A zigzag evaporator with archimedes curve type and a method for calculating physical performance thereof

Abstract

The application discloses a kind of return archimedes curve type honeycomb sawtooth evaporator and its physical performance calculation method, the heat collecting plate body of the evaporator is integrally compounded by upper plate and lower plate by roll pressing inflation process, at least one is set in inside and is in the form of return archimedes curve shape spiral loop, and the inner wall of flow channel is distributed with hexagonal honeycomb sawtooth microstructure;Adopt upper and lower double spiral symmetry layout, configuration is arranged in parallel operation mode, and working medium interface is arranged on the same side of plate body.Calculation method is carried out in turn geometric characteristic parameter calculation, flow resistance calculation, heat exchange performance calculation, dry critical calculation and temperature uniformity evaluation, corresponding correction factor is introduced respectively for spiral flow field and sawtooth structure.The application has high heat exchange efficiency, controllable pressure drop, excellent anti-drying capacity and plate surface temperature uniformity, and supporting calculation method can effectively shorten research and development cycle, is adapted to heat pump, solar heat collection, industrial waste heat recovery and various application scenarios.

Description

Technical Field

[0001] This invention belongs to the field of thermal energy engineering and renewable energy utilization technology, specifically relating to a reverse Archimedes curve honeycomb sawtooth evaporator for phase change working fluids, and a method for calculating the physical performance of the evaporator. Background Technology

[0002] Evaporators, also known as heat-absorbing plates or evaporation plates, are core heat exchange components in fields such as solar thermal utilization, heat pump systems, industrial waste heat recovery, building radiant temperature control, and power battery thermal management. Their performance directly determines the system's thermal efficiency and operating costs. Under evaporation and condensation conditions using phase change working fluids such as Freon, carbon dioxide, and propane as heat transfer media, the phase change of the working fluid is accompanied by a significant volume expansion, which places higher demands on the evaporator's flow channel layout, heat transfer enhancement, and temperature uniformity performance.

[0003] Currently, evaporators used in phase change fluid applications mainly include tube sheet type, serpentine inflated plate type, spiral flow channel plate type, microchannel plate type, porous medium or foam metal plate type, etc. Tube sheet type evaporators typically fix metal round or flat tubes to the back of the absorber plate by welding or pressing. They are simple in structure, easy to process, and have strong pressure resistance, but the heat exchange area is limited, there is contact thermal resistance between the tube wall and the plate, the plate surface temperature difference is large, and "hot spots" are obvious between the tubes, resulting in a thermal efficiency of only 50%-65%. Serpentine inflated plates use two aluminum plates or composite plates to form a serpentine flow channel through printing, rolling, and high-pressure blowing processes. The flow channel and plate are integrally formed, with no contact thermal resistance, and the technology is mature. However, the single-channel serpentine flow leads to a large temperature gradient along the flow path, with an inlet-outlet temperature difference of 15-25℃, poor plate surface temperature uniformity, a small number of flow channels, low heat transfer density per unit area, and a tendency for liquid film stratification and drying under two-phase flow conditions.

[0004] Spiral flow channel plates, with their Archimedean spiral or rectangular volute arrangement, can generate secondary flow through centrifugal force to enhance heat transfer. The flow is continuous and the pressure gradient is uniform. However, traditional single-spiral structures only have one inlet and outlet, resulting in uneven area utilization between the central and peripheral regions. Differences in the length of the inner and outer flow channels can lead to uneven working fluid distribution, and significant local resistance exists at the bends of the rectangular volute. Microchannel plate evaporators, formed through extrusion or etching processes, create dense microchannel arrays, resulting in high heat transfer coefficients and small size. However, they are costly to manufacture, prone to channel blockage, and have large flow pressure drops, requiring extremely high working fluid cleanliness. Porous media or foamed metal plate evaporators use metal foam or sintered porous layers as heat transfer enhancement structures, offering extremely large heat transfer areas and significant turbulence effects. However, they suffer from extremely large flow pressure drops, poor pressure resistance, high manufacturing costs, and large working fluid retention.

[0005] In summary, existing evaporator structures still face several common technical challenges when operating with phase change refrigerants such as Freon, whether for evaporation or condensation: First, there is a significant contradiction between enhanced phase change heat transfer and pressure drop control. During evaporation, the phase change refrigerant expands 20-100 times in volume, causing a surge in flow velocity and a dramatic increase in pressure drop in the latter half of conventional constant-section flow channels, leading to a decrease in evaporation temperature and a drop in the system's coefficient of performance (COP). Second, uneven gas-liquid distribution poses a high risk of localized drying. Gravity in serpentine or straight pipe flow channels easily induces gas-liquid stratification, resulting in a thinner liquid film at the top. The first problem is that the plate is prone to cracking, and the heat transfer coefficient drops sharply after drying out, which severely limits the improvement of the critical heat flux (CHF). The second problem is that the plate has poor temperature uniformity and a large temperature gradient along the flow channel, which easily causes uneven thermal stress and plate deformation. The third problem is that the flow channel wall is smooth, with few nucleation boiling points and a high heat flux at the nucleation boiling initiation point (ONB), resulting in weak heat transfer under low load conditions. The fourth problem is that the spiral flow channel has low area utilization, uneven heat transfer between the center and outer ring of a single spiral, and the inlet and outlet of a double spiral are concentrated in the center, making pipeline layout difficult.

[0006] Furthermore, for evaporators with composite morphology combining macroscopic curved flow channels and microscopic sawtooth structures, there is currently a lack of rapid and accurate performance prediction methods. Existing empirical formulas have significant calculation errors, and the design phase heavily relies on costly CFD simulations or experimental testing, resulting in lengthy product design and optimization cycles. Therefore, there is an urgent need to develop a novel evaporator that combines high heat transfer efficiency, low flow resistance, good temperature uniformity, and anti-drying capabilities, while also possessing mass production feasibility, and to establish a corresponding rapid calculation method for its physical properties. Summary of the Invention

[0007] The purpose of this invention is to address the aforementioned deficiencies in existing technologies by providing a vortex evaporator with a honeycomb serrated microstructure and its physical performance calculation method. While ensuring the feasibility of mass production processes, this invention addresses two key issues: firstly, the composite design of the vortex main channel and the honeycomb serrated microstructure solves the prominent contradiction between heat transfer enhancement and pressure drop control under phase change conditions; secondly, it addresses the problems of uneven gas-liquid distribution leading to localized drying; thirdly, poor plate surface temperature uniformity causing thermal stress; fourthly, it addresses the high heat flux density at the nucleation boiling initiation point on smooth walls; and fifthly, it addresses the low area utilization and inconvenient pipe layout of traditional spiral flow channels. Finally, it establishes a performance calculation system adapted to this composite structure, solving the problems of large performance prediction errors, reliance on costly simulations and experiments, and long design cycles associated with evaporators combining macroscopic curved flow channels and microscopic serrated structures. This achieves comprehensive optimization of evaporator heat transfer efficiency, flow resistance, temperature uniformity, and anti-drying capability.

[0008] To achieve the above objectives, the first aspect of the present invention provides a reverse Archimedes curve honeycomb sawtooth evaporator, comprising a heat collection plate body (1).

[0009] The heat collection plate body (1) includes an upper plate and a lower plate. The upper plate and the lower plate are interlocked and connected. The area where the two plates are not joined protrudes outward to form a channel protrusion (22), and a hollow working fluid flow channel is formed between the upper plate and the lower plate.

[0010] The flow channel includes at least one spiral loop. The spiral loop is in the shape of a reverse Archimedean curve and is composed of an inlet section, a spiral section, a central reversal section and a return section connected in sequence. The working fluid flows in from the inlet section, spirals from the outside to the inside through the spiral section according to the spiral trajectory to the central area of ​​the plate. After being reversed through the central reversal section, it connects to the return section and flows back from the inside to the outside of the heat collection plate body (1) through the gaps between each loop.

[0011] The inner wall of the flow channel is provided with serrated microstructures (23), which include staggered serrated protrusions (30) and serrated recesses (31).

[0012] The edge of the heat collection plate body (1) is connected to a working fluid interface corresponding to the vortex circuit.

[0013] Furthermore, the flow channel includes an upper vortex circuit (2) located in the upper half of the heat collection plate body (1) and a lower vortex circuit (3) located in the lower half of the heat collection plate body (1).

[0014] The upper spiral circuit (2) is provided with a working fluid inlet (upper circuit) (4) and a working fluid outlet (upper circuit) (5), and the lower spiral circuit (3) is provided with a working fluid inlet (lower circuit) (6) and a working fluid outlet (lower circuit) (7).

[0015] The working fluid inlet (upper loop) (4), the working fluid outlet (upper loop) (5), the working fluid inlet (lower loop) (6), and the working fluid outlet (lower loop) (7) are all centrally located on the same top edge of the heat collector plate body (1).

[0016] Furthermore, the upper spiral loop (2) is composed of an upper circulation inlet section (10), an upper circulation outer ring (11), an upper circulation middle ring (12), an upper circulation inner ring (13), an upper circulation center reversal section (14), and an upper circulation return section (15) connected in sequence; the working fluid flows in from the upper circulation inlet section (10), coils from the outside to the inside to the upper circulation center reversal section (14) and achieves a 180° reverse bend, and then flows back from the inside to the working fluid outlet (upper loop) (5) through the upper circulation return section (15);

[0017] The lower spiral loop (3) consists of a lower circulation inlet section (16), a lower circulation outer ring (17), a lower circulation middle ring (18), a lower circulation inner ring (19), a lower circulation center reversal section (20), and a lower circulation return section (21) connected in sequence. The working fluid flows in from the lower circulation inlet section (16), coils from the outside to the inside to the lower circulation center reversal section (20) and achieves a 180° reverse bend, and then flows back from the inside to the working fluid outlet (lower loop) (7) through the lower circulation return section (21).

[0018] Furthermore, the upper plate and the lower plate are integrally composited through a roll forming process. The unprinted barrier area of ​​the plate bulges outward to form the channel protrusion (22), while the plate in the printed barrier area remains in contact, forming a support and separation structure between the flow channels.

[0019] Furthermore, the cross-section of the channel protrusion (22) is a semi-elliptical arch structure, and the serrated microstructure (23) is distributed on the inner wall and / or base surface of the protrusion of the semi-elliptical arch structure.

[0020] Furthermore, the sawtooth microstructures (23) are arranged in a hexagonal honeycomb pattern on the inner wall surface of the flow channel;

[0021] The serrated protrusions (30) are supported between the inner walls of the channel and also serve as structural support points during blow molding; the serrated recesses (31) are recessed into the wall surface to form an activated vaporization core.

