  of the condenser of  FIG. 39 . 
         FIGS. 41-43  are detail cross-sectional views of a condenser having a laminated construction. 
         FIG. 44  is a detail cross-sectional view of a condenser having an extruded construction. 
         FIG. 45  is a perspective detail and cross-sectional view of a condenser having an extruded construction. 
         FIG. 46  is a cross-sectional view of one side of a heat transfer system packaging around a cyclical heat exchange system. 
     
    
    
     Like reference symbols in the various drawings indicate like elements. 
     DETAILED DESCRIPTION 
     As discussed above, in a loop heat pipe (LHP), the reservoir is co-located with the evaporator, thus, the reservoir is thermally and hydraulically connected with the reservoir 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 also 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 that liquid be present in the reservoir prior to start-up, that is, application of power to the evaporator of the LHP. However, if the working fluid in the LHP is in a supercritical state prior to start-up of the LHP, liquid will not be present in the reservoir prior to start-up. A supercritical state is a 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 that liquid returning to the evaporator is 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. 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 is 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  may also 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 heat transport system  100 . 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. patent 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 aluminum or any heat conductive material. 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 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 reaches 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. 
     If the set point temperature has been reached (step  335 ), the heat transport 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 set point temperature of the reservoir  155  is reduced. The rate at which the heat transfer system  105  cools down during the main mode depends 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 required, 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 controlled 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 . 
     In this particular implementation, the main mode is triggered by the determination that the set point temperature has been reached (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 is wet (step  310 ) or after the reservoir has been cold biased (step  305 ). 
     At any time during operation, the heat transfer system  105  can experience heat conditions such as those resulting from heat conduction across the primary wick  140  and parasitic heat applied to the liquid line  125 . Both conditions cause formation of vapor on the liquid side of the evaporator. 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 supply to the primary wick  140 , thus causing the main evaporator  115  to fail. Parasitic heat input into the liquid line  125  (referred to as “parasitic heat gains”) can cause liquid within the liquid line  125  to form vapor. 
     To reduce the adverse impact of heat conditions discussed above, the priming system  110  operates at a power level 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 around the bayonet tube  142  directly into the fluid outlet port  139 . Vapor that forms within the first vapor passage  144  makes it way 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 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 and 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 secondary 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 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 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    450  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  operates at a temperature of −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. patent application Ser. No. 09/896,561, filed Jun. 29, 2001, now U.S. Pat. No. 6,889,754, issued May 10, 2005. 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 tube  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 also to  FIG. 7 , a heat transport system  700  is shown in which the main evaporator is a four-port evaporator  600 . The heat transport 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 . 
     Design considerations of the heat transport system  100  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 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 secondary 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 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 (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  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) such as a sensor  1016  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 115 K 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 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. 
     Evaporator Design 
     Evaporators are integral components in two-phase heat transfer systems. For example, as shown above in  FIGS. 5A and 5B , the evaporator  500  includes an evaporator body or container  515  that is in contact with the primary wick  540  that surrounds the core  510 . The core  510  defines a flow passage for the working fluid. The primary wick  540  is surrounded at its periphery by a plurality of peripheral flow channels or vapor grooves  545 . The channels  545  collect vapor at the interface between the wick  540  and the evaporator body  515 . The channels  545  are in contact with the vapor outlet  550  that feeds into the vapor line  130  that feeds into the condenser to enable evacuation of the vapor formed within the main evaporator  115 . 
     The evaporator  500  and the other evaporators discussed above often have a cylindrical geometry, that is, the core of the evaporator forms a cylindrical passage through which the working fluid passes. The cylindrical geometry of the evaporator is useful for cooling applications in which the heat acquisition surface is cylindrically hollow. Many cooling applications require that heat be transferred away from a heat source having a flat surface. In these sort of applications, the evaporator can be modified to include a flat conductive saddle to match the footprint of the heat source having the flat surface. Such a design is shown, for example, in U.S. Pat. No. 6,382,309. 
     The cylindrical geometry of the evaporator facilitates compliance with thermodynamic constraints of LHP operation (that is, the minimization of heat leaks into the reservoir). The constraints of LHP operation stem from the amount of subcooling an LHP needs to produce for normal equilibrium operation. Additionally, the cylindrical geometry of the evaporator is relatively easy to fabricate, handle, machine, and process. 
     However, as will be described hereinafter, an evaporator can be designed with a planar form to more naturally attach to a flat heat source. 
     Planar Design 
     Referring to  FIG. 10 , an evaporator  1000  for a heat transfer system includes a heated wall  1005 , a liquid barrier wall  1011 , a primary wick  1015  between the heated wall  1005  and the inner side of the liquid barrier wall  1011 , vapor removal channels  1020 , and liquid flow channels  1025 . 
     The heated wall  1005  is in intimate contact with the primary wick  1015 . The liquid barrier wall  1011  contains working fluid on an inner side of the liquid barrier wall  1011  such that the working fluid flows only along the inner side of the liquid barrier wall  1011 . The liquid barrier wall  1011  closes the evaporator&#39;s envelope and helps to organize and distribute the working fluid through the liquid flow channels  1025 . The vapor removal channels  1020  are located at an interface between a vaporization surface  1017  of the primary wick  1015  and the heated wall  1005 . The liquid flow channels  1025  are located between the liquid barrier wall  1011  and the primary wick  1015 . 
