Abstract:
A heat sink assembly for cooling an electronic device comprises a fan housed in a shroud, the fan including a hub and fan blades extending therefrom for causing an axially directed airflow through the shroud upon rotation of the fan blades. A thermosiphon comprises an evaporator defining an evaporating chamber containing a working fluid therein and further including a condenser mounted thereabove. The thermosiphon is positioned at one end of the shroud such that the fan is aligned with the condenser for directing the axial airflow therethrough. The condenser includes a plurality of tubes forming a tube grouping. Each tube having an opening in fluid communication with the evaporator and for receiving and condensing vapor of the working fluid received from the evaporator. The tubes are axially aligned with the airflow and are laterally positioned such that a lateral width of the tube grouping is approximately equal to a width of the hub and substantially in lateral alignment therewith.

Description:
TECHNICAL FIELD 
     The present invention relates to heat sinks in general, and more particularly to heat sinks for use in dissipating waste heat generated by electrical or electronic components and assemblies. 
     BACKGROUND OF THE INVENTION 
     Research activities have focused on developing heat sinks to efficiently dissipate heat from highly concentrated heat sources such as microprocessors and computer chips. These heat sources typically have power densities in the range of about 5 to 35 W/cm 2  (4 to 31 Btu/ft 2 s) and relatively small available space for placement of fans, heat exchangers, heat sinks and the like. 
     At the component level, various types of heat exchangers and heat sinks have been used that apply natural or forced convection or other cooling methods. The most commonly existing heat sinks for microelectronics cooling have generally used air to directly remove heat from the heat source. However, air has a relatively low heat capacity. Such heat sinks are suitable for removing heat from relatively low power heat sources with power density in the range of 5 to 15 W/cm 2  (4 to 13 Btu/ft 2 s). Increases in computing speed resulted in corresponding increases in the power density of the heat sources in the order of 20 to 35 W/cm 2  (18 to 31 Btu/ft 2 s) thus requiring more effective heat sinks. Liquid-cooled heat sinks employing high heat capacity fluids like water and water-glycol solutions are more particularly suited to remove heat from these types of high power density heat sources. One type of liquid cooled heat sink circulates the cooling liquid so that the liquid removes heat from the heat source and is then transferred to a remote location where the heat is easily dissipated into a flowing air stream with the use of a liquid-to-air heat exchanger. These types of heat sinks are characterized as indirect heat sinks. 
     As computing speeds continue to increase even more dramatically, the corresponding power densities of the devices rise up to 100 W/cm 2 . The constraints of the necessary cooling system miniaturization coupled with high heat flux calls for extremely efficient, compact, simple and reliable heat sinks such as a thermosiphon. A typical thermosiphon comprises an evaporating section and a condensing section. The heat-generating device is mounted to the evaporating section. In some thermosiphons, the heat-generating device is affixed to the internal surface of the evaporating section where it is submerged in the working fluid. Alternatively, the heat-generating device can also be affixed to the external surface of the evaporating section. The working fluid of the thermosiphon is generally a halocarbon fluid, which circulates in a closed-loop fashion between the evaporating and condensing sections. The captive working fluid changes its state from liquid-to-vapor in the evaporating section as it absorbs heat from the heat-generating device. Reverse transformation of the working fluid from vapor-to-liquid occurs as it rejects heat to a cooling fluid like air flowing on an external finned surface of the condensing section. The thermosiphon relies exclusively on gravity for the motion of the working fluid between the evaporating and condensing sections. As for the motion of the cooling fluid on the external surface of the condensing section, a fluid moving device like an axial fan is employed. 
     Most electronics devices have high degree of non-uniformity built into them. Thermal management of these devices is subject to two constraints that the thermal engineer must address. First, the heat flux generated by the electronics device is highly non-uniform. Second, the air circulated by the air-moving device like an axial fan is very non-uniformly distributed. Most computer chips have their heat generation concentrated in a very small region in the core of the chip. For example, a typical 40×40 mm 2  computer chip has almost 80% of its total heat flux concentrated in its central 10×10 mm 2  surface. The heat flux distribution in a typical electronics device is shown schematically in FIG.  4 . The second non-uniformity is attributed to the attachment of the air-moving device like an axial fan attached to the exterior of the thermosiphon. Axial fan has a large hub which acts as blockage to airflow. The airflow exiting the axial fan is highly concentrated in the peripheral region of the fan blades as shown in FIG.  5 . The maximum air velocity is in the tip region of fan blades. The velocity falls off sharply to zero in the central hub region. Under certain flow conditions and blade angle, the local velocity at the root of the fan blade may even become negative, i.e., opposite to the direction of the predominant airflow. 
