A method for designing a high-temperature and high-pressure fuel centrifugal pump ring structure

Abstract

The application provides a high-temperature and high-pressure fuel centrifugal pump ring structure design method, which comprises the following steps: performing water conservancy design on a closed impeller and a volute, twisted blades are used for the blades of the impeller; an impeller ring is arranged at the outer edge of the impeller inlet, a pump body ring is arranged at the corresponding position of the inner wall surface of the volute and the impeller ring, an annular sealing gap is formed between the impeller ring and the pump body ring, the gap comprises at least two sealing teeth with different diameters or axial misalignments; a finite element thermal-structure coupling model of the impeller, the volute and the ring assembly is established, the radial deformation amount of the outer diameter of the impeller ring and the inner diameter of the pump body ring in the steady-state process from the assembly temperature to the highest working temperature is calculated, and the thermal gap is obtained according to the radial deformation amount of the outer diameter of the impeller ring and the inner diameter of the pump body ring, and the safety of the thermal gap is checked; the centrifugal pump can simultaneously achieve high sealing performance, high efficiency and low axial force under the high-temperature and high-pressure fuel working condition.

Description

Technical Field

[0001] This invention relates to the field of high-temperature and high-pressure fuel centrifugal pumps, and in particular to a design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump. Background Technology

[0002] Centrifugal fuel pumps are core power components in aircraft engines, gas turbines, and high-pressure common rail fuel systems. As modern power equipment develops towards higher power density and higher reliability, the operating conditions of fuel pumps are becoming increasingly demanding, typically requiring long-term stable operation under high-temperature and high-pressure fuel environments. The inlet ring (also known as a sealing ring) is an annular gap component installed between the outer edge of the impeller inlet and the pump body. Its main function is to reduce volumetric leakage from the high-pressure area to the low-pressure area at the impeller inlet, while also affecting the pressure distribution of the impeller's front and rear cover plates, thus impacting pump efficiency and axial force balance.

[0003] The existing mouth ring structure is usually a common cylindrical mouth ring: the structure is the simplest, but there is only one throttling gap, which leads to serious leakage under high pressure differential and no thermal compensation capability; Single-bend mouth ring (single labyrinth): Although there is a turning throttling, the gap change is uncontrollable at high temperatures, and the sealing effect decreases sharply with the increase of temperature; Multi-tooth labyrinth mouth ring: Although it has multiple levels of clearance, the size of each clearance is fixed, which cannot adapt to thermal deformation over a wide temperature range, and it does not take into account the coupling design with axial force. Contact-type end face sealing ring: Zero leakage is achieved by pushing the end face of the sealing ring to fit together with a spring. However, the friction pair is prone to burning and wear under high speed and high temperature, and the self-lubricating properties of fuel are poor, resulting in a limited life.

[0004] As a result, existing design methods mostly rely on empirical formulas or single-physics field simulations, lacking a system design process that couples multiple fields of heat, structure, and fluid, leading to low reliability of the ring under high temperature and high pressure conditions. Summary of the Invention

[0005] The purpose of this invention is to provide a design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump that enables the centrifugal pump to simultaneously achieve high sealing performance, high efficiency, and low axial force under high-temperature and high-pressure fuel conditions.

[0006] To solve the above technical problems, the present invention provides a design method for a high-temperature and high-pressure fuel centrifugal pump inlet ring structure, comprising the following steps: Step 1: Conduct hydraulic design of closed impeller and volute, wherein the impeller blades are twisted blades; Step 2: Set an impeller inlet ring at the outer edge of the impeller inlet and a pump body inlet ring at the corresponding position on the inner wall of the volute. A ring assembly with an annular sealing gap is formed between the impeller inlet ring and the pump body inlet ring. The gap includes at least two sealing teeth with different diameters or axial misalignment. Step 3: For the high-temperature fuel operating temperature range, establish a finite element thermal-structural coupling model of the impeller, volute, and inlet ring assembly, calculate the radial deformation of the outer diameter of the impeller inlet ring and the inner diameter of the pump body inlet ring during the steady-state process from the assembly temperature to the highest operating temperature, and obtain the hot-state gap based on the radial deformation of the outer diameter of the impeller inlet ring and the inner diameter of the pump body inlet ring, and perform a safety check on the hot-state gap.

[0007] Furthermore, it also includes step four, opening a thermal compensation groove along the circumferential direction on the sealing surface of the pump body inlet ring, wherein the depth and width of the thermal compensation groove are determined according to the radial deformation amount in step three.

