Abstract:
A vaporizing heat exchanger ( 10, 68 ) is provided for vaporizing a fluid flow using a thermal energy containing flow. Flow paths ( 24, 26, 28 ) for each of the fluid flow and the thermal energy containing flow are located relative to each other such that the fluid flow makes multiple passes wherein the fluid is vaporized and subsequently superheated.

Description:
CROSS REFERENCE TO RELATED APPLICATION 
     This application claims the benefit of the earlier filed provisional application Ser. No. 60/709,556 entitled “Water Vaporizer With Intermediate Steam Superheating Pass” and naming the same inventors as the present application, the entire disclosure of which is incorporated herein by reference. 
    
    
     FIELD OF THE INVENTION 
     This invention relates to heat exchangers, and in more particular applications, to vaporizing heat exchangers for vaporizing a liquid flow. 
     BACKGROUND OF THE INVENTION 
     Water vaporizers are frequently employed in steam reformer based fuel processor systems. In such systems, a hydrocarbon fuel such as natural gas, propane, methanol, gasoline, diesel, etc. is combined with steam and reacted over a catalyst at elevated temperature in order to create a hydrogen-rich gas (reformate) which can be used as a fuel source for a fuel cell anode or as a source of impure hydrogen which can be purified through membrane separation or pressure swing adsorption (PSA) to yield high-purity hydrogen. The water vaporizer serves to vaporize a liquid water source and create superheated steam, which can then be mixed with the gaseous or liquid hydrocarbon fuel source to form the reactants for the steam reforming process. In order to maximize system efficiency, the heat source utilized for vaporization of the liquid water is frequently a high temperature exhaust gas created by combusting unreacted off-gas from the fuel cell anode or PSA or hydrogen separation membrane. 
     Three distinct regions of heat transfer can typically be identified in such vaporizers. The first region is where the water exists as a subcooled liquid, receiving sensible heating from the heat source fluid; the second region is where the water undergoes vaporization, existing as a two-phase liquid-vapor mixture receiving latent heat from the heat source fluid; the third region is where the water exists as a superheated vapor, again receiving sensible heating from the heat source fluid. The area of sudden transition from the second region to the third region, referred to as the “dryout” location, is typically characterized by a sharp increase in the temperature of the wall separating the heat source fluid and the water flow. This sharp increase is due to the two-phase heat transfer coefficient being substantially higher than the single-phase vapor heat transfer coefficient, resulting in a wall temperature which is relatively close to the vaporizing temperature in the two-phase region and relatively close to the heat source fluid temperature in the superheat region. The temperature gradient is especially pronounced in vaporizers where the fluids flow in a direction counter to one another, and where the inlet temperature of the heat source fluid is substantially higher than the vaporizing temperature of the water. Such a steep temperature gradient over a localized region of the heat exchanger can result in high thermal stresses in that region, leading to the eventual failure of the vaporizer due to thermal fatigue. This problem can be further exacerbated in cases where the water is at a high pressure relative to the heat source fluid, as is frequently the case, since it will subject the wall to large mechanical stresses in addition to the thermal stresses. 
     Furthermore, fuel cells generally require the operating and cooling fluids to be within specified temperature ranges for each fluid. For example, reformate which is used as fuel at the anode side of the fuel cell generally must be within a specified temperature range for optimal fuel cell operation and also to minimize catalyst degradation. Generally the temperature of a reformate flow is much higher than the maximum input temperature specified for the fuel cell and therefore, the flow must be cooled. 
     SUMMARY OF THE INVENTION 
     In one form, an exhaust gas-heated water vaporizer designed to vaporize high pressure liquid water and deliver high temperature, high pressure superheated steam is provided. The vaporizer has a novel construction and flow circuiting which can provide dramatically reduced thermal stresses at the dryout region, eliminate pressure-induced stresses at the dryout region, and provides a thermally unconstrained “floating” design, thereby greatly reducing the likelihood of thermal fatigue failure in comparison to known constructions. 