     The non-uniformity of airflow is far more pronounced in push mode wherein the fan blows relatively cooler ambient air into the heat exchanger. In pull mode, on the other hand, the fan sucks relatively hotter air from the heat exchanger. For a high heat load push mode is advantageous when airflow rate is low. In order to attain flatter airflow profile entering the heat exchanger face a standoff distance of at least three times the hub diameter is preferable between the fan and the heat exchanger. However, because of packaging constraints only about one-fifth to one-quarter of the hub diameter standoff distance is typically available between the fan and heat exchanger. This is because the airflow at the heat exchanger face is non-uniform. 
     A limitation of the axial fan relating to smallness of the pressure rise across the fan needs to be borne in mind. The curve of the pressure head developed by the fan falls off very rapidly as the volumetric flow rate of air increases. In other words, the air exiting an axial fan cannot sustain a high-pressure drop through the fins. Therefore, managing the airflow through the heat sink at a low-pressure drop is a very important consideration in the design of a thermosiphon. 
     It is apparent from the foregoing considerations that from a system&#39;s point of view, the computer chip, heat sink and fan assembly are constrained not only by very non-uniform heat flux but also by non-uniform airflow capable of sustaining small pressure drop across the heat exchanger. Ideally, the airflow should be high in regions of high heat flux and low in regions of low heat flux. Overlaying FIGS. 4 and 5 in push mode clearly reveals that the airflow distribution is opposite to that ideally desired for better heat transfer. This is detrimental to the functioning of a computer chip, as the chip junction temperature becomes high because of inadequate heat removal locally from the core of the chip. The thermal performance penalty attributed to these non-uniformities can be of the order of 25 to 50% compared to the case with uniform heat flux and uniform airflow. Thus thermal solution becomes considerably more challenging when the heat flux as well as the airflow is non-uniform. The difficulty is compounded when the available airflow rate is small. Therefore, careful attention must be paid to the fluid flow and heat transfer boundary conditions when developing the thermal solutions for the computer chips. 
     The compact thermosiphons intended to fit in a computer case require boiling and condensing processes to occur in close proximity to each other thereby imposing conflicting thermal conditions in a relatively small volume. This poses significant challenges to the process of optimizing the thermosiphon performance. 
     Thus, what is desired is a thermosiphon optimization process to intensify the processes of boiling, condensation and convective heat transfer at the external surface of the condenser while maintaining low airside pressure drop. 
     SUMMARY OF THE INVENTION 
     One aspect of the present invention is a heat sink assembly for cooling an electronic device. The heat sink assembly comprises a fan housed in a shroud, the fan having a hub and fan blades extending therefrom for causing an axially directed airflow through the shroud upon rotation of the fan blades. A thermosiphon is positioned at one end of the shroud such that the fan is aligned with the condenser for directing the axial airflow therethrough. The thermosiphon comprises an evaporator defining an evaporating chamber containing a working fluid therein and further including a condenser mounted thereabove. The condenser includes a plurality of tubes forming a tube grouping. Each tube having an opening in fluid communication with the evaporator and for receiving and condensing vapor of the working fluid received from the evaporator. The tubes are axially aligned with the airflow and are laterally positioned such that a lateral width of the tube grouping is approximately equal to a width of the hub and substantially in lateral alignment thereto. 
     Another aspect of the present invention is a condenser for a heat sink assembly for cooling an electronic device. The heat sink assembly comprises a base having an upper housing affixed thereto wherein the upper housing has open ends. A fan is mounted at one of the open ends, the fan having a hub and fan blades extending therefrom for causing an axially directed airflow through the housing upon rotation of the fan blades. A plurality of tubes is positioned within the housing for transmitting therethrough a vapor of a working fluid. The tubes define a tube grouping such that the tubes are arranged in axial alignment with the fan and laterally positioned such that a lateral width of the tube grouping is approximately equal to a width of the hub and substantially in lateral alignment thereto. 