     The heated wall  1005  acts as a heat acquisition surface for a heat source. The heated wall  1005  is made from a heat-conductive material, such as, for example, sheet metal. Material chosen for the heated wall  1005  typically is able to withstand internal pressure of the working fluid. 
     The vapor removal channels  1020  are designed to balance the hydraulic resistance of the vapor removal channels  1020  with the heat conduction through the heated wall  1005  into the primary wick  1015 . The vapor removal channels  1020  can be electro-etched, machined, or formed in a surface with any other convenient method. 
     The vapor removal channels  1020  are shown as grooves in the inner side of the heated wall  1005 . However, the vapor removal channels  1020  can be designed and located in several different ways, depending on the design approach chosen. For example, according to other implementations, the vapor removal channels  1020  are grooved into an outer surface of the primary wick  1015  or embedded into the primary wick  1015  such that they are under the surface of the primary wick  1015 . The design of the vapor removal channels  1020  is selected to increase the ease and convenience of manufacturing and to closely approximate one or more of the following guidelines. 
     First, the hydraulic diameter of the vapor removal channels  1020  should be sufficient to handle a vapor flow generated on the vaporization surface  1017  of the primary wick  1015  without a significant pressure drop. Second, the surface of contact between the heated wall  1005  and the primary wick  1015  should be maximized to provide efficient heat transfer from the heat source to vaporization surface  1017  of the primary wick  1015 . Third, a thickness  1030  of the heated wall  1005 , which is in contact with the primary wick  1015 , should be minimized. As the thickness  1030  increases, vaporization at the surface of the primary wick  1015  is reduced and transport of vapor through the vapor removal channels  1020  is reduced. 
     The evaporator  1000  can be assembled from separate parts. Alternatively, the evaporator  1000  can be made as a single part by in-situ sintering of the primary wick  1015  between two walls having special mandrels to form channels on both sides of the primary wick  1015 . 
     The primary wick  1015  provides the vaporization surface  1017  and pumps or feeds the working fluid from the liquid flow channels  1025  to the vaporization surface  1017  of the primary wick  1015 . 
     The size and design of the primary wick  1015  involves several considerations. The thermal conductivity of the primary wick  1015  should be low enough to reduce heat leak from the vaporization surface  1017 , through the primary wick  1015 , and to the liquid flow channels  1025 . Heat leakage can also be affected by the linear dimensions of the primary wick  1015 . For this reason, the linear dimensions of the primary wick  1015  should be properly optimized to reduce heat leakage. For example, an increase in a thickness  1019  of the primary wick  1015  can reduce heat leakage. However, increased thickness  1019  can increase hydraulic resistance of the primary wick  1015  to the flow of the working fluid. In working LHP designs, hydraulic resistance of the working fluid due to the primary wick  1015  can be significant and a proper balancing of these factors is important. 
     The force that drives or pumps the working fluid of a heat transfer system is a temperature or pressure difference between vapor and liquid sides of a primary wick. The pressure difference is supported by the primary wick and it is maintained by proper management of the incoming working fluid thermal balance. 
     The liquid returning to the evaporator from the condenser passes through a liquid return line and is slightly subcooled. The degree of subcooling offsets the heat leak through the primary wick and the heat leak from the ambient into the reservoir within the liquid return line. The subcooling of the liquid maintains a thermal balance of the reservoir. However, there exist other useful methods to maintain thermal balance of the reservoir. 
     One method is an organized heat exchange between reservoir and the environment. For evaporators having a planar design, such as those often used for terrestrial applications, the heat transfer system includes heat exchange fins on the reservoir and/or on the liquid barrier wall  1011  of the evaporator  1000 . The forces of natural convection on these fins provide subcooling and reduce stress on the condenser and the reservoir of the heat transfer system. 
     The temperature of the reservoir or the temperature difference between the reservoir and the vaporization surface  1017  of the primary wick  1015  supports the circulation of the working fluid through the heat transfer system. Some heat transfer systems may require an additional amount of subcooling. The required amount may be greater than what the condenser can produce, even if the condenser is completely blocked. 
     In designing the evaporator  1000 , three variables need to be managed. First, the organization and design of the liquid flow channels  1025  needs to be determined. Second, the venting of the vapor from the liquid flow channels  1025  needs to be accounted for. Third, the evaporator  1000  should be designed to ensure that liquid fills the liquid flow channels  1025 . These three variables are interrelated and thus should be considered and optimized together to form an effective heat transfer system. 
     As mentioned, it is important to obtain a proper balance between the heat leak into the liquid side of the evaporator and the pumping capabilities of the primary wick. This balancing process cannot be done independently from the optimization of the condenser, which provides subcooling, because the greater heat leak allowed in the design of the evaporator, the more subcooling needs to be produced in the condenser. The longer the condenser, the greater are the hydraulic losses in a fluid line, which may require different wick material with better pumping capabilities. 
