A Dynamic Design Method for Gas Generator Rotor Structure

Abstract

This application discloses a dynamic design method for a gas generator rotor structure, belonging to the field of aero-engine technology. The method includes: finite element modeling of the gas generator rotor system; modification of the finite element model based on the end-tooth connection structure of the gas generator rotor system, followed by calculations of the rotor critical speed, bending mode shape, and strain energy distribution; design of the end-tooth radial height based on the calculation results of the bending mode shape and the strain energy distribution of the end-tooth connection structure; design of the assembly structure of the first-stage rotor journal, second-stage rotor disk, third-stage rotor disk, and central tie rod based on the strain energy distribution calculation results of the rotor system's support bearings and the first-stage rotor disk; and design of the radial height of the central tie rod and the clearance between the central tie rod and the power turbine shaft based on the distribution calculation results of the dual-rotor coupled mode shapes. This application proposes criteria for support layout design, central tie rod clearance design, end-tooth radial height design, and strain energy distribution design, providing a basis for rotor structure dynamic design.

Description

Technical Field

[0001] This application relates to the field of aero-engine technology, and in particular, to a method for dynamic design of a gas generator rotor structure. Background Technology

[0002] The gas generator rotor is a core component of a turboshaft engine, and the stability of its structural dynamics design directly affects engine performance, vibration, and lifespan. Due to its operating environment of high temperature, high pressure, and high speed, the rotor system's structural dynamics design places extremely high demands on it and is one of the main problems restricting further engine development. Current gas generator rotor system structural dynamics designs mainly rely on experience from previous models, failing to systematically propose design ideas and rationale, and suffer from problems such as unreasonable connection structure design and failure to consider support strain energy.

[0003] Existing gas generator rotor systems (see reference) Figure 1 The dynamic design is mainly based on the calculation of critical speed and unbalanced response using finite element software. During the calculation process, the connecting structures such as end teeth are treated as a whole. The dynamic stability of the rotor system is evaluated by means of critical margin, rotor response, etc. However, it does not provide strong guidance for the dynamic optimization of the rotor system.

[0004] Therefore, the dynamic design of gas generator rotor structure under existing traditional methods is difficult to guarantee its stability, which can easily lead to problems such as excessive vibration. Moreover, it does not indicate the optimization direction, and requires repeated structural modifications for trial and error, which is not conducive to improving design efficiency and optimizing design quality. Summary of the Invention

[0005] In view of at least one of the above technical problems, this application provides a dynamic design method for a gas generator rotor structure, which proposes criteria such as support layout design, central tie rod clearance design, end tooth radial height design, and strain energy distribution design, and provides a basis for rotor structure dynamic design through rotor dynamic simulation analysis and experimental verification.

[0006] According to one aspect of this application, a method for dynamic design of a gas generator rotor structure is provided, comprising the following steps:

[0007] S100: Finite element modeling of gas generator rotor system;

[0008] S200: The finite element model is modified according to the end tooth connection structure of the gas generator rotor system, and then the rotor critical speed, bending mode shape and strain energy distribution are calculated.

[0009] S300: Based on the calculation results of the bending vibration mode and the strain energy distribution of the end tooth connection structure, the radial height of the end teeth is designed; based on the calculation results of the strain energy distribution of the rotor system's support bearings and the rotor's first-stage disk, the assembly structure of the rotor's first-stage disk journal, rotor's second-stage disk, rotor's third-stage disk, and central tie rod is designed; based on the distribution calculation results of the dual-rotor coupled vibration modes, the radial height of the central tie rod and the clearance between the central tie rod and the power turbine shaft are designed.

[0010] In some embodiments of this application, in step S200, when correcting the finite element model based on the end tooth connection structure of the gas generator rotor system, the end tooth connection position is modeled separately. The end tooth thickness is set to h, and the end tooth stiffness loss correction zone is set to 3h. The formula for calculating the elastic modulus of the connection between the two parts within the end tooth stiffness loss correction zone is as follows:

[0011] ;

[0012] In the formula, This represents the corrected elastic modulus; This represents the original elastic modulus of the connection structure; This is a correction factor; For contact stiffness loss; This is due to assembly stiffness loss; This represents the loss of bending stiffness.

