Negative pressure plate heat exchanger for superfluid helium system and optimization design method of negative pressure plate heat exchanger

Abstract

The invention relates to the technical field of design and manufacturing of low-temperature heat exchangers, and provides a negative-pressure plate heat exchanger for a superfluid helium system and an optimization design method of the negative-pressure plate heat exchanger. The negative pressure plate heat exchanger for the superfluid helium system comprises a heat exchanger body, the heat exchanger body comprises a plurality of heat exchange plates arranged in a stacked mode, and a flow channel is formed between every two adjacent heat exchange plates; the two wall faces of each heat exchange plate are each provided with a fluid core heat exchange unit and two fluid distribution units, each fluid core heat exchange unit comprises a plurality of first herringbone corrugated structures, and each fluid distribution unit comprises a plurality of second herringbone corrugated structures. The ratio Lco / W0 of the length of each fluid core heat exchange unit to the width of each heat exchange plate ranges from 0.72 to 4.75, and the ratio p / b of the corrugation pitch to the corrugation depth of each first herringbone corrugation structure and the ratio p / b of the corrugation pitch to the corrugation depth of each second herringbone corrugation structure range from 1.4 to 6.7. The defects that in the prior art, a negative pressure heat exchanger for a superfluid helium system is large in size and high in cost are overcome.

Description

Technical Field

[0001] This invention relates to the field of cryogenic heat exchanger design and manufacturing technology, and in particular to a negative pressure plate heat exchanger for superfluid helium systems and its optimized design method. Background Technology

[0002] Superfluid helium systems are an indispensable infrastructure in superconducting accelerators. As a key component of superfluid helium systems, the performance of the negative pressure heat exchanger directly affects the overall system's operating efficiency and stability. Currently, commonly used types of negative pressure heat exchangers in superfluid helium systems include plate-fin heat exchangers, laminated finned tube heat exchangers, spiral tube-fin heat exchangers, and perforated plate heat exchangers, but none of these meet the development requirements of superfluid helium systems for compact and low-cost equipment.

[0003] Brazed plate heat exchangers have been widely used in industry due to their mature manufacturing process, high heat exchange efficiency, compact size, and low cost. However, because the superfluid helium temperature range has special operating conditions such as extreme low temperatures and negative pressures, the thermal-hydraulic performance of brazed plate heat exchangers in this temperature range lacks clear design basis, and their heat transfer and pressure drop characteristics have not been systematically and thoroughly studied. As a result, conventional brazed plate heat exchangers cannot be directly and reliably applied to superfluid helium systems. Summary of the Invention

[0004] This invention provides a negative pressure plate heat exchanger for superfluid helium systems and its optimized design method, in order to solve the defects of existing negative pressure heat exchangers for superfluid helium systems, such as large size and high cost.

[0005] One aspect of the present invention provides a negative pressure plate heat exchanger for a superfluid helium system, comprising: a heat exchanger body, the heat exchanger body comprising a plurality of stacked heat exchange plates, and flow channels formed between adjacent heat exchange plates.

[0006] Along the thickness direction of the heat exchange plate, each of the two walls of the heat exchange plate is provided with a fluid core heat exchange unit and two fluid distribution units. Along the length direction of the heat exchange plate, the fluid core heat exchange unit is located between two fluid distribution units. The fluid core heat exchange unit includes multiple adjacent first herringbone corrugated structures, and the fluid distribution unit includes multiple adjacent second herringbone corrugated structures. On the same wall, the first herringbone corrugated structures and the second herringbone corrugated structures face the same direction, and the first herringbone corrugated structures and the second herringbone corrugated structures on adjacent heat exchange plates face opposite directions. The ratio of the length of the fluid core heat exchange unit to the width of the heat exchange plate is... L co / W 0The ratio of the corrugation pitch to the corrugation depth of the first and second herringbone corrugated structures is between 0.72 and 4.75. p / b All ranged from 1.4 to 6.7.

[0007] According to the negative pressure plate heat exchanger for a superfluid helium system provided by the present invention, the length of the fluid core heat exchange unit is 130 mm to 380 mm, and the width of the heat exchange plates is 80 mm to 180 mm; and / or, the corrugation pitch of the first herringbone corrugated structure and the second herringbone corrugated structure is... p Both are 5mm to 10mm deep, and the corrugation depth of the first herringbone corrugated structure and the second herringbone corrugated structure are... b All range from 1.5mm to 3.5mm.

[0008] According to the negative pressure plate heat exchanger for superfluid helium systems provided by the present invention, the corrugation angle β of the first herringbone corrugated structure is 30° to 65°; and / or, the corrugation angle γ of the second herringbone corrugated structure is 45° to 65°.

[0009] According to the present invention, the negative pressure plate heat exchanger for a superfluid helium system has a heat exchange plate length of 250 mm to 500 mm; and / or, the thickness of the heat exchange plate... δ The thickness ranges from 0.4 mm to 0.6 mm.

[0010] According to the negative pressure plate heat exchanger for a superfluid helium system provided by the present invention, the fluid distribution unit is provided with corner holes, and the corner holes on the plurality of heat exchange plates form a fluid distribution channel, the fluid distribution channel communicating with the flow channel, and the diameter of the corner hole is... D p The diameter is 20mm to 35mm.

[0011] According to the present invention, a negative pressure plate heat exchanger for a superfluid helium system is provided on the heat exchanger body, wherein the heat exchanger body is provided with a first fluid interface, a second fluid interface, a third fluid interface and a fourth fluid interface, the first fluid interface and the third fluid interface are used to connect to a supercritical helium / liquid helium pipeline, and the second fluid interface and the fourth fluid interface are used to connect to a negative pressure helium pipeline.

[0012] According to the negative pressure plate heat exchanger for a superfluid helium system provided by the present invention, the first fluid interface, the second fluid interface, the third fluid interface and the fourth fluid interface are all provided with ultra-high vacuum flanges.

[0013] Another aspect of the present invention provides an optimized design method for a negative pressure plate heat exchanger for a superfluid helium system, comprising the following steps.

[0014] Based on the structural features of the first and second herringbone corrugated structures on the heat exchange plates, including the corrugation pitch and depth of the first and second herringbone corrugated structures, as well as the length of the fluid core heat exchange unit and the width of the heat exchange plates, a flow and heat transfer numerical model of the fluid core heat exchange unit is established. Using computational fluid dynamics methods, the heat transfer performance of the fluid core heat exchange unit is obtained. j Factor criterion relationship, characterizing the flow resistance performance of the fluid core heat exchange unit. f The factor criterion relationship, and the empirical relationship for obtaining the pressure drop of the fluid distribution unit of the heat exchange plate.

[0015] Key geometric parameters of the heat exchanger body of the negative pressure plate heat exchanger are selected, and a dimensionless optimization objective function is constructed with the heat exchanger efficiency, the total pressure drop on the negative pressure side of the heat exchanger body, and the volume of the heat exchanger body as objectives.

