A centripetal impeller and a method of designing the same

Abstract

The application relates to the technical field of aero-engines, and discloses a centripetal impeller and a design method thereof, the centripetal impeller comprising an impeller disc, blades and a damping structure; the blade roots of the blades are connected with the impeller disc, and a plurality of blades are arranged at intervals along the circumference of the impeller disc; the damping structure is embedded in the sandwich space of the blades, the distance between the damping structure and the edge of the blade is L1, 0 < L1 < 2 mm is satisfied, the blade height of the blade is H, the distance between the damping structure and the blade root is L2, and 0.05H <= L2 <= 0.3H is satisfied. The damping structure is embedded in the sandwich space of the blade, the damping structure efficiently consumes vibration energy during the vibration process of the blade, vibration of the blade itself is reduced, and the aerodynamic performance of the impeller is ensured; the size of the damping structure is limited, the damping structure is close to the strong vibration area of the blade, the vibration reduction effect is ensured, the aerodynamic key area and the structural high stress area of the blade are avoided, and the aerodynamic performance and the overall strength of the blade are ensured.

Description

Technical Field

[0001] This invention relates to the field of aero-engine technology, specifically to a centripetal impeller and its design method. Background Technology

[0002] Centripetal turbines, with their advantages of simple structure, high reliability, and high efficiency even at low flow rates, are widely used in small and medium-sized aero-engines. When the centripetal impeller of an aero-engine rotates at high speed, it is subjected to alternating loads of centrifugal force, aerodynamic force, and thermal stress. The superposition of vibration stresses can easily lead to high-cycle fatigue of the impeller. A significant portion of structural failures in aero-engines originates from blade vibration fatigue fracture. Therefore, in the design of centripetal turbines for aero-engines, it is necessary to control the vibration of the centripetal impeller to ensure its lifespan and reliability.

[0003] In existing technologies, vibration reduction techniques for centripetal impellers involve hollowing out the impeller rim area to create a sawtooth structure, thereby altering the local stiffness and adjusting the natural frequency to avoid resonance with the excitation frequency. However, this method has a relatively small impact on the vibration of the blade area of ​​the centripetal impeller, making it difficult to effectively suppress the vibration of the centripetal impeller blades. Summary of the Invention

[0004] This invention provides a centripetal impeller and its design method to solve the problem of effectively suppressing vibration of centripetal impeller blades.

[0005] In a first aspect, the present invention provides a centripetal impeller, comprising an impeller disk, blades, and a damping structure; the blades have blade roots connected to the impeller disk, and multiple blades are provided, the multiple blades being spaced apart circumferentially along the impeller disk, and each blade having a sandwich space; the damping structure is embedded in the sandwich space, the distance between the damping structure and the edge of the blade is L1, satisfying 0 < L1 < 2 mm, the blade height is H, and the distance between the damping structure and the blade root is L2, satisfying 0.05H ≤ L2 ≤ 0.3H.

[0006] Beneficial effects: By embedding a damping structure within the blade interlayer space, the damping characteristics of the damping structure are utilized to efficiently dissipate vibration energy during blade vibration without altering the original blade profile, thereby reducing vibration of the blade itself and ensuring the aerodynamic performance of the impeller. Furthermore, by limiting the size of the damping structure, it is positioned close to the blade's strong vibration zone to ensure the vibration reduction effect, while avoiding the blade's critical aerodynamic areas and high-stress areas, thus guaranteeing the blade's aerodynamic performance and overall strength.

[0007] In one optional embodiment, the static strength reserve coefficient at the leaf root is n. 0.2 , satisfying n 0.2 =σ 0.2(T) / σ max , where σ 0.2 (T) represents the yield strength of the material at the operating temperature, σ max This represents the maximum equivalent stress at that location; when n 0.2 When n < 1.2, L2 = 0.3H; when n 0.2 When n > 1.5, L2 = 0.05H; when 1.2 ≤ n 0.2 When ≤1.5, L2 = 0.3H - (n 0.2 -1.2)×(5 / 6)H.

