Abstract:
A heat exchanger for exchanging heat between a first fluid and a second fluid is disclosed; wherein the heat exchanger has improved thermal efficiency and low fluid back pressure. The heat exchanger comprises conduits for conveying the first fluid, wherein the conduits include a plurality of flow passages. The flow passages are defined by a plurality of fins that are continuous along the direction of flow of the first fluid. Each fin includes a wave-shaped region, and adjacent fins sub-divide each flow passage into first and second sections that are interposed by a third section. The wave-shape of the fins creates a continuously varying cross-sectional area for each third section. The variation of the cross-sectional area of the third section, coupled with the wave-shape of the fins, induces a swirl flow between the third section and each of the first and second sections. This swirl flow improves the efficiency of the overall convection heat transfer in each conduit. Further, the overall cross-sectional area of each conduit remains constant even as the cross-sectional areas of individual flow passages changes, which mitigates the development of back pressure in the flow of the first fluid.

Description:
FIELD OF THE INVENTION 
       [0001]    The present invention relates to energy conversion in general, and, more particularly, to heat exchangers. 
       BACKGROUND OF THE INVENTION 
       [0002]    The Earth&#39;s oceans are continually heated by the sun and cover nearly 70% of the Earth&#39;s surface. The temperature difference between deep and shallow waters contains a vast amount of solar energy that can potentially be harnessed for human use. In fact, it is estimated that the thermal energy contained in the temperature difference between the warm ocean surface waters and deep cold waters within ±10° of the Equator represents a 3 Tera-watt (3×10 12  W) resource. 
         [0003]    The total energy available is one or two orders of magnitude higher than other ocean-energy options such as wave power, but the small magnitude of the temperature difference makes energy extraction comparatively difficult and expensive, due to low thermal efficiency. 
         [0004]    Ocean thermal energy conversion (“OTEC”) is a method for generating electricity which uses the temperature difference that exists between deep and shallow waters to run a heat engine. A heat engine is a thermodynamic device placed between a high temperature reservoir and a low temperature reservoir. As heat flows from one reservoir to the other, the engine converts some of the heat to work. This principle is used in steam turbines and internal combustion engines. Rather than using heat energy from the burning of fuel, OTEC power draws on temperature differences caused by the sun&#39;s warming of the ocean surface. 
         [0005]    One heat cycle suitable for OTEC is the Rankine cycle, which uses a low-pressure turbine. Systems may be either closed-cycle or open-cycle. Closed-cycle systems use a fluid with a low boiling point, such as ammonia, to rotate the turbine to generate electricity. Warm surface seawater is pumped through a heat exchanger where the low-boiling-point fluid is vaporized. The expanding vapor turns the turbo-generator. Then, cold, deep seawater—pumped through a second heat exchanger—condenses the vapor back into a liquid, which is then recycled through the system. Open-cycle engines use the water heat source as the working fluid. 
         [0006]    As with any heat engine, the greatest efficiency and power is produced with the largest temperature difference. This temperature difference generally increases with decreasing latitude (i.e., near the equator, in the tropics), but evaporation prevents the surface temperature from exceeding 27° C. Also, the subsurface water rarely falls below 5° C. Historically, the main technical challenge of OTEC was to generate significant amounts of power, efficiently, from this very small temperature ratio. But changes in the efficiency of modern heat exchanger designs enables performance approaching the theoretical maximum efficiency. 
         [0007]    OTEC systems have been shown to be technically viable, but the high capital cost of these systems has thwarted commercialization. Heat exchangers are the second largest contributor to OTEC plant capital cost (the largest is the cost of the offshore moored vessel or platform). The optimization of the enormous heat exchangers that are required for an OTEC plant is therefore of great importance and can have a major impact on the economic viability of OTEC technology. 
         [0008]    The primary function of a heat exchanger is to efficiently transfer thermal energy from one media to another. Heat exchangers of various types are widely used across many applications such as in air conditioners, industrial chemical plants and power generation. For OTEC systems, while there are many conventional heat-exchanger designs that can be considered, there are, as a practical matter, no good choices. 
         [0009]    Almost all heat exchangers can be classified by two types of geometries: Shell and Tube Heat Exchangers or Compact and Extended Surface Heat Exchangers. Of the Compact and Extended Surface Heat Exchanger geometry, two common types are Plate-Fin Heat Exchangers and Extruded-Fin Heat Exchangers. Plate-fin heat exchangers are typically used in the Cryogenics Industry and extruded-fin heat exchangers are most commonly used in the Automotive Industry. 
         [0010]    Conventional shell and tube heat exchangers are widely available for marine use. But the overall heat transfer coefficient, U, that is associated with reasonable pressure drops for OTEC is typically below 2000 W/m 2 K. This drives the size and cost for this type of heat exchanger and reduces its economic viability. 