[0008] Furthermore, step five involves establishing a fluid dynamics model that includes the fluid domain of the mouth ring gap, the impeller front and rear cover plate gap, and the volute flow channel. The hot gap in step three and the depth and width of the thermal compensation groove in step four are used as design variables. Multiple sets of different values ​​of the design variables are arranged and combined to form multiple calculation schemes. For each scheme, the flow and pressure distribution of high-temperature and high-pressure fuel passing through the sealing teeth and thermal compensation groove are simulated. The design parameters of the impeller mouth ring and pump body mouth ring that minimize the energy loss of the fuel centrifugal pump and maximize the efficiency and head are selected.

[0009] Furthermore, in step three, the process of establishing a finite element thermal-structural coupling model of the impeller, volute, and nozzle ring assembly includes applying convective boundary conditions to the surfaces of the impeller, volute, and nozzle ring assembly that are in contact with fuel, and setting the heat transfer coefficients of the impeller surface, the inner wall of the volute, and the gap, respectively; solving the three-dimensional steady-state heat conduction equation using the finite element method to obtain the nodal temperature field; and performing static structural analysis based on the nodal temperature field to calculate the radial displacement of the impeller nozzle ring and the pump body nozzle ring caused by thermal expansion.

[0010] Furthermore, the heat transfer coefficient of the impeller surface is 1000-5000 W / (m²·K), the heat transfer coefficient of the inner wall of the volute is 500-2000 W / (m²·K), and the heat transfer coefficient of the mouth ring gap is 5000-8000 W / (m²·K).

[0011] Furthermore, in step three, when performing a safety check on the hot gap, the hot gap at the highest operating temperature is greater than the minimum allowable gap and less than the maximum allowable gap; the minimum allowable gap is 0.05-0.1mm, and the maximum allowable gap is 0.3-0.5mm.

[0012] Furthermore, in step four, the method for determining the depth of the heat compensation groove includes calculating the thickness of the heat-affected layer. The thickness of the heat-affected layer is calculated based on the radial depth of the average temperature drop to the inside and outside of the pump body inlet ring or the inner and outer radii of the pump body inlet ring. The depth of the heat compensation groove is 1-1.5 times the thickness of the heat-affected layer.

[0013] Furthermore, in step four, the width of the thermal compensation groove is determined as follows: the width of the thermal compensation groove is not less than 2-3 times the radial deformation of the inner diameter of the pump body inlet ring.

[0014] Furthermore, in step five, when screening the design parameters of the impeller wear ring and the pump body wear ring, the screening criteria include, in order, excluding schemes with an efficiency of no more than 80%, prioritizing schemes with low energy loss, and prioritizing schemes with high efficiency.

[0015] Furthermore, the impeller inlet ring and the impeller, and the pump body inlet ring and the pump body are configured to be in bonded contact or frictional contact, with the friction coefficient of the frictional contact being 0.1-0.2.

[0016] The beneficial effects of this invention are as follows: This invention provides a design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump. Through systematic steps (determination of an efficient hydraulic model, selection of a labyrinthine inlet ring foundation structure, calculation of gap compensation through thermo-structural coupling simulation, design of thermally adaptable variable gap characteristics, and iteration of CFD parameters for coupled axial force balance), a set of inlet ring geometric parameters is finally obtained to optimize the flow channel shape and size of the fuel centrifugal pump, the design of the sealing ring gap structure, and the matching design of the impeller and volute. This reduces internal flow losses, improves the performance of the fuel centrifugal pump, reduces energy loss, and increases the overall efficiency of the system, enabling the centrifugal pump to simultaneously achieve the goals of high sealing performance, high efficiency, and low axial force under high-temperature and high-pressure fuel conditions. Attached Figure Description

[0017] Figure 1 This is a flowchart illustrating the present invention.

[0018] Figure 2 This is a schematic diagram of the impeller structure in this invention.

[0019] Figure 3 This is a schematic diagram of the thermal compensation groove in this invention.

[0020] Attached figure label: 1. Thermal compensation groove. Detailed Implementation

[0021] The technical solutions of the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. All other embodiments obtained by those skilled in the art based on the embodiments of the present invention are within the scope of protection of the present invention.

[0022] Those skilled in the art should understand that, in the disclosure of this invention, the terms "longitudinal," "lateral," "upper," "lower," "front," "rear," "left," "right," "vertical," "horizontal," "top," "bottom," "inner," and "outer," etc., indicate the orientation or positional relationship based on the orientation or positional relationship shown in the accompanying drawings. They are only for the convenience of describing this invention and simplifying the description, and do not indicate or imply that the device or element referred to must have a specific orientation, or be constructed and operated in a specific orientation. Therefore, the above terms should not be construed as limiting this invention.