[0013] In some embodiments of this application, contact stiffness loss The calculation formula is as follows:

[0014] ;

[0015] In the formula, This represents the area of ​​the contact surface of the end teeth. Let be the cross-sectional area of ​​the power turbine shaft. The angle between the end tooth contact surface and the center line of the power turbine shaft;

[0016] Assembly stiffness loss The calculation formula is as follows:

[0017] ;

[0018] In the formula, The average contact stress of the pre-set end tooth contact surface; The average contact stress on the end tooth contact surface during operation;

[0019] Bending stiffness loss The calculation formula is as follows:

[0020] ;

[0021] In the formula, The angle between the position of the lower tooth under bending deformation and its tangent.

[0022] In some embodiments of this application, during step S200, when calculating the rotor critical speed, bending mode shape, and strain energy distribution, the strain energy of different structural positions in the rotor system is distinguished, and the calculation formula is as follows:

[0023] ;

[0024] In the formula, Indicates the proportion of strain energy; This indicates structures in different positions, including blade disks, central tie rods, end tooth connection structures, and pivot bearings; The strain energy of each structure under different bending modes; This represents the strain energy of the entire rotor system under this bending mode.

[0025] In some embodiments of this application, in steps S200 and S300, when calculating the rotor critical speed and bending mode shape, the critical speed margin for the bending mode shape should satisfy the following formula:

[0026] ;

[0027] In the formula, This indicates the critical rotational speed corresponding to the bending mode. This indicates the maximum operating speed of the rotor; if the requirement is not met, the radial height of the end teeth is increased until the formula is satisfied.

[0028] In some embodiments of this application, in step S300, based on the strain energy distribution calculation results of the pivot bearing and the first-stage rotor disk of the rotor system, the length of the front journal of the first-stage rotor disk is increased, and the front pivot bearing of the rotor is supported on the outer periphery of the front journal of the first-stage rotor disk. The central tie rod is connected to the inner hole of the first-stage rotor disk through a thread, and the connection position is located on the side of the front pivot bearing facing the second-stage rotor disk. The second-stage rotor disk and the third-stage rotor disk are designed as an integrated machine, and the end tooth connection structure between the second and third-stage disks is eliminated.

[0029] In some embodiments of this application, step S300 further includes optimizing the clamping member at the front end of the central tie rod and the assembly structure of the centrifugal impeller and the gas turbine: an adapter plate is provided between the clamping member and the centrifugal impeller, the clamping member is threadedly connected to the central tie rod and, after being tightened, abuts against the rear end of the centrifugal impeller through the adapter plate, the clamping member is used to lock the central tie rod and the centrifugal impeller; the rear journal of the centrifugal impeller is connected to the front journal of the gas turbine through end teeth, and the centrifugal impeller and the gas turbine are clamped together by the central tie rod.

[0030] In some embodiments of this application, in step S300, the conditions for designing the radial height of the central tie rod and the clearance between the central tie rod and the power turbine shaft based on the distribution calculation results of the dual-rotor coupled vibration modes of the gas generator rotor and the power turbine rotor include: S301, when the power turbine rotor passes the second critical speed, it is necessary to avoid the midpoint of the power turbine shaft rubbing against the central tie rod; S302, in the power turbine blade shedding test, when the power turbine rotor overclocks past the third critical speed, the clearance between the balance boss of the power turbine shaft and the central tie rod should be designed to be greater than the sum of the maximum value of the third critical response at the balance boss at the rear section of the power turbine shaft under the maximum unbalance and the calculated maximum value of the response of the central tie rod under the allowable maximum unbalance.

[0031] In some embodiments of this application, when both the central tie rod and the power turbine shaft are hollow shaft structures with uniform axial thickness, the gap value is selected from the calculation obtained when the power turbine rotor over-rotates past the third critical level in the power turbine blade shedding test in S302.

[0032] In some embodiments of this application, when the central tie rod or the power turbine shaft is a hollow shaft with different axial thicknesses, the gaps calculated twice, in S301 when the power turbine rotor passes the second critical speed and in S320 when the power turbine rotor overshoots the third critical speed, are combined to ensure that each segment of the power turbine shaft does not rub against the central tie rod.