         [0011]    Plate-fin heat exchangers are assembled as an array of stacked aluminum plates with thin corrugated sheets of fins between the plates. The entire array is then brazed together to form a heat exchanger core. Unfortunately, plate-fin heat exchangers are undesirable for many applications because of high material and fabrication costs. Second, brazed joints are poorly suited to applications in which corrosive media are used. In OTEC applications, brazed joints are particularly susceptible to galvanic corrosion when exposed to seawater. Third, plate-fin heat exchangers are typically characterized by low thermal and/or flow efficiency. Such designs suffer from varying amounts of fluid flow resistance, and still do not eliminate manufacturing cost and corrosion issues for very large scale assemblies, however. Fourth, the small passages found in typical plate-fin heat exchangers are prone to biofouling. Fifth, maintenance, such as refitting, repair, and refurbishment, on plate-fin heat exchangers is challenging due to the difficulty of accessing their internal regions. 
         [0012]    Conventional extruded-fin heat exchangers are made from standard aluminum extrusions with thin aluminum sections to increase thermal efficiency. Extruded-fin heat exchangers are typically much less expensive to produce than plate-fin heat exchanger cores because of the elimination of the additional fin component and brazing operations. The trade-off with this design is that they typically have lower thermal performance due to a boundary layer created in the fluid(s) being used. This boundary layer is a physical phenomenon that can only be eliminated by disrupting the fluid flow within the heat exchanger core to create stirring turbulences. This issue has been studied for decades and has been solved for various applications in different ways with limited success. 
         [0013]    Typically, plain fins are mounted on the top and bottom of the extruded channels. These fins are much like the rectangular fins in the plate-fin heat exchanger because they are straight and uninterrupted along the full length of flow. Fins that are straight along the flow length tend to develop fluid boundary layers that are quite thick, which results in lower values of the heat transfer coefficient. A plain fin arrangement will exhibit relatively low pressure drop but have relatively low heat transfer. More complex fin designs that provide disruptions to fluid flow can improve heat transfer; however, complex fin designs suffer from a higher pressure drop through the heat exchanger. 
         [0014]    With today&#39;s growing need for energy, using a renewable constant source is a desirable solution. As a consequence, there is a renewed interest in OTEC power plants. But development of an OTEC heat exchanger that accommodates high flow rates while minimizing pumping parasitic losses and offering long life in the ocean environment remains elusive. 
       SUMMARY OF THE INVENTION 
       [0015]    The present invention provides a heat exchanger having higher heat transfer efficiency than heat exchangers known in the prior art. Heat exchangers in accordance with the present invention are particularly well-suited for use in OTEC systems. 
         [0016]    An embodiment of the present invention comprises first conduits for conveying a first fluid through the heat exchanger and second conduits for conveying a second fluid through the heat exchanger, wherein the first fluid and second fluid are thermally coupled by the heat exchanger. The first conduits include flow passages that induce turbulence in the flowing first fluid without significantly increasing fluid back pressure. The turbulence is induced by wave-shaped fins that project into each first conduit to form a plurality of flow passages. The wave-shaped fins are continuous along the direction of fluid flow. Adjacent pairs of wave-shaped fins define three sections in each flow passage: first and second sections that are interposed by a third section. The wave-shape of the fins results in a continuous variation of the cross-sectional area of the third section along the direction of fluid flow. As this cross-sectional area changes, the first fluid flowing through the flow passage is forced to exchange between the third section and each of the two remaining sections of the flow passage. This exchange of fluid between the three sections induces a swirl, or vortex, flow in the first and second sections, which increases the overall convection heat transfer in the flow passage. 
         [0017]    The fins are arranged within each first conduit such alternating fins project into the first conduit from opposite surfaces and so that the wave shapes of adjacent fin pairs are offset by a phase difference. This phase difference leads to a continuously periodic change on the cross-sectional area of the third section along the length of the flow passage. As the cross-sectional area shrinks, first fluid is “squeezed” from each third section into the first and second sections of each flow passage. As the cross-sectional area of the third section increases, first fluid is drawn back into the third section from the other two sections. Further, the wave-shape of the fins defines a shape of the third section that induces the first fluid to swirl as it enters and exits the first and second sections. This swirl flow creates turbulence that enhances heat transfer between the first fluid and walls of the flow passages. 
         [0018]    It is a further aspect of the invention that the first conduits avoid inducing a significant fluidic back pressure while conveying the first fluid through the heat exchanger. Increased back pressure of the fluid is mitigated by the fact that the overall cross-sectional area of the first conduits remains the same even while the cross-sectional areas of individual flow passages within it change. The consistency of overall cross-sectional area of the conduits results from the complimentary nature of adjacent flow passages within them. Specifically, as the cross-sectional area of a first flow passage is shrinking, the cross-sectional area of its adjacent flow passages is increasing by a commensurate amount. As a result, the sum of the cross-sectional areas of all flow passages in a given first conduit remains constant. 