[0023] It is understood that the term "a" should be understood as "at least one" or "one or more", that is, in one embodiment, the number of an element can be one, while in another embodiment, the number of the element can be multiple, and the term "a" should not be understood as a limitation on the number.

[0024] like Figures 1-3 The present invention provides a design method for a high-temperature and high-pressure fuel centrifugal pump inlet ring structure, comprising the following steps: Step 1: Based on the design specifications of the fuel-fired centrifugal pump (flow rate Q, head H, speed n, and medium properties), complete the hydraulic design of the closed impeller and volute. Specifically, the impeller blades are twisted to achieve a uniform outlet velocity and a higher head coefficient. The matching between the impeller and volute is optimized using CFD to ensure that the efficiency at the design operating point is not lower than the preset target value (e.g., ≥80%). The hydraulic design of the impeller and volute is completed using the centrifugal pump velocity coefficient method.

[0025] Step 2: Install an impeller inlet ring at the outer edge of the impeller inlet and a pump body inlet ring on the corresponding inner wall of the volute. An annular sealing gap is formed between the impeller inlet ring and the pump body inlet ring. This gap employs a labyrinthine structure: that is, it contains at least two sealing teeth of different diameters or axially misaligned to increase the local resistance to leakage flow.

[0026] Step 3: For the high-temperature fuel operating temperature range T_min-T_max, establish a finite element thermo-structural coupling model of the impeller-volute-rim ring assembly. Calculate the outer diameter r of the impeller rim ring during the steady-state process from the assembly temperature T_assembly to the maximum operating temperature T_max. outer The change in the inner diameter of the pump body inlet ring Δr inner .

[0027] Step 3.1: Use ANSYS Mechanical to calculate the temperature distribution inside the component.

[0028] 3.1.1 Apply convective boundaries to all surfaces of the impeller, volute, and fuel contact ring, with the heat transfer coefficient set as follows: Impeller surface: h1≈1000-5000W / (m)2 K) (depends on rotational speed and turbulence); Circular casing wall surface: h2≈500-2000W / (m 2 K); Annular gap: Due to the narrow gap and high flow velocity, h3 can be taken as a relatively high value, 5000-8000 W / (m²). 2 K).

[0029] h1: Impeller surface heat transfer coefficient (unit: W / (m²)) 2 K); h2: volute wall surface: heat transfer coefficient (unit: W / (m²)) 2 K); h3: Heat transfer coefficient of the annular gap (unit: W / (m²)) 2 K).

[0030] 3.1.2 Ambient temperature setting: fuel mainstream temperature Tfluid: i.e. the temperature under design conditions (e.g., 150℃).

[0031] 3.1.3 Contact thermal resistance setting: Contact surface between impeller inlet ring and impeller body: Contact thermal resistance can be set (ignoring gap, considered as bonding or defining thermal conductivity).

[0032] 3.1.4 Assembly temperature Tassembly setting: usually room temperature (20℃), at which point the components are in a free state (no thermal stress, only assembly stress can be selected).

[0033] 3.1.5 Calculate the nodal temperature field T(x,y,z).

[0034] The nodal temperature field T(x,y,z) was obtained by solving the three-dimensional steady-state heat conduction equation using the finite element method (FEM) and ANSYS Mechanical software. The underlying governing equation is as follows: ; k: thermal conductivity of the material (unit: W / (m·K)); Temperature gradient in the x-direction; Temperature gradient in the y-direction; Temperature gradient in the z-direction.

[0035] Step 3.2: Based on the temperature field obtained from the thermal analysis, use the Workbench static structure module of the simulation software to perform static structural analysis and calculate the displacement and stress caused by thermal expansion.

[0036] 3.2.1 Displacement constraints, impeller: Hub inner bore surface: Apply cylindrical constraints (allowing only axial and circumferential freedom, radially fixed), or simulate the equivalent stiffness of bearing support; Pump body (volute): Mounting flange face: Apply full constraints (fixed); 3.2.2 Temperature load: The nodal temperature obtained from thermal analysis is applied as a volume load.

[0037] 3.2.3 Contact and connection, between impeller inlet ring and pump body inlet ring: set to frictionless or gap contact (initial gap exists, no force is transmitted), only their deformation is analyzed; assembly relationship between impeller inlet ring and impeller, and between pump body inlet ring and pump body: set to bonded or frictional contact (coefficient 0.1-0.2).