     In operation, as power from a heat source is applied to the evaporator  1000 , liquid from the liquid flow channels  1025  enters the primary wick  1015  and evaporates, forming vapor that is free to flow along the vapor removal channels  1020 . Liquid flow into the evaporator  1000  is provided by the liquid flow channels  1025 . The liquid flow channels  1025  supply the primary wick  1015  with enough liquid to replace liquid that is vaporized on the vapor side of the primary wick  1015  and to replace liquid that is vaporized on the liquid side of the primary wick  1015 . 
     The evaporator  1000  may include a secondary wick  1040 , which provides phase management on a liquid side of the evaporator  1000  and supports feeding of the primary wick  1015  in critical modes of operation (as discussed above). The secondary wick  1040  is formed between the liquid flow channels  1025  and the primary wick  1015 . The secondary wick  1040  can be a mesh screen (as shown in the  FIG. 10 ), or an advanced and complicated artery, or a slab wick structure. Additionally, the evaporator  1000  may include a vapor vent channel  1045  at an interface between the primary wick  1015  and the secondary wick  1040 . 
     Heat conduction through the primary wick  1015  may initiate vaporization of the working fluid in a wrong place, on a liquid side of the evaporator  1000  near or within the liquid flow channels  1025 . The vapor vent channel  1045  delivers the unwanted vapor away from the primary wick  1015  into the two-phase reservoir. 
     The fine pore structure of the primary wick  1015  can create a significant flow resistance for the liquid. Therefore, it is important to optimize the number, the geometry, and the design of the liquid flow channels  1025 . The goal of this optimization is to support a uniform, or close to uniform, feeding flow to the vaporization surface  1017 . Moreover, as the thickness  1019  of the primary wick  1015  is reduced, the liquid flow channels  1025  can be spaced farther apart. 
     The evaporator  1000  may require significant vapor pressure to operate with a particular working fluid within the evaporator  1000 . Use of a working fluid with a high vapor pressure can cause several problems with pressure containment of the evaporator envelope. Traditional solutions to the pressure containment problem, such as thickening the walls of the evaporator, are not always effective. For example, in planar evaporators having a significant flat area, the walls become so thick that the temperature difference is increased and the evaporator heat conductance is degraded. Additionally, even microscopic deflection of the walls due to the pressure containment results in a loss of contact between the walls and the primary wick. Such a loss of contact impacts heat transfer through the evaporator. And, microscopic deflection of the walls creates difficulties with the interfaces between the evaporator and the heat source and any external cooling equipment. 
     Annular Design 
     Referring to  FIGS. 11-13 , an annular evaporator  1100  is formed by effectively rolling the planar evaporator  1000  such that the primary wick  1015  loops back into itself and forms an annular shape. The evaporator  1100  can be used in applications in which the heat sources have a cylindrical exterior profile, or in applications where the heat source can be shaped as a cylinder. The annular shape combines the strength of a cylinder for pressure containment and the curved interface surface for best possible contact with the cylindrically shaped heat sources. 
     The evaporator  1100  includes a heated wall  1105 , a liquid barrier wall  1110 , a primary wick  1115  positioned between the heated wall  1105  and the inner side of the liquid barrier wall  1110 , vapor removal channels  1120 , and liquid flow channels  1125 . The liquid barrier wall  1110  is coaxial with the primary wick  1115  and the heated wall  1105 . 
     The heated wall  1105  intimately contacts the primary wick  1115 . The liquid barrier wall  1110  contains working fluid on an inner side of the liquid barrier wall  1110  such that the working fluid flows only along the inner side of the liquid barrier wall  1110 . The liquid barrier wall  1110  closes the evaporator&#39;s envelope and helps to organize and distribute the working fluid through the liquid flow channels  1125 . 
     The vapor removal channels  1120  are located at an interface between a vaporization surface  1117  of the primary wick  1115  and the heated wall  1105 . The liquid flow channels  1125  are located between the liquid barrier wall  1110  and the primary wick  1115 . The heated wall  1105  acts a heat acquisition surface and the vapor generated on this surface is removed by the vapor removal channels  1120 . 
     The primary wick  1115  fills the volume between the heated wall  1105  and the liquid barrier wall  1110  of the evaporator  1100  to provide reliable reverse menisci vaporization. 
     The evaporator  1100  can also be equipped with heat exchange fins  1150  that contact the liquid barrier wall  1110  to cold bias the liquid barrier wall  1110 . The liquid flow channels  1125  receive liquid from a liquid inlet  1155  and the vapor removal channels  1120  extend to and provide vapor to a vapor outlet  1160 . 
     The evaporator  1100  can be used in a heat transfer system that includes an annular reservoir  1165  adjacent the primary wick  1115 . The reservoir  1165  may be cold biased with the heat exchange fins  1150 , which extend across the reservoir  1165 . The cold biasing of the reservoir  1165  permits utilization of the entire condenser area without the need to generate subcooling at the condenser. The excessive cooling provided by cold biasing the reservoir  1165  and the evaporator  1100  compensates the parasitic heat leaks through the primary wick  1115  into the liquid side of the evaporator  1100 . 
