Centripetal impeller with curved trailing edge blades and design method of centripetal impeller

Abstract

The invention relates to the technical field of centripetal impellers, and discloses a centripetal impeller with curved trailing edge blades and a design method of the centripetal impeller. A centripetal impeller with curved trailing edge blades comprises a disc body, a blade body and a blade body, the multiple blades are evenly distributed on the disc body in the circumferential direction, each blade comprises a front edge, a blade tip, a tail edge and a blade root which are sequentially arranged, the blade roots are connected with the disc body, and each tail edge comprises a first linear part connected with the corresponding blade tip and a first curve part connected with the corresponding blade root so as to form a curve type tail edge. The problems that an existing large-expansion-ratio and high-load centripetal impeller is large in size and weight, high in blade root stress, high in flow loss and low in working efficiency, and the service life is affected are solved.

Description

Technical Field

[0001] The present invention relates to the technical field of centripetal impellers, and in particular to a centripetal impeller with curved trailing edge blades and a design method thereof. Background Art

[0002] To improve power, the expansion ratio of centrifugal impellers used in aircraft gas turbines must be increased. This significantly increases the inlet and outlet areas of the impeller compared to conventional low-expansion-ratio turbines, necessitating a longer blade chord length to ensure uniform airflow acceleration. Furthermore, due to the large expansion ratio, the volumetric flow rate at the outlet is significantly greater than at the inlet, increasing the outlet flow area of ​​the impeller. This increases the blade height at the impeller outlet, with the ratio exceeding 4.0 to the inlet blade height. This directly increases the size and weight of the impeller, impacting the compactness of the aircraft engine structure and the improvement of the power-to-weight ratio. Existing centrifugal impellers typically use a linear root-to-tip distribution along the trailing edge. This increased size and weight can lead to increased steady-state stress at the blade root, reducing the blade's ability to resist upstream excitation forces. In certain circumstances, this can also cause the blade and disk to vibrate in conjunction with the excitation force, leading to fatigue cracks in high-stress areas of the blade, such as the trailing edge of the blade root, and impacting engine safety. Moreover, due to the large blade outlet height and long axial chord length of the blade tip, the friction area between the blade tip and the wheel cover is large, and the radial pressure gradient at the root tip is high. Under the conditions of large expansion ratio and high load, there is a secondary flow from the blade tip to the blade root at the trailing edge of the impeller. This separated flow will mix with the vortex in the end channel at the blade root, resulting in an increase in the flow loss at the impeller outlet. The larger the expansion ratio, the higher the flow loss. Summary of the Invention

[0003] In view of this, the present invention provides a centripetal impeller with curved trailing edge blades and a design method thereof to solve the problems of existing large expansion ratio and high load centripetal impellers such as large size and weight, high stress at the blade root, affecting service life, strong flow loss and low efficiency.

[0004] In a first aspect, the present invention provides a centripetal impeller having curved trailing edge blades, comprising:

[0005] disc body;

[0006] Multiple blades are evenly distributed on the disk body along the circumferential direction. The blades include a leading edge, a blade tip, a trailing edge and a blade root arranged in sequence. The blade root is connected to the disk body. The trailing edge includes a first linear portion connected to the blade tip and a first curved portion connected to the blade root to form a curved trailing edge.

[0007] The centripetal impeller with curved trailing edge blades provided by the present invention has the same axial chord length for the blades above the intersection of the first linear portion and the first curved portion. The axial chord length of the blades below the intersection gradually increases as the radius decreases. This allows the concave surface of the first curved portion of the trailing edge to face the center of the disk, thereby reducing the size of the upper and middle portions of the blades and minimizing the overall weight of the blades, thereby reducing the steady-state stress at the blade root. Simultaneously, the increased axial chord length of the blades below the intersection increases the area at the blade root, improving blade stiffness and significantly increasing their resistance to unsteady aerodynamic excitation. This avoids fatigue cracks at the high stress point at the trailing edge blade root, thereby increasing the strength and life margin of the centripetal impeller. On the other hand, due to the shortening of the axial chord length of the blade at the tip, the tip scraping vortex can be broken down and reduced in advance, which increases the mixing distance between the gap leakage flow and the channel vortex, making the flow at the blade outlet smoother. The appropriate lengthening of the chord length of the blade root ensures the area and flow organization ability of the lower part of the blade, eliminates the secondary flow from the high radius to the low radius at the trailing edge of the impeller under high load, and its mixing between the blade root and the end channel vortex, thereby reducing the flow loss at the impeller outlet.