[0016] Based on the above j Factor criterion relationship, the stated f Using the factor criterion relationship, the pressure drop empirical relationship of the fluid distribution unit, and the dimensionless optimization objective function, the particle swarm optimization algorithm is employed to perform multi-objective optimization on the overall structure of the heat exchanger body, obtaining the Pareto front solution set.

[0017] Based on the preset heat exchanger performance constraints, negative pressure side pressure drop constraints, and volume constraints, the final structural geometric parameters of the heat exchanger body are screened and determined from the Pareto front solution set.

[0018] According to the optimization design method for a negative pressure plate heat exchanger for a superfluid helium system provided by the present invention, the dimensionless optimization objective function is: ; ; ; in, To average heat exchanger efficiency; The mean Euler number; The volume is dimensionless. Re It is the Reynolds number; Pr It is a Prandtl number; N The number of flow channels in the main body of the heat exchanger; T For fluid temperature; c p The specific heat capacity at constant pressure of the fluid; ρ For fluid density; μ For fluid viscosity; λ The fluid's thermal conductivity; mThe mass flow rate in the channel; L The length of the flow channel; W The width of the flow channel; β The corrugation angle of the first herringbone corrugated structure; p For corrugated pitch; b Ripple depth; δ The thickness of the heat exchange plate; Δp The pressure drop in the flow channel ；V The volume of the heat exchanger body is denoted by ; u is the fluid velocity; subscript h represents the hot fluid in the heat exchanger body; subscript c represents the cold fluid in the heat exchanger body; subscript in represents the fluid inlet of the heat exchanger body; and subscript out represents the fluid outlet of the heat exchanger body.

[0019] The optimized design method for a negative pressure plate heat exchanger for a superfluid helium system provided by the present invention, wherein... j The factor criterion relationship is as follows: ; ; ; in, Re It is the Reynolds number; Pr It is a Prandtl number; Nu For Nusselt numbers; The average convective heat transfer coefficient; Q Total heat exchange; A The heat exchange area of ​​the heat exchanger body; D e The equivalent diameter of the flow channel of the negative pressure plate heat exchanger; λ The fluid's thermal conductivity; T The value represents the fluid temperature; the subscript w represents the heat exchange wall surface of the heat exchanger body; the subscript in represents the fluid inlet of the heat exchanger body; and the subscript out represents the fluid outlet of the heat exchanger body. And / or, the f The factor criterion relationship is as follows: ; in, ρ For fluid density, G The mass flow rate in the channel is [missing information]. L Δ is the length of the flow channel. p The pressure drop in the flow channel; And / or, the empirical relationship for the pressure drop of the fluid distribution unit of the heat exchange plates is: ; ; in, P 1 represents the inlet pressure of the inlet fluid distribution unit. P 2 represents the outlet pressure of the inlet fluid distribution unit. P 3 represents the inlet pressure of the outlet fluid distribution unit. P 4 represents the outlet pressure of the outlet fluid distribution unit, Δ P dis The pressure drop of the fluid distribution unit.

[0020] The negative pressure plate heat exchanger for superfluid helium systems provided by this invention employs a compact heat exchanger body composed of multiple stacked heat exchange plates with flow channels formed between adjacent plates. Each heat exchange plate has a fluid core heat exchange unit with a first herringbone corrugated structure and a fluid distribution unit with a second herringbone corrugated structure on two walls along its thickness direction. The length-to-width ratio of the fluid core heat exchange unit to the heat exchange plate is defined. L co / W 0 The depth ratio is 0.72 to 4.75, and the nodal depth ratios of the first herringbone corrugated structure and the second herringbone corrugated structure are defined. p / b With a heat transfer efficiency ranging from 1.4 to 6.7, it can fully utilize the inherent advantages of brazed plate heat exchangers, such as mature manufacturing processes, compact size, and low cost. At the same time, through the specific corrugated structure layout and optimized design of key geometric parameter ratios, it effectively improves the heat transfer efficiency and fluid distribution uniformity of the plate surface under the extreme low temperature negative pressure conditions of superfluid helium. Furthermore, the structured fluid distribution unit enhances the auxiliary heat exchange function of the distribution zone, further improving the overall compactness and heat exchange efficiency. This solves the problems of existing technologies such as plate-fin and coiled tube heat exchangers, which are difficult to balance in terms of reducing size and cost, and conventional brazed plate heat exchangers, which cannot be directly and reliably applied due to the lack of reliable thermal-hydraulic design basis for the superfluid helium temperature range.

[0021] The present invention provides an optimization design method for negative pressure plate heat exchangers in superfluid helium systems. Based on the specific herringbone corrugated structure of the heat exchange plates, a numerical model of the flow and heat transfer of the core fluid heat exchange unit is established. Computational fluid dynamics methods are then used to obtain key criterion relationships for heat transfer and flow resistance under superfluid helium conditions, as well as empirical relationships for pressure drop in the fluid distribution unit. This provides a precise and specific theoretical basis for the design. Furthermore, by selecting key geometric parameters, a dimensionless optimization objective function is constructed, focusing on heat exchanger efficiency, negative pressure side pressure drop, and volume. This simplifies the complex engineering problem into a computable mathematical solution. The model is then used; subsequently, the particle swarm optimization algorithm is employed to efficiently solve the multi-objective problem and obtain a Pareto front solution set representing the performance trade-offs; finally, the optimal structural parameters are selected from the solution set based on actual engineering constraints, thus forming a complete and systematic design process from theoretical modeling, objective quantification, intelligent optimization to engineering selection. This method can significantly reduce the design difficulty and cycle of plate heat exchangers for superfluid helium negative pressure conditions, effectively avoiding the traditional design mode that relies on experience or repeated trial and error, and providing a scientific, reliable and efficient design means for obtaining optimized heat exchanger products with compact structure, high heat exchange efficiency and low pressure drop. Attached Figure Description

[0022] To more clearly illustrate the technical solutions in this invention or the prior art, the drawings used in the description of the embodiments or the prior art will be briefly introduced below. Obviously, the drawings described below are some embodiments of this invention. For those skilled in the art, other drawings can be obtained from these drawings without creative effort.

[0023] Figure 1 This is a schematic diagram of a negative pressure plate heat exchanger for a superfluid helium system provided in an embodiment of the present invention.

[0024] Figure 2 This is a schematic diagram illustrating the working principle of a negative pressure plate heat exchanger for a superfluid helium system provided in an embodiment of the present invention.

[0025] Figure 3 This is one of the schematic diagrams of heat exchange plates in a negative pressure plate heat exchanger for a superfluid helium system provided in an embodiment of the present invention.

[0026] Figure 4 This is the second schematic diagram of the heat exchange plates in the negative pressure plate heat exchanger for the superfluid helium system provided in the embodiments of the present invention.