         [0019]    An embodiment of the present invention comprises a heat exchanger that thermally couples a first fluid and a second fluid, wherein the heat exchanger comprises: (1) a first plate comprising a first material that is thermally conductive, wherein the first plate comprises; (i) a first conduit for conveying the first fluid along a first direction, wherein the first conduit comprises at least one first channel; and (ii) a first plurality of first fins, wherein each first fin is continuous along the first direction, and wherein each first fin comprises a first fin portion that has a first periodic shape having a first periodicity along the first direction; and (2) a second plate comprising the first material, wherein the second plate comprises; (i) a second conduit for conveying the first fluid along the first direction, wherein the second conduit comprises at least one second channel; a (ii) second plurality of second fins, wherein each second fin is continuous along the first direction, and wherein each second fin comprises a second fin portion that has the first periodic shape having the first periodicity along the first direction; wherein the first plurality of first fins and the second plurality of second fins collectively define a plurality of passages for conveying the second fluid along the first direction, and wherein each of the plurality of passages comprises a first section, second section, and third section; and wherein the cross-sectional area of the first section varies along the first direction based on the first periodicity, and wherein the variation of the cross-sectional area of the first section induces flow of the second fluid between the first section and each of the second section and third section. 
     
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         [0020]      FIG. 1  depicts a schematic diagram of an OTEC power generation system in accordance with an illustrative embodiment of the present invention. 
           [0021]      FIG. 2A  depicts a schematic drawing of a first heat exchanger in accordance with the prior art. 
           [0022]      FIG. 2B  depicts a perspective view of a plate  202 . 
           [0023]      FIG. 3A  depicts a schematic drawing of a second heat exchanger in accordance with the prior art. 
           [0024]      FIG. 3B  depicts a perspective view of a section of tube  302 . 
           [0025]      FIG. 3C  depicts a schematic of a portion of heat exchanger  300 . 
           [0026]      FIG. 4  depicts a perspective view of a heat exchanger in accordance with the illustrative embodiment of the present invention. 
           [0027]      FIG. 5A  depicts a top view of a portion of plate  402  prior to the formation of a wave-shape on its fins in accordance with the illustrative embodiment of the present invention. 
           [0028]      FIG. 5B  depicts a top view of a portion of plate  402 , after the formation of a wave-shape on its fins. 
           [0029]      FIG. 5C  depicts a bottom view of a portion of plate  402 , after the formation of a wave-shape on its fins. 
           [0030]      FIG. 6A  depicts a representational cross-sectional view of conduit  410 - 2  in accordance with the illustrative embodiment of the present invention. 
           [0031]      FIG. 6B  depicts a top view, through line a-a of  FIG. 6A , of conduit  410 - 2  in accordance with the illustrative embodiment of the present invention. 
           [0032]      FIG. 6C  depicts a detailed cross-sectional view of a portion of conduit  410 - 2 . 
           [0033]      FIG. 7  provides operations of a method for thermally coupling a first fluid and a second fluid in accordance with the illustrative embodiment of the present invention. 
           [0034]      FIG. 8A  depicts a flow passage  602 - i  within in conduit  410 - 2 , at point A depicted in  FIG. 6B . 
           [0035]      FIG. 8B  depicts a flow passage  602 - i  within in conduit  410 - 2 , at point B depicted in  FIG. 6B . 
           [0036]      FIG. 8C  depicts a flow passage  602 - i  within in conduit  410 - 2 , at point C depicted in  FIG. 6B . 
           [0037]      FIG. 9A  depicts a representational cross-sectional view through a portion of a heat exchanger at a first point along the direction of fluid flow, in accordance with an alternative embodiment of the present invention. 
           [0038]      FIG. 9B  depicts a representational cross-sectional view through a portion of a heat exchanger at a second point along the direction of fluid flow, in accordance with the alternative embodiment of the present invention. 
       
    
    
     DETAILED DESCRIPTION 
       [0039]      FIG. 1  depicts a schematic diagram of an OTEC power generation system in accordance with an illustrative embodiment of the present invention. OTEC system  100  comprises turbogenerator  104 , closed-loop conduit  106 , heat exchanger  110 - 1 , heat exchanger  110 - 2 , pumps  114 ,  116 , and  124 , and conduits  120 ,  122 ,  128 , and  130 . OTEC system  100  is deployed in water body  136  wherein a suitable temperature difference exists between water near the surface and water located at a deep level of water body  136 . 
         [0040]    Turbo-generator  104  is a conventional turbine-driven generator. Turbogenerator  104  is mounted on floating platform  102 , which is a conventional floating energy-plant platform. Platform  102  is anchored to the ocean floor by mooring line  132  and anchor  134 , which is embedded in the ocean floor. In some instances, platform  102  is not anchored to the ocean floor but is allowed to drift. Such a system is sometimes referred to as a “grazing plant.” 
         [0041]    In typical operation, pump  114  pumps a primary fluid (i.e., working fluid  108 ), in liquid form, through closed-loop conduit  106  to heat exchanger  110 - 1 . Ammonia is often used as working fluid  108  in OTEC systems; however, it will be clear to one skilled in the art that any fluid that evaporates at the temperature of the water in surface region  118  and condenses at the temperature of the water in deep water region  126  is suitable for use as working fluid  108  (subject to material compatibility requirements). 