[0033] This application has the following beneficial effects:

[0034] This application presents a dynamic design method for a gas generator rotor structure. First, a finite element model of the gas generator rotor system is created, and simulation calculations are easily performed using existing finite element software. Then, the finite element model is modified based on the end-tooth connection structure of the gas generator rotor system, fully considering the stiffness variations of the end teeth and other connection structures. Next, the rotor's critical speed, bending mode shape, and strain energy distribution are calculated based on the modified finite element model. Then, the radial height of the end teeth is designed based on the calculated bending mode shape and strain energy distribution of the end-tooth connection structure. No design criteria are proposed for the radial height of the end-tooth connection structure. Based on the calculated strain energy distribution of the rotor system's pivot bearings and the rotor's first-stage disk, the assembly structure of the rotor's first-stage disk journal, second-stage disk, third-stage disk, and central tie rod is designed, and an optimized assembly structure for the rotor pivot and rotor disk-journal is proposed. Based on the calculated distribution of the dual-rotor coupled mode shapes, the radial height of the central tie rod and the clearance between the central tie rod and the power turbine shaft are designed, and requirements are proposed for the clearance design between the central tie rod and the power turbine shaft. Through the aforementioned optimization designs, the stability of the gas generator rotor structure's dynamic design is effectively guaranteed, avoiding problems such as excessive vibration. Furthermore, it eliminates the need for repeated structural modifications for trial and error, clearly defining the direction for rotor system dynamic optimization. This application, through rotor dynamics simulation analysis and experimental verification, provides a basis for rotor structure dynamics design.

[0035] Of course, any product implementing this application does not necessarily need to achieve all the advantages described above simultaneously. In addition to the purposes, features, and advantages described above, this application also has other purposes, features, and advantages. The following will provide a more detailed description of this application with reference to figures. Attached Figure Description

[0036] The accompanying drawings, which form part of this application, are used to provide a further understanding of this application. The illustrative embodiments and descriptions of this application are used to explain this application and do not constitute an undue limitation of this application. In the drawings:

[0037] Figure 1 This is a schematic diagram of a conventional gas generator rotor system.

[0038] Figure 2 This is a flowchart illustrating the design method of a preferred embodiment of this application;

[0039] Figure 3 This is a schematic diagram of the end tooth stiffness loss assessment according to a preferred embodiment of this application;

[0040] Figure 4 This is a schematic diagram showing the location of the end tooth stiffness loss correction in a preferred embodiment of this application;

[0041] Figure 5This is a schematic diagram of the strain energy under the bending vibration mode of the combustion rotor according to a preferred embodiment of this application;

[0042] Figure 6 This is a schematic diagram of the optimized assembly relationship at the end teeth according to a preferred embodiment of this application;

[0043] Figure 7 This is a schematic diagram of the optimized assembly relationship at the rotor front support point according to a preferred embodiment of this application;

[0044] Figure 8 This is a schematic diagram of the structure of the dual rotor according to a preferred embodiment of this application;

[0045] Figure 9 This is a schematic diagram showing the angle between the position of the lower tooth of the bending deformation and its tangent in a preferred embodiment of this application.

[0046] Figure 10 This is a schematic diagram of the second-order critical response mode of a power turbine according to a preferred embodiment of this application;

[0047] Figure 11 This is a schematic diagram of the third-order critical response mode of the power turbine according to a preferred embodiment of this application.

[0048] Legend: 1000, end tooth connection structure; 2000, end tooth stiffness loss correction zone; 3000, end tooth with maximum strain energy; 100, rotor system; 101, front support bearing; 102, rotor first stage disk; 1021, rotor first stage disk front journal; 103, central tie rod; 104, centrifugal impeller; 105, gas turbine; 106, clamping component; 107, transfer plate. Detailed Implementation

[0049] The following description provides specific application scenarios and requirements for this specification, intended to enable those skilled in the art to make and use the contents of this specification. Various partial modifications to the disclosed embodiments will be apparent to those skilled in the art, and the general principles defined herein can be applied to other embodiments and applications without departing from the spirit and scope of this specification. Therefore, this specification is not limited to the embodiments shown, but rather to the widest scope consistent with the claims.

[0050] Figure 1 This is a schematic diagram of a conventional gas generator rotor system. Figure 2 This is a flowchart illustrating the design method of a preferred embodiment of this application; Figure 3 This is a schematic diagram of the end tooth stiffness loss assessment according to a preferred embodiment of this application; Figure 4 This is a schematic diagram showing the location of the end tooth stiffness loss correction in a preferred embodiment of this application; Figure 5 This is a schematic diagram of the strain energy under the bending vibration mode of the combustion rotor according to a preferred embodiment of this application; Figure 6 This is a schematic diagram of the optimized assembly relationship at the end teeth according to a preferred embodiment of this application; Figure 7 This is a schematic diagram of the optimized assembly relationship at the rotor front support point according to a preferred embodiment of this application; Figure 8 This is a schematic diagram of the structure of the dual rotor according to a preferred embodiment of this application; Figure 9 This is a schematic diagram showing the angle between the position of the lower tooth of the bending deformation and its tangent in a preferred embodiment of this application. Figure 10 This is a schematic diagram of the second-order critical response mode of a power turbine according to a preferred embodiment of this application; Figure 11 This is a schematic diagram of the third-order critical response mode of the power turbine according to a preferred embodiment of this application.