     Other objects, features, and advantages of the invention will become apparent from a review of the entire specification, including the appended claims and drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a perspective view of the exterior of a vaporizer; 
         FIG. 2  is a diagrammatic representation of the flow paths and separating walls of an embodiment of a vaporizer; 
         FIG. 3  is a partial cut-away view showing exhaust gas heat transfer surfaces of an embodiment of a vaporizer; 
         FIG. 3A  is an enlarged view of the indicated part of  FIG. 3 ; 
         FIG. 4  is a partial cut-away view showing a super-heated steam transfer surface of an embodiment of a vaporizer; 
         FIG. 4A  is an enlarged view of the indicated part of  FIG. 4 ; 
         FIG. 5  is a partial cut-away view showing a first flow path of an embodiment of a vaporizer; 
         FIG. 6  is a cross-sectional view of an embodiment of a vaporizer depicting water flow along a first flow path; 
         FIG. 7  is a cross-sectional view of an embodiment of a vaporizer depicting a steam flow path; 
         FIG. 8  is a cross-sectional view of an embodiment of a vaporizer depicting an exhaust gas flow path; 
         FIG. 9  is a graph depicting the temperature profiles of fluids flowing in the vaporizer of  FIG. 1 ; 
         FIG. 10  is a graph depicting the temperature profiles of fluids flowing in a prior art counter-flow vaporizer; 
         FIGS. 11 and 11A  are graphs comparing the temperature profiles of fluids flowing in the vaporizer of  FIG. 1  to the fluids in a prior art counter-flow vaporizer, each operating at 2:1 turndown; 
         FIG. 12  is a partial cut-away view of an embodiment of a vaporizer and combined reformate cooler; 
         FIG. 13  is a partial cut-away view of the vaporizer section of the vaporizer and combined reformate cooler of  FIG. 12 ; 
         FIG. 13A  is an enlarged view of the indicated part of  FIG. 13 ; 
         FIG. 14  is a partial cut-away view of the superheater section of the vaporizer and combined reformate cooler of  FIG. 12 ; 
         FIG. 14A  is a diagrammatic representation of the flows in the structure of  FIG. 14 ; 
         FIG. 15  is a graph depicting the temperature profiles of fluids flowing in the superheater section of  FIG. 14 ; 
         FIG. 16  is a partial cut-away view of the reformate cooler section of the vaporizer and combined reformate cooler of  FIG. 12 ; 
         FIG. 16A  is a diagrammatic representation of the flows in the structure of  FIG. 16 ; 
         FIG. 17  is a graph depicting the temperature profiles of fluids flowing in the reformate cooler section of  FIG. 16 ; 
         FIG. 18  is an exploded view of an embodiment of a vaporizer and combined reformate cooler; and 
         FIG. 19  is a diagrammatic representation of a vaporizer and combined reformate cooler. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
     In one embodiment, referring to  FIG. 1 , a water vaporizer  10  is shown which includes a long tubular housing  12  which reduces down to smaller diameter tubes at both ends, these being the exhaust gas inlet and outlet ports  14  and  16 . At the exhaust inlet end, water inlet and steam outlet ports  18  and  20  enter the water vaporizer  10  in a direction perpendicular to the axis  22  of the vaporizer tube. 
       FIG. 2  is a diagrammatic representation of the flow paths through the vaporizer. The vaporizer  10  has a first flow path  24  connected to the water inlet port  18 , a second flow path  26  connected to the steam outlet port  20 , and a third flow path  28  connected to the exhaust inlet and exhaust outlet ports  14  and  16 . The first and second flow paths  24  and  26  are separated by a first cylindrical wall  30  (referred to as separating wall  1 ); the second and third flow paths  26  and  28  are separated by a second cylindrical wall  32  (referred to as separating wall  2 ). These two cylindrical separating walls  30  and  32  are concentric to each other and to the outer housing  12  of the vaporizer  10 , with the second cylindrical wall  32  being of a larger diameter than the first cylindrical wall  30 . The ends  34  and  36  of the first and second flow paths  24  and  26  located at the exhaust outlet end of the water vaporizer  10  are connected together, so that water can flow into the water vaporizer  10  through the water inlet port  18 , travel through the first flow path  24 , then travel through the second flow path  26  in a direction counter to its flow through the first flow path, and then flow out of the water vaporizer  10  through the steam outlet port  20 . 