[0022] The parallel mode is that the working fluid inlet (upper loop) (4) and the working fluid inlet (lower loop) (6) are connected to each other, and the working fluid outlet (upper loop) (5) and the working fluid outlet (lower loop) (7) are connected to each other;

[0023] Furthermore, the heat collection plate body (1) retains a sealing edge (9) around its perimeter that continuously seals the upper plate and the lower plate.

[0024] The heat collection plate body (1) has mounting positioning holes / wiring holes (8) that pass through the upper plate and the lower plate for installation, fixing and wiring.

[0025] The working fluid interface is a copper tube interface, which is fixedly connected to the flow channel port of the heat collection plate body (1) by brazing.

[0026] A second aspect of this invention provides a method for calculating the physical properties of a folded Archimedes curve type honeycomb sawtooth evaporator, comprising the following steps:

[0027] Input the structural parameters and operating condition parameters of the evaporator to be calculated;

[0028] Geometric characteristic parameter calculation: Based on the structural parameters, the flow cross-sectional area, hydraulic diameter, and area magnification factor determined by the sawtooth microstructure are calculated through the geometric relationship of the flow channel profile, and then the total heat exchange area and the total flow channel length are solved;

[0029] Flow resistance calculation: Based on the total flow length and hydraulic diameter, the frictional resistance along the flow path is calculated using the sawtooth resistance correction factor determined by the sawtooth parameters. The local resistance of the bend is calculated based on the centrifugal force effect determined by the curvature radius of each bend. The total flow resistance is obtained by superimposing the local resistance at the inlet and outlet.

[0030] Heat transfer performance calculation: Determine the working fluid flow condition; Under single-phase flow condition, calculate the single-phase heat transfer coefficient based on the basic Nusselt number, superimposed with the vortex secondary flow enhancement factor and the sawtooth turbulence enhancement factor; Under two-phase evaporation condition, calculate the two-phase evaporation heat transfer coefficient based on the superimposed model, combined with the nucleus boiling heat transfer coefficient with the introduction of the sawtooth nucleus enhancement factor and the convective evaporation heat transfer coefficient with the introduction of the vortex two-phase enhancement factor.

[0031] Critical drying calculation: The critical drying heat flux density is calculated using a model that combines centrifugal force factor and sawtooth factor.

[0032] Temperature uniformity assessment: The flow channel is divided into multiple micro-segments along the flow channel, and a one-dimensional thermal balance equation is established for the micro-segment. The two-dimensional thermal conductivity of the aluminum plate substrate between the rings is coupled to solve the two-dimensional temperature distribution on the plate surface. The plate surface uniformity coefficient is calculated based on the highest and lowest temperatures.

[0033] The output includes physical performance evaluation results such as heat transfer coefficient, total flow resistance, two-dimensional temperature distribution on the plate surface, temperature uniformity coefficient, and drying margin.

[0034] Furthermore, in the calculation of the geometric characteristic parameters, the reversal curve is approximated as a rectangular spiral. The total length of the spiral on one side is calculated by determining the arithmetic progression of the circumference of each turn, and the total flow path length is derived by superimposing the length of the intermediate reversal section. The specific calculation formula includes: calculating the flow cross-sectional area using a semi-elliptical arched cross-section.

[0035]

[0036] Calculate the sawtooth area magnification factor based on sawtooth height and tooth spacing:

[0037]

[0038] Hydraulic diameter and total heat transfer area calculated based on wetted perimeter:

[0039] ,

[0040] According to the The total length of a single-sided spiral is calculated by summing the circumferences of the spirals, and then the total flow path length is obtained.

[0041]

[0042]

[0043] In the formula, For channel width, For the channel height, For the height of the serrations, The tooth spacing is the serration. For the wet perimeter of the flow channel, For a single flow channel, The hydraulic diameter of the flow channel. This is the magnification factor for the sawtooth area. For the total heat exchange area, This is the total length of a single-sided spiral. This represents the total length of a single loop. This represents the number of spiral turns on one side. The spacing between the circles, , The initial length and width of the innermost spiral are given. This is the length of the intermediate turnaround section.

[0044] Furthermore, in the flow resistance calculation, the friction coefficient along the bend is obtained by multiplying the turbulent basic friction coefficient by a sawtooth resistance correction factor determined by the relative tooth height, tooth pitch-to-height ratio, and Reynolds number; the local resistance of the bend is calculated based on the elbow resistance coefficient corrected by centrifugal force introduced based on the radius of curvature of each bend; the specific calculation formula includes: friction resistance along the bend:

[0045]

[0046] The basic coefficient of friction is:

[0047] Sawtooth resistance correction factor:

[0048]

[0049] The friction coefficient along the friction path satisfies: ;

[0050] Local resistance of the elbow:

[0051] Among them, the elbow resistance coefficient is:

[0052]

[0053] Total flow resistance:

[0054] In the formula, Here is the friction coefficient along the friction path. The friction coefficient of the straight pipe turbulent flow system. The total flow path length is _____. The hydraulic diameter of the flow channel. The density of the working fluid. The average axial velocity of the working fluid. The working fluid Reynolds number, These are the resistance fitting coefficients. For the height of the serrations, The tooth spacing is the serration. This is the sawtooth resistance correction factor. For the first The number of bends corresponding to the spiral spiral. Let be the local resistance coefficient of the bend in the i-th rotation of the spiral. For the first The working fluid velocity within the closed flow channel, The radius of curvature of the spiral is denoted as . For frictional resistance along the path, For the local resistance of the elbow, This creates localized obstacles to import and export. This represents the total flow resistance.

[0055] Furthermore, in the single-phase flow condition for the heat transfer performance calculation, the single-phase Nusselt number is obtained by superimposing a secondary flow enhancement factor determined by the Dean number and a sawtooth turbulence enhancement factor determined by the sawtooth geometry parameters and Reynolds number on the basis of the basic Nusselt number; the specific calculation formula includes:

[0056] Basic Nusselt number:

[0057]

[0058] Calculation of vortex secondary flow enhancement factor based on Dean number:

[0059]

[0060] Serrated turbulence enhancement factor:

[0061]

[0062] Corrected single-phase Nusselt number:

[0063]

[0064] Single-phase heat transfer coefficient:

[0065]

[0066] In the formula, For the straight-pipe basic Nusselt number, The working fluid Reynolds number, Let De be the Prandtl number of the working fluid, and De be the Dean number. The hydraulic diameter of the flow channel. The radius of curvature of the spiral is denoted as . It is a vortex secondary flow enhancement factor. For the height of the serrations, The tooth spacing is the serration. These are the heat transfer fitting coefficients. As a sawtooth turbulence enhancement factor, This is the corrected single-phase Nusselt number. The thermal conductivity of the working fluid is The single-phase heat transfer coefficient is denoted as .

[0067] Furthermore, in the two-phase evaporation condition for heat transfer performance calculation, a boiling superposition model is adopted. A sawtooth nucleus boiling enhancement factor is introduced into the pool boiling heat transfer coefficient to calculate the nucleus boiling heat transfer coefficient, and a vortex two-phase enhancement factor is introduced into the convective evaporation heat transfer coefficient based on Lockhart-Martinelli parameters to calculate the convective evaporation heat transfer coefficient. Specific calculation formulas include:

[0068] Serrated nucleus boiling enhancement factor:

[0069]

[0070] nucleus boiling heat transfer coefficient:

[0071]

[0072] Lockhart-Martinelli parameters:

[0073]

[0074] Convection-evaporation heat transfer coefficient:

[0075]

[0076] Two-phase evaporation heat transfer coefficient:

[0077] In the formula, It is a serrated nucleus boiling enhancement factor. These are the fitting coefficients for nucleate boiling. For the height of the serrations, The hydraulic diameter of the flow channel. The tooth spacing is the serration. The nucleation boiling heat transfer coefficient, For heat flux density, The molar mass of the working fluid. For the working fluid comparison pressure, For the surface roughness of the wall, For Lockhart-Martinelli parameters, For the dryness of the working fluid, , These are the densities of the liquid phase and the vapor phase, respectively. , These are the dynamic viscosity of the liquid phase and the vapor phase, respectively. This is the heat transfer coefficient when the liquid phase flows alone. It is a spiral two-phase enhancement factor. The convective-evaporative heat transfer coefficient is... The overall heat transfer coefficient for the two phases is... , These are the correction factors for nucleate boiling and convective evaporation, respectively.

[0078] Furthermore, in the aforementioned dry-out criticality calculation, based on the critical heat flux density fundamental model, the dry-out heat flux density is corrected by multiplying the fundamental dry-out heat flux density by the vortex centrifugal force factor and the serrated factor, respectively; the specific calculation formula includes: centrifugal Froude number and vortex centrifugal force factor:

[0079]

[0080] Sawtooth factor:

[0081]

[0082] Dry heat flux density:

[0083]

[0084] In the formula, For the eccentric Froude number, The axial velocity of the working fluid. It is the acceleration due to gravity. The radius of curvature of the spiral is denoted as . This is a correction factor for the centrifugal force of the vortex. For the axial Froude number, This is a correction factor for the sawtooth structure. For the height of the serrations, The hydraulic diameter of the flow channel. This refers to the number of saw teeth per unit area. For the dry heat flux density of the straight pipe foundation, This is the corrected dry heat flux density.

[0085] Furthermore, in the temperature uniformity assessment, the flow channel is divided into multiple micro-segments along the flow direction. A one-dimensional thermal balance differential equation for each fluid micro-segment is established and coupled with a two-dimensional thermal conductivity differential equation for the aluminum plate substrate between the rings. The two-dimensional temperature distribution on the plate surface is solved numerically using the finite difference method, and the plate surface temperature uniformity coefficient is calculated based on the temperature extremes. The specific calculation formulas include:

[0086] One-dimensional heat balance equation for a fluid micro-element:

[0087]

[0088] Plate surface temperature coefficient:

[0089]

[0090] In the formula, For the working fluid mass flow rate, The specific heat capacity of the working fluid at constant pressure The heat transfer coefficient, For the wet perimeter of the flow channel, The flow channel wall temperature, The temperature of the working fluid. , These are the highest and lowest temperatures on the board surface, respectively. , These are the inlet temperature and outlet temperature of the working fluid, respectively. The uniform temperature coefficient of the plate surface.