[0008] In an optional embodiment, the intersection of the first linear portion and the first curved portion is located at 30%-40% of the blade height.

[0009] The specific location of the intersection can be determined based on structural strength and aerodynamic performance simulation results. This configuration minimizes the chord length of the upper blade portion, thereby reducing the overall blade weight. It also ensures the area and flow organization capabilities of the lower blade portion, resulting in smoother flow at the blade root and increased strength reserves at the blade root.

[0010] In an optional embodiment, the ratio of the axial chord length at the blade root to the axial chord length at the blade tip is 1.1-1.3.

[0011] The adoption of an axial chord ratio of 1.1 to 1.3 between the blade root and the blade tip is mainly to ensure that the aerodynamic load on the blade tip does not increase too much, resulting in an increase in tip leakage loss, while minimizing the axial chord length of the blade tip and thus reducing the size and weight of the impeller blades. It is the result of a comprehensive balance between tip leakage loss, size, and weight.

[0012] In an optional embodiment, the trailing edge thickness gradually increases from the blade tip to the blade root, and the ratio of the trailing edge thickness at the blade root to the trailing edge thickness at the blade tip is 2.5-3.5.

[0013] The variable trailing edge thickness distribution with large blade root thickness and small blade tip thickness is adopted mainly to improve the strength reserve of the trailing edge blade root, while minimizing the wake loss, thereby taking into account the aerodynamic performance and structural strength of the impeller.

[0014] In an optional embodiment, the thickness of the leading edge gradually increases from the blade tip to the blade root or remains unchanged, and the ratio of the leading edge thickness at the blade tip to the leading edge thickness at the blade root is 0.75-1.0.

[0015] The leading edge adopts a thickness design with a larger blade root and a smaller blade tip, mainly to simplify the design and save processing costs.

[0016] In an optional embodiment, the blade tip includes a second curved portion connected to the leading edge and a second linear portion connected to the trailing edge, the second linear portion is parallel to the engine axis, and the angle between the first linear portion and the second linear portion is 90 degrees, and / or

[0017] The ratio of the projected length of the second linear portion to the blade tip in the axial direction parallel to the disk body is 0.9-1.0.

[0018] The angle between the first and second linear sections is 90°. This ensures that the blade tip outlet is parallel to the engine axis, maintaining a straight blade tip outlet. This facilitates tip clearance control and reduces tip leakage during engine operation. The projected length ratio of 0.9-1.0 minimizes excessive change in tip curvature from blade inlet to outlet, resulting in more uniform airflow.

[0019] In an optional embodiment, the angle between the tangent line at the intersection of the first curved portion of the trailing edge and the blade root and the axis of the disk body is 20°-35°, and / or

[0020] The connection between the blade root and the disc body is provided with a fillet, and the radius of the fillet gradually decreases from the leading edge to the trailing edge.

[0021] The purpose of setting the fillet is to enhance the thickness at the blade root. A variable fillet radius design is adopted from the blade inlet to the trailing edge outlet. This is mainly due to the high tangential velocity and high strength stress at the blade inlet. A larger fillet radius can enhance the blade rigidity. However, since the radius at the blade outlet is smaller, the distance between adjacent blades is much smaller than that at the inlet. A smaller fillet radius can avoid interference between adjacent blade fillets and increase the airflow area at the blade root outlet, making the flow smoother.

[0022] In a second aspect, the present invention further provides a method for designing a centripetal impeller having curved trailing edge blades, comprising the following steps:

[0023] Based on the design principles of low-consistency impellers, the number of impeller blades and the average axial chord length were initially selected, while maintaining the flow path geometry and rotor speed. A linear trailing edge distribution impeller blade scheme was generated, and a calculation model was constructed to determine the number of blades and throat area based on the efficiency, power, and inlet flow of the centripetal impeller.

[0024] On the basis of ensuring that the number of impeller blades and throat area remain unchanged, by adjusting the end wall angle and spline curve control points, changing the blade modeling parameters, constructing the first linear part and the first curved part of the trailing edge, thus generating a curved trailing edge centripetal impeller blade scheme;

[0025] Calculate the efficiency, power and inlet flow of the centrifugal impeller in each state. When the first preset requirements are met, perform structural modeling and design of the centrifugal impeller disc and blades.