[0027] Figure 5 This is a schematic diagram of the herringbone corrugated structure of the heat exchange plates in the negative pressure plate heat exchanger for the superfluid helium system provided in the embodiments of the present invention.

[0028] Figure 6This is a flowchart illustrating the optimized design method for a negative pressure plate heat exchanger for a superfluid helium system provided in an embodiment of the present invention.

[0029] Figure 7 This is a schematic diagram of the modeling and solution of the boundary conditions for the herringbone corrugated structure in the optimization design method for a negative pressure plate heat exchanger for a superfluid helium system provided in an embodiment of the present invention.

[0030] Figure 8 This is a Pareto front diagram of the first herringbone corrugated structure at 45° in the optimized design method of the negative pressure plate heat exchanger for the superfluid helium system provided in the embodiments of the present invention.

[0031] Figure 9 This is a Pareto front diagram of the first herringbone corrugated structure at 65° in the optimized design method of the negative pressure plate heat exchanger for the superfluid helium system provided in the embodiments of the present invention.

[0032] Figure label: 100. Heat exchanger body; 110. Heat exchange plates; 111. Fluid core heat exchange unit; 1111. First herringbone corrugated structure; 112. Fluid distribution unit; 1121. Second herringbone corrugated structure; 1122. Corner hole; 120. First fluid interface; 130. Second fluid interface; 140. Third fluid interface; 150. Fourth fluid interface; 200. Ultra-high vacuum flange; 300. Reducer. Detailed Implementation

[0033] To make the objectives, technical solutions, and advantages of this invention clearer, the technical solutions of this invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some, not all, of the embodiments of this invention. All other embodiments obtained by those skilled in the art based on the embodiments of this invention without creative effort are within the scope of protection of this invention.

[0034] The following is combined with Figures 1 to 9 This invention describes a negative pressure plate heat exchanger for a superfluid helium system and its optimized design method.

[0035] See Figures 1 to 5 As shown, the negative pressure plate heat exchanger for a superfluid helium system provided in this embodiment of the invention includes: a heat exchanger body 100, the heat exchanger body 100 including multiple stacked heat exchange plates 110, and flow channels formed between adjacent heat exchange plates 110.

[0036] Along the thickness direction of the heat exchange plate 110, each of the two walls of the heat exchange plate 110 is provided with a fluid core heat exchange unit 111 and two fluid distribution units 112. Along the length direction of the heat exchange plate 110, the fluid core heat exchange unit 111 is located between the two fluid distribution units 112. The fluid core heat exchange unit 111 includes multiple adjacent first herringbone corrugated structures 1111, and the fluid distribution unit 112 includes multiple adjacent second herringbone corrugated structures 1121. On the same wall, the first herringbone corrugated structures 1111 and the second herringbone corrugated structures 1121 have the same orientation, and the first herringbone corrugated structures 1111 and the second herringbone corrugated structures 1121 on adjacent heat exchange plates 110 have opposite orientations. The ratio of the length of the fluid core heat exchange unit 111 to the width of the heat exchange plate 110 is... L co / W 0 The ratio of the corrugation pitch to the corrugation depth of the first herringbone corrugated structure 1111 to the second herringbone corrugated structure 1121 is between 0.72 and 4.75. p / b All ranged from 1.4 to 6.7.

[0037] The negative pressure plate heat exchanger for a superfluid helium system provided by the present invention employs a compact heat exchanger body 100 composed of multiple stacked heat exchange plates 110 with flow channels formed between adjacent plates. Each heat exchange plate 110 has a fluid core heat exchange unit 111 composed of a first herringbone corrugated structure 1111 and a fluid distribution unit 112 composed of a second herringbone corrugated structure 1121 on two walls along its thickness direction. The length-to-width ratio of the fluid core heat exchange unit 111 to the heat exchange plate 110 is defined. L co / W 0 The ratio is 0.72 to 4.75, and the depth ratio of the first herringbone corrugated structure 1111 and the second herringbone corrugated structure 1121 is defined. p / b With a heat transfer efficiency ranging from 1.4 to 6.7, it can fully utilize the inherent advantages of brazed plate heat exchangers, such as mature manufacturing processes, compact size, and low cost. At the same time, through the specific corrugated structure layout and optimized design of key geometric parameter ratios, it effectively improves the heat transfer efficiency and fluid distribution uniformity of the plate surface under the extreme low temperature negative pressure conditions of superfluid helium. Furthermore, the structured fluid distribution unit 112 enhances the auxiliary heat exchange function of the distribution zone, further improving the overall compactness and heat exchange efficiency. This solves the problems in existing technologies, such as plate-fin and coiled tube heat exchangers, which are difficult to balance in terms of reducing size and cost, and the problems in conventional brazed plate heat exchangers, which cannot be directly and reliably applied due to the lack of reliable thermal-hydraulic design basis for the superfluid helium temperature range.

[0038] Specifically, the heat exchanger body 100 is used to realize heat exchange between hot and cold fluids. Its core is a plate bundle formed by multiple heat exchange plates 110 sealed and stacked by brazing. Each heat exchange plate 110 has a specific herringbone corrugated structure on its two main walls. These herringbone corrugated structures are divided into two functional areas: the main area located in the middle of the heat exchange plate 110 constitutes the fluid core heat exchange unit 111, which is densely covered with the first herringbone corrugated structure 1111 and mainly undertakes the core heat exchange task; the area located at both ends of the plate and surrounding the corner holes 1122 constitutes the fluid distribution unit 112, which is densely covered with the second herringbone corrugated structure 1121. Its main function is to realize the uniform distribution of fluid in the width direction of the plate, and it can also participate in auxiliary heat exchange. In order to form a stable and unobstructed channel between the stacked plates, the corrugations on the same plate are aligned in the same direction, while the corrugations of two adjacent stacked plates are set to opposite directions. By adjusting the ratio of the length of the fluid core heat exchange unit 111 to the width of the heat exchange plate 110... L co / W 0 Limited to 0.72 to 4.75, and the pitch depth ratio for corrugations. p / b The pressure is limited to between 1.4 and 6.7, ensuring that the heat exchanger achieves an optimal balance between heat transfer efficiency, flow resistance, and overall volume under superfluid helium negative pressure conditions. Furthermore, corner holes 1122 at the ends of the plates are used for fluid collection and distribution, and connections to external pipelines are achieved through fluid inlet and outlet interfaces.

[0039] See Figure 3 and Figure 4 As shown, according to some embodiments of the present invention, the length of the fluid core heat exchange unit 111 is 130 mm to 380 mm, and the width of the heat exchange plate 110 is 80 mm to 180 mm.

[0040] By optimizing the length of the fluid core heat exchange unit 111 within the range of 130mm to 380mm and limiting the width of the heat exchange plates 110 to between 80mm and 180mm, the ratio of the effective heat exchange area to the fluid flow cross-section can be precisely controlled. This ensures sufficient heat exchange capacity while effectively managing the fluid velocity and pressure drop. The synergistic limitation of these two key dimensions provides a direct geometric basis for the heat exchanger to achieve the comprehensive performance goals of high efficiency, compactness, and low pressure loss under the extreme conditions of superfluid helium. This allows the heat exchanger body 100 to better adapt to the stringent requirements of superfluid helium systems for equipment miniaturization and high efficiency.