         [0042]    Heat exchanger  110 - 1  and  110 - 2  are configured for operation as an evaporator and condenser, respectively. One skilled in the art will recognize that the operation of a heat exchanger as evaporator or condenser is dependent upon the manner in which it is configured within system  100 . Heat exchanger  110  is described in detail below and with respect to  FIGS. 4 through 7C . 
         [0043]    In order to enable its operation as an evaporator, pump  116  draws warm secondary fluid (i.e., seawater from surface region  118 ) into heat exchanger  110 - 1  via conduit  120 . At heat exchanger  110 - 1  heat from the warm water is absorbed by working fluid  108 , which induces working fluid  108  to vaporize. After passing through heat exchanger  110 - 1 , the warm water is ejected back into water body  136  via conduit  122 . In a typical OTEC deployment, the water is surface region  118  is at a substantially constant temperature of approximately 25 degrees centigrade (subject to weather and sunlight conditions). 
         [0044]    The expanding working fluid  108  vapor is forced through turbogenerator  104 , thereby driving the turbogenerator to generate electrical energy. The generated electrical energy is provided on output cable  112 . Once it has passed through turbogenerator  104 , the vaporized working fluid enters heat exchanger  110 - 2 . 
         [0045]    At heat exchanger  110 - 2 , pump  124  draws cold secondary fluid (i.e., seawater from deep water region  126 ) into heat exchanger  110 - 2  via conduit  128 . The cold water travels through heat exchanger  110 - 2  where it absorbs heat from the vaporized working fluid. As a result, working fluid  108  condenses back into liquid form. After passing through heat exchanger  110 - 2 , the cold water is ejected into water body  136  via conduit  130 . Typically deep water region  126  is 1000+ meters below the surface of water body  136 , at which depth water is at a substantially constant temperature of a few degrees centigrade. 
         [0046]    Pump  114  pumps the condensed working fluid  108  back into heat exchanger  110 - 1  where it is again vaporized; thereby continuing the Rankine cycle that drives turbogenerator  104 . 
         [0047]      FIG. 2A  depicts a schematic drawing of a first heat exchanger in accordance with the prior art. Heat exchanger  200  comprises a plurality of plates  202 , and housing  208 . Heat exchanger  200  is in accordance with a heat exchanger described by T. F. Brise in U.K. Patent Application GB2424265, which was published Sep. 9, 2006, and which is incorporated by reference herein. 
         [0048]    Each plate  202  is an extruded body that comprises a plurality of channels  204 , which convey a first fluid along the z-direction as shown. In some cases, each of channels  204  comprises projections directed inward from the surface of the channel. These projections increase the heat transference between the first fluid and the plate material. 
         [0049]    Plates  202  are stacked and welded together to form heat exchanger core  210 . Adjacent plates in heat exchanger core  210  collectively define conduits  212  for conveying a second fluid along the z-direction. Each plate  202  further comprises a plurality of fin segments  206 , which depend from the upper and lower surfaces of the body of plate  202  and project into conduits  212 . 
         [0050]    Each fin segment  206  is a discrete element that is formed from fins that are initially continuous along the z-direction. To form fin segments  206 , these initially continuous fins are “segmented and the fin segments twisted so that at least the upper parts of the fin segments are at an angle” with respect to the z-direction (i.e., the longitudinal axis of the tube). According to Brise, “Segmenting the fins and twisting the fin segments creates lateral flow paths across the surface of the tube along which coolant can flow.” 
         [0051]    Heat exchanger core  210  is located within heat exchanger  200  via housing  208 . Housing  208  locates heat exchanger core  210  by means of seats  214 , which receive the uppermost and lowermost fin segments that project from the heat exchanger core. Once heat exchanger core  210  is positioned within sleeve  208 , weld joints are formed at seats  214  to firmly fix the heat exchanger core in place. 
         [0052]      FIG. 2B  depicts a perspective view of a plate  202 . As is readily seen from the figure, the top and bottom outer surfaces of plate  202  are covered by rows of individual fin segments  206 . These fin segments are arranged along the z-direction in rows. 
         [0053]    As the second fluid flows along the z-direction, the discontinuities between fin segments  206  create turbulence that enhances heat transfer between the second fluid and plate  202 . 
         [0054]    Unfortunately, heat exchangers such as heat exchanger  200  have several disadvantages. First, in many cases the formation of fin segments  206  shears off much of the base of each fin segment. As a result, fin segments  206  are highly susceptible to fracture during use. Fractured fins are likely to lodge in conduits  212  and restrict the flow of the second fluid through the heat exchanger. 
         [0055]    Second, when the plates  202  are stacked together to form conduits  212 , fin segments  206  are interlocked such that they nearly completely obstruct the flow path for the second fluid. As a result, fin segments  206  cause a large pressure-drop through conduits  212 . 
         [0056]    Third, it is well-known to those skilled in the art that discontinuities, such as those between fin segments  206 , significantly increase back pressure. 