     These and other advantages of the invention will be further understood and appreciated by those skilled in the art by reference to the following written specification, claims and appended drawings. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a perspective view of a heat sink assembly embodying the present invention, wherein an axial fan is arranged to draw cooling air through a thermosiphon. 
     FIG. 2 is an elevational cross-section view of the thermosiphon shown in FIG.  1  and taken along the line  2 — 2   
     FIG. 3 is an enlarged segment of the cross-sectional view of the boilerplate shown in FIG.  2 . 
     FIG. 4 is a typical heat flux distribution of an electronic device requiring cooling. 
     FIG. 5A is a typical air velocity distribution just downstream of the axial fan used in conjunction with a thermosiphon in push mode. 
     FIG. 5B is a typical air velocity distribution just upstream of the axial fan used in conjunction with a thermosiphon in pull mode 
     FIG. 6 is an elevational cross-section view of a second embodiment thermosiphon. 
     FIG. 7 is an elevational cross-section view of a third embodiment thermosiphon. 
     FIG. 8 is an elevational cross-section view of a fourth embodiment thermosiphon. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
     For purposes of description herein, the terms “upper”, “lower”, “left”, “rear”, “right”, “front”, “vertical”, “horizontal”, and derivatives thereof shall relate to the invention as oriented in FIG.  2 . However, it is to be understood that the invention may assume various alternative orientations and step sequences, except where expressly specified to the contrary. It is also to be understood that the specific devices and processes illustrated in the attached drawings, and described in the following specification, are simply exemplary embodiments of the inventive concepts defined in the appended claims. Hence, specific dimensions and other physical characteristics relating to the embodiments disclosed herein are not to be considered as limiting, unless the claims expressly state otherwise. 
     Turning to the drawings, FIG. 1 shows an air-cooled thermosiphon heat sink  10 , which is one of the preferred embodiments of the present invention and illustrates its various components. 
     As illustrated in FIG. 1 a single axial fan  14  is housed in shroud  16  and coupled to thermosiphon  12  through duct  18 . The fan  14  could be a pull or push type fan, however, a pull type of fan is preferred to minimize shadowing of the thermosiphon  12  by the fan hub  15 . The shadowing effect of hub  15  occurs over a lateral width  55  denoted by dimension “H” and substantially at a center of thermosiphon  12 . The shadowing effect of hub  15  is greater with a push type fan than a pull type fan and reduces the airflow behind the hub and thereby interferes with the heat transfer from thermosiphon  12  to the cooling air stream. 
     Although axial fan  14  is configured as a pull type fan, and thereby minimizes the shadowing effect, FIGS. 4,  5 A and  5 B illustrate the differences between the heat distribution of the device  8  to be cooled and the areas of maximum airflow of fan  14 . As shown in FIG. 4, the heat distribution of device  8  approximates a bell curve with the greatest heat at the area above the center of device  8 . Conversely, the area of maximum airflow in push mode, as illustrated in FIG. 5A, appears as an inverse of the heat distribution, namely minimal airflow in the middle and maximum airflow at the outmost portion of the fan. In like manner, FIG. 5B illustrates the airflow in pull mode being similar to the push mode airflow illustrated in FIG.  5 A. Therefore, without any enhancements, the fan generates maximum airflow over the minimum heat regions and low airflow over the regions of maximum heat. 
     FIG. 2 shows a sectional view of a preferred embodiment of the thermosiphon  12 . Thermosiphon  12  comprises an evaporator  20  and a condenser  22  mounted thereabove. 
     The evaporator  20  comprises a baseplate  26  having a thickness  25  denoted by dimension “t” and sidewalls  24  about a periphery of baseplate  26 . The thickness  25  “t” of the evaporator base plate  26  is suitably chosen based on an analysis of the particular boiling and heat transfer considerations for a desired application. Electronic device  8  having a mean width  9  denoted by dimension “z” is attached to a bottom surface  27  of baseplate  26  using a heat conductive adhesive, also known as “thermal grease”. Bottom surface  27  is preferably polished for attachment of electronic device  8  to enhance the thermal contact from device  8  to baseplate  26 . 