[0041] As an example, the length of the fluid core heat exchange unit 111 can be 130mm, 200mm, 250mm, 300mm, or 380mm, etc. The width of the heat exchange plate 110 can be 80mm, 120mm, 150mm, or 180mm, etc.

[0042] See Figures 3 to 5 As shown, according to some embodiments of the present invention, the corrugation pitch of the first herringbone corrugated structure 1111 and the second herringbone corrugated structure 1121 is... p Both are 5mm to 10mm deep, with the first herringbone corrugated structure 1111 and the second herringbone corrugated structure 1121 having corrugation depths of 5mm to 10mm. b All range from 1.5mm to 3.5mm.

[0043] By adjusting the corrugation pitch of the first herringbone corrugated structure 1111 and the second herringbone corrugated structure 1121 p All are limited to the range of 5mm to 10mm, and the corrugation depth is also limited. b Both are limited to the range of 1.5mm to 3.5mm, which can precisely control the hydraulic diameter and turbulence intensity of the flow channel. These two parameters together determine the degree of fluid disturbance and contact area in the flow channel. Their optimized combination enhances the heat transfer surface while avoiding unnecessary surges in flow resistance caused by excessively tortuous or narrow flow channels. Thus, under the special fluid properties of superfluid helium low temperature and low viscosity, an ideal balance between enhanced heat transfer and pressure drop control is achieved, providing a key microscale structural guarantee for obtaining excellent comprehensive thermal-hydraulic performance.

[0044] As an example, the corrugation pitch of the first herringbone corrugated structure 1111 and the second herringbone corrugated structure 1121 p All can be 5mm, 8mm, or 10mm, etc. The corrugation depth of the first herringbone corrugated structure 1111 and the second herringbone corrugated structure 1121... b All can be 1.5mm, 2mm, 3mm or 3.5mm, etc.

[0045] See Figure 3 and Figure 4 As shown, according to some embodiments of the present invention, the corrugation angle β of the first herringbone corrugated structure 1111 is 30° to 65°.

[0046] By optimizing and limiting the corrugation angle β of the first herringbone corrugated structure 1111 to the range of 30° to 65°, the fluid can be effectively guided to form specific three-dimensional rotation and secondary flow within the flow channel, thereby significantly enhancing the momentum and heat exchange between the fluid and the plate wall. Furthermore, this angle range is optimized for the heat transfer mechanism under superfluid helium cryogenic and low Prandtl number conditions, synergistically balancing the contradictory relationship between promoting turbulence development to enhance heat transfer and controlling the increase in flow resistance. It is one of the key geometric parameters for improving the heat transfer efficiency of the fluid core heat exchange unit 111.

[0047] As an example, the corrugation angle β of the first herringbone corrugated structure 1111 can be 30°, 45° or 65°, etc.

[0048] See Figure 3 and Figure 4 As shown, according to some embodiments of the present invention, the corrugation angle γ of the second herringbone corrugated structure 1121 is 45° to 65°.

[0049] By optimizing the corrugation angle γ of the second herringbone corrugated structure 1121 to 45° to 65°, the corrugations in the fluid distribution unit 112 can have a steeper guiding angle. This not only ensures that the fluid can be quickly and evenly dispersed to the core heat exchange area across the entire plate width after entering through the corner hole 1122, effectively improving the uniformity of flow distribution and reducing distribution pressure loss, but also significantly enhances the heat exchange capacity of the corrugated surface in this area. This allows the distribution area, which traditionally mainly performs the function of guiding flow, to also contribute considerable heat exchange, thereby further improving the compactness and heat transfer efficiency of the heat exchanger as a whole, and realizing the integrated enhancement of distribution and heat exchange functions.

[0050] As an example, the corrugation angle γ of the second herringbone corrugated structure 1121 can be 45°, 60° or 65°, etc.

[0051] See Figure 3 and Figure 4 As shown, according to some embodiments of the present invention, the length of the heat exchange plate 110 is 250 mm to 500 mm.

[0052] By optimizing and limiting the length of the heat exchange plates 110 to the range of 250mm to 500mm, sufficient flow channel length can be provided for the fluid to ensure the necessary heat exchange stroke, allowing enough time for heat exchange between the hot and cold fluids to achieve the expected temperature change. At the same time, this length range works in conjunction with the optimized values ​​of the plate width and core area length mentioned above to ensure that the overall dimensions of the heat exchanger body 100 can be reasonably controlled while meeting the requirements of high-efficiency heat exchange performance. This avoids volume redundancy caused by excessively long plates or insufficient heat exchange caused by excessively short plates, thus meeting the compact installation space requirements of the superfluid helium system for key equipment.

[0053] As an example, the length of the heat exchange plate 110 can be 250mm, 300mm, 400mm or 500mm, etc.

[0054] See Figure 5 As shown, according to some embodiments of the present invention, the thickness of the heat exchange plate 110 is... δ The thickness ranges from 0.4 mm to 0.6 mm.

[0055] By optimizing and limiting the thickness δ of the heat exchange plate 110 to 0.4 mm to 0.6 mm, the thermal resistance of the plate itself can be minimized and the overall weight of the heat exchanger body 100 can be reduced while ensuring sufficient mechanical strength of the plate structure and reliability of the brazing process. This is an optimization result that takes into account material cost, manufacturability, and heat transfer performance. Furthermore, the thinner heat exchange plate 110 helps to reduce the temperature drop caused by heat conduction through the plate, thereby improving the overall heat transfer coefficient. It also helps to achieve further lightweight and compact design of the heat exchanger.

[0056] As an example, the thickness of heat exchange plate 110 δ It can be 0.4mm, 0.5mm or 0.6mm, etc.

[0057] See Figure 3 and Figure 4 As shown, according to some embodiments of the present invention, the fluid distribution unit 112 is provided with corner holes 1122, and the corner holes 1122 on the plurality of heat exchange plates 110 form a fluid distribution channel, which communicates with the flow channel. The diameter of the corner hole 1122 is... D p The diameter is 20mm to 35mm.

[0058] The corner hole 1122 forms the main inlet and outlet for fluid entering and exiting the flow channels between the plates, and is a key structure for achieving uniform distribution and efficient collection of fluid among multiple channels. By adjusting the diameter of the corner hole 1122... D pThe optimized design, within the range of 20mm to 35mm, provides a flow cross-section that matches the plate size and design flow rate for the fluid distribution and collection process. This size range effectively coordinates the flow velocity and local resistance within the distribution channel, avoiding excessive distribution pressure drop and uneven flow distribution caused by excessively small corner holes 1122, and also preventing excessive encroachment on the valuable heat exchange area at the plate edge due to excessively large corner holes 1122. Thus, while ensuring uniform fluid distribution, it also ensures the full utilization of the overall heat transfer performance of the heat exchanger.