         [0057]    Fourth, it is well known that discontinuities in a flow path are prone to impurity build-up, which leads to fouling during use. For example, the use of serrated fins, such as those included in heat exchanger  200 , is discouraged in many heat exchanger design handbooks, such as “ Heat Exchanger Design Handbook ,” written by T. Kuppan. In general, it is difficult and expensive to clean fouled fluid conduits of a heat exchanger, and nearly impossible for compact heat exchangers. 
         [0058]      FIG. 3A  depicts a schematic drawing of a second heat exchanger in accordance with the prior art. Heat exchanger  300  comprises tube  302 , ruffled fins  304 , inlet  306 , outlet  308 , and frame  310 . Heat exchanger  300  is in accordance with a heat exchanger described by W. E. McCullough in U.S. Pat. No. 2,347,957, which was issued May 2, 1944, and which is incorporated by reference herein. 
         [0059]    Tube  302  is a tubular-shaped conduit for conveying a first fluid along the z-direction as shown. Tube  302  is formed by flattening larger extruded tubing stock to form a central tube (i.e., tube  302 ) having attached flat projections. 
         [0060]    These flat projections are run through rollers, or other suitable apparatus, to ruffle the projections into ruffled fins  304 . As a result, ruffled fins  304  are wavy projections that extend laterally from tube  302 . 
         [0061]      FIG. 3B  depicts a perspective view of a section of tube  302 . Tube  302  comprises conduit  312 , which conveys the first fluid along the z-direction. To form heat exchanger core  312 , tube  302  is bent multiple times to form a continuous serpentine-shaped coil. In order to facilitate bending tube  302 , portions of ruffled fins  304  are cut away near the regions of the bends. Heat exchanger core  312  is mounted into frame  310  such that each of the plurality of straight runs of tube  302  is oriented along the z-direction, and ruffled fins  304  project along the x-direction, as shown. 
         [0062]      FIG. 3C  depicts a schematic of a portion of heat exchanger  300 . As evidenced by  FIG. 3C , the bends of tube  302  are judiciously placed so that the crests  312  of ruffled fins  304  collectively define funnel-shaped “air scoops  314 ” that collectively form a plurality of “Venturi throats.” In other words, tube  302  is bent such that the wave of ruffled fins  304  of a first straight section is 180 degrees out of phase with the waves of adjacent straight sections. For example, referring to FIGS. 7 and 11 of the patent, McCullough discloses that for adjacent straight runs, the crests of the ruffled fins  304  are located “extending away from each other and beyond the width of the tube section 11. Due to this construction each of these ruffles acts as an air scoop and between them is an opening having a general funnel shape which narrows down to a flat opening 24 between the runs of tube section . . . ” And, “The cross section through these ruffles (FIG. 16) shows that the air stream first impinges forcefully against the fins 22 and 23, then rushes with increased velocity through the Venturi throat, 24, and expands to make contact with the ruffles 25 and 28.” See e.g., McCullough, page 2, second column, lines 3-20. 
         [0063]    It is clear from this description that heat exchanger  300  is suitable only for exchanging heat between the first fluid and air. Further, heat exchanger  300  conveys the first fluid along the z-direction and the second fluid (i.e., air) along an orthogonal direction (i.e., the x-direction, as shown). As a result, the interaction length of heat two fluids in heat exchanger  300  is extremely short, which limits the efficiency of heat transfer. Still further, the 180 degree bends in the flow-path of the first fluid lead to a large pressure drop that reduces the efficiency of the heat exchanger and limits the size of such heat exchangers. 
         [0064]      FIG. 4  depicts a perspective view of a heat exchanger in accordance with the illustrative embodiment of the present invention. Heat exchanger  110  comprises plates  402 - 1  and  402 - 2  and end plates  408 - 1  and  408 - 2 . Although the illustrative embodiment comprises two plates  402 , it will be clear to one skilled in the art, after reading this specification, how to specify, make, and use alternative embodiments of the present invention that comprise any practical number of plates  402 .  FIG. 4  is described with continuing reference to  FIG. 1 . 
         [0065]    Heat exchanger  110  thermally couples working fluid  108  and seawater taken from a region of the water body  136 . For example, heat exchanger  110 - 1  acts as an evaporator that heats working fluid  108  by transferring heat from warm seawater of surface region  118 . In similar fashion, heat exchanger  110 - 2  acts as a condenser that cools vaporized working fluid  108  by thermally coupling it with cold seawater of deep water region  126 . 
         [0066]    Each of plates  402 - 1  and  402 - 2  (collectively referred to as plates  402 ) conveys seawater along the z-direction via conduits  404 - 1  and  404 - 2 , respectively. Each of conduits  404 - 1  and  404 - 2  comprises a plurality of channels  406 . In the illustrative embodiment, plates  402  are extruded plates of aluminum alloy. In some embodiments, plates  402  are made of another suitable material, such as aluminum, composite materials, graphite, graphite foam, and the like. It is preferable that plates  402  be made of a material that is substantially corrosion-free when exposed to seawater and/or common working fluids. Channels  406  are fluidically coupled to closed-loop conduit  106  via manifolds (not shown for clarity). 