     In another implementation, the evaporator design can be inverted and vaporization features can be placed on an outer perimeter and the liquid return features can be placed on the inner perimeter. 
     The annular shape of the evaporator  1100  may provide one or more of the following or additional advantages. First, problems with pressure containment may be reduced or eliminated in the annular evaporator  1100 . Second, the primary wick  1115  may not need to be sintered inside, thus providing more space for a more sophisticated design of the vapor and liquid sides of the primary wick  1115 . 
     Referring also to  FIGS. 14A-14H , an annular evaporator  1400  is shown having a liquid inlet  1455  and a vapor outlet  1460 . The annular evaporator  1400  includes a heated wall  1700  ( FIGS. 14G ,  14 H,  15 A, and  15 B), a liquid barrier wall  1500  ( FIGS. 14G ,  14 H, and  17 A- 17 D), a primary wick  1600  ( FIGS. 14G ,  14 H, and  16 A- 16 D) positioned between the heated wall  1700  and the inner side of the liquid barrier wall  1500 , vapor removal channels  1465  ( FIGS. 14H ,  15 A, and  15 B), and liquid flow channels  1505  ( FIGS. 14H ). The annular evaporator  1400  also includes a ring  1800  (FIGS.  14 G and  18 A- 18 D) that ensures spacing between the heated wall  1700  and the liquid barrier wall  1500  and a ring  1900  ( FIGS. 14G ,  14 H, and  19 A- 19 D) at a base of the evaporator  1400  that provides support for the liquid barrier wall  1500  and the primary wick  1600 . The heated wall  1700 , the liquid barrier wall  1500 , the ring  1800 , the ring  1900 , and the primary wick  1600  are preferably formed of stainless steel. 
     The upper portion of the evaporator  1400  (that is, above the primary wick  1600 ) includes an expansion volume  1470  ( FIG. 14H ). The liquid flow channels  1505 , which are formed in the liquid barrier wall  1500 , are fed by the liquid inlet  1455 . The primary wick  1600  separates the liquid flow channels  1505  from the vapor removal channels  1465  that lead to the vapor outlet  1460  through a vapor annulus  1475  ( FIG. 14H ) formed in the ring  1900 . The vapor removal channels  1465  may be photo-etched into the surface of the heated wall  1700 . 
     The evaporators disclosed herein can operate in any combination of materials, dimensions and arrangements, so long as they embody the features as described above. There are no restrictions other than criteria mentioned here; the evaporator can be made of any shape size and material. The only design constraints are that the applicable materials be compatible with each other and that the working fluid be selected in consideration of structural constraints, corrosion, generation of noncondensable gases, and lifetime issues. 
     Many terrestrial applications can incorporate an LHP with an annular evaporator  1100 . The orientation of the annular evaporator in a gravity field is predetermined by the nature of application and the shape of the hot surface. 
     Cyclical Heat Exchange System 
     Cyclical heat exchange systems may be configured with one or more heat transfer systems to control a temperature at a region of the heat exchange system. The cyclical heat exchange system may be any system that operates using a thermodynamic cycle, such as, for example, a cyclical heat exchange system, a Stirling heat exchange system (also known as a Stirling engine), or an air conditioning system. 
     Referring to  FIG. 20 , a Stirling heat exchange system  2000  utilizes a known type of environmentally friendly and efficient refrigeration cycle. The Stirling system  2000  functions by directing a working fluid (for example, helium) through four repetitive operations; that is, a heat addition operation at constant temperature, a constant volume heat rejection operation, a constant temperature heat rejection operation and a heat addition operation at constant volume. 
     The Stirling system  2000  is designed as a Free Piston Stirling Cooler (FPSC), such as Global Cooling&#39;s model M100B (Available from Global Cooling Manufacturing, 94 N. Columbus Rd., Athens, Ohio). The FPSC  2000  includes a linear motor portion  2005  housing a linear motor (not shown) that receives an AC power input  2010 . The FPSC  2000  includes a heat acceptor  2015 , a regenerator  2020 , and a heat rejecter  2025 . The FPSC  2000  includes a balance mass  2030  coupled to the body of the linear motor within the linear motor portion  2005  to absorb vibrations during operation of the FPSC  2000 . The FPSC  2000  also includes a charge port  2035 . The FPSC  2000  includes internal components, such as those shown in the FPSC  2100  of  FIG. 21 . 
     The FPSC  2100  includes a linear motor  2105  housed within the linear motor portion  2110 . The linear motor portion  2110  houses a piston  2115  that is coupled to flat springs  2120  at one end and a displacer  2125  at another end. The displacer  2125  couples to an expansion space  2130  and a compression space  2135  that form, respectively, cold and hot sides. The heat acceptor  2015  is mounted to the cold side of the expansion space  2130  and the heat rejector is mounted to the hot side of the compression space  2135 . The FPSC  2100  also includes a balance mass  2140  coupled to the linear motor portion  2110  to absorb vibrations during operation of the FPSC  2100 . 