     An upper surface of baseplate  26  defines boiling surface  31  and can have a plurality of stud fins  28  formed thereon. Stud fins  28  are preferably machined as an integral part of baseplate  26  for maximum heat transfer. The boiling surface  31  of baseplate  26  can also have a surface coating  30  deposited thereon to enhance the boiling properties of boiling surface  31 . Surface coating  30  can comprise a sintered metal powder of aluminum or copper. 
     Sidewalls  24  have a height  37  denoted by dimension “h” and have a bottom affixed to baseplate  26 . Sidewalls  24  can also be integrally formed with base  26  as a single structure. An upper surface of sidewalls  24  defines an upper horizontal flange  29  about the periphery of evaporator  20  to which the base  32  of condenser  22  is attached thereby defining evaporating chamber  36 . The height of evaporating chamber  36  is also represented by dimension “h”. Base  32  is preferably affixed to flange  29  by one of brazing, welding or diffusion bonding to form a leakproof chamber  36 . The flange joint between the sidewalls  24  and base  32  can be enhanced by means of a trunion groove type mating of the protruding and recessed side of the flange prior to brazing or welding. A good joint can also be enhanced by means of peripheral screws (not shown) fastening base  32  to sidewalls  24 . The screws provide additional reinforcement and prevent leakage at high pressure. Evaporation chamber  36  is charged with a working fluid  38  through charging port  40  in base  32 . Chamber  36  also functions as a manifold to distribute saturated or super-heated vapor into the hairpin condenser tubes  44 . 
     Condenser  22  comprises base  32  and two hairpin condenser tubes  44 . Hairpin tubes  44  are formed in an inverted “U” shape wherein each leg thereof has a respective inlet end  43  extending through base  32  into evaporation chamber  36 . Inlet ends  43  are open and place an interior of tubes  44  in fluid communication with evaporation chamber  36 . In this manner, working fluid vapor formed as a result of boiling on the boiling surface  31  can enter either end of hairpin tubes  44  and rise therein for the ultimate dissipation of heat. Each of hairpin tubes  44  has a width  45  denoted by the dimension “a”; a bend radius  48  at an upper end thereof denoted by the dimension “R”; and is positioned above the area of high heat flux {dot over (q)}″ of device  8 . Radius  48  (R) is selected such that tubes  44  and their respective legs form a tube grouping behind fan hub  15  within hub  15  diameter  55  as denoted by dimension “H”. Thus, hairpin tubes  44  reside in the wake of hub  15  in the middle of the thermosiphon  12 , and serve primarily as conduits for vapor flow between the evaporator  20  and the condenser  22 . 
     Tubes  44  have a minimal lateral tube spacing  46  denoted by dimension “e”. The properties of base  32 , and the minimum distance permissible for forming slots to receive the tube ends therein govern tube spacing  46 . The spacing  46  (e) between the tubes  44  serves as a high aspect ratio rectangular duct  60  of cross section e×D where D is the depth of tubes along the direction of the axial airflow through thermosiphon  12 . Central duct  60  has a low airside pressure drop compared to fins and high heat transfer coefficient approaching that of two infinite parallel plates. The airflow through the central duct  60  serves to condense some of the vapor on the bare side of tubes  44  though most of the condensation is skewed on the finned side of tubes  44 . 
     Two types of fins are used in condenser  22  of thermosiphon  12 . First fins  50  having a height  51 , denoted by dimension “q”, are intentionally placed outside of tubes  44  and are substantially inline with the fan blades of axial fan  14  where the airflow is high. Second fins  52  having a height  53 , denoted by dimension “p”, are placed in the low flow region directly behind the hub  15  between the legs of each hairpin tube  44 . Fins  50  and  52  are generally of a convoluted accordion configuration and have their apexes bonded to the surface of tubes  44  or housing  34  they contact. Housing  34  encases tubes  44 , first fins  50  and second fins  52  to direct and maintain the airflow from fan  14  thereover. 