[0091] Compared with the prior art, the present invention has the following significant advantages:

[0092] This invention employs a symmetrically arranged double spiral loop structure. The working fluid spirals inward from the inlet section, reaches the central region, and then turns back, flowing back outward to the outer side of the plate, forming a complete zigzag Archimedes curve flow channel. This design retains the advantages of spiral flow channels in generating secondary flow and enhancing heat transfer through centrifugal force, while solving the problems of uneven utilization of the inner and outer ring areas and large working fluid distribution deviations in traditional single spiral flow channels by evenly distributing the plate surface area through dual loops. All working fluid interfaces are concentrated on the same side edge of the plate, avoiding the defects of complex piping layout and high assembly difficulty caused by the inlet and outlet of traditional spiral flow channels being located in the center. It can be directly adapted to standardized piping connections, making it more feasible for engineering implementation.

[0093] This invention constructs a dual heat transfer enhancement mechanism through the combined effects of macroscopic vortex secondary flow and microscopic sawtooth turbulence: At the macroscopic level, the centrifugal force generated by the vortex channel induces secondary flow, breaking the flow boundary layer and enhancing radial mixing of the fluid; at the microscopic level, the hexagonal honeycomb sawtooth microstructure on the inner wall of the channel effectively increases the heat transfer area, while periodically disturbing the boundary layer, further enhancing the convective heat transfer intensity. Compared to pure microchannel or porous media enhancement structures, this scheme, through a semi-elliptical arched channel cross-section combined with a sawtooth structure design, effectively controls the flow pressure drop while enhancing heat transfer, alleviating the contradiction between enhanced heat transfer and increased pressure drop. This allows for improved heat transfer coefficients at similar pressure drops or reduced system pump power at similar heat transfer performance, ultimately improving the overall system energy efficiency.

[0094] This invention improves the anti-drying capability of two-phase evaporation from two dimensions: nucleation boiling activation and liquid film stabilization. On the one hand, the serrated recesses form a large number of discrete activated vaporization nuclei, effectively reducing the heat flux density at the nucleation boiling initiation point. Nucleation boiling heat exchange can be initiated under low load conditions, improving the heat exchange performance in the low load range. On the other hand, the centrifugal force of the vortex flow channel causes radial separation of the gas and liquid phases. The liquid phase adheres to the outer wall surface to form a stable liquid film. Combined with the disturbance effect of the serrated protrusions on the liquid film, it avoids local thinning and rupture of the liquid film, significantly increasing the critical dry-out heat flux density, reducing the risk of local dry-out, broadening the safe operating heat flux range of the evaporator, and improving the reliability of system operation.

[0095] The calculation method of this invention is custom-designed for composite morphologies combining zigzag flow channels with honeycomb serrated microstructures. It introduces correction factors corresponding to the structural features in six calculation modules: geometric characteristics, flow resistance, single-phase heat transfer, two-phase evaporation, drying criticality, and isothermal performance. This accurately characterizes the synergistic mechanism between the macroscopic curved flow channel and the microscopic wall structure. Compared to empirical formulas for general straight pipes and ordinary serpentine flow channels, this method significantly reduces the performance prediction error of composite structures, and the calculation results are closer to actual operating data, providing a reliable quantitative basis for structural optimization.

[0096] This method employs a hierarchical and progressive analytical calculation logic, from solving geometric parameters to calculating resistance, heat transfer, and temperature uniformity performance. It is based entirely on closed mathematical formulas and modified models, eliminating the need for time-consuming and labor-intensive CFD numerical simulations or prototype experimental tests. It can complete the performance evaluation of a single set of structural parameters in a short time, supports batch iterative optimization of multiple parameters, and can significantly shorten the structural design and scheme selection cycle of evaporators, significantly reduce R&D costs, and improve product iteration efficiency.

[0097] This method covers all operating conditions of single-phase flow and two-phase evaporation, supports performance calculations for different operating modes, and can output comprehensive performance indicators such as heat transfer coefficient, total flow resistance, two-dimensional temperature distribution on the plate surface, temperature uniformity coefficient, drying heat flux density, and drying margin. This method can not only quantitatively evaluate the heat transfer capacity of the evaporator, but also quantitatively assess operational safety and temperature uniformity, providing comprehensive quantitative support for structural parameter optimization and operating condition matching. It can be directly applied to scheme verification and performance prediction in the engineering design phase.

[0098] This invention achieves a comprehensive enhancement effect superior to the sum of the individual effects of the macroscopic vortex flow channel and the turbulent vaporization effect of the microscopic serrated structure in five aspects: flow heat transfer, two-phase boiling, anti-drying, plate surface temperature uniformity, and pressure drop balance. Specifically:

[0099] Firstly, regarding single-phase flow heat transfer, the deflected Archimedes vortex channel generates Dean's secondary flow, enhancing radial mixing of the fluid and disrupting the near-wall boundary layer. The symmetrical flow path with one inlet and one outlet ensures uniform curvature distribution along the flow path, avoiding the uneven velocity distribution between the inner and outer rings of a single helical structure. The hexagonal staggered sawtooth structure periodically disturbs the boundary layer, increasing the heat transfer area; under low Reynolds number conditions, it can trigger turbulence earlier, improving insufficient heat transfer under low load conditions. The macroscopic fluid mixing and microscopic sawtooth turbulence work together to increase the single-phase heat transfer coefficient by 35% to 60% compared to a smooth vortex channel, an improvement exceeding the simple superposition of the effects of the two structures individually.

[0100] Secondly, regarding two-phase evaporation and nucleation boiling: The serrated recesses form numerous stable vaporization nuclei, significantly reducing the heat flux density required for nucleation boiling initiation, allowing for continuous steam generation even under low load conditions. The serrated protrusions divide the wall surface into multiple regions, suppressing the insulation effect caused by gas film adhesion. The centrifugal force of the vortex flow channel causes radial separation of the gas and liquid phases. The liquid phase adheres to the wall surface and continuously replenishes the serrated vaporization nuclei, preventing heat transfer failure due to insufficient liquid in smooth flow channels under low dryness conditions. The continuous liquid replenishment driven by centrifugal force, combined with the stable vaporization nuclei provided by the serrated structure, enables the working fluid to maintain stable nucleation boiling at 20% of the rated heat flux density, doubling the heat transfer capacity under low load conditions.

[0101] Third, regarding anti-drying performance. The root cause of drying is the rupture of the liquid film on the wall, leading to localized liquid shortage. The two structures of this invention work together at different levels to delay the occurrence of drying. The centrifugal force of the vortex channel compresses the liquid phase to adhere to the wall, forming a continuous liquid film, eliminating the gas-liquid stratification phenomenon caused by gravity, and reducing the overall liquid shortage area. The hexagonal serrated protrusions guide and spread the liquid film, while the serrated recesses can store trace amounts of liquid phase, forming localized liquid storage micro-regions, inhibiting the thinning or even breakage of the localized liquid film. The continuous supply of liquid phase at the overall level and the uniform spreading of the liquid film at the localized level provide dual protection, increasing the critical drying heat flux density by 42% to 58% compared to a smooth vortex channel, and significantly widening the safe operating heat flux range of the evaporator.

[0102] Fourth, regarding the plate surface temperature uniformity. The symmetrically arranged double spiral loops evenly distribute the heat load on the plate surface, and the zigzag flow path allows the hot and cold fluids to be distributed alternately on the plate surface, eliminating the defects of the single spiral structure where the outer ring temperature is too high and the central area temperature is too low; the four interfaces arranged on the same side ensure uniform flow distribution in each loop. The serrated structure of the inner wall of the flow channel increases the lateral thermal conductivity cross-sectional area of ​​the aluminum plate substrate, which is conducive to the rapid diffusion of local hot spots; the increased heat exchange area reduces the heat flux density per unit area and reduces the local temperature rise. The maximum temperature difference on the plate surface of the evaporator of this invention is only 3℃ to 6℃, which is far lower than the 15℃ to 25℃ of the traditional serpentine blown plate and the 10℃ to 18℃ of the single spiral plate, significantly reducing the risk of thermal stress and plate deformation caused by uneven temperature.

[0103] Fifth, regarding the balance of flow pressure drop. Single heat exchange enhancement structures generally present a contradiction between improved heat exchange performance and a sharp increase in flow pressure drop. This invention achieves separate optimization of heat exchange performance and flow resistance through the combination of macroscopic flow channel morphology and microscopic sawtooth structure. The semi-elliptical arched large-section flow channel adapts to the volume expansion during the evaporation process of the working fluid, and the gradually changing curvature of the bend reduces local flow resistance; the dual-loop parallel operation mode increases the total flow area and reduces the working fluid velocity. The open hexagonal sawtooth structure lacks closed slits, and its flow blocking effect is far weaker than that of rectangular straight teeth or dense microchannel structures. The increase in resistance can be precisely controlled by adjusting the tooth height and tooth pitch parameters. Under the same heat transfer coefficient conditions, the flow pressure drop of the evaporator of this invention is reduced by more than 60% compared to the microchannel structure; under the same pressure drop conditions, the heat transfer coefficient is increased by 40% compared to the smooth vortex flow channel, which can effectively reduce the system's circulating pump consumption. Attached Figure Description

[0104] Figure 1 This is a top view of the overall structure of the reverse Archimedes curve honeycomb sawtooth evaporator described in this invention;

[0105] Figure 2 This is an enlarged schematic diagram of the upper single vortex channel of the evaporator, showing the foldback flow channel structure of the upper vortex circuit;

[0106] Figure 3 This is an enlarged schematic diagram of the lower single vortex channel of the evaporator, showing the foldback flow channel structure of the lower vortex circuit;

[0107] Figure 4 for Figure 1 The AA cross-sectional view shows the cross-sectional shape of the flow channel and the arrangement of the hexagonal serrated microstructures;

[0108] Figure 5 This is a partially enlarged schematic diagram of the hexagonal sawtooth microstructure. Detailed Implementation

[0109] Example 1

[0110] like Figure 1 As shown, this embodiment provides a folded Archimedes curve honeycomb sawtooth evaporator, which includes a heat collection plate body (1).

[0111] The heat collection plate body (1) includes an upper plate and a lower plate. The upper plate and the lower plate are interlocked and connected. The area where the two plates are not joined protrudes outward to form a channel protrusion (22), and a hollow working fluid flow channel is formed between the upper plate and the lower plate.

[0112] The flow channel includes at least one spiral loop. The spiral loop is in the shape of a reverse Archimedean curve and is composed of an inlet section, a spiral section, a central reversal section and a return section connected in sequence. The working fluid flows in from the inlet section, spirals from the outside to the inside through the spiral section according to the spiral trajectory to the central area of ​​the plate. After being reversed through the central reversal section, it connects to the return section and flows back from the inside to the outside of the heat collection plate body (1) through the gaps between each loop.

[0113] The inner wall of the flow channel is provided with serrated microstructures (23), which include staggered serrated protrusions (30) and serrated recesses (31).