[0026] Analyze the strength and vibration of the centripetal impeller. If the strength and vibration calculation results meet the second preset requirements, the design is completed. If not, readjust the blade shaping parameters until the second preset requirements are met.

[0027] The design method of the centrifugal impeller of the present invention can effectively reduce the secondary flow from the blade tip to the blade root caused by the trailing edge of the original impeller, thereby reducing the vortex intensity of the end channel of the blade root, reducing the flow loss at the impeller outlet, and promoting the improvement of turbine efficiency. At the same time, the improved centrifugal impeller strengthens the blade root, and the maximum equivalent stress in the root area of ​​the centrifugal impeller blade trailing edge of the centrifugal impeller blade is reduced by 54.6%. Under this steady-state stress, the vibration stress resistance of the trailing edge root is increased by 33.3%. The maximum stress position is adjusted from the trailing edge root to the blade tip, avoiding the problem of the steady-state high stress range and the vibration high stress range overlapping, and reducing the probability of the trailing edge crack failure common in centrifugal impellers.

[0028] In an optional embodiment, the specific steps of determining the number of blades and the throat area are:

[0029] According to the design principle of low-consistency impeller and vibration requirements, the number of blades is selected, and the average axial chord length is obtained according to the impeller consistency coefficient formula, and then the axial chord length of the blade tip section and the axial chord length of the blade root section are obtained to determine the linear distribution of the trailing edge.

[0030] Where, the impeller consistency coefficient γ=Z b L ave / D1, axial chord length of blade tip section Axial chord length of blade root section L t =εL h ,

[0031] The value of γ is 4.0-4.5, Z b is the number of impeller blades, D1 is the impeller inlet diameter, L ave is the average axial chord length, and ε is the root-apex meridian chord length ratio, which is 1.1-1.3.

[0032] In an optional embodiment, the specific steps of constructing the first linear portion and the first curved portion of the trailing edge are:

[0033] The linear distribution of the trailing edge is parameterized using a fourth-order cubic B-spline curve. By adjusting the position of the control points of the B-spline curve, the relative position of the trailing edge profile is determined so that the distribution of the blade trailing edge perpendicular to the disk axis conforms to the following relationship:

[0034]

[0035] Where L is the relative chord length, defined as the local axial chord length L local and the blade tip axial chord length L t The ratio of the leaf height to the trailing edge height is h. local Ratio to blade outlet height H.

[0036] The impeller blade shaping parameters are adjusted to appropriately reduce the exit geometry angles of the blade root and blade tip sections, while increasing the exit geometry angles at the middle of the blade. This reduces the degree to which the exit airflow angles at the blade root and blade tip deviate from the axial direction, increases the channel width at the blade root and blade tip, and simultaneously increases the aerodynamic load in the mid-blade area. This design, on the one hand, maintains the impeller outlet throat area value unchanged during the trailing edge profile adjustment process, and on the other hand, avoids the problems of increased root outlet density, slow airflow acceleration, end-zone boundary layer accumulation, and large end-zone flow losses caused by an increase in the blade root chord length. Furthermore, the airflow acceleration capacity of the mid-blade area is fully utilized, reducing the load at the blade tip, which is also conducive to suppressing gap leakage losses. BRIEF DESCRIPTION OF THE DRAWINGS

[0037] In order to more clearly illustrate the specific embodiments of the present invention or the technical solutions in the prior art, the following briefly introduces the drawings required for use in the specific embodiments or the description of the prior art. Obviously, the drawings described below are some embodiments of the present invention. For ordinary technicians in this field, other drawings can be obtained based on these drawings without paying any creative work.

[0038] Figure 1 Schematic diagram of a centripetal impeller according to an embodiment of the present invention;

[0039] Figure 2 for Figure 1 sectional view of

[0040] Figure 3 is a schematic diagram of a blade according to an embodiment of the present invention;

[0041] Figure 4 is a schematic diagram of the blade from another angle;

[0042] Figure 5 Schematic diagram comparing the trailing edge of a blade according to an embodiment of the present invention and a conventional blade;

[0043] Figure 6 This is a schematic diagram of the flow at the trailing edge of a blade according to an embodiment of the present invention;

[0044] Figure 7 Schematic diagram of the flow at the trailing edge of an existing blade.