[0008] Beneficial effects: The distance between the damping structure and the blade root is selected based on the static strength reserve coefficient at the blade root, and the structure of the damping structure at different positions is optimized.

[0009] In one optional embodiment, the centripetal impeller has an inlet and an outlet, the blade root friction length from the inlet to the outlet is x% of the total blade root length, and the distance L2 between the damping structure and the blade root is y% of the blade height H, satisfying the following expression: y = (3.2323x) / (2π) 4 -6.4647x 3 +3.5264x 2 -0.294x+0.102)×100%; and 8%≤y≤25%.

[0010] Beneficial effects: By limiting the curve expression, the damping structure is brought closer to the blade root in the low-stress areas at the blade inlet and outlet, providing better damping effect. In the high-stress area of ​​the blade's middle chord, the distance between the damping structure and the blade root is increased, providing better strength, thus achieving a balance between blade strength maintenance and damping effect.

[0011] In one alternative embodiment, the blade has amplitude contour lines, and multiple amplitude contour lines are formed. The damping structure intersects with each of the amplitude contour lines at each intersection point, and the extension direction of the damping structure at each intersection point is perpendicular to the extension direction of the intersecting amplitude contour line.

[0012] Beneficial effects: At the intersection of the damping structure with each amplitude contour line, its extension direction is perpendicular to the amplitude contour line, so that the direction of the damping structure is completely consistent with the vibration bending direction of the blade. It can follow the vibration mode of the blade to the maximum extent to bend and deform, give full play to the high damping characteristics, maximize the consumption of blade vibration energy, and ensure the vibration reduction effect.

[0013] In one optional embodiment, the damping structure includes a plurality of damping units, with adjacent damping units spaced apart. Each damping unit intersects with each amplitude contour line at a point, and the extension direction of the damping unit at each intersection point is perpendicular to the extension direction of the amplitude contour line it intersects with.

[0014] Beneficial effects: The damping structure is divided into several damping units with intervals, and each damping unit intersects the amplitude contour line perpendicularly, ensuring that the damping structure follows the vibration and bending of the blade and achieves the basis for efficient vibration reduction.

[0015] In one optional implementation, the length of the amplitude contour line is A, and the sum of the widths of the plurality of damping elements along the extension direction of the amplitude contour line is B, satisfying 0.4A≤B≤A.

[0016] Beneficial effects: Along the extension direction of the amplitude contour line, the total width of several damping units is not less than 40% of the length of the amplitude contour line. This reduces the proportion of the damping structure in the overall blade structure, thereby reducing the impact on the blade strength, while ensuring the damping effect of the damping unit during the bending vibration of the blade.

[0017] In one alternative embodiment, the edge of the blade includes a leading edge, the damping structure has a top edge and a bottom edge, the top edge being arranged parallel to the leading edge, and the bottom edge being arranged parallel to the amplitude contour lines at its location.

[0018] Beneficial effects: The top edge of the damping structure is parallel to the leading edge of the blade, and the bottom edge is parallel to the amplitude contour line at the location, so that the outline of the damping structure is adapted to the structural shape and vibration characteristics of the blade, and the arrangement of the damping structure in the blade is more in line with the actual structure of the blade, avoiding local stress concentration.

[0019] In one optional embodiment, the thickness of the blade is W1, and the thickness of the damping structure is W2, satisfying 0.3W1≤W2≤0.5W2.

[0020] Beneficial effects: By limiting the thickness of the damping structure, it is possible to ensure that the damping structure has sufficient damping performance, avoid insufficient vibration reduction due to excessive thickness, and avoid the adverse effects of excessive thickness on the overall strength of the blade, thus preventing the blade from failing in strength when rotating at high speed and subjected to alternating loads.