[0051] A method for dynamic design of a gas generator rotor structure includes the following steps:

[0052] S100: Perform finite element modeling of the gas generator rotor system 100;

[0053] S200: Based on the end tooth connection structure 1000 of the gas generator rotor system 100, the finite element model is modified, and then the rotor critical speed, bending mode shape and strain energy distribution are calculated.

[0054] S300: Based on the bending vibration mode and strain energy distribution calculation results of the end tooth connection structure 1000, the radial height of the end teeth is designed; based on the strain energy distribution calculation results of the pivot bearing of the rotor system 100 and the rotor first-stage disk 102, the assembly structure of the rotor first-stage disk front journal 1021, rotor second-stage disk, rotor third-stage disk and central tie rod 103 is designed; based on the distribution calculation results of the dual rotor coupled vibration mode, the radial height of the central tie rod 103 and the clearance between the central tie rod 103 and the power turbine shaft are designed.

[0055] The term "dual-rotor coupled vibration mode" here refers to the structure connected to the coupled vibration of the gas generator rotor and the power turbine rotor. Please refer to [the relevant documentation / reference]. Figure 8 As shown, the dual-rotor configuration of the gas generator rotor and the power turbine rotor is a mature existing configuration, and will not be elaborated further here.

[0056] It should be noted that the existing rotor structure dynamics design method has the following shortcomings: the stiffness variation of the end teeth and other connecting structures is not considered in the finite element simulation calculation; no design criteria are proposed for the radial height of the 1000-degree end tooth connecting structure; no requirements are proposed for the clearance design between the central tie rod and the power turbine shaft; and no optimization method is proposed for the assembly of the rotor support and the rotor disk-journal.

[0057] The dynamic design method for the gas generator rotor structure in this application first involves finite element modeling of the gas generator rotor system 100, which is then easily simulated using existing finite element software. Subsequently, the finite element model is modified based on the end tooth connection structure 1000 of the gas generator rotor system 100, fully considering the stiffness variations of the end teeth and other connection structures. Then, the rotor critical speed, bending mode shape, and strain energy distribution are calculated based on the modified finite element model. Finally, the radial height of the end teeth can be designed based on the calculation results of the bending mode shape and the strain energy distribution of the end tooth connection structure 1000. The paper does not propose design criteria for the radial height of the end-tooth connection structure 1000. Based on the strain energy distribution calculation results of the pivot bearings and rotor primary disk 102 of the rotor system 100, the assembly structure of the rotor primary disk front journal 1021, rotor secondary disk, rotor tertiary disk, and central tie rod 103 is designed, and an optimized assembly structure for the rotor pivot and rotor disk-journal is proposed. Based on the distribution calculation results of the dual-rotor coupled vibration modes, the radial height of the central tie rod 103 and the clearance between the central tie rod 103 and the power turbine shaft are designed, and requirements are proposed for the clearance design between the central tie rod 103 and the power turbine shaft. Through the above-mentioned multiple optimization designs, the stability of the gas generator rotor structure dynamics design is effectively guaranteed, avoiding problems such as excessive vibration, and eliminating the need for repeated structural modifications for trial and error and improvement. The dynamic optimization direction of the rotor system 100 is clearly proposed. This application has been verified by rotor dynamics simulation analysis and experiments, providing a basis for rotor structure dynamics design.

[0058] Preferably, in step S200, when correcting the finite element model based on the end tooth connection structure 1000 of the gas generator rotor system 100, the end tooth connection position is modeled separately. The end tooth thickness is set to h, and the end tooth stiffness loss correction zone 2000 is set to 3h. The formula for calculating the elastic modulus of the connection between the two parts within the end tooth stiffness loss correction zone 2000 is as follows:

[0059] ;

[0060] In the formula, This represents the corrected elastic modulus; This represents the original elastic modulus of the connection structure; This is a correction factor; For contact stiffness loss; This is due to assembly stiffness loss; For bending stiffness loss; in some embodiments, the final correction factor of the end tooth connection structure 1000 is preset to a certain value. If the value is between 0.6 and 0.9, the radial height of the end teeth can be optimized to improve and reduce the correction factor. .