     Referring also to  FIG. 22 , in one implementation, a FPSC  2200  includes heat rejector  2205  made of a copper sleeve and a heat acceptor  2210  made of a copper sleeve. The heat rejector  2205  has an outer diameter (OD) of approximately 100 mm and a width of approximately 53 mm to provide a 166 cm 2  heat rejection surface capable of providing a flux of −6 W/cm 2  when operating in a temperature range of 20° C. to 70° C. The heat acceptor  2210  has an OD of approximately 100 mm and a width of approximately 37 mm to provide a 115 cm 2  heat accepting surface capable of providing a flux of −5.2 W/cm 2  in a temperature range of −30° C. to 5° C. 
     Briefly, in operation an FPSC is filled with a coolant (such as, for example, helium gas) that is shuttled back and forth by combined movements of the piston and the displacer. In an ideal system, thermal energy is rejected to the environment through the heat rejector while the coolant is compressed by the piston and thermal energy is extracted from the environment through the heat acceptor while the coolant expands. 
     Referring to  FIG. 23 , a thermodynamic system  2300  includes a cyclical heat exchange system such as a cyclical heat exchange system  2305  (for example, the systems  2000 ,  2100 ,  2200 ) and a heat transfer system  2310  thermally coupled to a portion  2315  of the cyclical heat exchange system  2305 . The cyclical heat exchange system  2305  is cylindrical and the heat transfer system  2310  is shaped to surround the portion  2315  of the cyclical heat exchange system  2305  to reject heat from the portion  2315 . In this implementation, the portion  2315  is the hot side (that is, the heat rejector) of the cyclical heat exchange system  2305 . The thermodynamic system  2300  also includes a fan  2320  positioned at the hot side of the cyclical heat exchange system  2305  to force air over a condenser of the heat transfer system  2310  and thus to provide additional convection cooling. 
     A cold side  2335  (that is, the heat acceptor) of the cyclical heat exchange system  2305  is thermally coupled to a CO 2  refluxer  2340  of a thermosyphon  2345 . The thermosyphon  2345  includes a cold-side heat exchanger  2350  that is configured to cool air within the thermodynamic system  2300  that is forced across the heat exchanger  2350  by a fan  2355 . 
     Referring to  FIG. 24 , in another implementation, a thermodynamic system  2400  includes a cyclical heat exchange system such as a cyclical heat exchange system  2405  (for example, the systems  2000 ,  2100 ,  2200 ) and a heat transfer system  2410  thermally coupled to a hot side  2415  of the cyclical heat exchange system  2405 . The thermodynamic system  2400  includes a heat transfer system  2420  thermally coupled to a cold side  2425  of the cyclical heat exchange system  2405 . The thermodynamic system  2400  also includes fans  2430 ,  2435 . The fan  2430  is positioned at the hot side  2415  of the thermodynamic system to force air through a condenser of the heat transfer system  2410 . The fan  2435  is positioned at the cold side  2425  of the thermodynamic system  2400  to force air through a condenser of the heat transfer system  2420 . 
     Referring to  FIG. 25 , in one implementation, a thermodynamic system  2500  includes a heat transfer system  2505  coupled to a cyclical heat exchange system such as a cyclical heat exchange system  2510 . The heat transfer system  2505  is used to cool a hot side  2515  of the cyclical heat exchange system  2510 . The heat transfer system  2505  includes an annular evaporator  2520  that includes an expansion volume (or reservoir)  2525 , a liquid return line  2530  providing fluid communication between liquid outlets  2535  of a condenser  2540  and a liquid inlet of the evaporator  2520 . The heat transfer system  2505  also includes a vapor line  2545  providing fluid communication between a vapor outlet of the evaporator  2520  and vapor inlets  2550  of the condenser  2540 . 
     The condenser  2540  is constructed from smooth-wall tubing and is equipped with heat exchange fins  2555  or fin stock to intensify heat exchange on the outside of the tubing. 
     The evaporator  2520  includes a primary wick  2560  sandwiched between a heated wall  2565  and a liquid barrier wall  2570  and separating the liquid and the vapor. The liquid barrier wall  2570  is cold-biased by heat exchange fins  2575  formed along the outer surface of the heated wall  2565 . The heat exchange fins  2575  provide subcooling for the reservoir  2525  and the entire liquid side of the evaporator  2520 . The heat exchange fins  2575  of the evaporator  2520  may be designed separately from the heat exchange fins  2555  of the condenser  2540 . 
     The liquid return line  2530  extends into the reservoir  2525  located above the primary wick  2560 , and vapor bubbles, if any, from the liquid return line  2530  and the vapor removal channels at the interface of the primary wick  2560  and the heated wall  2565  are vented into the reservoir  2525 . Typical working fluids for the heat transfer system  2505  include (but are not limited to) methanol, butane, CO 2 , propylene, and ammonia. 
     The evaporator  2520  is attached to the hot side  2515  of the cyclical heat exchange system  2510 . In one implementation, this attachment is integral in that the evaporator  2520  is an integral part of the cyclical heat exchange system  2510 . In another implementation, attachment can be non-integral in that the evaporator  2520  can be clamped to an outer surface of the hot side  2510 . The heat transfer system  2505  is cooled by a forced convection sink, which can be provided by a simple fan  2580 . Alternatively, the heat transfer system  2505  is cooled by a natural or draft convection. 