[0045] Description of reference numerals:

[0046] 1. Plate body;

[0047] 2. Blade; 201. Leading edge; 202. Blade tip; 203. Trailing edge; 2031. First linear portion; 2032. First curved portion; 204. Blade root;

[0048] 3. Scallop-shaped notch;

[0049] 4. Weight reduction hole;

[0050] 5. Rounding;

[0051] 6. Middle arc;

[0052] 7. Suction surface;

[0053] 8. Pressure surface. DETAILED DESCRIPTION

[0054] To make the purpose, technical solutions, and advantages of the embodiments of the present invention more clear, the technical solutions in the embodiments of the present invention will be clearly and completely described below in conjunction with the accompanying drawings in the embodiments of the present invention. Obviously, the described embodiments are part of the embodiments of the present invention, not all of the embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without making creative efforts shall fall within the scope of protection of the present invention.

[0055] The following combination Figures 1 to 6 , describing embodiments of the present invention.

[0056] According to an embodiment of the present invention, a centripetal impeller having curved trailing edge blades is provided. Figures 1 to 4 As shown, it is applicable to centripetal turbines with a single-stage expansion ratio of 5.0 or above, including:

[0057] Plate 1;

[0058] Multiple blades 2 are evenly distributed on the disk body along the circumferential direction. The blades include a leading edge 201, a blade tip 202, a trailing edge 203 and a blade root 204 arranged in sequence. The blade root is connected to the disk body, and the trailing edge includes a first linear portion 2031 connected to the blade tip and a first curved portion 2032 connected to the blade root to form a curved trailing edge.

[0059] The disc body and blades are formed by integral precision casting. The disc body is a solid structure. At the impeller inlet, scallop-shaped notches 3 are evenly arranged on the disc body between two adjacent blades. The number of scallop-shaped notches is the same as the number of blades. The width of the scallop-shaped notches can be determined based on a comprehensive evaluation of strength, weight and turbine performance. You can either cut off only part of the disc body between the two blades, or cut off all of the disc body except the blades. The main function of the scallop-shaped notch is to reduce the weight of the impeller and reduce the stress level of the disc body. Specifically, different shapes can be set according to the structural strength, feasibility of the processing technology and the need to suppress wheel back leakage. In addition, a circular weight-reducing hole 4 is arranged at the axis of the disc body outlet. The center of the weight-reducing hole coincides with the engine axis. Figure 2 As shown, to ensure the rigidity of the disk, the radius of the weight-reducing hole R3 is not greater than 0.7R 2h The ratio of the weight-reducing hole depth to R3 is between 0.7 and 0.9. The specific result can be determined according to the structural strength and vibration assessment results, so as to reduce the impeller weight as much as possible while ensuring that the wall thickness of the impeller outlet disc meets the strength and vibration requirements.

[0060] The point where the lower end of the impeller blade connects to the impeller disc is defined as the blade root, and the flat surface at the upper end of the impeller blade is defined as the blade tip. The leading edge, located near the impeller inlet, is generally rectangular, while the trailing edge, located near the impeller outlet, is curved—a combination of partially straight and partially curved sections, with the first curved section concave toward the disc axis. The leading edge, blade tip, trailing edge, and blade root are connected end to end to form a closed, nearly fan-shaped structure.

[0061] The centripetal impeller with curved trailing edge blades provided by the present invention has the same axial chord length for the blades above the intersection of the first linear portion and the first curved portion. The axial chord length of the blades below the intersection gradually increases as the radius decreases. This allows the concave surface of the first curved portion of the trailing edge to face the center of the disk, thereby reducing the size of the upper and middle portions of the blades and minimizing the overall weight of the blades, thereby reducing the steady-state stress at the blade root. Simultaneously, the increased axial chord length of the blades below the intersection increases the area at the blade root, improving blade stiffness and significantly increasing their resistance to unsteady aerodynamic excitation. This avoids fatigue cracks at the high stress point at the trailing edge blade root, thereby increasing the strength and life margin of the centripetal impeller. On the other hand, due to the shortening of the axial chord length of the blade at the tip, the tip scraping vortex can be broken down and reduced in advance, which increases the mixing distance between the gap leakage flow and the channel vortex, making the flow at the blade outlet smoother. The appropriate lengthening of the chord length of the blade root ensures the area and flow organization ability of the lower part of the blade, eliminates the secondary flow from the high radius to the low radius at the trailing edge of the impeller under high load, and its mixing between the blade root and the end channel vortex, thereby reducing the flow loss at the impeller outlet.