[0114] The edge of the heat collection plate body (1) is connected to a working fluid interface corresponding to the vortex circuit. This basic structure, through the composite design of macroscopic vortex flow channel and microscopic sawtooth structure, can achieve heat exchange enhancement while reasonably controlling pressure drop, and has good temperature uniformity and anti-drying ability. It is suitable for the evaporation of various phase change working fluids such as Freon, carbon dioxide, and propane.

[0115] As one implementation, the flow channel in this embodiment includes an upper vortex circuit (2) located in the upper half of the heat collection plate body (1) and a lower vortex circuit (3) located in the lower half of the heat collection plate body (1).

[0116] The upper spiral circuit (2) is provided with a working fluid inlet (upper circuit) (4) and a working fluid outlet (upper circuit) (5), and the lower spiral circuit (3) is provided with a working fluid inlet (lower circuit) (6) and a working fluid outlet (lower circuit) (7).

[0117] The working fluid inlet (upper loop) (4), the working fluid outlet (upper loop) (5), the working fluid inlet (lower loop) (6), and the working fluid outlet (lower loop) (7) are all centrally located on the same top edge of the heat collector plate body (1). The symmetrical double spiral layout can generate secondary flow by centrifugal force to enhance heat transfer, and can also evenly distribute the plate area, improving the problems of uneven utilization of the inner and outer ring areas and large working fluid distribution deviation in the single spiral flow channel; the arrangement of the working fluid interfaces on the same side facilitates the connection of system pipelines, reduces assembly complexity, and makes the project more feasible.

[0118] As one implementation method, the upper spiral loop (2) in this embodiment consists of an upper circulation inlet section (10), an upper circulation outer ring (11), an upper circulation middle ring (12), an upper circulation inner ring (13), an upper circulation center reversal section (14), and an upper circulation return section (15) connected in sequence. The working fluid flows in from the upper circulation inlet section (10), coils from the outside to the inside to the upper circulation center reversal section (14) and achieves a 180° reverse bend, and then flows back from the inside to the working fluid outlet (upper loop) (5) through the upper circulation return section (15).

[0119] The lower vortex loop (3) consists of a lower circulation inlet section (16), a lower circulation outer ring (17), a lower circulation middle ring (18), a lower circulation inner ring (19), a lower circulation center reversal section (20), and a lower circulation reflux section (21) connected in sequence. The working fluid flows in from the lower circulation inlet section (16), coils from the outside to the inside to the lower circulation center reversal section (20) and achieves a 180° reverse bend, and then flows back from the inside to the working fluid outlet (lower loop) (7) through the lower circulation reflux section (21). The reversible flow path allows the working fluid to complete a complete flow from the outside to the inside and then from the inside to the outside. The flow is continuous and the pressure gradient is uniform, which can effectively extend the heat exchange flow per unit plate area and increase the heat exchange density. At the same time, the gradual change of the vortex curvature can match the volume expansion trend of the working fluid evaporation process, alleviating the problem of the flow velocity soaring and the pressure drop increasing sharply in the second half.

[0120] In one implementation, the upper plate and the lower plate are integrally composited by a roll forming process. The unprinted barrier area of ​​the plate protrudes outward to form the channel protrusion (22), while the plate in the printed barrier area remains in contact, forming a support and separation structure between the flow channels. The roll forming structure eliminates the contact thermal resistance of the tube sheet structure, resulting in higher heat transfer efficiency. The support structure formed by the printed barrier area can improve the overall pressure resistance of the plate and ensure the stability of the flow channel shape. At the same time, the process is mature and has the feasibility of mass production.

[0121] In one embodiment, the cross-section of the channel protrusion (22) in this embodiment is a semi-elliptical arch structure, and the serrated microstructure (23) is distributed on the inner wall and / or base surface of the semi-elliptical arch structure. The semi-elliptical arch cross-sectional shape can adapt to the flow characteristics of the two-phase working fluid, and while ensuring the effective heat exchange area, it can reasonably control the flow resistance, alleviate the contradiction between heat exchange enhancement and pressure drop increase, and help improve the overall energy efficiency of the system.

[0122] In one embodiment, the sawtooth microstructure (23) described in this embodiment is arranged in a hexagonal honeycomb pattern on the inner wall surface of the flow channel;

[0123] The serrated protrusions (30) are supported between the inner walls of the channel and also serve as structural support points during blow molding; the serrated recesses (31) are recessed into the wall surface to form activated vaporization nuclei. The hexagonal honeycomb arrangement of serrations can achieve three functions simultaneously: first, periodically disturbing the flow boundary layer to enhance fluid mixing and improve convective heat transfer intensity; second, providing a large number of discrete activated vaporization nuclei through the recessed structure to reduce the initial heat flux density of nucleation boiling and improve heat transfer performance under low load conditions; and third, acting as internal support points during blow molding to control the consistency of channel height and improve molding yield and structural pressure bearing capacity.

[0124] As one implementation method, in this embodiment, the upper spiral circuit (2) and the lower spiral circuit (3) are configured to operate in parallel.

[0125] In parallel mode, the working fluid inlet (upper loop) (4) and the working fluid inlet (lower loop) (6) are interconnected, and the working fluid outlet (upper loop) (5) and the working fluid outlet (lower loop) (7) are interconnected.

[0126] As one implementation method, in this embodiment, the heat collection plate body (1) has a sealing edge (9) around its perimeter that continuously seals the upper plate and the lower plate.

[0127] The heat collection plate body (1) has mounting positioning holes / wiring holes (8) that pass through the upper plate and the lower plate for installation, fixing and wiring.

[0128] The working fluid interface is a copper pipe interface, which is fixedly connected to the flow channel port of the heat collection plate body (1) by brazing. The continuous sealing edge on all four sides can ensure the sealing reliability and pressure resistance of the plate body and reduce the risk of leakage; the integrated installation positioning hole and wiring hole take into account both assembly and fixing and wiring functions, and are suitable for installation needs of multiple scenarios such as photovoltaic thermal, heat pump, and power battery thermal management; the copper pipe brazed joint has high connection strength and good sealing performance, and can be directly adapted to standardized system pipelines.

[0129] Example 2

[0130] This embodiment provides a method for calculating the physical properties of a folded Archimedes curve type honeycomb sawtooth evaporator, including the following steps:

[0131] Input the structural parameters and operating condition parameters of the evaporator to be calculated;

[0132] Geometric characteristic parameter calculation: Based on the structural parameters, the flow cross-sectional area, hydraulic diameter, and area magnification factor determined by the sawtooth microstructure are calculated through the geometric relationship of the flow channel profile, and then the total heat exchange area and the total flow channel length are solved;

[0133] Flow resistance calculation: Based on the total flow length and hydraulic diameter, the frictional resistance along the flow path is calculated using the sawtooth resistance correction factor determined by the sawtooth parameters. The local resistance of the bend is calculated based on the centrifugal force effect determined by the curvature radius of each bend. The total flow resistance is obtained by superimposing the local resistance at the inlet and outlet.

[0134] Heat transfer performance calculation: Determine the working fluid flow condition; Under single-phase flow condition, calculate the single-phase heat transfer coefficient based on the basic Nusselt number, superimposed with the vortex secondary flow enhancement factor and the sawtooth turbulence enhancement factor; Under two-phase evaporation condition, calculate the two-phase evaporation heat transfer coefficient based on the superimposed model, combined with the nucleus boiling heat transfer coefficient with the introduction of the sawtooth nucleus enhancement factor and the convective evaporation heat transfer coefficient with the introduction of the vortex two-phase enhancement factor.

[0135] Critical drying calculation: The critical drying heat flux density is calculated using a model that combines centrifugal force factor and sawtooth factor.

[0136] Temperature uniformity assessment: The flow channel is divided into multiple micro-segments along the flow channel, and a one-dimensional thermal balance equation is established for the micro-segment. The two-dimensional thermal conductivity of the aluminum plate substrate between the rings is coupled to solve the two-dimensional temperature distribution on the plate surface. The plate surface uniformity coefficient is calculated based on the highest and lowest temperatures.

[0137] The output includes physical performance evaluation results such as heat transfer coefficient, total flow resistance, two-dimensional temperature distribution on the plate surface, temperature uniformity coefficient, and drying margin. The entire method is custom-designed for the composite morphology of the vortex flow channel combined with the honeycomb serrated microstructure, covering the full-dimensional performance calculation of geometry, resistance, heat transfer, drying, and temperature uniformity. It can achieve rapid quantitative evaluation of evaporator performance without relying on costly simulation and experimental testing, and can significantly shorten the product development cycle.

[0138] As one implementation method, in the calculation of the geometric characteristic parameters in this embodiment, the reversal curve is approximated as a rectangular spiral. The total length of the spiral on one side is calculated by determining the arithmetic progression of the circumference of each turn, and the total flow path length is derived by superimposing the length of the intermediate reversal section. The specific calculation formula includes: calculating the flow cross-sectional area using a semi-elliptical arched cross-section.

[0139]

[0140] Calculate the sawtooth area magnification factor based on sawtooth height and tooth spacing:

[0141]

[0142] Hydraulic diameter and total heat transfer area calculated based on wetted perimeter:

[0143] ,

[0144] According to the The total length of a single-sided spiral is calculated by summing the circumferences of the spirals, and then the total flow path length is obtained.

[0145]

[0146]

[0147] In the formula, For channel width, For the channel height, For the height of the serrations, The tooth spacing is the serration. For the wet perimeter of the flow channel, For a single flow channel, The hydraulic diameter of the flow channel. This is the magnification factor for the sawtooth area. For the total heat exchange area, This is the total length of a single-sided spiral. This represents the total length of a single loop. This represents the number of spiral turns on one side. The spacing between the circles, , The initial length and width of the innermost spiral are given. This refers to the length of the intermediate turnaround section. The analytical calculation method using the rectangular spiral approximation can quickly obtain the core geometric parameters of the flow channel without relying on 3D modeling and measurement. The calculation logic is clear and efficient, making it suitable for rapid iteration of solutions and batch calculation of multiple parameters.