[0021] In one optional embodiment, the operating temperature of the blade is T; when 0 < T < 500°C, the damping structure is a titanium-based alloy damping structure or a cobalt-nickel-based alloy damping structure; when 700°C < T < 800°C, the damping structure is an oxygen-doped MPEA alloy damping structure.

[0022] Secondly, the present invention also provides a design method for a centripetal impeller, wherein the centripetal impeller is the aforementioned centripetal impeller, comprising the following steps: Step S1, analyzing the shape of the blades, determining the dimensions of the damping structure, and establishing a model of the centripetal impeller; Step S2, calculating the static strength, high-cycle fatigue strength, and natural frequency of the model of the centripetal impeller, and comparing the calculation results with the evaluation criteria; if the requirements are met, the designed centripetal impeller is deemed usable; if the requirements are not met, the damping structure is adjusted until the model of the centripetal impeller meets the evaluation criteria requirements.

[0023] In an optional implementation, the adjustment of the damping structure in step S2 includes the following steps: Step S21: Calculate the amplitude of the blades in the above model, adjust the damping structure to be strip-shaped according to the amplitude contour map, and make the damping structure perpendicular to the amplitude contour lines to complete the establishment of the centripetal impeller model of the strip-shaped damping structure; Step S22: Perform static strength, high-cycle fatigue strength and natural frequency calculations on the centripetal impeller model of the strip-shaped damping structure, and compare the calculation results with the evaluation criteria; if the requirements are met, it proves that the designed centripetal impeller is usable; if the requirements are not met, proceed to step S23; Step S23: Readjust the dimensional parameters of the strip-shaped damping structure until the centripetal impeller model of the strip-shaped damping structure meets the evaluation criteria requirements. Attached Figure Description

[0024] To more clearly illustrate the specific embodiments of the present invention or the technical solutions in the prior art, the drawings used in the description of the specific embodiments or the prior art will be briefly introduced below. Obviously, the drawings described below are some embodiments of the present invention. For those skilled in the art, other drawings can be obtained from these drawings without creative effort.

[0025] Figure 1 This is a cross-sectional view of the centripetal impeller according to an embodiment of the present invention; Figure 2 This is a cross-sectional view of the blade according to an embodiment of the present invention; Figure 3 This is a schematic diagram of the amplitude contour lines of the blade in an embodiment of the present invention; Figure 4 This is a schematic diagram of the distribution structure of the damping structure according to an embodiment of the present invention; Figure 5 This is a schematic diagram of the inlet and outlet of the centripetal impeller and the friction trajectory points along the blade root in an embodiment of the present invention. Figure 6 This is a flowchart illustrating the steps of the centripetal impeller design method of the present invention.

[0026] Explanation of reference numerals in the attached figures: 10. Impeller disk; 20. Blade; 21. Blade root; 22. Edge; 221. Leading edge; 222. Upper edge; 23. Amplitude contour lines; 30. Damping structure; 31. Damping unit; 32. Top edge; 33. Bottom edge; 41. Inlet; 42. Outlet. Detailed Implementation

[0027] To make the objectives, technical solutions, and advantages of the embodiments of the present invention clearer, 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, not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of the present invention.

[0028] The following is combined Figures 1 to 6 The following describes embodiments of the present invention.

[0029] According to an embodiment of the present invention, in a first aspect, a centripetal impeller is provided, comprising an impeller disk 10, blades 20, and a damping structure 30; the blades 20 have blade roots 21 connected to the impeller disk 10, and multiple blades 20 are provided, which are spaced apart circumferentially along the impeller disk 10, and the blades 20 have a sandwich space; the damping structure 30 is embedded in the sandwich space, the distance between the damping structure 30 and the edge 22 of the blade 20 is L1, satisfying 0 < L1 < 2 mm, the blade height of the blade 20 is H, and the distance between the damping structure 30 and the blade root 21 is L2, satisfying 0.05H ≤ L2 ≤ 0.3H.