[0062] In one embodiment, the intersection of the first linear portion and the first curved portion is located at 30%-40% of the blade height.

[0063] That is, the ratio of the projected length of the first linear portion in the axial direction perpendicular to the disk body to the projected length of the trailing edge in this direction is 0.6-0.7, preferably 0.6, that is, the intersection of the first linear portion and the first curved portion is located at 40% of the blade height.

[0064] like Figure 3 As shown, the blade tip adopts a flat tip configuration at the outlet, namely, the blade tip includes a portion parallel to the disk axial direction. One endpoint of this portion is P1, and the other endpoint, namely the intersection of the blade tip and the trailing edge, is P2. The intersection of the first linear portion and the first curved portion of the trailing edge is P3, and the intersection of the trailing edge and the blade root is P4. The projected length of the first linear portion along the axis perpendicular to the disk is the length between P2 and P3, and the projected length of the trailing edge along the axis perpendicular to the disk is H. In this embodiment, P3 is located near the middle and lower portion of the blade, and the specific length can be determined based on the structural strength and aerodynamic performance simulation results.

[0065] On the one hand, the above setting reduces the chord length of the upper part of the blade to the greatest extent, thereby reducing the overall weight of the blade as much as possible, and reducing the friction area between the blade tip and the wheel cover, reducing the scraping vortex intensity; at the same time, it ensures the area and flow organization ability of the lower part of the blade, making the flow at the root of the blade smoother, while improving the strength reserve of the root of the blade.

[0066] In one embodiment, a ratio of the axial chord length of the blade root portion to the axial chord length of the blade tip portion is 1.1-1.3.

[0067] That is, the ratio of the projected length of the blade root in the axial direction parallel to the disk body to the projected length of the blade tip in the same direction is 1.1-1.3.

[0068] like Figure 3 As shown, the projected length of the blade root in the axial direction parallel to the disk body is L h The projected length of the blade tip along the axis parallel to the disk is L t .

[0069] The adoption of an axial chord ratio of 1.1 to 1.3 between the blade root and the blade tip is mainly to ensure that the aerodynamic load on the blade tip does not increase too much, resulting in an increase in tip leakage loss, while minimizing the axial chord length of the blade tip and thus reducing the size and weight of the impeller blades. It is the result of a comprehensive balance between tip leakage loss, size, and weight.

[0070] In one embodiment, the trailing edge thickness gradually increases from the blade tip to the blade root, and a ratio of the trailing edge thickness at the blade root to the trailing edge thickness at the blade tip is 2.5-3.5.

[0071] The thickness of the trailing edge follows a generally parabolic distribution from tip to root, i.e., a variable trailing edge thickness distribution is adopted along the blade height direction R1. The specific thickness distribution can be determined based on the results of strength and vibration calculations. The variable trailing edge thickness distribution, with greater thickness at the root and less thickness at the tip, is primarily intended to improve the strength reserve at the trailing edge blade root while minimizing wake losses, thereby balancing the impeller's aerodynamic performance and structural strength.

[0072] In one embodiment, the thickness of the leading edge gradually increases or remains constant from the blade tip to the blade root, and the ratio of the leading edge thickness at the blade tip to the leading edge thickness at the blade root is 0.75-1.0.

[0073] The leading edge of the blade is roughly rectangular, and a constant thickness design can be adopted along the blade height direction. The specific thickness can be determined according to the strength and vibration calculation results. The leading edge adopts a thickness design with a larger blade root and a smaller blade tip, mainly to simplify the design and save processing costs.

[0074] In one embodiment, the blade tip includes a second curved portion connected to the leading edge and a second linear portion connected to the trailing edge, the second linear portion is parallel to the engine axis, and the angle between the first linear portion and the second linear portion is 90 degrees, and / or

[0075] The ratio of the projected length of the second linear portion to the blade tip in the axial direction parallel to the disk body is 0.9-1.0.