         [0067]    End plates  408 - 1  and  408 - 2  (collectively referred to as end plates  408 ) mate with plates  402 - 1  and  402 - 2 , respectively. Plates  402  and end plates  408  collectively define a heat exchanger core that conveys both working fluid  108  and seawater along the z-direction while also fluidically isolating the fluids from one another. 
         [0068]    Plates  402  are stacked together with end plates  408  to collectively define conduits  410 - 1 ,  410 - 2 , and  410 - 3  (collectively referred to as conduits  410 ). Each of conduits  410  conveys working fluid  108  through heat exchanger  110  along the z-direction. Conduits  410  are described in more detail below and with respect to  FIG. 6 . 
         [0069]    It should be noted that although the illustrative embodiment conveys seawater through conduits  404  and working fluid through conduits  410 , in some embodiments seawater is conveyed through conduits  410  and working fluid is conveyed through conduits  404 . In some alternative embodiments, a secondary fluid other than seawater is conveyed through the heat exchanger. 
         [0070]    Plates  402  and end plates  408  are joined with a substantially galvanic corrosion-free joint, such as a friction-stir welding joint. In some embodiments, plates  402  and end plates  408  are joined using a different joining technology. 
         [0071]      FIG. 5A  depicts a top view of a portion of plate  402  prior to the formation of a wave-shape on its fins in accordance with the illustrative embodiment of the present invention. Plate  402  comprises body  502 , straight fins  506 , and partition  512 . Plate  402  is representative of each of plates  402 - 1  and  402 - 2 . 
         [0072]    Body  502  is the central portion of plate  402 , which comprises conduit  404 . On each end (left and right ends, as depicted in  FIG. 4 ), body  502  comprises a sidewall  508 . Sidewall  508  projects above and below body  502  a distance necessary to provide a desired amount of clearance along the y-direction above fins  510  when plate  402  is joined with another plate  402  or an end plate  408 . Sidewall  508  also provides a surface for joining plate  402  with another plate  402  or an end plate  408 . 
         [0073]    Straight fins  506  are projections that project from surface  504  of plate  402 . Straight fins  506  are normal to surface  504 . 
         [0074]    Partition  512  is a straight wall that forms a sidewall for an end flow passage of conduit  410 , as described below and with respect to  FIGS. 6A and 6B . 
         [0075]    Straight fins  506 , sidewall  508 , and body  502  are contiguous portions of a single extrusion that forms plate  402 . 
         [0076]      FIG. 5B  depicts a top view of a portion of plate  402 , after the formation of a wave-shape on its fins. 
         [0077]    Straight fins  506  are formed into fins  510  through a conventional forming process, such as stamping, rolling, crimping, or hot forming. For example, a pair of stamping dies, having a desired wave shape, can be used to press this desired wave shape into straight fins  506 . Alternatively, rollers or cams having a suitable forming surface can be rolled along the sides of straight fins  506  to deform them into fins  510 . Such a rolling process offers an ability to form fins  510  in a continuous manner. Further parallel sets of rollers or cams enable the formation of a plurality of fins  510  at the same time. 
         [0078]    Once straight fins  506  have been shaped, fins  510  are characterized by the desired wave shape along the z-direction, wherein the wave shape is that of a sinusoid having amplitude a and wavelength λ. In some embodiments, fins  510  are characterized by a periodic shape other than a sinusoid, such as triangular or chevron-shaped patterns. 
         [0079]      FIG. 5C  depicts a bottom view of a portion of plate  402 , after the formation of a wave-shape on its fins. Plate  402  further comprises fins  516 , which project from bottom surface  514  of body  502 . Fins  516  are analogous to fins  510 ; however, fins  516  are formed such that, when plates  402 - 1  is flipped over and plates  402 - 1  and  402 - 2  are joined, the wave shapes of fins  510  and  514  nest such that there is a relative phase, φ, between them along the z-direction. In the illustrative embodiment, φ has a magnitude of approximately 180 degrees. In some embodiments, φ has a magnitude of approximately 90 degrees. In some embodiments, φ can have any magnitude within the range of 0 degrees to 360 degrees. The relationship of φ to the performance of heat exchanger  110  is discussed in more detail below and with respect to FIGS.  7  and  8 A-C. 
         [0080]    In similar fashion to conduit  410 - 2 , conduit  410 - 1  is collectively defined by plate  402 - 1  and end plate  408 - 1 . The top surface of end plate  408 - 1  is analogous to surface  504 . Further, end plate  408 - 1  comprises fins  510  that project into conduit  410 - 1  from its top surface. The fins  510  of end plate  408 - 1  nest with fins  516  that project from surface  514  of plate  402 - 1  to form a plurality of flow passages in conduit  410 - 1 . Still further, the top surface of end plate  408 - 2  is analogous to surface  514  and end plate  408 - 2  comprises fins  516  that project into conduit  410 - 3  from its bottom surface. The fins  516  of end plate  408 - 2  nest with fins  510  that project from surface  504  of plate  402 - 2  to form a plurality of flow passages in conduit  410 - 3 . As a result, conduits  410 - 1  and  410 - 3  are analogous to conduit  410 - 2 . 