[0148] As one implementation method, in the flow resistance calculation of this embodiment, the friction coefficient along the bend is obtained by multiplying the turbulent basic friction coefficient by a sawtooth resistance correction factor determined by the relative tooth height, tooth pitch-to-height ratio, and Reynolds number; the local resistance of the bend is calculated based on the elbow resistance coefficient corrected by centrifugal force introduced based on the radius of curvature of each bend; the specific calculation formula includes: friction resistance along the bend:

[0149]

[0150] The basic coefficient of friction is:

[0151] Sawtooth resistance correction factor:

[0152]

[0153] The friction coefficient along the friction path satisfies: ;

[0154] Local resistance of the elbow:

[0155] Among them, the elbow resistance coefficient is:

[0156]

[0157] Total flow resistance:

[0158] In the formula, Here is the friction coefficient along the friction path. The friction coefficient of the straight pipe turbulent flow system. The total flow path length is _____. The hydraulic diameter of the flow channel. The density of the working fluid. The average axial velocity of the working fluid. The working fluid Reynolds number, These are the resistance fitting coefficients. For the height of the serrations, The tooth spacing is the serration. This is the sawtooth resistance correction factor. For the first The number of bends corresponding to the spiral spiral. Let be the local resistance coefficient of the bend in the i-th rotation of the spiral. For the first The working fluid velocity within the closed flow channel, The radius of curvature of the spiral is denoted as . For frictional resistance along the path, For the local resistance of the elbow, This creates localized obstacles to import and export. The total flow resistance is given. Customized correction factors are introduced for sawtooth walls and spiral bends, which can more accurately predict the flow pressure drop of composite structures compared to the general empirical formula for straight pipe resistance. The calculation results are closer to actual operating data, providing a reliable basis for system resistance matching and pump selection.

[0159] As one implementation method, in this embodiment, in the single-phase flow condition for calculating the heat transfer performance, the single-phase Nusselt number is obtained by superimposing a secondary flow enhancement factor determined by the Dean number and a sawtooth turbulence enhancement factor determined by the sawtooth geometry parameters and Reynolds number on the basis of the basic Nusselt number; the specific calculation formula includes:

[0160] Basic Nusselt number:

[0161]

[0162] Calculation of vortex secondary flow enhancement factor based on Dean number:

[0163]

[0164] Serrated turbulence enhancement factor:

[0165]

[0166] Corrected single-phase Nusselt number:

[0167]

[0168] Single-phase heat transfer coefficient:

[0169]

[0170] In the formula, For the straight-pipe basic Nusselt number, The working fluid Reynolds number, Let De be the Prandtl number of the working fluid, and De be the Dean number. The hydraulic diameter of the flow channel. The radius of curvature of the spiral is denoted as . It is a vortex secondary flow enhancement factor. For the height of the serrations, The tooth spacing is the serration. These are the heat transfer fitting coefficients. As a sawtooth turbulence enhancement factor, This is the corrected single-phase Nusselt number. The thermal conductivity of the working fluid is The coefficient is the single-phase heat transfer coefficient. Considering both the enhanced effects of the vortex secondary flow and the sawtooth turbulence, the fundamental Nusselt number is corrected based on the field synergy principle. The calculation accuracy is significantly better than the commonly used empirical formula for straight-tube heat transfer, and it can accurately characterize the synergistic heat transfer effect of the macroscopic flow channel and the microscopic structure.

[0171] As one implementation method, in this embodiment, in the two-phase evaporation condition for calculating the heat transfer performance, a boiling superposition model is adopted. A sawtooth nucleus boiling enhancement factor is introduced into the pool boiling heat transfer coefficient to calculate the nucleus boiling heat transfer coefficient, and a vortex two-phase enhancement factor is introduced into the convective evaporation heat transfer coefficient based on the Lockhart-Martinelli parameters to calculate the convective evaporation heat transfer coefficient. The specific calculation formulas include:

[0172] Serrated nucleus boiling enhancement factor:

[0173]

[0174] nucleus boiling heat transfer coefficient:

[0175]

[0176] Lockhart-Martinelli parameters:

[0177]

[0178] Convection-evaporation heat transfer coefficient:

[0179]

[0180] Two-phase evaporation heat transfer coefficient:

[0181] In the formula, It is a serrated nucleus boiling enhancement factor. These are the fitting coefficients for nucleate boiling. For the height of the serrations, The hydraulic diameter of the flow channel. The tooth spacing is the serration. The nucleation boiling heat transfer coefficient, For heat flux density, The molar mass of the working fluid. For the working fluid comparison pressure, For the surface roughness of the wall, For Lockhart-Martinelli parameters, For the dryness of the working fluid, , These are the densities of the liquid phase and the vapor phase, respectively. , These are the dynamic viscosity of the liquid phase and the vapor phase, respectively. This is the heat transfer coefficient when the liquid phase flows alone. It is a spiral two-phase enhancement factor. The convective-evaporative heat transfer coefficient is... The overall heat transfer coefficient for the two phases is... , These are the nucleation boiling correction factor and the convection evaporation correction factor, respectively. Based on the classical boiling superposition model, a structure-customized correction factor is introduced, solving the problem of rapid prediction of two-phase heat transfer under the composite morphology of curved flow channels combined with serrated microstructures. This allows for rapid evaluation of evaporation heat transfer performance under different operating conditions, providing quantitative support for structural optimization.

[0182] Furthermore, in the aforementioned dry-out criticality calculation, based on the critical heat flux density fundamental model, the dry-out heat flux density is corrected by multiplying the fundamental dry-out heat flux density by the vortex centrifugal force factor and the serrated factor, respectively; the specific calculation formula includes: centrifugal Froude number and vortex centrifugal force factor:

[0183]

[0184] Sawtooth factor:

[0185]

[0186] Dry heat flux density:

[0187]

[0188] In the formula, For the eccentric Froude number, The axial velocity of the working fluid. It is the acceleration due to gravity. The radius of curvature of the spiral is denoted as . This is a correction factor for the centrifugal force of the vortex. For the axial Froude number, This is a correction factor for the sawtooth structure. For the height of the serrations, The hydraulic diameter of the flow channel. This refers to the number of saw teeth per unit area. For the dry heat flux density of the straight pipe foundation, This is to correct the dried heat flux density. By considering both the liquid film adhesion effect of the vortex centrifugal force and the liquid film spreading effect of the serrated structure, the critical dried heat flux density can be predicted more accurately. This provides a quantitative basis for assessing the safe operating range of the evaporator and verifying the dried margin, thereby improving the reliability of system operation.

[0189] As one implementation method, in the temperature uniformity assessment in this embodiment, the flow channel is divided into multiple micro-segments along the flow direction. A one-dimensional thermal balance differential equation for the fluid micro-segment is established and coupled with a two-dimensional thermal conductivity differential equation for the aluminum plate substrate between the rings. The two-dimensional temperature distribution on the plate surface is solved numerically using the finite difference method, and the plate surface temperature uniformity coefficient is calculated based on the temperature extreme values. The specific calculation formula includes:

[0190] One-dimensional heat balance equation for a fluid micro-element:

[0191]

[0192] Plate surface temperature coefficient:

[0193]

[0194] In the formula, For the working fluid mass flow rate, The specific heat capacity of the working fluid at constant pressure The heat transfer coefficient, For the wet perimeter of the flow channel, The flow channel wall temperature, The temperature of the working fluid. , These are the highest and lowest temperatures on the board surface, respectively. , These are the inlet temperature and outlet temperature of the working fluid, respectively. This represents the temperature uniformity coefficient of the plate surface. The calculation method, which combines one-dimensional thermal balance along the path with two-dimensional thermal conduction coupling between loops, can quickly solve for the temperature field and temperature uniformity coefficient of the plate surface. Its computational efficiency is far higher than that of full-scale CFD simulation. It can output temperature uniformity indicators to assess the level of thermal stress, providing comprehensive quantitative support for structural safety assessment and scheme optimization.

[0195] Example 3

[0196] This embodiment is a supplementary description of Embodiments 1 and 2 above. This embodiment is a high-temperature evaporator adapted for industrial flue gas waste heat recovery scenarios. Depending on the waste heat grade, heat transfer oil or R245fa type medium-high temperature organic working fluid can be selected as the heat transfer medium, balancing heat exchange efficiency, pressure resistance reliability, and long-term operational stability under medium-high temperature operating conditions. The following, in conjunction with the accompanying drawings, provides a detailed description of the structural composition, processing technology, and performance calculation methods of this embodiment.

[0197] like Figure 1As shown, the main body of the evaporator is the heat collection plate body 1. The upper spiral loop 2 and lower spiral loop 3 are symmetrically arranged along the height of the plate surface. The two loops can be flexibly configured to operate according to the operating conditions. Four interfaces are located on the same side of the top of the plate: working fluid inlet (upper loop) 4, working fluid outlet (upper loop) 5, working fluid inlet (lower loop) 6, and working fluid outlet (lower loop) 7. In parallel mode, the inlets of the two loops are interconnected, and the two outlets are interconnected, allowing the working fluid to enter both loops simultaneously. This results in a larger total flow area and a lower overall pressure drop, making it suitable for high-flow-rate heat exchange applications using thermal oil.

[0198] Three mounting / wiring holes 8 are provided in the lower middle part of the plate, with a diameter ranging from 10mm to 20mm, which can simultaneously meet the needs of plate installation and fixing as well as the installation of temperature measurement and control lines. A continuous sealing edge 9 is maintained around the plate, with a width of 15mm to 30mm, which facilitates the assembly and fixing of the matching frame and ensures the overall pressure-bearing and sealing performance of the flow channel. In this embodiment, the size of a single plate is 1500mm × 3000mm, the base material is 3003 aluminum alloy, and the finished plate thickness is 1.2mm; a single vortex circuit has 5 flow channels, falling within the applicable range of 3 to 8 channels, with uniform spacing between the channels. The vortex curvature gradually decreases from the outer circle to the inner circle, forming an inverse match with the volume expansion trend during the working fluid evaporation process: the outer circle has a large flow cross-section and moderate flow velocity, while the inner circle has a short flow path and controllable pressure drop, which can alleviate the problem of soaring flow velocity and drastic pressure drop in the latter half of the flow channel under phase change conditions.

[0199] like Figure 2 , Figure 3 As shown, a single spiral loop adopts a reverse Archimedean curve shape, approximately a rectangular spiral structure, with an outer ring, middle ring, and inner ring arranged sequentially from the outside to the inside. After reaching the central region, it flows back to the outside of the plate via a reversal section. Taking the spiral loop as an example, the working fluid enters the flow channel from the upper circulation inlet section 10, flows sequentially through the upper circulation outer ring 11, upper circulation middle ring 12, and upper circulation inner ring 13, and after reaching the upper circulation center reversal section 14, it completes a 180-degree reverse bend, and then flows back outward along the gaps between the flow channels, finally flowing out of the plate via the upper circulation return section 15. The lower spiral loop is correspondingly equipped with a lower circulation inlet section 16, a lower circulation outer ring 17, a lower circulation middle ring 18, a lower circulation inner ring 19, a lower circulation center reversal section 20, and a lower circulation return section 21, with the flow channel arrangement maintaining vertical symmetry with the upper spiral loop.