     Initially, the liquid phase of the working fluid is collected in a lower part of the evaporator  2520 , the liquid return line  2530 , and the condenser  2540 . The primary wick  2560  is wet because of capillary forces. As soon as heat is applied (for example, the cyclical heat exchange system  2510  is turned on), the primary wick  2560  begins to generate vapor, which travels through vapor removal channels (similar to vapor removal channels  1120  of evaporator  1100 ) of the evaporator  2520 , through the vapor outlet of the evaporator  2520 , and into the vapor line  2545 . 
     The vapor then enters the condenser  2540  at an upper part of the condenser  2540 . The condenser  2540  condenses the vapor into liquid and the liquid is collected at a lower part of the condenser  2540 . The liquid is pushed into the reservoir  2525  because of the pressure difference between the reservoir  2525  and the lower part of the condenser  2540 . Liquid from the reservoir  2525  enters liquid flow channels of the evaporator  2520 . The liquid flow channels of the evaporator  2520  are configured like the vapor removal channels  1125  of the evaporator  1100  and are properly sized and located to provide adequate liquid replacement for the liquid that vaporized. Capillary pressure created by the primary wick  2560  is sufficient to withstand the overall LHP pressure drop and to prevent vapor bubbles from travelling through the primary wick  2560  toward the liquid flow channels. 
     The liquid flow channels of the evaporator  2520  can be replaced by a simple annulus, if the cold biasing discussed above is sufficient to compensate the increased heat leak across the primary wick  2560 , which is caused by the increase in surface area of the heat exchange surface of annulus versus the surface area of the liquid flow channels. 
     Referring to  FIGS. 26-28 , a heat transfer system  2600  includes an evaporator  2605  coupled to a cyclical heat exchange system  2610  and an expansion volume  2615  coupled to the evaporator  2605 . The vapor channels of the evaporator  2605  feed to a vapor line  2620  that feed a series of channels  2625  of a condenser  2630 . The condensed liquid from the condenser  2630  is collected in a liquid return channel  2635 . The heat transfer system  2600  also includes fin stock  2640  thermally coupled to the condenser  2630 . 
     The evaporator  2605  includes a heated wall  2700 , a liquid barrier wall  2705 , a primary wick  2710  positioned between the heated wall  2700  and an inner side of the liquid barrier wall  2705 , vapor removal channels  2715 , and liquid flow channels  2720 . The liquid barrier wall  2705  is coaxial with the primary wick  2710  and the heated wall  2700 . The liquid flow channels  2720  are fed by a liquid return channel  2725  and the vapor removal channels  2715  feed into a vapor outlet  2730 . 
     The heated wall  2700  intimately contacts the primary wick  2710 . The liquid barrier wall  2705  contains working fluid on an inner side of the liquid barrier wall  2705  such that the working fluid flows only along the inner side of the liquid barrier wall  2705 . The liquid barrier wall  2705  closes the evaporator&#39;s envelope and helps to organize and distribute the working fluid through the liquid flow channels  2720 . 
     In one implementation, the evaporator  2605  is approximately 2″ tall and the expansion volume  2615  is approximately 1″ in height. The evaporator  2605  and the expansion volume  2615  are wrapped around a portion of the cyclical heat exchange system  2610  having a 4″ outer diameter. The vapor line  2620  has a radius of ⅛″. The cyclical heat exchange system  2610  includes approximately 58 condenser channels  2625 , with each condenser channel  2625  having a length of 2″ and a radius of 0.012″, the channels  2625  being spread out such that the width of the condenser  2630  is approximate 40″. The liquid return channel  2725  has a radius of 1/16″. The heat exchanger  2800  (which includes the condenser  2630  and the fin stock  2640  is approximately 40″ long and is wrapped into an inner and outer loop (see  FIGS. 30 ,  33 , and  34 ) to produce a cylindrical heat exchanger having an outer diameter of approximately 8″. The evaporator  2605  has a cross-sectional width  2750  of approximately ⅛″, as defined by the heated wall  2700  and the liquid barrier wall  2705 . The vapor removal channels  2715  have widths of approximately 0.020″ and depths of approximately 0.020″ and are separated from each other by approximately 0.020″ to produce 25 channels per inch. 
     As mentioned above, the heat transfer system (such as system  2310 ) is thermally coupled to the portion (such as portion  2315 ) of the cyclical heat exchange system. The thermal coupling between the heat transfer system and the portion can be by any suitable method. In one implementation, if the evaporator of the heat transfer system is thermally coupled to the hot side of the cyclical heat exchange system, the evaporator may surround and contact the hot side and the thermal coupling may be enabled by a thermal grease compound applied between the hot side and the evaporator. In another implementation, if the evaporator of the heat transfer system is thermally coupled to the hot side of the cyclical heat exchange system, the evaporator may be constructed integrally with the hot side of the cyclical heat exchange system by forming vapor channels directly into the hot side of the cyclical heat exchange system. 