[0059] As an example, the diameter of corner hole 1122 D p It can be 20mm, 30mm or 35mm, etc.

[0060] See Figure 1 As shown, according to some embodiments of the present invention, the heat exchanger body 100 is provided with a first fluid interface 120, a second fluid interface 130, a third fluid interface 140 and a fourth fluid interface 150. The first fluid interface 120 and the third fluid interface 140 are used to connect to supercritical helium / liquid helium pipelines, and the second fluid interface 130 and the fourth fluid interface 150 are used to connect to negative pressure helium pipelines.

[0061] By providing a first fluid interface 120 and a third fluid interface 140 for connecting supercritical helium / liquid helium pipelines, and a second fluid interface 130 and a fourth fluid interface 150 for connecting negative pressure helium pipelines on the heat exchanger body 100, the fluid loops on the positive and negative pressure sides can be clearly defined and physically isolated, meeting the requirements of independent circulation and heat exchange of helium media at different pressures and in different phases in the superfluid helium system. This interface configuration provides clear and standard connection points for the functional integration of the heat exchanger in the system, facilitating docking with complex external cryogenic piping systems.

[0062] join Figure 1 As shown, according to some embodiments of the present invention, ultra-high vacuum flanges 200 are connected to the first fluid interface 120, the second fluid interface 130, the third fluid interface 140 and the fourth fluid interface 150.

[0063] By connecting ultra-high vacuum flanges 200 to the first fluid interface 120, the second fluid interface 130, the third fluid interface 140, and the fourth fluid interface 150, the connection interface between the heat exchanger and the external piping system can be ensured to have extremely high sealing performance, meeting the stringent requirements of the ultrafluid helium system for extremely high vacuum and prevention of micro-leakage. This standardized flange connection method not only provides reliable vacuum sealing, but also greatly simplifies the installation, disassembly, and maintenance process of the heat exchanger in complex cryogenic pipelines, improving the system's engineering applicability and maintainability.

[0064] Specifically, the ultra-high vacuum flange 200 can be modeled as CF35 and CF63, etc.

[0065] join Figure 1 As shown, according to some embodiments of the present invention, a reducing pipe 300 is connected to the first fluid interface 120, the second fluid interface 130, the third fluid interface 140 and the fourth fluid interface 150.

[0066] By connecting reducers 300 to the first fluid interface 120, the second fluid interface 130, the third fluid interface 140, and the fourth fluid interface 150, the differences between the heat exchanger interface size and the external system piping size can be flexibly adapted, effectively solving the installation difficulties caused by interface specification mismatch. The reducer 300 serves as a universal adapter solution, enabling the same model of heat exchanger to be easily and reliably connected to ultra-high vacuum pipes of various diameters through reducers 300 of different specifications, significantly improving the adaptability and installation convenience of this negative pressure plate heat exchanger in diverse ultrafluid helium system engineering applications.

[0067] See Figure 2 As shown, during operation, supercritical helium or liquid helium flows into the heat exchanger body 100 from the positive pressure side pipe through the first fluid interface 120 and flows within the positive pressure side channel. Simultaneously, negative pressure helium flows into the heat exchanger body 100 from the negative pressure side pipe through the second fluid interface 130 and flows within the negative pressure side channel. The two fluids flow in opposite directions or cross each other within closely adjacent channels separated by plates, achieving efficient heat exchange through the plate walls. The cooled positive pressure side helium medium flows out from the third fluid interface 140 and returns to the positive pressure side system, while the heated negative pressure side helium flows out from the fourth fluid interface 150 and returns to the negative pressure side system, thus completing a complete heat exchange cycle and achieving the functions of cooling the negative pressure helium or heating the positive pressure helium medium, meeting the specific temperature control requirements of the superfluid helium system.

[0068] The following describes the optimized design method of the negative pressure plate heat exchanger for the superfluid helium system provided by the present invention. The optimized design method of the negative pressure plate heat exchanger for the superfluid helium system described below can be referred to in correspondence with the negative pressure plate heat exchanger for the superfluid helium system described above.

[0069] See Figures 6 to 9 As shown in the figure, the optimized design method for a negative pressure plate heat exchanger for a superfluid helium system provided by the embodiments of the present invention includes the following steps.

[0070] S610. Based on the structural features of the first herringbone corrugated structure 1111 and the second herringbone corrugated structure 1121 on the heat exchange plate 110, the structural features include the corrugation pitch and corrugation depth of the first herringbone corrugated structure 1111 and the second herringbone corrugated structure 1121, as well as the length of the fluid core heat exchange unit 111 and the width of the heat exchange plate 110. A flow and heat transfer numerical model of the fluid core heat exchange unit 111 is established. Through computational fluid dynamics methods, the heat transfer performance of the fluid core heat exchange unit 111 is obtained. j Factor criterion relationship, characterizing the flow resistance performance of fluid core heat exchange unit 111 f Factor criterion relationship, and empirical relationship for obtaining pressure drop of fluid distribution unit 112 of heat exchange plate 110.

[0071] S620. Select the key geometric parameters of the heat exchanger body of the negative pressure plate heat exchanger, and construct a dimensionless optimization objective function with the heat exchanger efficiency, the total pressure drop on the negative pressure side of the heat exchanger body, and the volume of the heat exchanger body as objectives.

[0072] S630, based on j Factor criterion relationship f The factor criterion relationship, the pressure drop empirical relationship of fluid distribution unit 112, and the dimensionless optimization objective function are used to perform multi-objective optimization of the overall structure of the heat exchanger body using the particle swarm optimization algorithm, and the Pareto front solution set is obtained.

[0073] S640. Based on the preset heat exchanger performance constraints, negative pressure side pressure drop constraints, and volume constraints, the final structural geometric parameters of the heat exchanger body are selected and determined from the Pareto front solution set.

[0074] The present invention provides an optimized design method for a negative pressure plate heat exchanger in a superfluid helium system. Based on the specific herringbone corrugated structure of the heat exchange plate 110, a numerical model of the flow and heat transfer of the core fluid heat exchange unit 111 is established. Computational fluid dynamics is then used to obtain key criterion relationships for heat transfer and flow resistance under superfluid helium conditions, as well as empirical relationships for the pressure drop of the fluid distribution unit 112. This provides a precise and specific theoretical basis for the design. Furthermore, by selecting key geometric parameters, a dimensionless optimization objective function is constructed, focusing on heat exchanger efficiency, negative pressure side pressure drop, and volume. This simplifies the complex engineering problem into a manageable solution. The mathematical model for calculation is then used; subsequently, the particle swarm optimization algorithm is employed to efficiently solve the multi-objective problem and obtain the Pareto front solution set representing the performance trade-offs; finally, the optimal structural parameters are selected from the solution set based on actual engineering constraints, thus forming a complete and systematic design process from theoretical modeling, objective quantification, intelligent optimization to engineering selection. This method can significantly reduce the design difficulty and cycle of plate heat exchangers for superfluid helium negative pressure conditions, effectively avoiding the traditional design mode that relies on experience or repeated trial and error, and providing a scientific, reliable and efficient design means for obtaining optimized heat exchanger products with compact structure, high heat exchange efficiency and low pressure drop.