[0030] The radial impeller of this embodiment uses a damping structure 30 embedded in the interlayer space of the blade 20. Without changing the original blade shape of the blade 20, the damping characteristics of the damping structure 30 are used to efficiently consume vibration energy during the vibration of the blade 20, thereby reducing the vibration of the blade 20 itself. While effectively reducing vibration, the aerodynamic performance of the impeller is guaranteed. Furthermore, by limiting the size of the damping structure 30, the damping structure 30 is placed close to the strong vibration area of ​​the blade 20 to ensure the vibration reduction effect, while avoiding the aerodynamic critical area and high stress area of ​​the blade 20, thus ensuring the aerodynamic performance and overall strength of the blade 20.

[0031] Specifically, in this embodiment, L1=1mm, L2=0.2H=16mm.

[0032] Of course, in other alternative implementations, the dimensions of the damping structure 30 can also be adjusted.

[0033] It should be noted that, within the limits of the strength of the blade 20, the distribution area of ​​the damping structure 30 on the blade 20 should be as large as possible to improve the vibration reduction effect on the blade 20.

[0034] It should be noted that the vibration amplitude of the blades 20 of the radial impeller gradually increases from the blade root 21 to the blade tip. The edge 22 of the blade 20 is the part with the most significant vibration displacement and bending deformation, and it is also the area where vibration energy is most concentrated. If L1 ≥ 2 mm, the damping structure 30 will have a significant gap with the strong vibration area of ​​the edge 22 of the blade 20, and will not be able to fully follow the large-amplitude bending vibration of the edge 22 of the blade 20. It can only suppress the small-amplitude vibration of the middle part of the blade 20, resulting in insufficient vibration energy consumption, a significant decrease in the vibration reduction effect, and an unsatisfactory vibration reduction effect.

[0035] It should be noted that the blade root 21 is the connection point between the blade 20 and the impeller disk 10. It is the main bearing area of ​​centrifugal force and alternating load when the centrifugal impeller rotates at high speed. The distance L2 between the damping structure 30 and the blade root 21 is limited to ensure the effective bearing area of ​​the base material of the blade root 21, so as to have sufficient static strength and fatigue strength, and ensure that the blade 20 can be reliably connected to the impeller disk 10, and that the blade root 21 will not break or loosen due to centrifugal force.

[0036] In one embodiment, the static strength reserve coefficient at the leaf root 21 is n 0.2 , satisfying n 0.2 =σ 0.2 (T) / σ max , where σ 0.2 (T) represents the yield strength of the material at the operating temperature, σ max This represents the maximum equivalent stress at that location; when n 0.2 When n < 1.2, L2 = 0.3H; when n 0.2 When n > 1.5, L2 = 0.05H; when 1.2 ≤ n 0.2 When ≤1.5, L2 = 0.3H - (n 0.2 -1.2)×(5 / 6)H.

[0037] It is worth noting that the distance between the damping structure 30 and the blade root 21 is selected based on the static strength reserve coefficient at the blade root 21, and the structure of the damping structure 30 at different positions is optimized.

[0038] In one embodiment, refer to Figure 5 As shown, the centripetal impeller has an inlet 41 and an outlet 42. The length of the blade root 21 from the inlet 41 to the outlet 42 is x, and the distance L2 between the damping structure 30 and the blade root is y, which is the percentage of the blade height H of the blade 20. The following expression is satisfied: y = (3.2323x) 4 -6.4647x 3 +3.5264x 2 -0.294x+0.102)×100%; and 8%≤y≤25%.

[0039] It should be noted that the reference Figure 5 As shown, the trajectory of the blade root 21 of the blade 20 from the inlet 41 to the outlet 42 is arc-shaped. At point D1 in the figure, that is, when the blade 20 is at the inlet 41, x=0. At point D2 in the figure, that is, when the blade 20 is at the midpoint of the arc trajectory, x=0.5. At point D3 in the figure, that is, when the blade 20 is at the outlet 42, x=1.