     The preferred working fluid of thermosiphon  12  is a fluid such as demineralized water, methanol or a halocarbon such as R134a (C 2 H 2 F 4 ). For a thermosiphon  12  utilizing R134a as working fluid  38 , both the evaporator and condenser can be fabricated out of aluminum. However, an aluminum evaporator or condenser cannot be used when water is the working fluid in view of the corrosive effect of water on aluminum over time. An all-aluminum construction has the benefit of reduced manufacturing costs. Because of its low thermal conductivity, aluminum presents a higher thermal resistance in comparison to copper. Therefore, an evaporator  20  constructed from aluminum is not suitable when the heat flux generated by the electronics device  8  is very high. Copper is the preferred material of construction for evaporator  20  when the heat flux generated by the electronics device  8  is very high. Copper also has the benefit of usability for both R134a and water based working fluids  38 , while aluminum is generally suitable only for an R- 134 a working fluid. 
     Based on theoretical and experimental study, the following dimensions of thermosiphon  12  found to be optimal: the ratio of the width  45  of tubes  44  to hub diameter  55  of fan  14  is expressed by the relationship 0.08≦a/H≦0.25; the ratio of the height  53  of second fins  52  to hub diameter  15  of fan  14  is expressed by the relationship 0.125≦p/H≦0.5; the ratio of the height  51  of first fins  50  to diameter  55  of hub  15  of fan  14  is expressed by the relationship 0.15≦q/H≦0.5; and the ratio of the height  37  of evaporating chamber  36  to the height  57  of tubes  44  is expressed by the relationship 0.075≦h/L≦0.25. The linear fin density of each fin strip ranges from 8 fins per inch to 20 fins per inch. 
     In use, as device  8  generates power and thus, heat, the heat so generated is transferred to baseplate  26 . As baseplate and especially fins  28  increase in temperature, surface  30  becomes sufficiently hot to cause the working liquid covering the baseplate  26  to nucleate or boil. The working fluid vapor rises and enters hairpin condenser tubes  44 . The heated vapor contacts the sidewalls of tubes  44  and transfers the thermal energy in the vapor to the walls of tubes  44  and thereafter by conduction to convoluted fins  50  and  52 . Axial fan  14  causes cooling air to flow primarily through convoluted first fins  50  and secondarily through second fins  52  and duct  60 , convectively drawing heat therefrom. By removing thermal energy from the vapor, the vapor is cooled below its condensation temperature and condenses on the interior walls of tubes  44 . The condensed liquid congregates and falls back through tubes  44  to the pool of working fluid in vapor chamber  38  whereupon the process is repeated. 
     Turning now to FIG. 6 another embodiment  112  of a thermosiphon is illustrated wherein like features according to the previous embodiment are identified with like numbers preceded by the numeral “1”. In describing thermosiphon  112  of FIG. 6, only the components that differ from the components of thermosiphon  12  of FIG. 2 will be described below since the common components are already described with reference to FIG.  2 . 
     In the embodiment of FIG. 6, two different tube heights are used in order to utilize the region shadowed by fan hub  15  for vapor flow. A hairpin tube  144  has a width  145  denoted by dimension “a” and is bent to a small radius  148  denoted by dimension “R”. Hairpin tube  144  is placed substantially above the highest heat flux {dot over (q)}″ region (the center) of device  108 . The intervening space between the innermost tube segments of hairpin tube  144  is filled with third convoluted fins  172  having a height  173  denoted by dimension “n”. Wide tube  170  has a height slightly greater than hairpin tube  144  and is formed to envelop the hairpin tube  144  within its inverted U-shape. Ends  169  of tube  170  extend through base  132  such that an interior of tube  170  is in fluid communication with vapor chamber  136  through either end  169 . Wide tube  170  has a width  171  denoted by dimension “b” which is generally larger, and thus less restrictive, than width  145  of tube  144 . Second fins  152  having a height  153  denoted by the dimension “p” extend between adjacent legs of tubes  144  and  170 . Enveloping the tube  144  by tube  170  in this fashion helps to maintain structural integrity at high internal pressure and also facilitates manufacturing. 