[0075] Specifically, in step S610, the flow channel structure can be simplified in Fluent software based on the structural characteristics of the herringbone corrugated flow channel to construct a simplified flow and heat transfer model. Simplification conditions may include: using three-dimensional variable radius fillets to represent the geometry of the brazing joint; selecting the negative pressure side fluid domain as the research object; the negative pressure side fluid being an incompressible fluid; ignoring the influence of gravity on the fluid; ignoring the solid domain and only studying the fluid domain; and ignoring the effect of changes in the Pr number on the fluid's properties. j The influence of factors is ignored; the temperature field change on the positive pressure side is ignored, and the boundary condition is set to constant wall temperature.

[0076] The computational domain geometry model can be meshed using FluentMeshing, employing a polyhedral mesh. In Fluent software, the energy and momentum conservation equations for a second-order upwind scheme are used to numerically calculate the cryogenic helium gas within a herringbone corrugated channel, yielding the pressure and temperature fields. Furthermore, based on these pressure and temperature fields, different calculations can be performed. Re The number j Factors and f Factors were fitted to obtain the fluid core heat exchange unit 111. j Factors and f The criterion relationship of the factor, and the empirical relationship of the pressure drop of the fluid distribution unit 112 with respect to the dynamic head VP.

[0077] The specific data processing methods are as follows: The inlet temperature is taken as the temperature at the velocity inlet boundary, the outlet temperature is taken as the mass-weighted average temperature at the outlet section, and the wall temperature is based on the constant wall temperature boundary condition.

[0078] Total heat exchange area Determined by the following formula: ; in, A p For the projected area, m 2 .

[0079] Total heat exchange Determined by the following formula: ; Average convective heat transfer coefficient Determined by the following formula: ; The Nu number is determined by the following formula: ; j The factor is determined by the following formula: ; in, T Let K be the fluid temperature. Q The total heat exchange is expressed in W. c p Specific heat capacity at constant pressure of the fluid, J / (kg·K); ρ The fluid density is expressed in kg / m³. λ Where W is the thermal conductivity of the fluid, W / (m·K); m The mass flow rate in the flow channel is kg / s; The cross-sectional area of ​​the flow channel is represented by the subscript w, which represents the heat exchange wall surface of the heat exchanger body. The subscripts in and out represent the fluid inlet and fluid outlet of the heat exchanger body, respectively.

[0080] f The factor is determined by the following formula: ; in, ρ The fluid density is expressed in kg / m³. D e is the equivalent diameter of the flow channel of the negative pressure plate heat exchanger, in meters; G The mass flow rate in the channel is kg / (m·s); L Δ is the length of the flow channel, in meters (m); p Let Pa be the pressure drop in the flow channel.

[0081] Pressure drop Δ of fluid distribution unit 112 Pdis Determined by the following formula: ; ; in, P 1 represents the inlet pressure of the inlet fluid distribution unit 112; P 2 represents the outlet pressure of the inlet fluid distribution unit 112; P 3 represents the inlet pressure of the outlet fluid distribution unit 112; P 4 represents the outlet pressure of the outlet fluid distribution unit 112; Δ P dis The pressure drop of the fluid distribution unit 112.

[0082] In step S620, the key geometric parameters most sensitive to heat transfer, pressure drop and volume are selected from the herringbone corrugations and plate macroscopic dimensions as design variables. The heat exchanger efficiency, total pressure drop on the negative pressure side and volume are listed as optimization objectives. Through dimensionless processing, the multiple objectives are transformed into a unified and comparable optimization objective function, providing a comprehensive performance evaluation basis for subsequent algorithm search that takes into account efficient heat exchange, low pressure loss and compact structure.

[0083] Key geometric parameters may include the corrugation angle, corrugation height, and corrugation pitch of the herringbone corrugated structure, as well as the length of the fluid core heat exchange unit 111 of the heat exchange plate 110 and the width of the heat exchange plate 110, which have a significant impact on the heat exchanger efficiency, negative pressure side pressure drop, and overall volume.

[0084] Heat exchanger hot fluid outlet temperature and its independent variable parameters: ; in, N Number of flow channels; T Let K be the fluid temperature. c p Specific heat capacity at constant pressure of the fluid, J / (kg·K); ρ The fluid density is expressed in kg / m³. μ Where is the fluid viscosity, Pa·s; λ Where W is the thermal conductivity of the fluid, W / (m·K); m The mass flow rate in the flow channel is kg / s; the subscripts h and c represent the hot fluid and cold fluid in the heat exchanger body, respectively; the subscripts in and out represent the fluid inlet and fluid outlet of the heat exchanger body, respectively.

[0085] After dimensionless transformation, a dimensionless relationship can be obtained regarding the average heat exchanger efficiency: ; Pressure drop on the negative pressure side and its independent variable parameters: ; in, L The length of the flow channel is in meters (m). W The width of the flow channel is in meters (m). β The corrugation angle of the first herringbone corrugated structure 1111; p The pitch is the corrugation pitch, in meters (m). b The depth of the ripple is m.

[0086] After dimensionless transformation, we can obtain a dimensionless relation regarding the mean Euler number: ; Heat exchanger volume and its independent variable parameters: ; in, L The length of the flow channel is in meters (m). W The width of the flow channel is in meters (m). b The depth of the ripple is m.

[0087] After dimensionless transformation, we can obtain dimensionless relations regarding volume characteristics: .

[0088] in, To average heat exchanger efficiency; The mean Euler number; The volume is dimensionless. T For fluid temperature, m This refers to the mass flow rate in the flow channel; Re It is the Reynolds number; Pr It is a Prandtl number; N The number of flow channels in the main body of the heat exchanger; c p The specific heat capacity at constant pressure of the fluid; ρ For fluid density; μ For fluid viscosity; λ The fluid's thermal conductivity; L The length of the flow channel; W The width of the flow channel; β The corrugation angle of the first herringbone corrugated structure; p For corrugated pitch; b Ripple depth; Δp Pressure drop in the flow channel ； V denoted by , where is the volume of the heat exchanger body; u is the fluid velocity; subscript h represents the hot fluid in the heat exchanger body; subscript c represents the cold fluid in the heat exchanger body; subscript in represents the fluid inlet of the heat exchanger body; and subscript out represents the fluid outlet of the heat exchanger body.