[0076] like Figure 3 As shown, the angle between the first linear portion P2P3 and the second linear portion P1P2 is 90°, meaning that the second linear portion P1P2 is parallel to the engine axis. This primarily aims to create a structure at the blade tip outlet that is parallel to the engine axis, maintaining a straight tip outlet. This facilitates tip clearance control and reduces tip clearance leakage during engine operation. The projected length ratio of 0.9-1.0 is designed to minimize excessive change in tip curvature from blade inlet to outlet, thereby ensuring more uniform airflow. The specific ratio can be determined based on turbine performance calculations.

[0077] In one embodiment, the angle between the tangent line at the intersection of the first curved portion of the trailing edge and the blade root and the axis of the disk body is 20°-35°, and / or

[0078] A fillet 5 is provided at the connection position between the blade root and the disc body, and the radius of the fillet gradually decreases from the leading edge to the trailing edge.

[0079] like Figure 5 As shown in , the angle between the tangent line at the intersection of the trailing edge and the blade root and the axis of the disc is β, and the specific value can be determined based on the calculation results of aerodynamic performance and strength performance. Figure 4As shown in the figure, the fillet is set at the connection position between the blade root and the disk body in order to enhance the thickness of the blade root. A variable fillet radius design is adopted from the blade inlet to the trailing edge outlet. In the impeller inlet area, the fillet radius is 3mm, and near the trailing edge of the impeller outlet, the fillet radius is 2mm. This is mainly due to the high tangential velocity and strength stress at the blade inlet. A larger fillet radius can enhance the blade rigidity. Since the radius at the blade outlet is small, the distance between adjacent blades is much smaller than that at the inlet. A smaller fillet radius can avoid interference between adjacent blade fillets and increase the airflow area at the blade root outlet, making the flow smoother.

[0080] According to an embodiment of the present invention, on the other hand, a design method for a centripetal impeller with curved trailing edge blades is provided, wherein the blade profile cross section is constructed in the manner of a median camber line 6 + thickness distribution, wherein the blade profile thickness of the suction surface 7 and the pressure surface 8 are equal, such as Figure 4 As shown, the following steps are included:

[0081] Based on the design principles of low-consistency impellers, the number of impeller blades and the average axial chord length were initially selected, while maintaining the flow path geometry and rotor speed. A linear trailing edge distribution impeller blade scheme was generated, and a calculation model was constructed to determine the number of blades and throat area based on the efficiency, power, and inlet flow of the centripetal impeller.

[0082] On the basis of ensuring that the number of impeller blades and throat area remain unchanged, by adjusting the end wall angle and spline curve control points, changing the blade modeling parameters, constructing the first linear part and the first curved part of the trailing edge, thus generating a curved trailing edge centripetal impeller blade scheme;

[0083] Calculate the efficiency, power and inlet flow of the centrifugal impeller in each state. When the first preset requirements are met, perform structural modeling and design of the centrifugal impeller disc and blades.

[0084] Analyze the strength and vibration of the centripetal impeller. If the strength and vibration calculation results meet the second preset requirements, the design is completed. If not, readjust the blade shaping parameters until the second preset requirements are met.

[0085] like Figure 6 and Figure 7As shown, the design method of the centrifugal impeller of the present invention can effectively reduce the secondary flow from the blade tip to the blade root caused by the trailing edge of the original impeller, thereby reducing the vortex intensity in the end channel of the blade root area, reducing the flow loss at the impeller outlet, and promoting the improvement of turbine efficiency. At the same time, the improved centrifugal impeller strengthens the blade root, and the maximum equivalent stress in the root area of ​​the centrifugal impeller blade trailing edge of the centrifugal impeller blade is reduced by 54.6%. Under this steady-state stress, the vibration stress resistance of the trailing edge root is increased by 33.3%. The maximum stress position is adjusted from the trailing edge root to the blade tip, avoiding the problem of the steady-state high stress range and the vibration high stress range overlapping, and reducing the probability of the trailing edge crack failure common in centrifugal impellers.

[0086] In one embodiment, the specific steps for determining the number of blades and the throat area are:

[0087] According to the design principle of low-consistency impeller and vibration requirements, the number of blades is selected, and the average axial chord length is obtained according to the impeller consistency coefficient formula, and then the axial chord length of the blade tip section and the axial chord length of the blade root section are obtained to determine the linear distribution of the trailing edge.