[0040] It should be noted that the damping structure 30 cannot be set at the outlet 42 and the inlet 41. That is, in the actual calculation of the expression, x≠0 or x≠1, and the range of the proportional x satisfies L1 / S≤x≤(S-L1) / S, where S is the total length of the blade root 21 along the blade 20 from the inlet 41 to the outlet 42 (that is, the total length of the trajectory of the blade root 21 from the inlet 41 to the outlet 42).

[0041] Specifically, in this embodiment, the relevant parameters of the damping structure 30 are shown in the table below.

[0042]

[0043] It is worth noting that, by limiting the curve expression, the damping structure 30 is brought closer to the blade root in the low-stress areas of the inlet 41 and outlet 42 of the blade 20, providing a better damping effect. In the high-stress area of ​​the middle chord of the blade 20, the distance between the damping structure 30 and the blade root 21 is increased to provide better strength, thus achieving a balance between the strength maintenance and damping effect of the blade 20.

[0044] In one embodiment, such as Figure 3 As shown, the blade 20 has amplitude contour lines 23, and multiple amplitude contour lines 23 are formed. The damping structure 30 intersects with each amplitude contour line 23. The extension direction of the damping structure 30 at each intersection point is perpendicular to the extension direction of the amplitude contour line 23 it intersects with.

[0045] Specifically, the amplitude contour line 23 is a closed (or semi-closed) curve formed by connecting points with equal vibration amplitude on the surface of the blade 20 after the vibration characteristic simulation test of the blade 20. It is a geometric line that characterizes the spatial distribution law of the vibration amplitude of the centripetal impeller blade 20 under working conditions. It is also an intuitive and visual expression of the vibration characteristics of the blade 20, and is used to accurately locate the strong vibration area and weak vibration area of ​​the blade 20.

[0046] Specifically, in this embodiment, in order to improve the vibration reduction effect of the centripetal turbine blades 20 while reducing the distribution area of ​​the damping structure 30 and reducing the impact of the damping structure 30 on the strength of the blades 20, the damping structure 30 adopts a strip structure.

[0047] It is worth noting that at the intersection of the damping structure 30 with each amplitude contour line 23, its extension direction is perpendicular to the amplitude contour line 23, so that the direction of the damping structure 30 is completely consistent with the vibration bending direction of the blade 20. This allows it to follow the vibration mode of the blade 20 to the maximum extent, fully utilize its high damping characteristics, maximize the consumption of the vibration energy of the blade 20, and ensure the vibration reduction effect.

[0048] In one embodiment, such as Figure 4 As shown, the damping structure 30 includes a plurality of damping units 31, with adjacent damping units 31 spaced apart. Each damping unit 31 intersects with each amplitude contour line 23. The extension direction of the damping unit 31 at each intersection point is perpendicular to the extension direction of the amplitude contour line 23 it intersects with.

[0049] Specifically, such as Figure 4 As shown, in this embodiment, the damping structure 30 includes five strip-shaped damping units 31 arranged at intervals between each other.

[0050] It is worth noting that the damping structure 30 is divided into several damping units 31 arranged at intervals, and each damping unit 31 intersects perpendicularly with the amplitude contour line 23, which ensures that the damping structure 30 follows the vibration and bending of the blade 20 and achieves the basis for efficient vibration reduction.

[0051] In one embodiment, the length of the amplitude contour line 23 is A, and the sum of the widths of the plurality of damping elements 31 along the extension direction of the amplitude contour line 23 is B, satisfying 0.4A≤B≤A.

[0052] It should be noted that the length of each amplitude contour line 23 is different. When calculating the total width of several damping elements 31, it is necessary to calculate along the extension direction of different amplitude contour lines 23.