[0061] Understandably, the optimization design approach of this application involves evaluating the component body and the connecting end teeth separately. The strain energy proportion of the component body can be higher, while the strain energy proportion of the modified area structure must be very low. This application increases the radial height of the end teeth, which actually improves the correction coefficient. Changes in radial height can improve the end teeth , and Compared to the traditional method of directly thickening the end teeth, the rotor weight increase in this application is relatively small. However, increasing the radial height will lead to a larger end tooth linear velocity and a larger stress at the end tooth root. This can be optimized by rounding the end tooth root to ensure the stress level. In some embodiments, the rounding of the end tooth root can be optimized by moving from the original single arc to multiple arcs, and then to cubic spline curves to make the rounding transition smoother and gradually improve the stress concentration phenomenon. (Because in conventional designs, the rounding is the tangent circle of two straight lines. Generally, only the radius of the tangent circle is changed, or multiple arcs are used to adjust the chamfer design. However, adjusting the radius will change the position of the tangent point. At the same time, if the stress at the tangent point is large (because the transition from the straight line to the arc at the tangent point is not smooth enough), multiple arcs are also difficult to handle. Therefore, this application makes the arc of the rounding smoother by changing the arc of the rounding to a spline curve.)

[0062] It should be noted that, compared with existing methods that find the location of maximum strain, this application focuses on the end tooth region (a 3h-width correction region, in which the strain energy is calculated by replacing the original end tooth element with an integrated element with a corrected elastic modulus in the finite element model). Here, the proportion of end tooth strain energy will provide different thresholds depending on the location of the end tooth (different temperatures) (for example, the proportion of end tooth strain energy in the first half of the compressor is no more than 10%, and the proportion of end tooth strain energy in the centrifugal impeller is no more than 5%, and the specific value will be adjusted according to the temperature field, etc.). If the proportion is greater than the threshold, optimization is required; if it is less than the threshold, optimization is not necessary.

[0063] Preferably, contact stiffness loss The calculation formula is as follows:

[0064] ;

[0065] In the formula, This represents the area of ​​the contact surface of the end teeth. Let be the cross-sectional area of ​​the power turbine shaft. The angle between the end tooth contact surface and the center line of the power turbine shaft;

[0066] Assembly stiffness loss The calculation formula is as follows:

[0067] ;

[0068] In the formula, The average contact stress of the pre-set end tooth contact surface; The average contact stress on the end tooth contact surface during operation;

[0069] Bending stiffness loss The calculation formula is as follows:

[0070] ;

[0071] In the formula, To determine the angle between the position of the lower tooth during bending deformation and its tangent, please refer to... Figure 11 As shown.

[0072] Understandably, the correction method in this application is relatively simple, and the calculation formula for the correction coefficient is precise and straightforward, making it convenient for engineering designers to quickly carry out calculations. Combined with the formula that the critical speed margin should satisfy, a design criterion is proposed for the radial height of the end tooth connection structure 1000.

[0073] Preferably, in step S200, when calculating the rotor critical speed, bending mode shape, and strain energy distribution, the strain energy of different structural positions in the rotor system 100 is distinguished, and the calculation formula is as follows:

[0074] ;

[0075] In the formula, Indicates the proportion of strain energy; This indicates structures in different positions, including blade disk, central tie rod 103, end tooth connection structure 1000, and pivot bearing; The strain energy of each structure under different bending modes; Let 100 be the strain energy of the entire rotor system under this bending mode.

[0076] It is understandable that Formula 5 can be used to easily and efficiently calculate the rotor critical speed, bending mode and strain energy distribution. In the calculation, the structures at different positions of the rotor system 100, mainly including the blade disk, the central tie rod 103, the end tooth connection structure 1000 and the support bearing, can be used to calculate the strain energy and its proportion, providing a basis for subsequent key structural optimization.

[0077] Preferably, in steps S200 and S300, when calculating the rotor critical speed and bending mode shape, the critical speed margin for the bending mode shape should satisfy the following formula:

[0078] ;

[0079] In the formula, This indicates the critical rotational speed corresponding to the bending mode. This indicates the maximum operating speed of the rotor; if the requirement is not met, the radial height of the end teeth is increased until the formula is satisfied.