     Referring to  FIGS. 30-32 , a heat transfer system  3000  is packaged around a cyclical heat exchange system  3005 . The heat transfer system  3000  includes a condenser  3010  surrounding an evaporator  3015 . Working fluid that has been vaporized exits the evaporator  3015  through a vapor outlet  3020  connected to the condenser  3010 . The condenser  3010  loops around and doubles back inside itself at junction  3025 . 
     The cyclical heat exchange system  3005  is surrounded about its heat rejection surface  3100  by the evaporator  3015 . The evaporator  3015  is in intimate contact with the heat rejection surface  3100 . The refrigeration assembly (which is the combination of the cyclical heat exchange system  3005  and the heat transfer system  3000 ) is mounted in a tube  3205 , with a fan  3210  mounted at the end of the tube  3205  to force air through fins  3030  of the condenser  3010  to exhaust channels  3035 . 
     The evaporator  3015  has a wick  3215  in which working fluid absorbs heat from the heat rejection surface  3100  and changes phase from liquid to vapor. The heat transfer system  3000  includes a reservoir  3220  at the top of the evaporator  3015  that provides an expansion volume. For simplicity of illustration, the evaporator  3015  has been illustrated in this view as a simple hatched block that shows no internal detail. Such internal details are discussed elsewhere in this description. 
     The vaporized working fluid exits the evaporator  3015  through the vapor outlet  3020  and enters a vapor line  3040  of the condenser  3010 . The working fluid flows downward from the vapor line  3040 , through channels  3045  of the condenser  3010 , to a liquid return line  3050 . As the working fluid flows through the channels  3045  of the condenser  3010  it loses heat, through the fins  3030  to the air passing between the fins, to change phase from vapor to liquid. Air that has passed through the fins  3030  of the condenser  3010  flows away through the exhaust channel  3035 . Liquefied working fluid (and possibly some uncondensed vapor) flows from the liquid return line  3050  back into the evaporator  3015  through the liquid return port  3055 . 
     Referring to  FIGS. 33 and 34 , a heat transport system  3300  surrounds a portion of a cyclical heat exchange system  3302  that is surrounded, in turn, by exhaust channels  3305 . The heat transport system  3300  includes an evaporator  3310  having an upper portion that surrounds the cyclical heat exchange system  3302 . A vapor port  3315  connects the evaporator  3310  to a vapor line  3312  of a condenser  3320 . The vapor line  3312  includes an outer region that circles around the evaporator  3310  and then doubles back on itself at junction  3325  to form an inner region that circles back around the evaporator  3310  in the opposite direction. The heat transport system  3300  also includes cooling fins  3330  on the condenser  3320 . 
     The heat transport system  3300  also includes a liquid return port  3400  that provides a path for condensed working fluid from a liquid line  3405  of the condenser  3320  to return to the evaporator  3310 . 
     As mentioned above, the interface between the evaporator  3310  and the heat rejection surface of the cyclical heat exchange system  3302  may be implemented according one of several alternative implementations. 
     Referring to  FIG. 35 , in one implementation, an evaporator  3500  slips over a heat rejection surface  3502  of a cyclical heat exchange system  3505 . The evaporator  3500  includes a heated wall  3510 , a liquid barrier wall  3515 , and a wick  3520  sandwiched between the heated wall  3510  and the liquid barrier wall  3515 . The wick  3520  is equipped with vapor channels  3525  and liquid flow channels  3530  are formed at the liquid barrier wall  3515  in simplified form for clarity. 
     The evaporator  3500  is slipped over the cyclical heat exchange system  3505  and may be held in place with the use of a clamp  3600  (shown in  FIG. 36 ). To aid heat transfer, thermally conductive grease  3535  is disposed between the cyclical heat exchange system  3505  and heated wall  3510  of the evaporator  3500 . In an alternative implementation, the vapor channels  3525  are formed in the heated wall  3510  instead of in the wick  3520 . 
     Referring to  FIG. 37 , in another implementation, an evaporator  3700  is fit over a heat rejection surface  3702  of a cyclical heat exchange system  3705  with an interference fit. The evaporator  3700  includes a heated wall  3710 , a liquid barrier wall  3715 , and a wick  3720  sandwiched between the heated wall  3710  and the liquid barrier wall  3715 . The evaporator  3700  is sized to have an interference fit with the heat rejection surface  3702  of the cyclical heat exchange system  3705 . 
     The evaporator  3700  is heated so that its inner diameter expands to permit it to slip over the unheated heat rejection surface  3702 . As the evaporator  3700  cools, it contracts to fix onto the cyclical heat exchange system  3705  in an interference fit relationship. Because of the tightness of the fit, no thermally conductive grease is needed to enhance heat transfer. The wick  3720  is equipped with vapor channels  3725 . In an alternative implementation, the vapor channels are formed in the heated wall  3710  instead of in the wick  3720 . Liquid flow channels  3730  are formed at the liquid barrier wall  3715  in a simplified form for clarity. 
     Referring to  FIG. 38 , in another implementation, an evaporator  3800  is fit over a heat rejection surface  3802  of a cyclical heat exchange system  3805  and features previously designed within the evaporator  3800  are now integrally formed within the heat rejection surface  3802 . In particular, the evaporator  3800  and the heat rejection surface  3802  are constructed together as an integrated assembly. The heat rejection surface  3802  is modified to have vapor channels  3825 ; in this way, the heat rejection surface  3802  acts as a heated wall for the evaporator  3800 . 