     Thus, in operation, water enters the vaporizer  10  through the water inlet port  18 , and flows through the first flow path  24 . Heat is transferred into the water through separating wall  30  at a rate Q 1 , the magnitude of which varies with location along the flow path  24 . The first flow path  24  consists of a subcooled liquid region  46  closest to the water inlet  18 , followed by a vaporizing region  42 , followed by a superheated vapor region  44  closest to the end  34  of the first flow path  24 . The water exits the first flow path  24  as a slightly superheated vapor, and flows back through the second flow path  26  towards the steam outlet  20 . Along the second flow path  26 , heat is transferred into the water vapor through separating wall  32  at a rate Q 2 , the magnitude of which varies with location along the flow path, and heat is transferred out of the water vapor through separating wall  30  at the previously mentioned location-dependant rate Q 1 . It should be understood that the location and existence of the sub-cooled, vaporizing and superheated regions  40 ,  42  and  44  may vary depending upon the temperature flow rates of the various fluids. The location of these regions shown in  FIG. 2  is merely one embodiment. 
       FIG. 3  is a partial cut-away view showing the exhaust gas heat transfer surfaces in the third flow path  28 . As best seen in enlarged detail in  FIG. 3A , a convoluted fin structure, such as for example a serpentine louvered fin  46 , is wrapped around and brazed to a cylinder  48  which forms separating wall  32 . However, it should be understood by those skilled in the art that other forms of heat transfer surfaces are also contemplated. For example, plain fins, slots, or the like are also suitable. The cylinder  48  is capped at both ends with heads  50  and  52  designed to withstand the internal pressure loads imposed by the elevated pressure of the water, relative to the exhaust gas. The water inlet and steam outlet ports  18  and  20  penetrate into the cylinder  48  which forms separating wall  32 . 
       FIG. 4  is a partial cut-away view showing the superheated steam heat transfer surfaces in the second flow path  26 . As best seen in enlarged  FIG. 4A , another convoluted fin structure  52  is wrapped around and brazed to a cylinder  54  which forms separating wall  30 , as well as being brazed to the inner wall of the cylinder  48  which forms separating wall  32 . However, it should be understood that the fin structure  52  may also be permitted to remain unbonded to one or more of the separating walls  30  and  32 . This may also allow the walls  30  and  32  to expand independent of one another. The beginning and ending locations of the fin structure  52  coincide with the beginning and ending locations of the previously mentioned fin structure  46  for the exhaust gas. The augmented fin  52  structure shown in  FIGS. 4 and 4A  is of the lanced and offset type, although other types of fin structures, such as for example serpentine louvered, would also work. The water inlet port  18  penetrates through the cylinder  54  which forms separating wall  30 , while the steam outlet port  20  is open to the annulus  56  between separating wall  30  and separating wall  32  so that it can receive the steam flow exiting the fin structure  52 . 
       FIG. 5  is a partial cut-away view showing the first flow path  24 , in which the incoming liquid water is sensibly heated, vaporized, and slightly superheated. The flow path  24  is a helical path which is bounded on one side by the inner surface of the cylinder  54  which forms separating wall  30 . Several manufacturing methods could be used to fabricate this flow path; as depicted in  FIG. 5 , the flow path  24  is created through the machining of a helical groove  58  into the outer diameter of a thick-walled cylinder  60 , the outside diameter of which is brazed to the inside diameter of the cylinder  54  which forms separating wall  30 . Again, these surfaces may also be left unattached to one another. The flow path  24  begins some distance in from the end  62  of the cylinder  60  nearest the water inlet and steam outlet ports  18  and  20 , so that a complete seal can be made at the end  62  to prevent any water from bypassing the first and second vaporizer flow paths  24  and  26 . The water inlet port  18  is open to the helical flow path  24  so that liquid water can enter the flow path through the port  18 . At the opposite end, the flow path  24  continues all the way to the end  64  of the cylinder  60  so that the partially superheated steam can flow from the end  34  of the first flow path  24  into the beginning  36  of the second flow path  26 . The cylinder  48  is capped at this end so that the steam flow is forced to return through the second flow path  26 . The evenness of distribution of the steam flow exiting the first flow path  24  and entering the second flow path  26  is improved by transitioning from the helical flow path  24  to the annular flow path  26  at the end of the first flow path  24 , as is shown in  FIGS. 4 and 5 . While not required, it may be advantageous to vary the width of the helical flow path  24  along the path&#39;s length, in order to best accommodate the large changes in density which occur as the water transitions from a high-density subcooled liquid to a low-density superheated vapor. 
     Additionally, it should be understood that other forms of this flow path are contemplated besides helical flow paths. The structure depicted in  FIG. 5  is merely one embodiment. 