[0088] Where, the impeller consistency coefficient γ=Z b L ave / D1, axial chord length of blade tip section Axial chord length of blade root section L t =εL h ,

[0089] The value of γ is 4.0-4.5, Z b is the number of impeller blades, D1 is the impeller inlet diameter, L ave is the average axial chord length, and ε is the root-apex meridian chord length ratio, which is 1.1-1.3.

[0090] According to the design principle of low-viscosity impeller, under the premise of keeping the flow channel geometry and rotor speed unchanged, the number of impeller blades and the average axial chord length are preliminarily selected so that the impeller consistency coefficient is between 4.0 and 4.5, which is 20%-30% lower than the current level. The preferred impeller consistency coefficient is 4.2. According to the performance and vibration requirements, the number of blades is usually between 12 and 14. In this embodiment, the preferred number of blades is 12. The average chord length L can be determined based on the relationship between the number of impeller blades, the impeller inlet radius (determined in the one-dimensional solution) and the consistency coefficient. ave ε is preferably 1.16. After the axial chord length of the blade tip section and the axial chord length of the blade root section are determined, the intersection points P2 and P4 of the trailing edge profile with the blade tip and blade root can be determined. Connecting points P2 and P4 with a line segment can determine the trailing edge profile distribution, such as Figure 5 shown.

[0091] A calculation model was built, and through three-dimensional CFD simulation analysis, the final number of impeller blades and throat area were determined based on whether the performance parameters such as turbine efficiency, power and inlet flow rate met the required values.

[0092] In one embodiment, the specific steps of constructing the first linear portion and the first curved portion of the trailing edge are:

[0093] The linear distribution of the trailing edge is parameterized using a fourth-order cubic B-spline curve. By adjusting the position of the control points of the B-spline curve, the relative position of the trailing edge profile is determined so that the distribution of the blade trailing edge perpendicular to the disk axis conforms to the following relationship:

[0094]

[0095] Where L is the relative chord length, defined as the local axial chord length L local and the blade tip axial chord length L t The ratio of the leaf height to the trailing edge height is h. local Ratio to blade outlet height H.

[0096] On the basis of ensuring that the number of impeller blades and throat area remain unchanged, with the goal of minimizing the axial chord length of the section above point P3 on the trailing edge and appropriately increasing the axial chord length of the blade root section, by adjusting the end wall angle and spline curve control points, and changing the impeller blade shaping parameters, a curved trailing edge centripetal impeller blade scheme is constructed, and the performance of this scheme is evaluated at both design points and non-design points. The B-spline curve parameterization method can be found in relevant public literature and will not be repeated here. From the above formula, it can be seen that when h = 1, the relative chord length of the blade tip is 1; when h = 0, the relative chord length of the blade root is 1.16.

[0097] The impeller blade shaping parameters are adjusted to appropriately reduce the exit geometry angles of the blade root and blade tip sections, while increasing the exit geometry angles at the middle of the blade. This reduces the degree to which the exit airflow angles at the blade root and blade tip deviate from the axial direction, increases the channel width at the blade root and blade tip, and simultaneously increases the aerodynamic load in the mid-blade area. This design, on the one hand, maintains the impeller outlet throat area value unchanged during the trailing edge profile adjustment process, and on the other hand, avoids the problems of increased root outlet density, slow airflow acceleration, end-zone boundary layer accumulation, and large end-zone flow losses caused by an increase in the blade root chord length. Furthermore, the airflow acceleration capacity of the mid-blade area is fully utilized, reducing the load at the blade tip, which is also conducive to suppressing gap leakage losses.

[0098] Although the embodiments of the present invention have been described with reference to the accompanying drawings, those skilled in the art may make various modifications and variations without departing from the spirit and scope of the present invention. Such modifications and variations are all within the scope defined by the appended claims.

Claims

1. A centripetal impeller with curved trailing edge blades, characterized in that: include: Plate (1); A plurality of blades (2) are uniformly distributed on the disk body (1) along the circumferential direction, wherein the blades (2) include a leading edge (201), a blade tip (202), a trailing edge (203) and a blade root (204) arranged in sequence, wherein the blade root (204) is connected to the disk body (1), and the trailing edge (203) includes a first linear portion (2031) connected to the blade tip (202) and a first curved portion (2032) connected to the blade root (204), so as to form a curved trailing edge.

2. The centripetal impeller with curved trailing edge blades according to claim 1, characterized in that: The intersection of the first linear portion (2031) and the first curved portion (2032) is located at 30%-40% of the blade height.