[0089] In step S630, the result obtained in step S610 is... j Factor criterion relationshipf The factor criterion relationship and the pressure drop empirical relationship of the fluid distribution unit 112 are embedded in the fitness calculation module, enabling the particle swarm optimization algorithm to quickly predict the heat exchanger efficiency, total pressure drop and volume on the negative pressure side under different combinations of geometric parameters in each iteration. Based on the dimensionless optimization objective function, the algorithm continuously converges to the optimal boundary by updating the velocity and position of particles in the multidimensional parameter space, and finally outputs a Pareto front solution set composed of a set of non-dominated solutions. This provides a set of candidate design schemes that take into account high-efficiency heat exchange, low pressure loss and compact structure for subsequent screening according to engineering constraints.

[0090] The outlet temperature of the heat exchanger hot fluid can be calculated using the following formula: ; ; ; in, Q The total heat exchange is expressed in W. m The mass flow rate in the flow channel is kg / s; c p Δ is the specific heat capacity of the fluid at constant pressure, J / (kg·K); T m The logarithmic mean temperature difference, in K; K The heat transfer coefficient is W / (m²). 2 ·K); α The convective heat transfer coefficient is W / (m²). 2 ·K); λ w The heat exchanger wall has a thermal conductivity of W / (m·K); the subscripts h and c represent the hot and cold fluids in the heat exchanger body, respectively; the subscript in represents the fluid inlet of the heat exchanger body; and the subscript out represents the fluid outlet of the heat exchanger body.

[0091] The total pressure drop on the negative pressure side is calculated by the following formula: ; ; Where, Δ P c Pressure drop of fluid core heat exchange unit 111, Pa; Δ P dis，in and Δ P dis，out The pressure drops of the inlet fluid distribution unit 112 and the outlet fluid distribution unit 112 are respectively, in Pa; ρ The fluid density is expressed in kg / m³. D e is the equivalent diameter of the flow channel of the negative pressure plate heat exchanger, in meters; G The mass flow rate in the channel is kg / (m·s); Let be the length of the fluid core heat exchange unit 111, in meters (m).

[0092] The heat exchanger volume is calculated using the following formula: ; in, L1 W1 is the total length of the heat exchanger body, in meters; W2 is the width of the heat exchanger body, in meters. δ The thickness of heat exchange plate 110 is in meters (m). b The depth of the ripple is in meters (m). N The number of flow channels in the heat exchanger body is denoted by h and c, which represent the hot fluid and cold fluid in the heat exchanger body, respectively.

[0093] The range of constraints for the optimization variables is shown in the table below:

[0094] In step S640, considering the extreme requirements of the superfluid helium system on heat exchange intensity, pumping power consumption, and installation space, the heat exchanger efficiency must not be lower than the minimum value required by the process, the total pressure drop on the negative pressure side must be less than the upper limit that the vacuum pump can withstand, and the overall volume of the unit must not exceed the allowable envelope of the compartment. The Pareto front solution set is sequentially subjected to hard constraint elimination and soft preference sorting, retaining the feasible subset that simultaneously satisfies the three boundary conditions. Then, according to the engineering weight of "efficiency first, pressure drop second, and volume smallest", the structural geometric parameter combination corresponding to the first place of the sort is selected as the final design scheme that takes into account thermal performance, operating energy consumption, and compactness, thus completing the final size and shape determination of the low-temperature negative pressure plate heat exchanger.

[0095] The following is a specific example illustrating the optimized design method of the negative pressure plate heat exchanger for the superfluid helium system provided by the present invention.

[0096] 1. Based on the structural characteristics of the herringbone corrugated flow channel in Fluent software, the flow channel structure was simplified, and a simplified flow and heat transfer model was constructed, such as... Figure 7 As shown.

[0097] 2. To ensure full development of the inlet flow, the inlet section of the computational domain is extended to 5 times the hydraulic diameter; to avoid the impact of outlet backflow on the flow characteristics of the simulated region and to increase computational stability, the outlet section is extended to 10 times the hydraulic diameter. The computational domain geometry model is meshed using Fluent Meshing, employing a polyhedral mesh.

[0098] The properties of helium are fitted to a polynomial of temperature, and then written into a user-defined file. During calculation, the properties of helium are called through the user-defined file.

[0099] like Figure 7As shown, the upper and lower walls of the fluid core heat exchange unit 111 adopt constant wall temperature boundary conditions, while the remaining walls adopt adiabatic wall boundary conditions. The inlet boundary condition adopts velocity inlet, the outlet boundary condition adopts pressure outlet, and the static pressure is 0.

[0100] 3. Calculate different pressure and temperature fields. Re The number j Factors and f Factors were fitted to obtain the fluid core heat exchange unit 111. j Factors and f The criterion relationship of the factor, and the empirical relationship of the pressure drop of the fluid distribution unit 112 with respect to the dynamic head VP: The criterion relationship between heat transfer and pressure drop in the fluid core heat exchange unit 111 is as follows: When the first herringbone corrugated structure 1111 has a corrugation angle of 45°: ; ; When the first herringbone corrugated structure 1111 has a corrugation angle of 65°: ; ; Empirical relationship for pressure drop in fluid distribution unit 112: ; ; in, G p ρ is the mass flow rate in orifice 1122, kg / (m·s); ρ is the fluid density, kg / m³; Δ P dis,in Pressure drop of inlet fluid distribution unit 112, Pa; Δ P dis,out The pressure drop of the outlet fluid distribution unit 112, in Pa; N The number of flow channels in the main body of the heat exchanger; Re 1 is the Reynolds number; VP is the dynamic head; subscript in is the fluid inlet of the heat exchanger body; subscript out is the fluid outlet of the heat exchanger body.

[0101] 4. Using Matlab software, write the objective function with the heat exchanger hot fluid outlet temperature, negative pressure side pressure drop, and heat exchanger volume as design objectives. Clarify the mathematical relationship between the optimization variables and the objective function, and then determine the constraints on the optimization variables based on engineering practice. Use the `particleswarm` function in Matlab to solve the multi-objective problem using the particle swarm optimization algorithm. Alternatively, a custom particle swarm optimization algorithm can be defined and adjusted. The Pareto front solutions are shown below. Figure 8 and Figure 9 As shown.

[0102] 5. Based on the solution results of the particle swarm optimization algorithm, determine the constraint range of the objective function according to the design requirements of the negative pressure heat exchanger, select suitable optimization results, and determine the final values ​​of the optimization variables in combination with the actual engineering situation.

[0103] The design requirements for negative pressure heat exchangers are shown in the table below:

[0104] The screening results are shown in the table below:

[0105] Finally, it should be noted that the above embodiments are only used to illustrate the technical solutions of the present invention, and not to limit them; although the present invention has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some of the technical features; and these modifications or substitutions do not cause the essence of the corresponding technical solutions to deviate from the spirit and scope of the technical solutions of the embodiments of the present invention.