[0200] like Figure 4As shown, the entire flow channel is integrally formed by the upper plate 20 and the lower plate 21 through a roll forming process. The area where the two plates are not metallurgically bonded bulges outward to form the channel protrusion 22. The channel cross-section has a semi-elliptical arch structure. In the figure, w is the channel width, h is the channel protrusion height, δ is the total thickness of the finished plate, and e is the tooth height of the serrated structure. In this embodiment, the channel width is 15mm, which is within the applicable range of 6mm to 20mm; the protrusion height is 3.0mm, which is within the applicable range of 1.5mm to 4.0mm; the cross-sectional width-to-height ratio is controlled between 3:1 and 6:1, and the corresponding channel hydraulic diameter is approximately 3mm to 8mm. This cross-sectional shape can ensure an effective heat exchange area while reasonably controlling the two-phase flow resistance, balancing the contradiction between heat exchange enhancement and pressure drop increase. The serrated microstructure 23 can be arranged alone on the inner wall of the semi-elliptical arch protrusion, or alone on the flow channel base surface, or simultaneously in two locations, and the arrangement position can be flexibly selected according to the actual heat exchange requirements.

[0201] like Figure 5 As shown, the serrated microstructure on the inner wall of the flow channel is arranged in a hexagonal honeycomb pattern, consisting of two parts: serrated protrusions 30 and serrated recesses 31. The distance between the centers of adjacent serrations is the tooth spacing, and the protrusion height of a single serration is the tooth height e. In this embodiment, the tooth height is 1.0 mm, which is within the applicable range of 0.3 mm to 1.5 mm; the tooth spacing is 5.0 mm, which is within the applicable range of 1.5 mm to 6 mm. This serrated microstructure has three functions: first, it enhances heat transfer by turbulence, as the periodically protruding structure can continuously disrupt the flow boundary layer, increase the degree of local turbulence, and enhance the radial mixing of the fluid; second, it increases the nucleation boiling sites, as a large number of activated vaporization nuclei can be formed at the serrated recesses 31, reducing the heat flux density at the nucleation boiling initiation point, enabling nucleation boiling heat transfer to start even under low load conditions; and third, it provides structural support, as the serrated protrusions 30 can act as internal support points during the blow molding process, limiting excessive deformation of the channel and improving the overall pressure bearing capacity and the consistency of the molding dimensions of the plate.

[0202] In this embodiment, the working fluid interface uses a copper tube, which is fixed to the flow channel port of the plate by brazing. The interface specification can be φ6.35mm or φ9.52mm; this embodiment uses a φ12.7mm copper tube to adapt to high flow conditions. The length of the welded section is controlled between 40mm and 80mm, and the brazing area is pre-treated to ensure weld strength and sealing performance. The finished plate has a lighter weight per unit area, resulting in significant weight reduction compared to traditional tube-plate structures.

[0203] The evaporator in this embodiment is manufactured using a roller-blowing process, and the complete manufacturing process is as follows:

[0204] The first step is the pre-treatment and cleaning / drying of aluminum plates. Two aluminum alloy plates with an initial thickness of 2.5mm to 3.5mm are selected, distinguishing between rough and smooth surfaces. The plates are first fed into a multi-roll leveling machine for rolling to eliminate warping, wavy edges, and internal metal stress generated during the rolling process, ensuring that the flatness of the plate surface meets the process requirements. After leveling, the plates are conveyed to a cleaning and drying integrated machine, where residual rolling oil, aluminum shavings, dust, and other impurities are removed from the plate surface through spray degreasing, preventing oil stains and water stains from damaging the adhesion of the subsequent graphite printing layer. After cleaning, the plates are dried for later use.

[0205] The second step is the screen printing of the flow channel pattern and its lamination with the slab. A cleaned and dried single aluminum plate is used, with its rough side as the printing surface. A fully automatic flatbed screen printing machine is used to print graphite material onto this rough surface according to the preset flow channel pattern, forming a barrier printing layer. After printing, a low-temperature drying process is performed to cure the graphite coating. Then, the patterned aluminum plate and another blank aluminum plate are fed into a laminating device, with the smooth sides facing outwards and the rough sides facing each other, precisely interlocking and overlapping to ensure the barrier printing layer is completely sandwiched between the two aluminum plates, completing the preparation of the composite slab.

[0206] The third step is hot rolling and quenching / leveling. The prepared composite slab is placed in a high-temperature environment of 500℃ to 550℃ and preheated to the rolling temperature. It is then rolled to the designed thickness using a roll forming machine, allowing the two aluminum plates to achieve metallurgical bonding in the area without graphite printing. After rolling, it undergoes quenching treatment at a temperature range of 450℃ to 550℃ to eliminate the internal stress generated during rolling and obtain a uniform and stable metallographic structure. Finally, it is sent again to a leveling process to correct the plate deformation caused by rolling, resulting in a rolled composite plate with a surface flatness that meets the process requirements.

[0207] The fourth step is high-pressure nitrogen blowing and forming. Based on the design pressure resistance of the finished evaporator, high-pressure nitrogen is introduced into the pre-reserved flow channel interface of the rolled composite plate. The blowing pressure is set to be 3 to 5 MPa higher than the design pressure resistance. Under the action of the high-pressure fluid, the area with the printed graphite barrier layer, lacking metallurgical bonding, undergoes plastic deformation and bulges outward, forming a hollow refrigerant flow channel that perfectly matches the printed pattern, resulting in the blown-formed evaporator blank.

[0208] The fifth step is forced cooling and shaping, followed by surface anti-corrosion protection treatment. The blown-formed evaporator blank is naturally cooled to room temperature. After cooling and cleaning, an epoxy resin coating is evenly sprayed onto the inner and outer surfaces of the blank to form a basic anti-corrosion and insulating layer. After the coating pre-dries, it is sent to a laminating equipment where a protective film is applied using hot rollers. This double protection isolates the evaporator from water vapor and acid / alkali corrosive media, while also preventing scratches on the anti-corrosion layer during transport and assembly, thus extending the evaporator's service life under humid and corrosive conditions.

[0209] Step 6: Channel airtightness test. Introduce 2MPa to 3.5MPa of nitrogen gas into the hollow channel of the evaporator after corrosion protection. After sealing the interface, maintain the pressure for the specified time according to the process specifications, monitoring the gas pressure drop throughout. After the pressure holding period, compare the pressure drop with the leakage rate standard. Products with leakage below the set threshold are considered airtight and qualified. Plates with leakage defects are sorted separately for rework and repair, while qualified products are transferred to the warehouse.

[0210] For the evaporator structure in this embodiment, the physical performance calculation method of this invention can be used for a comprehensive performance evaluation. The complete calculation process is as follows:

[0211] First, input the structural parameters and operating parameters of the evaporator to be calculated. The structural parameters include plate size, number of spiral coils, coil spacing, channel width, channel height, serration height and tooth spacing, etc.; the operating parameters include working fluid type, mass flow rate, inlet temperature or inlet dryness, working heat flux density, etc.

[0212] Geometric characteristic parameter calculation: The reversing Archimedes curve is approximated as a rectangular spiral model, with each flow channel having a rectangular profile that expands at equal intervals from the inside out. The profile dimensions of the i-th flow from the inside out satisfy the following recursive relationship:

[0213]

[0214]

[0215] In the formula, , These represent the initial length and initial width of the innermost spiral, respectively. The distance between the circles is denoted as .

[0216] The circumference of the i-th lap is calculated using the following formula:

[0217]

[0218] The total length of a single-sided spiral is obtained by accumulating the lengths of each spiral. The formula for calculating the closed spiral can be derived using the formula for the summation of an arithmetic series.

[0219]

[0220] In the formula, This represents the number of spiral rotations on one side.

[0221] The total flow length is obtained by superimposing the lengths of the intermediate turnaround connecting sections on both sides of the spiral:

[0222]

[0223] In the formula, This is the length of the intermediate turnaround section.

[0224] The cross-sectional area of ​​the flow path is calculated using a semi-ellipse approximation:

[0225]

[0226] Combined with the wet perimeter of the flow channel Calculate the hydraulic diameter:

[0227]

[0228] The magnification factor for the sawtooth area is calculated using the following formula:

[0229]

[0230] In the formula, For the height of the serrations, The pitch of the sawtooth teeth.

[0231] After taking into account the area increase of the sawtooth structure, the total heat exchange area is:

[0232]

[0233] This step also allows us to simultaneously obtain the number of serrations per unit length and the increase in total surface area.

[0234] Flow resistance calculation: The total flow resistance is divided into three parts: frictional resistance along the flow path, local resistance at the bend, and local resistance at the inlet and outlet. The total pressure drop satisfies:

[0235]

[0236] Frictional resistance along the friction path is calculated using the following formula:

[0237]

[0238] In the formula, The friction coefficient along the friction path is calculated using the Blasius formula modified by the sawtooth channel:

[0239]

[0240] The basic friction coefficient for straight pipe turbulence is applicable to the Reynolds number. The turbulent region is given by the formula:

[0241] The sawtooth resistance correction factor is determined by the relative tooth height, tooth pitch-to-tooth height ratio, and Reynolds number.

[0242]

[0243] In the formula, The recommended value for the hexagonal sawtooth pattern is 2.8, which is the experimental fitting coefficient.

[0244] For continuous bends with each spiral rotation, the local resistance of the bend is calculated using a centrifugal force-corrected bend resistance model, and the total local resistance of the bend is obtained by accumulating it round by round:

[0245]

[0246] In the formula, Let be the number of bends corresponding to the i-th spiral. Let be the working fluid velocity in the i-th flow channel. The bend resistance coefficient for the i-th lap is calculated using the modified Itō formula:

[0247]

[0248] In the formula, This represents the radius of curvature of the spiral corresponding to the number of revolutions.

[0249] The inlet and outlet local resistances are calculated based on conventional local resistance coefficients, with the inlet resistance coefficient being approximately 0.5 and the outlet resistance coefficient being approximately 1.0. The total flow resistance is obtained by adding the three resistances together.

[0250] Heat transfer performance calculation: Before the calculation, the flow conditions of the working fluid are determined, and the calculation is performed for both single-phase and two-phase conditions.

[0251] For single-phase flow conditions, the Diethus-Belt formula modified by the field cooperation principle is adopted, considering both the strengthening effects of vortex secondary flow and sawtooth turbulence. The total Nusselt number satisfies:

[0252]

[0253] The Nusselt number for straight pipe foundations is applicable to heating conditions, and the formula is:

[0254]

[0255] The vortex secondary flow enhancement factor is calculated based on the Dean vortex generated by centrifugal force. The Dean number $De$ and the enhancement factor are respectively:

[0256]

[0257]

[0258] The sawtooth turbulence enhancement factor is determined by both the sawtooth geometry and the Reynolds number:

[0259]

[0260] In the formula, The recommended value for the hexagonal serrations is 1.6, representing the experimental fit coefficient. This calculation method is applicable within the Reynolds number range. Prandtl numbers .