[0038] 3.2.4 Calculation of radial deformation: Impeller mouth ring outer diameter D imp，ring Select several circumferential nodes on the outer cylindrical surface of the ring and calculate its radial displacement u. r The change in outer diameter is obtained after averaging: ; : Change in the outer diameter of the impeller inlet ring (unit: mm); Average radial displacement of several circumferential nodes on the outer cylindrical surface of the ring (unit: mm).

[0039] Pump body inlet ring inner diameter D casing，ring Calculate the radial displacement at the inner cylindrical surface node, and then average it to obtain the change in outer diameter: ; Change in the outer diameter of the pump body inlet ring (unit: mm); : Average radial displacement of several circumferential nodes of the cylindrical surface node inside the ring (unit: mm).

[0040] Hot clearance: Let the initial clearance designed under cold conditions (assembly temperature) be... The hot gap at the highest operating temperature is: ; Thermal expansion of the pump body's inner diameter reduces the clearance (if the inner bore bulges inward), while expansion of the impeller's outer diameter also reduces the clearance; the directions of both must be defined according to actual geometry.

[0041] : Hot gap at the highest operating temperature (unit: mm); Initial clearance designed under cold conditions (assembly temperature) (unit: mm); 3.2.5 Safety Verification: make sure ( Minimum allowable clearance, typically 0.05-0.1mm, to prevent rubbing. make sure ( The maximum permissible gap is typically 0.3-0.5 mm to prevent excessive leakage.

[0042] Step 4: Based on the above basic clearance design, an annular thermal compensation groove 1 is further opened on the sealing surface of the pump body inlet ring. The depth ds and width ws of the thermal compensation groove 1 are determined according to the radial deformation calculated above. When the high temperature causes the pump body inlet ring to thermally expand and bulge inward, the thermal compensation groove 1 provides a space for deformation to be released, avoiding excessive shrinkage of the sealing surface.

[0043] The specific method for determining the depth and width of the annular thermal compensation groove 1 is as follows: 4.1 Determine the thickness δT of the heat-affected layer; The thickness of the heat-affected layer is defined as the temperature drop from the inner surface to the outermost surface. Or the radial depth of a certain threshold.

[0044] Approximate formula: ; in , , where are the inner and outer temperatures of the ring, respectively; k is the thermal conductivity of the material; and h is the heat transfer coefficient of the inner surface (obtained in step three). , These are the inner and outer radii of the ring, respectively.

[0045] The tank depth ds should be slightly greater than δT, usually ds=(1.0-1.5)δT, to ensure complete coverage of the heat-affected zone and make the bottom of the tank close to the isothermal zone.

[0046] 4.2 Determine the groove width w s The theoretical lower limit; The width of the groove must ensure that the thermal compensation groove does not significantly weaken the structural strength in the radial direction, while providing sufficient lateral space to allow for radial shrinkage of the inner wall.

[0047] From a stress analysis perspective, the groove width should be greater than the change in the inner diameter of the pump body inlet ring. (Calculated in step three) 2-3 times: ; Step 5: Establish a complete three-dimensional geometric model including the fluid domain of the inlet ring gap, the impeller front and rear cover plate gaps, and the volute flow channel; combine the calculations obtained in Steps 3 and 4. ds and ws are used as design variables. Multiple sets of different values ​​are combined based on the aforementioned basic design range to form several calculation schemes. For each scheme, ANSYS Fluent software is used to accurately simulate the flow and pressure distribution of high-temperature, high-pressure fuel as it passes through the labyrinth tooth and annular thermal compensation groove, calculating the pump efficiency, head, and energy loss under that scheme. This ensures... Under the premise that the pump is within the safe clearance range and ds and ws meet the lower limit of structural thermal compensation, select the group that minimizes pump energy loss and maximizes efficiency and head. Design parameters for the mouth rings of ds and ws.

[0048] Among them, Solidworks software was used to create a complete three-dimensional geometric model that includes the fluid domain of the mouth ring gap, the gap between the front and rear cover plates of the impeller, and the volute flow channel.

[0049] The specific screening criteria are as follows: First, exclude undesirable schemes that are inefficient (less than or equal to 80%), have insufficient head (calculated head is less than 95% of the design target head), or have an energy loss increase rate exceeding 20%; Second, prioritize the design with the lowest energy loss among the remaining schemes; Third, when the energy loss difference between different schemes does not exceed 5%, prioritize the scheme with higher efficiency.

[0050] This invention is not limited to the preferred embodiments described above. Anyone can derive other products in various forms under the guidance of this invention. However, regardless of any changes in shape or structure, any technical solution that is the same as or similar to this application falls within the protection scope of this invention.