         [0081]      FIGS. 6A and 6B  depict a representational cross-sectional and top view through line a-a, respectively, of conduit  410 - 2  in accordance with the illustrative embodiment of the present invention. When plates  402 - 1  and  402 - 2  are joined, fins  510  project from surface  504 - 1  into conduit  410 - 2 . Fins  510  nest with fins  516  that project from surface  514 - 2  into conduit  410 - 2 . As a result, fins  510  and  516  collectively define flow passages  602  within conduit  410 - 2 . In addition, flow passage  604  is defined by the right-most (as depicted in  FIG. 6A ) fin  516  and partition  512 , which projects from surface  504 - 1 . In similar fashion, although not shown for clarity, left-most fin  510  and partition  512  that projects from surface  514 - 2  collectively define a second flow passage  604 . 
         [0082]      FIG. 6C  depicts a detailed cross-sectional view of a portion of conduit  410 - 2 . Each of fins  510  and  516  has substantially the same dimensions. While the specific dimensions of fins  510  and  516  are dependent upon the desired characteristics of heat exchanger  110 , the dimensional relationships of fins  510  and  516  influence the manner in which working fluid  108  flows through conduit  410 - 2 . 
         [0083]    Each of fins  510  and  516  is characterized by a wave starting height of h 1 . The portion of each fin that comprises its wave shape is denoted as h 2 . In the illustrative embodiment, h 2  is greater than h 1  although it will be clear to one skilled in the art, after reading this specification, how to specify, make, and use alternative embodiments of the present invention wherein h 2  is not larger than h 1 . 
         [0084]    Fin-to-base clearance h 3  denotes the clearance between the fins and surfaces  504  and  514  of conduit  410 - 2 . The fin-to-base clearance is determined by the difference between the separation, s, between surfaces  504  and  514  and the sum of h 1  and h 2 . 
         [0085]    Each of the pluralities of fins  510  and  514  are characterized by the same fin-periodicity, d 1 . The wave-to-wave clearance, d 2 , is based on the fin-periodicity, d 1 , and the wave amplitude, a. 
         [0086]    Judicious selection of h 1 , h 2 , h 3 , s, d 1 , d 2 , a, λ, and φ enables the design of conduits  410  that have high heat transfer efficiency. 
         [0087]    It should be noted that prior art attempts to incorporate wavy channels (e.g., wavy fins) into conventional heat exchangers have utilized channels whose cross-sectional area remains constant along the direction of fluid flow. These channels are wavy in that they periodically abruptly change the flow direction of the fluid. Examples of such prior-art heat exchangers are found in “Forced Heat Convection in Wavy Fin Channel,” by Yang, et al., published in  The Journal of Thermal Science and Technology , Vol. 3, pp. 342-354, (2008). A boundary layer, which forms in the flowing fluid, will separate and reattach at alternative periods of the wave thereby resulting in vortices around the flow channels. While these vortices improve the heat transfer coefficient of the heat exchanger, such heat exchangers are subject to a significant pressure drop (i.e., fluidic back pressure) through the length of the flow channels. 
         [0088]      FIG. 7  provides operations of a method for thermally coupling a first fluid and a second fluid in accordance with the illustrative embodiment of the present invention. Method  700  begins with operation  701 , wherein heat exchanger  110  is provided.  FIG. 7  is described herein with continuing reference to  FIGS. 1 , and  4 - 6 C and further reference to  FIGS. 8A-C . 
         [0089]      FIGS. 8A-C  depict a flow passage  602 - i  within in conduit  410 - 2  at different points along the z-direction. The shape of flow passage  602 - i  is based on the relative orientations of shoulders  802  and wave portion  804  of fins  516 - i  and  510 - i , respectively. Flow passage  602 - i  comprises sections  806 ,  808 , and  810 . The shape and cross-sectional area of section  808  changes along the length of conduit  410 - 2  and is defined by the configuration of fins  510 - i  and  516 - i  at each point along the z-axis. 
         [0090]    At operation  702 , conduits  404  are fluidically coupled to a region of water body  136 . As a result, seawater is conveyed through heat exchanger  110  via conduits  404 . 
         [0091]    At operation  703 , conduits  410  are fluidically coupled to working fluid  108 . As a result, working fluid  108  is conveyed through heat exchanger  110  via conduits  410 . Both the seawater and working fluid are conveyed along the z-direction, as described above; therefore, the interaction length between them is substantially the length of heat exchanger  110 . A long interaction length enables a highly efficient transfer of thermal energy between the seawater and working fluid. 
         [0092]    At operation  704 , changes in the cross-sectional areas of sections  808  induce flow  812  of working fluid. These changes induce the flow of working fluid  108  between section  808  and sections  806  and  810  of each flow passage in the conduit. 