[0053] Specifically, in combination Figure 3 and Figure 4 As shown, the lengths of the four amplitude contour lines 23 in the diagram are denoted as A1, A2, A3, and A4 from left to right. The first amplitude contour line 23 on the left crosses three damping units 31, and the widths of the three damping units 31 along the extension direction of this amplitude contour line 23 are B, B, and A4, respectively. 11 B 12 B 13 B1=B 11 +B 12 +B 13 In this embodiment, B1 = 0.4A1; the second amplitude contour line 23 on the left crosses four damping units 31, and the widths of the four damping units 31 along the extension direction of this amplitude contour line 23 are respectively B 21 B 22 B 23 B24 B2=B 21 +B 22 +B 23 +B 24 In this embodiment, B2 = 0.4A2; the third amplitude contour line 23 on the left crosses five damping units 31, and the widths of the five damping units 31 along the extension direction of this amplitude contour line 23 are respectively B 31 B 32 B 33 B 34 B 35 B3=B 31 +B 32 +B 33 +B 34 +B 35 In this embodiment, B3 = 0.4A3; the rightmost amplitude contour line 23 crosses five damping units 31, and the widths of the five damping units 31 along the extension direction of this amplitude contour line 23 are respectively B 41 B 42 B 43 B 44 B 45 B4=B 41 +B 42 +B 43 +B 44 +B 45 In this embodiment, B4 = 0.4A4.

[0054] It is worth noting that, along the extension direction of the amplitude contour line 23, the total width of several damping units 31 is not less than 40% of the length of the amplitude contour line 23. This reduces the proportion of the damping structure 30 in the overall blade 20, thereby reducing the impact on the strength of the blade 20, while ensuring the vibration reduction effect of the damping unit 31 during the bending vibration of the blade 20.

[0055] In one embodiment, such as Figure 4 As shown, the edge 22 of the blade 20 includes a leading edge 221, and the damping structure 30 has a top edge 32 and a bottom edge 33. The top edge 32 is arranged parallel to the leading edge 221, and the bottom edge 33 is arranged parallel to the amplitude contour line 23 at its location.

[0056] Specifically, such as Figure 4 As shown, the edge 22 of the blade 20 also includes an upper edge 222, and the amplitude contour line 23 extends from the leading edge 221 of the blade 20 to the upper edge 222 of the blade 20.

[0057] It is worth noting that the top edge 32 of the damping structure 30 is parallel to the leading edge 221 of the blade 20, and the bottom edge 33 is parallel to the amplitude contour line 23 at its location. This makes the outline of the damping structure 30 match the structural shape and vibration characteristics of the blade 20, and makes the arrangement of the damping structure 30 within the blade 20 more in line with the actual structure of the blade 20, avoiding local stress concentration. In addition, the damping structure 30 always deforms along the optimal direction during the vibration of the blade 20, continuously and efficiently playing a vibration reduction role.

[0058] In one embodiment, such as Figure 2 As shown, the thickness of blade 20 is W1, and the thickness of damping structure 30 is W2, satisfying 0.3W1≤W2≤0.5W2.

[0059] Specifically, in this embodiment, the thickness W2 of the damping structure 30 is 0.4W1.

[0060] Of course, in other alternative implementations, the thickness of the damping structure 30 can be adjusted within a limited range according to the actual situation.

[0061] It should be noted that when W2 < 0.3W1, the thickness of the damping structure 30 is relatively thin, resulting in a poor damping effect on the blade 20; when W2 > 0.5W2, the thickness of the damping structure 30 is relatively thick, and the thicker damping structure 30 is embedded in the blade 20, affecting the overall structural strength of the radial impeller blade 20.

[0062] It is worth noting that by limiting the thickness of the damping structure 30, it is possible to ensure that the damping structure 30 has sufficient damping performance and avoid insufficient vibration reduction effect due to excessive thickness. At the same time, it avoids the adverse effects of excessive thickness of the damping structure 30 on the overall strength of the blade 20 and prevents the blade 20 from failing in strength when rotating at high speed and bearing alternating loads.