Claims

1. A method of designing a high temperature high pressure fuel centrifugal pump ring structure, characterized by, Includes the following steps: Step 1: Conduct hydraulic design of closed impeller and volute, wherein the impeller blades are twisted blades; Step 2: Set an impeller inlet ring at the outer edge of the impeller inlet and a pump body inlet ring at the corresponding position on the inner wall of the volute. A ring assembly with an annular sealing gap is formed between the impeller inlet ring and the pump body inlet ring. The gap includes at least two sealing teeth with different diameters or axial misalignment. Step 3: For the high-temperature fuel operating temperature range, establish a finite element thermal-structural coupling model of the impeller, volute, and inlet ring assembly, calculate the radial deformation of the outer diameter of the impeller inlet ring and the inner diameter of the pump body inlet ring during the steady-state process from the assembly temperature to the highest operating temperature, and obtain the hot-state gap based on the radial deformation of the outer diameter of the impeller inlet ring and the inner diameter of the pump body inlet ring, and perform a safety check on the hot-state gap.

2. The high temperature high pressure fuel centrifugal pump ring structure design method of claim 1, wherein: It also includes step four, opening a thermal compensation groove (1) circumferentially on the sealing surface of the pump body inlet ring, wherein the depth and width of the thermal compensation groove (1) are determined according to the radial deformation amount in step three.

3. The design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump according to claim 2, characterized in that: The process also includes step five, establishing a fluid dynamics model that includes the fluid domain of the mouth ring gap, the impeller front and rear cover plate gap, and the volute flow channel. The hot gap in step three and the depth and width of the thermal compensation groove in step four are used as design variables. Multiple sets of different values ​​of the design variables are arranged and combined to form multiple calculation schemes. For each scheme, the flow and pressure distribution of high-temperature and high-pressure fuel passing through the sealing teeth and thermal compensation groove are simulated. The design parameters of the impeller mouth ring and pump body mouth ring that minimize the energy loss of the fuel centrifugal pump and maximize the efficiency and head are selected.

4. The design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump according to claim 1, characterized in that: Step three involves establishing a finite element thermal-structural coupling model of the impeller, volute, and piston ring assembly. This includes applying convective boundary conditions to the surfaces of the impeller, volute, and piston ring assembly that are in contact with fuel, and setting the heat transfer coefficients for the impeller surface, the inner wall of the volute, and the gap, respectively. The three-dimensional steady-state heat conduction equation is solved using the finite element method to obtain the nodal temperature field. Static structural analysis is then performed based on the nodal temperature field to calculate the radial displacement of the impeller piston ring and the pump body piston ring caused by thermal expansion.

5. The design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump according to claim 4, characterized in that: The heat transfer coefficient of the impeller surface is 1000-5000W / (m²·K), the heat transfer coefficient of the inner wall of the volute is 500-2000W / (m²·K), and the heat transfer coefficient of the mouth ring gap is 5000-8000W / (m²·K).

6. The design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump according to claim 1, characterized in that: In step three, when performing a safety check on the hot gap, the hot gap at the highest operating temperature is greater than the minimum allowable gap and less than the maximum allowable gap; the minimum allowable gap is 0.05-0.1mm and the maximum allowable gap is 0.3-0.5mm.

7. The design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump according to claim 2, characterized in that: In step four, the method for determining the depth of the heat compensation groove (1) includes calculating the thickness of the heat-affected layer. The thickness of the heat-affected layer is calculated based on the radial depth of the average temperature of the pump body inlet ring or the inner and outer radii of the pump body inlet ring at the temperature drop. The depth of the heat compensation groove (1) is 1-1.5 times the thickness of the heat-affected layer.

8. The design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump according to claim 2, characterized in that: In step four, the width of the thermal compensation groove (1) is determined as follows: the width of the thermal compensation groove (1) is not less than 2-3 times the radial deformation of the inner diameter of the pump body inlet ring.

9. The design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump according to claim 3, characterized in that: In step five, when screening the design parameters of the impeller wear ring and the pump body wear ring, the screening criteria are as follows: exclude schemes with an efficiency of no more than 80%, prioritize schemes with low energy loss, and prioritize schemes with high efficiency.

10. The design method for the inlet ring structure of a high-temperature and high-pressure fuel centrifugal pump according to claim 1, characterized in that: The impeller inlet ring and the impeller, and the pump body inlet ring and the pump body are configured to be in bonded contact or frictional contact, with the friction coefficient of the frictional contact being 0.1-0.2.

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