         [0093]      FIG. 8A  depicts flow passage  602 - i  at point A, when wave portions  804  of fins  510 - i  and  516 - i  are separated by a maximum distance. At this point along conduit  410 - 2 , section  808  of flow passage  602 - i  has the largest cross-sectional area. 
         [0094]      FIG. 8B  depicts flow passage  602 - i  at point B, when fins  510 - i  and  516 - i  are straight. At this point along conduit  410 - 2 , section  808  of flow passage  602 - i  has a smaller cross-sectional area than that depicted in  FIG. 8A . 
         [0095]      FIG. 8C  depicts flow passage  602 - i  at point C, when fins  510 - i  and  516 - i  are separated by a minimum distance. At this point along conduit  410 - 2 , section  808  of flow passage  602 - i  has the smallest cross-sectional area. 
         [0096]    As working fluid  108  flows through conduit  410 - 2  from point A to point B to point C, the working fluid is squeezed from section  808  into each of sections  806  and  810 . This creates turbulence in and around sections  806  and  810  due to flow  812 . In some embodiments, flow  812  is a swirl flow. In some embodiments, flow  812  is a vortex flow. The turbulence in these sections significantly increases the thermal transfer efficiency of heat exchanger  110 . 
         [0097]    It should be noted that the combined cross-sectional area of passages  602  and  604  remains the same throughout the length of heat exchanger  110 . This is readily seen by the fact that as the cross-section of section  808 - i  becomes smaller, its reduction in size is offset by a commensurate increase in the cross-sectional area of  808 - i+ 1. The shift in cross-sectional area between neighboring flow passages is responsible for advantageously inducing turbulent flow in sections  806  and  810  of each flow passage. The conservation of the overall cross-sectional area of conduits  410  affords embodiments of the present invention additional advantage since the improved thermal transfer efficiency accrues without incurring an increase in flow resistance through the conduits. Still further, the lack of discontinuities in flow passages  602  and  604  enables heat exchanger  110  to avoid fluid flow interruptions inherent in prior-art heat exchangers, such as heat exchanger  200 . Such fluid flow interruptions greatly increase fluid back pressure in these prior-art systems. 
         [0098]      FIGS. 9A and 9B  depict representational cross-sectional views through a portion of a heat exchanger, at two points along the direction of fluid flow, in accordance with an alternative embodiment of the present invention. Heat exchanger  900  is substantially identical to heat exchanger  400 ; however, in heat exchanger  900 , plate  402 - 1  is replaced by plate  902 - 1 . 
         [0099]    Plate  902  is analogous to plate  402  prior to the formation of a wave-shape on its fins. In other words, plate  902  comprises fins  506 , rather than fins  510 , and a top view of plate  902  would be analogous to the top view of plate  402  depicted in  FIG. 5A . 
         [0100]    When plates  902 - 1  and  402 - 2  are joined as shown in  FIGS. 9A and 9B , fins  506  project from surface  504 - 1  into conduit  410 - 2 . Fins  506  nest with fins  516  that project from surface  514 - 2  into conduit  410 - 2 . As a result, fins  506  and  516  collectively define flow passages  904  within conduit  410 - 2 . 
         [0101]    Each of flow passages  904  comprises sections  906  and  910 , which are interposed by section  908 . Sections  904 ,  906 , and  910  are analogous to sections  806 ,  810 , and  812 , described above and with respect to  FIGS. 8A-C . 
         [0102]    In  FIG. 9A , the cross-sectional view is taken at a point analogous to point A of the heat exchanger, as described above and with respect to  FIG. 6B . In other words, the wave-shapes of fins  516  are depicted at their point of right-most deflection. 
         [0103]    In  FIG. 9B , the cross-sectional view is taken at a point analogous to point C of the heat exchanger, as described above and with respect to  FIG. 6B . As a result, fins  516  as depicted in  FIG. 9B  are 180 degrees out of phase with those depicted in  FIG. 9A . 
         [0104]    As working fluid  108  flows through conduit  410 - 2  from point A to point C, the working fluid is squeezed from section  908  into each of sections  906  and  910 . This creates turbulence in and around sections  906  and  910  due to flow into and out of section  908 . In some embodiments, this flow is a swirl flow. In some embodiments, this flow is a vortex flow. As for heat exchanger  110 , the turbulence in sections  906  and  910  significantly increases the thermal transfer efficiency of heat exchanger  900 . 
         [0105]    It should be noted that the combined cross-sectional area of passages  904 , like that of passages  602 , remains the same throughout the length of heat exchanger  110 . This is readily seen by the fact that as the cross-section of section  908 - i  becomes smaller, its reduction in size is offset by a commensurate increase in the cross-sectional area of section  908 - i+ 1. 
         [0106]    It is to be understood that the disclosure teaches just one example of the illustrative embodiment and that many variations of the invention can easily be devised by those skilled in the art after reading this disclosure and that the scope of the present invention is to be determined by the following claims.