     The evaporator  3800  includes a wick  3820  and a liquid barrier wall  3815  formed about the modified heat rejection surface  3802 , the wick  3820  and the liquid barrier wall  3815  being integrally bonded to the heat rejection surface  3802  to form the sealed evaporator  3800 . Liquid flow channels  3830  are portrayed in a simplified form for clarity. In this way, a hybrid cyclical heat exchange system with an integrated evaporator is formed. This integral construction provides enhanced thermal performance in comparison to the clamp-on construction and the interference fit construction because thermal resistance is reduced between the cyclical heat exchange system  3805  and the wick  3820  of the evaporator  3800 . 
     Referring to  FIG. 29 , graphs  2900  and  2905  show the relationship between a maximum temperature of the surface of the portion of the cyclical heat exchange system that is to be cooled by the heat transfer system and a surface area of the interface between the heat transfer system and the portion of the cyclical heat exchange system to be cooled. The maximum temperature indicates the maximum amount of heat rejection. In graph  2900 , the interface between the portion and the heat transfer system is accomplished with a thermal grease compound. In graph  2905 , the heat transfer system is made integral with the portion. 
     As shown, at an air flow of 300 CFM, if the interface is a thermal grease interface, then the maximum amount of heat rejection would fall within a maximum heat rejection surface temperature  2907  (for example, 70° C.) with a heat exchange surface area  2910  (for example, 100 ft 2 ). When the evaporator is constructed integrally with the portion by forming vapor channels directly in the heat rejection surface, that heat rejection surface would operate below the maximum heat rejection surface temperature of the thermal grease interface with significantly smaller heat exchange surface areas. 
     Referring to  FIG. 39 , a condenser  3900  is formed with fins  3905 , which provide thermal communication between the air or the environment and a vapor line  3910  of the condenser  3900 . The vapor line  3910  couples to a vapor outlet  3915  that connects an evaporator  3920  positioned within the condenser  3900 . 
     Referring to  FIGS. 40-43 , in one implementation, the condenser  3900  is laminated and is formed with flow channels that extend through a flat plate  4000  of the condenser  3900  between a vapor head  3925  and a liquid head  3930 . Copper is a suitable material for use in making a laminated condenser. The laminated structure condenser  3900  includes a base  4200  having fluid flow channels  4205  (shown in phantom) formed therein and a top layer  4210  is bonded to the base  4200  to cover and seal the fluid flow channels  4205 . The fluid flow channels  4205  are designed as trenches formed in the base  4200  and sealed beneath the top layer  4210 . The trenches for the fluid flow channels  4205  may be formed by chemical etching, electrochemical etching, mechanical machining, or electrical discharge machining processes. 
     Referring to  FIGS. 44 and 45 , in another implementation, the condenser  3900  is extruded and small flow channels  4400  extend through a flat plate  4405  of the condenser  3900 . Aluminum is a suitable material for use in such an extruded condenser. The extruded micro channel flat plate  4405  extends between a vapor header  4410  and a liquid header  4415 . Moreover, corrugated fin stock  4420  is bonded (for example, brazed or epoxied) to both sides of the flat plate  4405 . 
     Referring to  FIG. 46 , a cross-sectional view of one side of a heat transfer system  4600  that is coupled to a cyclical heat exchange system  4605 . This view shows relative dimensions that provide for particularly compact packaging of the heat transfer system. In this view, fins  4610  are portrayed as being 90 degrees out of phase for ease of illustration. To cool heat rejection surface  4615  of the cyclical heat exchange system  4605  having a 4-inch diameter, the evaporator  4620  has a thickness of 0.25 inch and the radial thickness of the condenser is 1.75 inches. This provides on overall dimension for the packaging (the combination of the heat transfer system  4600  and the cyclical heat exchange system  4605  of 8 inches. 
     As discussed, the evaporator used in the heat transfer system is equipped with a wick. Because a wick is employed within the evaporator of the heat transfer system, the condenser may be positioned at any location relative to the evaporator and relative to gravity. For example, the condenser may be positioned above the evaporator (relative to a gravitational pull), below the evaporator (relative to a gravitational pull), or adjacent the evaporator, thus experiencing the same gravitational pull as the evaporator. 
     Other implementations are within the scope of the following claims. 
     Notably, the terms Stirling engine, Stirling heat exchange system, and Free Piston Stirling Cooler have been referenced in several implementations above. However, the features and principals described with respect to those implementations also may be applied to other engines capable of conversions between mechanical energy and thermal energy. 
     Moreover, the features and principals described above may be applied to any heat engine, which is a thermodynamic system that can undergo a cycle, that is, a sequence of transformations which ultimately return it to its original state. If every transformation in the cycle is reversible, the cycle is reversible and the heat transfers occur in the opposite direction and the amount of work done switches sign. The simplest reversible cycle is a Carnot cycle, which exchanges heat with two heat reservoirs.