3. The centripetal impeller with curved trailing edge blades according to claim 2, characterized in that: The ratio of the axial chord length at the blade root (204) to the axial chord length at the blade tip (202) is 1.1-1.

3.

4. The centripetal impeller with curved trailing edge blades according to claim 1, characterized in that: The thickness of the trailing edge (203) gradually increases from the blade tip (202) to the blade root (204), and the ratio of the thickness of the trailing edge (203) at the blade root (204) to the thickness of the trailing edge (203) at the blade tip (202) is 2.5-3.

5.

5. The centripetal impeller with curved trailing edge blades according to claim 4, characterized in that: The thickness of the leading edge (201) gradually increases or remains unchanged from the blade tip (202) to the blade root (204), and the ratio of the thickness of the leading edge (201) at the blade tip (202) to the thickness of the leading edge (201) at the blade root (204) is 0.75-1.

0.

6. The centripetal impeller with curved trailing edge blades according to any one of claims 1 to 5, characterized in that: The blade tip (202) comprises a second curved portion connected to the leading edge (201) and a second linear portion connected to the trailing edge (203), the second linear portion is parallel to the engine axis, the angle between the first linear portion (2031) and the second linear portion is 90 degrees, and / or The ratio of the projected length of the second linear portion to the blade tip (202) in the axial direction parallel to the disk body (1) is 0.9-1.

0.

7. The centripetal impeller with curved trailing edge blades according to claim 6, characterized in that: The angle between the tangent line at the intersection of the first curved portion (2032) of the trailing edge (203) and the blade root (204) and the axis of the disk (1) is 20°-35°, and / or A chamfer (5) is provided at the connection position between the blade root (204) and the disk body (1), and the radius of the chamfer (5) gradually decreases from the leading edge (201) to the trailing edge (203).

8. A method for designing a centripetal impeller with curved trailing edge blades according to any one of claims 1 to 7, characterized in that: The following steps are involved: Based on the design principles of low-consistency impellers, the number of impeller blades and the average axial chord length were initially selected, while maintaining the flow path geometry and rotor speed. A linear trailing edge distribution impeller blade scheme was generated, and a calculation model was constructed to determine the number of blades and throat area based on the efficiency, power, and inlet flow of the centripetal impeller. On the basis of ensuring that the number of impeller blades and throat area remain unchanged, the blade modeling parameters are changed by adjusting the end wall angle and the spline curve control points, and the first linear part and the first curved part of the trailing edge are constructed, thereby generating a curved trailing edge centripetal impeller blade scheme; Calculate the efficiency, power and inlet flow of the centrifugal impeller in each state. When the first preset requirements are met, perform structural modeling and design of the centrifugal impeller disc and blades. Analyze the strength and vibration of the centripetal impeller. If the strength and vibration calculation results meet the second preset requirements, the design is completed. If not, readjust the blade shaping parameters until the second preset requirements are met.

9. The design method of a centripetal impeller with curved trailing edge blades according to claim 8, characterized in that: The specific steps to determine the number of blades and throat area are: According to the design principle of low-consistency impeller and vibration requirements, the number of blades is selected, and the average axial chord length is obtained according to the impeller consistency coefficient formula, and then the axial chord length of the blade tip section and the axial chord length of the blade root section are obtained to determine the linear distribution of the trailing edge. Where, the impeller consistency coefficient γ=Z b L ave / D1, axial chord length of blade tip section Axial chord length of blade root section L t =εL h , The value of γ is 4.0-4.5, Z b is the number of impeller blades, D1 is the impeller inlet diameter, L ave is the average axial chord length, and ε is the root-apex meridian chord length ratio, which is 1.1-1.

3.

10. The design method of a centripetal impeller with curved trailing edge blades according to claim 9, characterized in that: The specific steps of constructing the first linear portion and the first curved portion of the trailing edge are: The fourth-order cubic B-spline curve is used to parameterize the linear distribution of the trailing edge. By adjusting the position of the control point of the B-spline curve, the relative position of the trailing edge line is determined so that the distribution of the blade trailing edge perpendicular to the disk axis conforms to the following relationship: Where L is the relative chord length, defined as the local axial chord length L local and the blade tip axial chord length L t The ratio of h is the relative leaf height, defined as the local trailing edge height H local Ratio to blade outlet height H.

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