     By selecting third convoluted fin  172  having a height  173  and tube  144  having a small bend radius  148 , wide tube  170  can be kept relatively close to device  108 . The top of wide tube  170  can also be angled from the horizontal to prevent condensate build up and thus, always ensure the condensate return from the top of tube  170  to the chamber  136 . The size of the hairpin tube  144  having bend radius  145  and the short height  173  of fins  172  is selected specifically to utilize the low airflow in the region of hub  115 . Strategic placement of wide tube  170  on the outside of tube  144 , but within the width  155  of fan hub  115 , enables heat dissipation through first fins  150 . The majority of the vapor generated in vapor chamber  138  flows through the less restrictive wide tube  170  with larger cross-section and hence with lower flow resistance. First fins  150  are bonded to wide tube  170  and shroud  134  and are positioned in the wake of the fan blades of fan  114 , therefore ensuring good airflow and lower overall airside pressure drop. In this fashion, fins  150  are placed in the periphery of thermosiphon  112  and are utilized to dissipate the majority of the latent heat from the vapor carried by tube  170 . 
     For the embodiment illustrated in FIG. 6 as thermosiphon  112 , and through careful design and test iterations, it was established that the benefits of the present embodiment are best realized within the following ranges of the key dimensions: The ratio of the width  145  of tube  144  to hub diameter  155  of fan  114  is expressed by the relationship 0.08≦a/H≦0.2. The width of wide tube  170  to hub diameter  155  is expressed by the relationship 0.125≦b/H≦0.5. The ratio of the height  153  of third fins  152  to hub diameter  155  of fan  114  is expressed by the relationship 0.08≦n/H≦0.4. The ratio of the height  151  of first fins  150  to diameter  155  of hub  115  is expressed by the relationship 0.2≦q/H≦0.5. The ratio of the height  153  of second fins  152  to diameter  155  of hub  115  is expressed by the relationship 0.08≦p/H≦0.375. The ratio of the height  137  of evaporating chamber  136  to the height  157  of tubes  144  is expressed by the relationship 0.075≦h/L≦0.25. The linear fin density of each fin strip ranges from 8 fins per inch to 20 fins per inch. 
     Turning now to FIG. 7 another embodiment  212  of a thermosiphon is illustrated wherein like features according to the previous embodiment are identified with like numbers preceded by the numeral “2”. In describing thermosiphon  212  of FIG. 7, only the components that differ from the components of previous embodiments of FIGS. 2 and 6 will be described below since the common components are already described with reference to previous embodiments. 
     The embodiment illustrated in FIG. 7 employs two hairpin tubes  244  within a wide tube  270 . Hairpin tubes  244  are positioned directly over device  208  where the maximum heat flux {dot over (q)}″ region is realized. Wide tube  270  encompasses both hairpin tubes  244  and generally extends the width of fan  214 . Thermosiphon  212  fully addresses the non-uniformity of the airflow. 
     Fin sizes as well as the linear fin densities are varied to conform to the airflow induced by the fan. Third fins  272  having a height  273  denoted by dimension “n” are placed between two closely spaced hairpin tubes  244 . Second fins  252  are medium sized having a height  253  denoted by dimension “p” and are positioned interiorly of the legs of each hairpin tube  244 . First fins  250  having a height  251  denoted by dimension “q” are positioned between hairpin tubes  244  and wide tube  270  in the region corresponding to the maximum airflow from fan  214 . Outer fins  280  extend between tube  270  and shroud  234  outside of the primary airflow stream of fan  214 . Fins  280  are of medium size and have a height  281  denoted by dimension “r”. 
     This design is suitable for high heat load as well as for high heat flux. By employing non-uniform fins sizes, the pressure drop registered by the flowing air from fan  214  is utilized profitably for carrying waste heat. If the fins were of uniform size and density, the pressure drop would have still occurred, however, the heat pick up would have been less due to a reduced availability of vapor flow rate at the periphery. Selecting small bend radii and fins having a correspondingly relatively small height permits concentrating a maximum of tube space directly above the core of the heat-generating device  208 . In this way, the tube entrance losses are minimized for vapor flow and thereby maintaining an overall low vapor side pressure drop. As evident from FIG. 7, tubes  244  are bundled behind the fan hub  215  and significant portion of the finned area is placed behind the blades of fan  214 . Additional modulation of the airflow to qualitatively mimic the heat flux profile can be achieved by lowering the fin density in the middle and increasing the fin density at the periphery. 