[0261] After obtaining the corrected Nusselt number, the single-phase heat transfer coefficient is further calculated:

[0262]

[0263] In the formula, denoted as , where is the thermal conductivity of the working fluid.

[0264] For two-phase evaporation, the modified Steiner-Taberek superposition model is used, and the overall two-phase heat transfer coefficient is the superposition of the nucleus boiling term and the convective evaporation term:

[0265]

[0266] In the formula, , These are the correction factors for nucleate boiling and convection evaporation, respectively.

[0267] The nucleus boiling heat transfer coefficient is based on the Cooper pool boiling formula, and is modified by introducing a serrated nucleus boiling enhancement factor.

[0268]

[0269] In the formula, For heat flux density, The molar mass of the working fluid. For the working fluid comparison pressure, This refers to the surface roughness of the wall.

[0270] The serrated nucleus boiling enhancement factor is calculated using the following formula:

[0271]

[0272] In the formula, The fitting coefficient for nucleate boiling is 2.5, which is recommended to reflect the strengthening effect of the serrated depression providing additional vaporization nuclei.

[0273] The convective-evaporative heat transfer coefficient is calculated based on the Lockhart-Martinelli parameters, while also incorporating a vortex two-phase enhancement factor:

[0274]

[0275] In the formula, This is the heat transfer coefficient when the liquid phase flows alone. The vortex two-phase enhancement factor is typically set to 1.1 to 1.3, and is used to reflect the enhancing effect of centrifugal force on liquid film agitation.

[0276] Lockhart-Martinelli parameters are calculated using the following formula:

[0277]

[0278] In the formula, x is the dryness fraction of the working fluid. , These are the densities of the liquid phase and the vapor phase, respectively. , These are the dynamic viscosity of the liquid phase and the vapor phase, respectively.

[0279] Calculation of the critical dry heat flux density: Using the modified Kato formula, the dry heat flux density of the straight pipe foundation is corrected by multiplying it by the vortex centrifugal force factor and the sawtooth factor, respectively. The final dry heat flux density is:

[0280]

[0281] In the formula, The heat flux density is the dry heat flux density of the straight pipe foundation.

[0282] The centrifugal force factor is determined by the ratio of the centrifugal Froude number to the axial Froude number:

[0283]

[0284]

[0285] In the formula, Let be the centrifugal Froude number, u be the axial velocity of the working fluid, and g be the acceleration due to gravity. This is the axial Froude number. Centrifugal force can throw the liquid phase towards the wall, delaying liquid film rupture and increasing the critical drying heat flux density.

[0286] The sawtooth factor is determined by the relative tooth height and the number of sawtooths per unit area:

[0287]

[0288] In the formula, This represents the number of serrations per unit area. The serrated structure improves liquid film spreadability and delays drying. After calculating the drying heat flux density, the drying margin of the evaporator can be further determined, and operational safety can be assessed.

[0289] Temperature uniformity assessment: Defining the temperature uniformity coefficient of the panel surface As a quantitative evaluation index of temperature uniformity:

[0290]

[0291] In the formula, , These are the highest and lowest temperatures on the board surface, respectively. , These are the inlet and outlet temperatures of the working fluid, respectively.

[0292] In the specific calculation, the vortex flow channel is first divided into multiple infinitesimal elements along the flow direction, and a one-dimensional heat balance equation is established for each infinitesimal element:

[0293]

[0294] In the formula, For the working fluid mass flow rate, The specific heat capacity of the working fluid at constant pressure The flow channel wall temperature, The temperature of the working fluid.

[0295] Simultaneously considering the lateral heat conduction between adjacent flow channels through the aluminum plate substrate, the one-dimensional thermal balance along the flow path is coupled with the two-dimensional heat conduction between the flow channels. A numerical iterative solution using the finite difference method is employed to obtain the two-dimensional temperature distribution across the entire plate surface. By statistically analyzing the temperature extremes and substituting them into the uniform temperature coefficient formula, a quantitative uniform temperature evaluation result can be obtained. Furthermore, the thermal stress distribution can be derived to guide structural safety assessments.

[0296] After completing all the above calculations, the output includes a complete set of physical performance evaluation results, including heat transfer coefficient, total flow resistance, two-dimensional temperature distribution on the plate surface, uniform temperature coefficient, drying heat flux density, and drying margin. The entire algorithm adopts analytical calculation logic, and its computational efficiency is far higher than that of full-scale CFD simulation. It can realize rapid evaluation of schemes and parameter iteration during the design stage, significantly shortening the product development cycle.

[0297] The performance parameters of this embodiment under typical operating conditions are as follows: When using heat transfer oil as the working fluid, the operating temperature range is 120℃ to 180℃, the mass flow rate is 0.25 kg / s, the heat transfer coefficient is approximately 1400 W / (m²·K), the waste heat recovery capacity of a single plate is approximately 50 kW, and the structural pressure resistance can reach 2.5 MPa. If R245fa organic working fluid is used for phase change recovery, the waste heat recovery efficiency can be further improved. Compared with traditional serpentine blown plates, the evaporator of this embodiment, through the composite enhancement of vortex flow channels and serrated microstructures, significantly improves both heat transfer efficiency and anti-drying ability. The flow resistance is much lower than that of microchannel or porous media structures with the same heat transfer level, which can effectively improve the overall operating efficiency of the system.

[0298] The evaporator in this embodiment is formed in one step using a mature blow-in process, which has a processing cost far lower than that of microchannel products. It also has the advantages of being lightweight, having high pressure resistance, and high heat exchange efficiency, making it well-suited for the medium- and long-term stable operation requirements of industrial waste heat recovery scenarios.

[0299] Example 4: Heat pump system equipped with a zigzag Archimedean curve honeycomb serrated evaporator

[0300] This embodiment provides a zigzag Archimedes curve honeycomb evaporator for use with heat pump systems, suitable for heating equipment such as residential air-source heat pumps, commercial air-cooled heat pumps, and low-temperature air-source heating units. The working fluid can be R32, R410A, or other refrigerants, and it is also suitable for CO2 transcritical cycle conditions. This evaporator can improve problems such as uneven frost formation, localized dryness, reduced heat exchange efficiency, and incomplete defrosting caused by excessive temperature differences on the plate surfaces under low-temperature conditions.

[0301] The evaporator structural parameters of this embodiment are as follows: the plate size is 1200mm×800mm, the number of spiral coils on one side is 4, the channel width is 12mm, and the channel height is 2.5mm; the serration height is 0.8mm, and the tooth spacing is 4.0mm; the upper and lower dual circuits adopt a parallel operation mode, and the four working fluid interfaces are centrally arranged at the top of the short side of the plate; the plate has reserved installation positioning holes to facilitate assembly and fixation with the heat pump outdoor unit frame.

[0302] Taking a low-temperature heating condition as an example, the performance of the evaporator in this embodiment is calculated. The operating conditions are: ambient temperature -10℃, evaporation temperature -15℃, working fluid R410A, and mass flow rate 0.12kg / s. The core performance indicators obtained using the physical performance calculation method of this invention are as follows:

[0303] (1) The two-phase heat transfer coefficient is 1720 W / (m²·K), while the two-phase heat transfer coefficient of a conventional serpentine heat pump evaporator of the same specification is only 890 W / (m²·K). The heat transfer coefficient of this embodiment is increased by about 93%.

[0304] (2) The total flow pressure drop is 22.1 kPa, which meets the operating requirements of the heat pump compressor for low pressure loss of the evaporator;

[0305] (3) The critical drying heat flux density is 36.5 kW / m², and there is no risk of drying under low temperature frost and high load conditions;

[0306] (4) The maximum temperature difference on the board surface is 3.8℃. During defrosting, all areas of the board surface melt simultaneously, with no local residual ice layer, and defrosting energy consumption is reduced by about 18%.

[0307] The evaporator in this embodiment has the following advantages in heat pump system applications:

[0308] First, the parallel dual-loop circuit has a larger flow area, and the flow pressure drop is controllable under low temperature and high flow rate conditions, increasing the low temperature heating capacity of the heat pump by 12% to 16%.

[0309] Second, the honeycomb serrated structure provides a large number of vaporization nuclei, which can maintain stable nucleation boiling at low evaporation temperatures, avoiding the reduction in heating capacity caused by relying solely on convection heat transfer under low-temperature conditions.

[0310] Third, the plate surface exhibits excellent temperature uniformity and even frost formation, reducing the frequency of defrosting and improving the unit's annual coefficient of performance (COP) by 8% to 11%.

[0311] Fourth, it adopts a roll forming process to form an integral unit, eliminating the contact thermal resistance of the tube sheet structure, and the lightweight structure is adapted to the compact design requirements of the outdoor unit.

[0312] The above are merely preferred embodiments of the present invention and are not intended to limit the present invention. It should be noted that any modifications, equivalent substitutions, and improvements made within the spirit and principles of the present invention should be included within the protection scope of the present invention.

Claims

1. A zigzag Archimedes curve-shaped honeycomb evaporator, comprising a heat collection plate body (1), characterized in that: The heat collection plate body (1) includes an upper plate and a lower plate. The upper plate and the lower plate are interlocked and connected. The area where the two plates are not joined protrudes outward to form a channel protrusion (22), and a hollow working fluid flow channel is formed between the upper plate and the lower plate. The flow channel includes at least one spiral loop. The spiral loop is in the shape of a reverse Archimedean curve and is composed of an inlet section, a spiral section, a central reversal section and a return section connected in sequence. The working fluid flows in from the inlet section, spirals from the outside to the inside through the spiral section according to the spiral trajectory to the central area of ​​the plate. After being reversed through the central reversal section, it connects to the return section and flows back from the inside to the outside of the heat collection plate body (1) through the gaps between each loop. The inner wall of the flow channel is provided with serrated microstructures (23), which include staggered serrated protrusions (30) and serrated recesses (31). The edge of the heat collection plate body (1) is connected to a working fluid interface corresponding to the vortex circuit.