       FIGS. 6 ,  7 , and  8  are longitudinal cross-sectional views of the vaporizer  10  which illustrate the flow of the fluids through the first, second, and third flow paths  24 ,  26  and  28 , respectively. 
     In one example, using the vaporizer  10  of  FIG. 1 , the predicted bulk temperature profiles of the fluids and the predicted average separating wall temperatures along the length of the vaporizer are depicted in the graph in  FIG. 9 . The design case shown is for a 25° C. water flow at 15 bar absolute pressure which is heated to a superheated steam flow exiting at 550° C. The heat source is a combustor exhaust gas which enters at 875° C. and is cooled down to an exit temperature of 330° C. 
     Several observations can be made about the data shown in the graph. The predicted temperature profile of separating wall  32  is free of any steep temperature gradients, thus minimizing the potential for excessive thermal stress cycling of the wall  32 . In addition, the maximum predicted temperature of the wall  32  is approximately 700° C., at the exhaust gas inlet end of the vaporizer  10 . This temperature, while high, is substantially below the incoming exhaust temperature of 875° C. The predicted temperature profile of separating wall  30  shows substantially lower temperatures than was seen for separating wall  32 . While there are steep temperature gradients in the wall  30  at the beginning and end of the two-phase region, they are limited to a temperature range of only approximately 90° C. More importantly, there is no pressure-induced stress on the wall  30 , since it separates two flow paths  24  and  26  which are at essentially the same pressure. 
     As a comparison,  FIG. 10  shows the predicted fluid and wall temperature profiles for a more traditional single pass counterflow vaporizer which has been sized to meet the same conditions as used for  FIG. 9 . It should be immediately obvious that the steep temperature gradient at the end of the vaporizing region is much more severe than was seen in the embodiment of the vaporizer  10  shown in  FIG. 9  (approximately 452° C. vs. 90° C.). In addition, this wall is now subjected to a large pressure loading due to the pressure differential between the high pressure water flow and the near atmospheric exhaust gas flow. Thermal cyclic fatigue failure of the separating wall in this highly stressed region is expected to occur much sooner than would be expected for the embodiment of the vaporizer  10  shown in  FIG. 9 . It should also be noted that the peak temperature of the separating wall for this traditional design is approximately 800° C., or about 100° higher than is expected for the embodiment shown in  FIG. 9 . At these operating temperatures, such a difference may require the selection of a more expensive material for the separating wall in the traditional design than would be required for the embodiment of the vaporizer  10  shown in  FIG. 9 . 
     The disclosed embodiment of the vaporizer  10  can provide additional advantages over more traditionally known designs. For example, the cylindrical shapes of the pressure boundaries can help distribute the pressure loading, which can result in lower stress levels in the structure. Stress risers due to tube-header joints may be eliminated. The water inlet and steam outlet ports  18  and  20  are located at the same axial location, therefore the water ports  18  and  20  do not place a constraint on the axial expansion and contraction of the structure. 
     The present invention can also provide greater stability over turndown operation in comparison to more traditional designs. As the flow is reduced, a traditional vaporizer may show an increase in heat transfer effectiveness, leading to increased steam outlet temperatures. In this type of application, where the hot gas inlet temperature is substantially above the desired steam outlet temperature, the potential for overheating of the steam is especially great. In the present invention, this situation is largely mitigated because the hot gas and the incoming water flow in a concurrent direction. This results in a temperature “pinch” at the hot gas exit end of the vaporizer  10 , where the exhaust gas and the superheated steam exiting the first flow path  24  and entering the second flow path  26  reach approximately equal temperatures, with relatively little heat transfer occurring over the flow paths  24  and  26 , and  28  near that end of the heat exchanger  10 . This dramatically limits the increase in effectiveness which occurs as the flows are reduced.  FIGS. 11 and 11A  shows the predicted temperature profiles for both the traditional design and the disclosed design  10 , respectively, as the flows are reduced by 50% (2:1 turndown). It can be seen that the steam outlet temperature in the traditional design increases by 146°, from 550° C. to 696° C. In the disclosed design  10 , the steam exit temperature actually decreases by 23°, from 550° C. to 527° C. While the disclosed design does not maintain a constant steam exit temperature over turndown, it is substantially more stable than a traditional vaporizer design would be. 