[0063] In one embodiment, the operating temperature of the blade 20 is T; when 0 < T < 500°C, the damping structure 30 is a titanium-based alloy damping structure 30 or a cobalt-nickel-based alloy damping structure 30; when 700°C < T < 800°C, the damping structure 30 is an oxygen-doped MPEA alloy damping structure 30.

[0064] Specifically, in this embodiment, the operating temperature of the blade 20 is 750°C, and the material of the damping structure 30 is an oxygen-doped MPEA alloy (Multi-Principal Element Alloy).

[0065] Of course, in other alternative implementations, other materials can be selected as the damping structure 30 according to the operating temperature of the blade 20.

[0066] It should be noted that the centripetal impeller with damping structure 30 can be fabricated using powder feeding additive manufacturing method to produce blades 20.

[0067] According to an embodiment of the present invention, in a second aspect, a method for designing a centripetal impeller is also provided, wherein the centripetal impeller is as described above, comprising the steps of: Step S1: Analyze the shape of the blade 20, determine the dimensions of the damping structure 30, and establish a model of the centripetal impeller; Step S2: Calculate the static strength, high-cycle fatigue strength, and natural frequency of the centripetal impeller model, and compare the calculation results with the evaluation criteria. If the requirements are met, the designed centripetal impeller is deemed usable. If the requirements are not met, the damping structure 30 is adjusted until the centripetal impeller model meets the evaluation criteria.

[0068] Specifically, in step S1, the shape of the blade 20 of the centripetal impeller to be tuned is analyzed. Based on the shape and direction of the blade 20, a damping structure 30 with a similar shape to the blade 20 is designed in the blade 20. According to the size and shape constraints of the blade 20, the thickness of the damping structure 30 and the distance from the edge 22 and root 21 of the blade 20 are determined. The metal sandwich vibration reduction structure design and model establishment of the entire blade 20 are carried out.

[0069] Furthermore, such as Figure 6 As shown, the adjustment of the damping structure 30 in step S2 includes the following steps: Step S21: Calculate the amplitude of the blade 20 in the above model. According to the amplitude contour line 23, adjust the damping structure 30 to be strip-shaped and make the damping structure 30 pass perpendicularly through the amplitude contour line 23 to complete the establishment of the centripetal impeller model of the strip-shaped damping structure 30. Step S22: Perform static strength, high-cycle fatigue strength and natural frequency calculations on the radial impeller model of the belt damping structure 30, and compare the calculation results with the evaluation criteria; if the requirements are met, it proves that the designed radial impeller is usable; if the requirements are not met, proceed to step S23. Step S23: Readjust the dimensional parameters of the strip damping structure 30 until the centripetal impeller model of the strip damping structure 30 meets the evaluation criteria.

[0070] Although embodiments of the invention have been described in conjunction with the accompanying drawings, those skilled in the art can make various modifications and variations without departing from the spirit and scope of the invention, and such modifications and variations all fall within the scope defined by the appended claims.

Claims

1. A centripetal impeller, characterized in that, include: Impeller disk (10); The blade (20) has a blade root (21) connected to the impeller disk (10). Multiple blades (20) are provided, and the multiple blades (20) are spaced apart circumferentially along the impeller disk (10). The blade (20) has a sandwich space inside. The damping structure (30) is embedded in the interlayer space. The distance between the damping structure (30) and the edge (22) of the blade (20) is L1, which satisfies 0 < L1 < 2 mm. The blade height of the blade (20) is H. The distance between the damping structure (30) and the blade root (21) is L2, which satisfies 0.05H ≤ L2 ≤ 0.3H.

2. The radial impeller according to claim 1, characterized in that, The static strength reserve coefficient at the leaf root (21) is n 0.2 , satisfying n 0.2 =σ 0.2 (T) / σ max , where σ 0.2 (T) represents the yield strength of the material at the operating temperature, σ max This represents the maximum equivalent stress at that location; When n 0.2 When L < 1.2, L2 = 0.3H; When n 0.2 When L > 1.5, L2 = 0.05H; When 1.2≤n 0.2 When ≤1.5, L2 = 0.3H - (n 0.2 -1.2)×(5 / 6)H.