     For the embodiment illustrated in FIG. 7 as thermosiphon  212 , and through careful design and test iterations, it was established that the benefits of the present embodiment are best realized within the following ranges of the key dimensions. The ratio of the width  245  of tube  244  to hub diameter  255  of fan  214  is expressed by the relationship 0.08≦a/H≦0.25. The width  271  of wide tube  270  to hub diameter  255  is expressed by the relationship 0.08≦b/H≦0.2. The ratio of the height  253  of second fins  252  to hub diameter  255  of fan  214  is expressed by the relationship 0.1≦p/H≦0.3. The ratio of the height  281  of outer fins  280  to diameter  255  of hub  215  is expressed by the relationship 0.1≦r/H≦0.2. The ratio of the height  251  of first fins  250  to diameter  255  of hub  215  is expressed by the relationship 0.2≦q/H≦0.4. The ratio of the height  237  of evaporating chamber  236  to the height  257  of tubes  244  is expressed by the relationship 0.075≦h/L≦0.25. The linear fin density of each fin strip ranges from 8 fins per inch to 20 fins per inch. 
     Turning now to FIG. 8 another embodiment  312  of a thermosiphon is illustrated wherein like features according to the previous embodiment are identified with like numbers preceded by the numeral “3”. In describing thermosiphon  312  of FIG. 8, only the components that differ from the components of previous embodiments of FIGS.  2  and  6 - 7  will be described below since the common components are already described with reference to previous embodiments. 
     As illustrated in FIG. 8, the hairpin tube  144  of thermosiphon  112  as illustrated in FIG. 6 has been combined into a single central stem tube  396  in thermosiphon  312 . The single tube  396  reduces the number of brazing joints and thereby further reduces the potential for leakage of the working fluid from the thermosiphon  312  since tube  396  has only one inlet  395  extending through base  332  into evaporating chamber  336 . Thermosiphon  312  utilizes different tube and different fin sizes. The central stem tube  396  has a width  397  denoted by dimension “c” of wider cross-section than previous tube  44 . Central stem tube  396  is placed centrally behind fan hub  315  and directly above the high heat flux region of device  308 . Tube  396  is sealed at its top. Wide tube  370  has a width  371  denoted by dimension “b” and is formed to have a substantially flat top over the top of central stem tube  396 . First fins  350  at the periphery have a height  351 , denoted by dimension “q”, and are substantially in line with the airflow from fan  314 . First fins  350  are generally of the same height or taller than second fins  352  having a height  353  denoted by dimension “p”. 
     Thermosiphon  312  is particularly suited for high heat flux and very concentrated heat loads, and where spreading of heat is difficult and the vapor side pressure drop requirement is low. Additionally, central stem tube  396  significantly enhances heat transfer performance of the evaporator as a result of condensate dripping into the liquid pool  338  directly over the center of device  308 . This improves the performance of the boiling surface at very high heat flux. 
     For the embodiment illustrated in FIG. 8 as thermosiphon  312 , and through careful design and test iterations, it was established that the benefits of the present embodiment are best realized within the following ranges of the key dimensions. The ratio of the width  397  of tube  396  to hub diameter  355  of fan  314  is expressed by the relationship 0.125≦c/H≦0.3. The width  371  of wide tube  370  to hub diameter  355  is expressed by the relationship 0.08≦b/H≦0.2. The ratio of the height  353  of second fins  352  to hub diameter  355  is expressed by the relationship 0.1≦p/H≦0.325. The ratio of the height  351  of first fins  350  to diameter  355  of hub  315  is expressed by the relationship 0.2≦q/H≦0.5. The ratio of the height  337  of evaporating chamber  336  to the height  357  of wide tube  370  is expressed by the relationship 0.1≦h/L≦0.375. The linear fin density of each fin strip ranges from 8 fins per inch to 20 fins per inch. 
     In the foregoing description those skilled in the art will readily appreciate that modifications may be made to the invention without departing from the concepts disclosed herein. Such modifications are to be considered as included in the following claims, unless these claims expressly state otherwise.