     Yet another embodiment is illustrated in  FIG. 12 . In this embodiment, a vaporizer section  68  also integrates a superheater  70  for the vaporized flow and a reformate cooler portion  72  for cooling a reformate flow, such as a reformate flow in a steam reformer system. It should be understood that while this embodiment discloses the portion  72  to cool a reformate flow, other fluids may also be cooled in the portion  72 . In this embodiment, water is vaporized and superheated in the vaporizer section  68 , but the exhaust and superheated steam exit the vaporizer section  68  where the steam is superheated further by the reformate, and the reformate is then further cooled by the exhaust gas flow from the vaporizer section  68 . 
     The vaporizer section  68  of this embodiment is illustrated in more detail in  FIGS. 13 and 13A . Water enters through an inlet port  74  near the left hand side of the vaporizer section  68  and travels along a helical flow path  76  similar to the flow path  24  illustrated in  FIGS. 4-5 , and then turns and flows back along a second flow path  77 , as best seen in  FIG. 13A , that is similar to the flow path  26  shown in  FIGS. 4 and 5 . In fact, the relationship between the water and steam flows and the exhaust flow is similar to the flow relationships illustrated in  FIG. 2 , with the steam in the flow path  77  receiving heat from the exhaust gas and transferring heat to the water flow in the flow path  76 . The main difference is that initially, the exhaust enters at the opposite end  78 , of the vaporizer section  68  of  FIG. 13  and therefore, makes an initial adiabatic pass  80  which is not in contact with the steam flow before reversing direction and flowing through a path  81  in a counter-current flow heat exchange relationship with the steam flow in path  77 . 
     In the embodiment illustrated in  FIG. 13 , the exhaust gas makes a final pass  82  through the center of the vaporizing section  68 , exits the vaporizer section  68 , and enters the reformate cooler section  72  while the steam exits the vaporizer section  68  and enters the superheater section  70  which is concentric about the reformate cooler section  68 . 
     The structure of the superheater section  70  and the reformate cooler portion  72  are shown in  FIG. 14 , with a diagrammatic representation of the steam and reformate flows through the superheater  70  illustrated in  FIG. 14A . The steam flows through a pass  83  having a finned structure  84  which is bounded by a outermost cylindrical wall  86  and an adjacent cylindrical wall  87 . The reformate flows through a pass  88  having a finned structure  89  bounded by the adjacent wall  87  and an inner wall  90  in a counter-current direction to the flow of the steam. While the reformate enters at an elevated temperature, the steam enters the superheater section  70  with significant superheat and therefore the amount of heat that is transferred is reduced. The passes  83  and  88  are concentric annular flow passages, with the pass  83  being radially outboard of the pass  88 , and heat being transferred from the reformate to the steam through the cylindrical wall  87 . Once the steam passes through this section  70 , it exits the structure through an outlet port  92 .  FIG. 15  illustrates the temperature profiles of one embodiment wherein the reformate is used to add additional superheat to the steam flow in the superheater section  70 . 
     The reformate generally may require additional cooling and therefore the reformate cooler structure  72  is included. Once the reformate makes the first pass  88  in heat exchange relationship with the steam, it may be transferred to an additional inner, return pass  94  for further cooling. As illustrated in  FIGS. 16 and 16A , the reformate makes the return pass  94  through a finned structure  96  which is concentric with a annular flow path  98  for the exhaust gas. The reformate flow pass  94  is located in a radially outer annular passage relative to the exhaust flow path  98  and separated by a cylindrical wall  99 , with heat being transferred from the reformate to the exhaust through the wall  99 . As seen in  FIGS. 16-16A , the reformate and the exhaust gas flow in a concurrent-flow relationship and therefore the two flows exit the reformate cooler section at similar temperatures.  FIG. 17  illustrates the temperature profiles of one embodiment of the reformate flow and exhaust gas flow. As seen in this graph, the reformate flow enters the reformate cooler  72  at an elevated temperature, but as the reformate exits the structure  72 , its temperature has been greatly reduced and approaches a common outlet temperature with the exhaust gas. 
     Finally, as illustrated in  FIGS. 12 and 18  and diagrammatically in  FIG. 19 , the vaporizer  68 , superheater  70 , and reformate cooler structure  72  can be designed and assembled into a unitary structure  100 . The overall shape enhances ease of assembly. Furthermore, the cylindrical shapes of the components resist pressure differences that may exist between the various flows and the asymmetric design helps restrict or eliminate warping of the heat exchangers due to temperature differentials. Additionally, the structure  100  may be constructed to allow the individual components to expand and contract independently in response to thermal changes.