3. The centripetal impeller according to claim 1, characterized in that, The centripetal impeller has an inlet (41) and an outlet (42). The blade (20) has a blade root (21) from the inlet (41) to the outlet (42) as a percentage of the total blade root (21) length, x. The distance L2 between the damping structure (30) and the blade root (21) is a percentage of the blade height H of the blade (20), satisfying the following expression: y = (3.2323x) 4 -6.4647x 3 +3.5264x 2 -0.294x+0.102)×100%; And 8%≤y≤25%.

4. The centripetal impeller according to claim 1, characterized in that, The blade (20) has an amplitude contour line (23), and multiple amplitude contour lines (23) are formed. The damping structure (30) intersects with each amplitude contour line (23). The extension direction of the damping structure (30) at each intersection point is perpendicular to the extension direction of the amplitude contour line (23) it intersects with.

5. The radial impeller according to claim 4, characterized in that, The damping structure (30) includes a plurality of damping units (31), with adjacent damping units (31) spaced apart. Each damping unit (31) has an intersection point with each amplitude contour line (23). The extension direction of the damping unit (31) at each intersection point is perpendicular to the extension direction of the amplitude contour line (23) it intersects with.

6. The radial impeller according to claim 5, characterized in that, The length of the amplitude contour line (23) is A, and the sum of the widths of the damping units (31) along the extension direction of the amplitude contour line (23) is B, satisfying 0.4A≤B≤A.

7. The radial impeller according to claim 4, characterized in that, The edge (22) of the blade (20) includes a leading edge (221), and the damping structure (30) has a top edge (32) and a bottom edge (33). The top edge (32) is arranged parallel to the leading edge (221), and the bottom edge (33) is arranged parallel to the amplitude contour line (23) at its location.

8. The radial impeller according to any one of claims 1-7, characterized in that, The thickness of the blade (20) is W1, and the thickness of the damping structure (30) is W2, satisfying 0.3W1≤W2≤0.5W2; and / or, The operating temperature of the blade (20) is T; when 0 < T < 500℃, the damping structure (30) is a titanium-based alloy damping structure (30) or a cobalt-nickel-based alloy damping structure (30); when 700℃ < T < 800℃, the damping structure (30) is an oxygen-doped MPEA alloy damping structure (30).

9. A method for designing a centripetal impeller, wherein the centripetal impeller is the centripetal impeller according to any one of claims 1-8, characterized in that, Including the following steps: Step S1: Analyze the shape of the blade (20), determine the size of the damping structure (30), and establish a model of the centripetal impeller; Step S2: Calculate the static strength, high-cycle fatigue strength and natural frequency of the centripetal impeller model, and compare the calculation results with the evaluation criteria. If the requirements are met, the designed centripetal impeller is proven to be usable. If the requirements are not met, the damping structure (30) is adjusted until the centripetal impeller model meets the evaluation criteria.

10. The design method of the centripetal impeller according to claim 9, characterized in that, The adjustment of the damping structure (30) in step S2 includes the following steps: Step S21: Calculate the amplitude of the blade (20) in the above model. According to the amplitude contour line (23) diagram, adjust the damping structure (30) to be strip-shaped and make the damping structure (30) pass vertically through the amplitude contour line (23) to complete the establishment of the centripetal impeller model of the strip-shaped damping structure (30). Step S22: Calculate the static strength, high-cycle fatigue strength and natural frequency of the radial impeller model of the strip damping structure (30), and compare the calculation results with the evaluation criteria; if the requirements are met, it proves that the designed radial impeller is usable; if the requirements are not met, proceed to step S23. Step S23: Readjust the dimensional parameters of the strip damping structure (30) until the centripetal impeller model of the strip damping structure (30) meets the evaluation criteria.

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