2. The reverse Archimedes curve honeycomb sawtooth evaporator according to claim 1, characterized in that: The flow channel includes an upper vortex circuit (2) located in the upper half of the heat collector plate body (1) and a lower vortex circuit (3) located in the lower half of the heat collector plate body (1). The upper spiral circuit (2) is provided with a working fluid inlet (upper circuit) (4) and a working fluid outlet (upper circuit) (5), and the lower spiral circuit (3) is provided with a working fluid inlet (lower circuit) (6) and a working fluid outlet (lower circuit) (7). The working fluid inlet (upper loop) (4), the working fluid outlet (upper loop) (5), the working fluid inlet (lower loop) (6), and the working fluid outlet (lower loop) (7) are all centrally located on the same top edge of the heat collector plate body (1).

3. The folded Archimedes curve honeycomb sawtooth evaporator according to claim 2, characterized in that: The upper spiral loop (2) consists of an upper circulation inlet section (10), an upper circulation outer ring (11), an upper circulation middle ring (12), an upper circulation inner ring (13), an upper circulation center reversal section (14), and an upper circulation return section (15) connected in sequence. The working fluid flows in from the upper circulation inlet section (10), coils from the outside to the inside to the upper circulation center reversal section (14) and achieves a 180° reverse bend, and then flows back from the inside to the working fluid outlet (upper loop) (5) through the upper circulation return section (15). The lower spiral loop (3) consists of a lower circulation inlet section (16), a lower circulation outer ring (17), a lower circulation middle ring (18), a lower circulation inner ring (19), a lower circulation center reversal section (20), and a lower circulation return section (21) connected in sequence. The working fluid flows in from the lower circulation inlet section (16), coils from the outside to the inside to the lower circulation center reversal section (20) and achieves a 180° reverse bend, and then flows back from the inside to the working fluid outlet (lower loop) (7) through the lower circulation return section (21).

4. A method for calculating the physical properties of a zigzag Archimedes curve-type honeycomb evaporator as described in any one of claims 1 to 3, characterized in that, Includes the following steps: Input the structural parameters and operating condition parameters of the evaporator to be calculated; Geometric characteristic parameter calculation: Based on the structural parameters, the flow cross-sectional area, hydraulic diameter, and area magnification factor determined by the sawtooth microstructure are calculated through the geometric relationship of the flow channel profile, and then the total heat exchange area and the total flow channel length are solved; Flow resistance calculation: Based on the total flow length and hydraulic diameter, the frictional resistance along the flow path is calculated using the sawtooth resistance correction factor determined by the sawtooth parameters. The local resistance of the bend is calculated based on the centrifugal force effect determined by the curvature radius of each bend. The total flow resistance is obtained by superimposing the local resistance at the inlet and outlet. Heat transfer performance calculation: Determine the working fluid flow condition; Under single-phase flow condition, calculate the single-phase heat transfer coefficient based on the basic Nusselt number, superimposed with the vortex secondary flow enhancement factor and the sawtooth turbulence enhancement factor; Under two-phase evaporation condition, calculate the two-phase evaporation heat transfer coefficient based on the superimposed model, combined with the nucleus boiling heat transfer coefficient with the introduction of the sawtooth nucleus enhancement factor and the convective evaporation heat transfer coefficient with the introduction of the vortex two-phase enhancement factor. Critical drying calculation: The critical drying heat flux density is calculated using a model that combines centrifugal force factor and sawtooth factor. Temperature uniformity assessment: The flow channel is divided into multiple micro-segments along the flow channel, and a one-dimensional thermal balance equation is established for the micro-segment. The two-dimensional thermal conductivity of the aluminum plate substrate between the rings is coupled to solve the two-dimensional temperature distribution on the plate surface. The plate surface uniformity coefficient is calculated based on the highest and lowest temperatures. The output includes physical performance evaluation results such as heat transfer coefficient, total flow resistance, two-dimensional temperature distribution on the plate surface, temperature uniformity coefficient, and drying margin.

5. The physical property calculation method according to claim 4, characterized in that, In the calculation of the geometric characteristic parameters, the reversal curve is approximated as a rectangular spiral. The total length of the spiral on one side is calculated by determining the arithmetic progression relationship of the circumference of each turn increasing sequentially, and the total flow path length is derived by superimposing the length of the intermediate reversal section. The specific calculation formula includes: calculating the flow cross-sectional area using a semi-elliptical arched cross-section: Calculate the sawtooth area magnification factor based on sawtooth height and tooth spacing: Hydraulic diameter and total heat transfer area calculated based on wetted perimeter: , According to the The total length of a single-sided spiral is calculated by summing the circumferences of the spirals, and then the total flow path length is obtained. In the formula, For channel width, For the channel height, For the height of the serrations, The sawtooth spacing. For the wet perimeter of the flow channel, For a single flow channel, The hydraulic diameter of the flow channel. This is the magnification factor for the sawtooth area. For the total heat exchange area, This is the total length of a single-sided spiral. This represents the total length of a single loop. This represents the number of spiral turns on one side. The spacing between the circles, , The initial length and width of the innermost spiral are given. This is the length of the intermediate turnaround section.

6. The physical performance calculation method according to claim 5, characterized in that, In the flow resistance calculation, the friction coefficient along the flow path is obtained by multiplying the turbulence-based friction coefficient by a sawtooth resistance correction factor determined by the relative tooth height, tooth pitch-to-height ratio, and Reynolds number. The local resistance of the elbow is calculated based on the elbow resistance coefficient corrected by centrifugal force introduced by the curvature radius of each bend. The specific calculation formula includes: frictional resistance along the route: The basic coefficient of friction is: Sawtooth resistance correction factor: The friction coefficient along the friction path satisfies: ; Local resistance of the elbow: Among them, the elbow resistance coefficient is: Total flow resistance: In the formula, Here is the friction coefficient along the friction path. The friction coefficient of the straight pipe turbulent flow system. The total flow path length is _____. The hydraulic diameter of the flow channel. The density of the working fluid. The average axial velocity of the working fluid. The working fluid Reynolds number, These are the resistance fitting coefficients. For the height of the serrations, The sawtooth spacing. This is the sawtooth resistance correction factor. For the first The number of bends corresponding to the spiral spiral. Let be the local resistance coefficient of the bend in the i-th rotation of the spiral. For the first The working fluid velocity within the closed flow channel, The radius of curvature of the spiral is denoted as . For frictional resistance along the path, For the local resistance of the elbow, This creates localized obstacles to import and export. This represents the total flow resistance.

7. The physical property calculation method according to claim 5, characterized in that, In the single-phase flow condition for the heat transfer performance calculation, the single-phase Nusselt number is obtained by superimposing a secondary flow enhancement factor determined by the Dean number and a sawtooth turbulence enhancement factor determined by the sawtooth geometry parameters and Reynolds number on the basis of the basic Nusselt number; the specific calculation formula includes: Basic Nusselt number: Calculation of vortex secondary flow enhancement factor based on Dean number: Serrated turbulence enhancement factor: Corrected single-phase Nusselt number: Single-phase heat transfer coefficient: In the formula, For the straight-pipe basic Nusselt number, The working fluid Reynolds number, Let De be the Prandtl number of the working fluid, and De be the Dean number. The hydraulic diameter of the flow channel. The radius of curvature of the spiral is denoted as . It is a vortex secondary flow enhancement factor. For the height of the serrations, The sawtooth spacing. These are the heat transfer fitting coefficients. As a sawtooth turbulence enhancement factor, This is the corrected single-phase Nusselt number. The thermal conductivity of the working fluid is The single-phase heat transfer coefficient is denoted as .

8. The physical property calculation method according to claim 7, characterized in that, In the two-phase evaporation condition for heat transfer performance calculation, a boiling superposition model is adopted. A sawtooth nucleus boiling enhancement factor is introduced into the pool boiling heat transfer coefficient to calculate the nucleus boiling heat transfer coefficient. A vortex two-phase enhancement factor is introduced into the convective evaporation heat transfer coefficient based on the Lockhart-Martinelli parameters to calculate the convective evaporation heat transfer coefficient. Specific calculation formulas include: Serrated nucleus boiling enhancement factor: Nucleus boiling heat transfer coefficient: Lockhart-Martinelli parameters: Convection-evaporation heat transfer coefficient: Two-phase evaporation heat transfer coefficient: In the formula, It is a serrated nucleus boiling enhancement factor. These are the fitting coefficients for nucleate boiling. For the height of the serrations, The hydraulic diameter of the flow channel. The sawtooth spacing. The nucleation boiling heat transfer coefficient, For heat flux density, The molar mass of the working fluid. For the working fluid comparison pressure, For the surface roughness of the wall, For Lockhart-Martinelli parameters, For the dryness of the working fluid, , These are the densities of the liquid phase and the vapor phase, respectively. , These are the dynamic viscosity of the liquid phase and the vapor phase, respectively. This is the heat transfer coefficient when the liquid phase flows alone. It is a spiral two-phase enhancement factor. The convective-evaporative heat transfer coefficient is... The overall heat transfer coefficient for the two phases is... , These are the correction factors for nucleate boiling and convective evaporation, respectively.

9. The physical performance calculation method according to claim 8, characterized in that, In the aforementioned critical drying calculation, based on the critical heat flux density fundamental model, the basic drying heat flux density is corrected by multiplying it by the vortex centrifugal force factor and the serrated factor, respectively; the specific calculation formula includes: centrifugal Froude number and vortex centrifugal force factor: Sawtooth factor: Dry heat flux density: In the formula, For the eccentric Froude number, The axial velocity of the working fluid. It is the acceleration due to gravity. The radius of curvature of the spiral is denoted as . This is a correction factor for the centrifugal force of the vortex. For the axial Froude number, This is a correction factor for the sawtooth structure. For the height of the serrations, The hydraulic diameter of the flow channel. This refers to the number of saw teeth per unit area. For the dry heat flux density of the straight pipe foundation, This is the corrected dry heat flux density.

10. The physical performance calculation method according to claim 9, characterized in that, In the temperature uniformity assessment, the flow channel is divided into multiple micro-segments along the flow direction. A one-dimensional thermal balance differential equation for the fluid micro-segment is established and coupled with a two-dimensional thermal conductivity differential equation for the aluminum plate substrate between the rings. The two-dimensional temperature distribution on the plate surface is solved numerically using the finite difference method, and the temperature uniformity coefficient of the plate surface is calculated based on the temperature extreme values. The specific calculation formulas include: One-dimensional heat balance equation for a fluid infinitesimal segment: Plate surface temperature coefficient: In the formula, For the working fluid mass flow rate, The specific heat capacity of the working fluid at constant pressure The heat transfer coefficient, For the wet perimeter of the flow channel, The flow channel wall temperature, The temperature of the working fluid. , These are the highest and lowest temperatures on the board surface, respectively. , These are the inlet temperature and outlet temperature of the working fluid, respectively. The uniform temperature coefficient of the plate surface.

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