Claims

1. A negative pressure plate heat exchanger for a superfluid helium system, characterized in that, include: The heat exchanger body includes multiple stacked heat exchange plates, and flow channels are formed between adjacent heat exchange plates. Along the thickness direction of the heat exchange plate, each of the two walls of the heat exchange plate is provided with a fluid core heat exchange unit and two fluid distribution units. Along the length direction of the heat exchange plate, the fluid core heat exchange unit is located between the two fluid distribution units. The fluid core heat exchange unit includes a plurality of adjacent first herringbone corrugated structures, and the fluid distribution unit includes a plurality of adjacent second herringbone corrugated structures. On the same wall, the first herringbone corrugated structures and the second herringbone corrugated structures have the same orientation, and the first herringbone corrugated structures and the second herringbone corrugated structures on adjacent heat exchange plates have opposite orientations. The ratio of the length of the fluid core heat exchange unit to the width of the heat exchange plate L co / W 0 The ratio of the corrugation pitch to the corrugation depth of the first and second herringbone corrugated structures is between 0.72 and 4.

75. p / b All ranged from 1.4 to 6.

7.

2. The negative pressure plate heat exchanger for a superfluid helium system according to claim 1, characterized in that, The length of the fluid core heat exchange unit is 130mm to 380mm, and the width of the heat exchange plate is 80mm to 180mm. And / or, the corrugation pitch of the first herringbone corrugated structure and the second herringbone corrugated structure p Both are 5mm to 10mm deep, and the corrugation depth of the first herringbone corrugated structure and the second herringbone corrugated structure are... b All range from 1.5mm to 3.5mm.

3. The negative pressure plate heat exchanger for a superfluid helium system according to claim 1, characterized in that, The corrugation angle β of the first herringbone corrugated structure is 30° to 65°. And / or, the corrugation angle γ of the second herringbone corrugated structure is 45° to 65°.

4. The negative pressure plate heat exchanger for a superfluid helium system according to claim 1, characterized in that, The length of the heat exchange plates is 250mm to 500mm; And / or, the thickness of the heat exchange plates δ The thickness ranges from 0.4 mm to 0.6 mm.

5. The negative pressure plate heat exchanger for a superfluid helium system according to claim 1, characterized in that, The fluid distribution unit is provided with corner holes, and the corner holes on the multiple heat exchange plates form a fluid distribution channel. The fluid distribution channel communicates with the flow channel, and the diameter of the corner hole is... D p The diameter is 20mm to 35mm.

6. The negative pressure plate heat exchanger for a superfluid helium system according to claim 5, characterized in that, The heat exchanger body is provided with a first fluid interface, a second fluid interface, a third fluid interface and a fourth fluid interface. The first fluid interface and the third fluid interface are used to connect to a supercritical helium / liquid helium pipeline, and the second fluid interface and the fourth fluid interface are used to connect to a negative pressure helium pipeline.

7. The negative pressure plate heat exchanger for a superfluid helium system according to claim 6, characterized in that, The first fluid interface, the second fluid interface, the third fluid interface and the fourth fluid interface are all equipped with ultra-high vacuum flanges.

8. An optimized design method for a negative pressure plate heat exchanger for a superfluid helium system, characterized in that, include: Based on the structural features of the first and second herringbone corrugated structures on the heat exchange plates, including the corrugation pitch and depth of the first and second herringbone corrugated structures, as well as the length of the fluid core heat exchange unit and the width of the heat exchange plates, a flow and heat transfer numerical model of the fluid core heat exchange unit is established. Using computational fluid dynamics methods, the heat transfer performance of the fluid core heat exchange unit is obtained. j Factor criterion relationship, characterizing the flow resistance performance of the fluid core heat exchange unit. f Factor criterion relationship, and empirical relationship for obtaining the pressure drop of the fluid distribution unit of the heat exchange plate; The key geometric parameters of the heat exchanger body of the negative pressure plate heat exchanger are selected, and a dimensionless optimization objective function is constructed with the heat exchanger efficiency, the total pressure drop on the negative pressure side of the heat exchanger body, and the volume of the heat exchanger body as objectives. Based on the above j Factor criterion relationship, the stated f The factor criterion relationship, the pressure drop empirical relationship of the fluid distribution unit, and the dimensionless optimization objective function are used to perform multi-objective optimization on the overall structure of the heat exchanger body using the particle swarm optimization algorithm to obtain the Pareto front solution set. Based on the preset heat exchanger performance constraints, negative pressure side pressure drop constraints, and volume constraints, the final structural geometric parameters of the heat exchanger body are screened and determined from the Pareto front solution set.

9. The optimized design method for a negative pressure plate heat exchanger for a superfluid helium system according to claim 8, characterized in that, The dimensionless optimization objective function is: ； ； ； in, To average heat exchanger efficiency; The mean Euler number; The volume is dimensionless. Re It is the Reynolds number; Pr It is a Prandtl number; N The number of flow channels in the main body of the heat exchanger; T For fluid temperature; c p The specific heat capacity at constant pressure of the fluid; ρ For fluid density; μ For fluid viscosity; λ The fluid's thermal conductivity; m The mass flow rate in the channel; L The length of the flow channel; W The width of the flow channel; β The corrugation angle of the first herringbone corrugated structure; p For corrugated pitch; b Ripple depth; δ The thickness of the heat exchange plate; Δp The pressure drop in the flow channel; V The volume of the heat exchanger body is denoted by ; u is the fluid velocity; subscript h represents the hot fluid in the heat exchanger body; subscript c represents the cold fluid in the heat exchanger body; subscript in represents the fluid inlet of the heat exchanger body; and subscript out represents the fluid outlet of the heat exchanger body.

10. The optimized design method for a negative pressure plate heat exchanger for a superfluid helium system according to claim 8, characterized in that, The j The factor criterion relationship is as follows: ； ； ； in, Re It is the Reynolds number; Pr It is a Prandtl number; Nu For Nusselt numbers; The average convective heat transfer coefficient; Q Total heat exchange; A The heat exchange area of ​​the heat exchanger body; D e The equivalent diameter of the flow channel of the negative pressure plate heat exchanger; λ The fluid's thermal conductivity; T The value represents the fluid temperature; the subscript w represents the heat exchange wall surface of the heat exchanger body; the subscript in represents the fluid inlet of the heat exchanger body; and the subscript out represents the fluid outlet of the heat exchanger body. And / or, the f The factor criterion relationship is as follows: ； in, ρ For fluid density, G The mass flow rate in the channel is [missing information]. L Δ is the length of the flow channel. p The pressure drop in the flow channel; And / or, the empirical relationship for the pressure drop of the fluid distribution unit of the heat exchange plates is: ； ； in, P 1 represents the inlet pressure of the inlet fluid distribution unit. P 2 represents the outlet pressure of the inlet fluid distribution unit. P 3 represents the inlet pressure of the outlet fluid distribution unit. P 4 represents the outlet pressure of the outlet fluid distribution unit, Δ P dis The pressure drop of the fluid distribution unit.

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