[0080] Please refer to Figure 5 Table 1 shows the ratio of strain energy before and after end tooth optimization:

[0081] Table 1. Proportion of strain energy before and after end tooth optimization.

[0082] ;

[0083] It should be noted that increasing the radial height of the end teeth will increase the linear velocity of the end teeth. In order to reduce stress and improve service life, this can be solved by rounding the root of the end teeth.

[0084] Preferably, please refer to Figure 7 As shown, in step S300, based on the strain energy distribution calculation results of the pivot bearing and the first-stage rotor disk 102 of the rotor system 100, the length of the front journal 1021 of the first-stage rotor disk is increased, and the front pivot bearing 101 of the rotor is supported on the outer periphery of the front journal 1021 of the first-stage rotor disk. The central tie rod 103 is connected to the inner hole of the first-stage rotor disk 102 through threads, and the connection position is located on the side of the front pivot bearing 101 facing the second-stage rotor disk. The second-stage rotor disk and the third-stage rotor disk are designed as an integrated machine, and the end tooth connection structure 1000 between the second and third-stage disks is eliminated.

[0085] It should be noted that in the existing design, the first-stage rotor disc 102 is connected to the central tie rod 103, and the rotor front support bearing 101 rests on the central tie rod 103. This can easily lead to incoordination between the first-stage rotor disc 102 and the central tie rod 103 under bending deformation, resulting in insufficient stability of the connection structure. This application addresses this by lengthening the front journal 1021 of the first-stage rotor disc, with the rotor front support bearing 101 resting on the front journal 1021, and the central tie rod 103 connected to the first-stage rotor disc 102 via threads. This reduces the connection structure, enhances the stability of the assembly relationship under bending deformation, and improves the dynamic stability of the rotor's front support. Simultaneously, the second and third-stage rotor discs of the gas generator adopt an integrated machining design, eliminating the end-tooth connection structure 1000 between the second and third-stage discs. This also reduces the connection structure, improves the overall bending rigidity of the rotor, and enhances the rotor's robustness. In addition, the threaded section at the front end of the central tie rod 103 is behind the front support bearing 101. The preload applied to the central tie rod 103 will not affect the front journal 1021 of the rotor first stage disc, and will not cause the front journal 1021 of the rotor first stage disc to bend, thus not affecting the stability of the front support.

[0086] Preferably, please refer to Figure 6As shown, step S300 further includes optimizing the assembly structure of the clamping member 106 at the front end of the central tie rod 103 and the centrifugal impeller 104 and gas turbine 105: a transition plate 107 is provided between the clamping member 106 and the centrifugal impeller 104. The clamping member 106 is threadedly connected to the central tie rod 103 and, after being tightened, abuts against the rear end of the centrifugal impeller 104 through the transition plate 107. The clamping member 106 is used to lock the central tie rod 103 and the centrifugal impeller 104. The rear journal of the centrifugal impeller 104 is connected to the front journal of the gas turbine 105 through end teeth, and the centrifugal impeller 104 and the gas turbine 105 are pressed together by the central tie rod 103.

[0087] Optionally, the clamping component 106 adopts a clamping nut, which serves as a pre-tightening agent and uses a mature part that is easy to disassemble and adjust using conventional tools.

[0088] It should be noted that the existing clamping nut is connected to the centrifugal impeller disk / gas turbine disk via the adapter plate 107. The rear journal of the centrifugal impeller, the front journal of the gas turbine, and the adapter plate 107 are connected by bolts. The radial height of the connection structure is lower and it is closer to the gas turbine 105 end, resulting in higher temperatures, a harsh working environment, and concentrated strain energy. In contrast, the clamping component 106 of this application is directly clamped onto the centrifugal impeller 104 via the adapter plate 107. The rear journal of the centrifugal impeller 104 is connected to the front journal of the gas turbine 105 via end teeth and is pre-tightened by the central tie rod 103. The structure is simpler and more compact, and it can avoid possible rubbing between the centrifugal impeller 104 and the central tie rod 103 under heavy loads. Furthermore, by directly clamping the clamping component 106 onto the centrifugal impeller disk via the adapter plate 107, if the strain energy proportion of the rear end teeth of the centrifugal impeller is high, the radial height of the rear end teeth of the centrifugal impeller, i.e., the end teeth with the maximum strain energy 3000, can be adjusted without being limited by the clamping component 106 and the adapter plate 107.

[0089] Preferably, in step S300, the conditions for designing the radial height of the central tie rod 103 and the clearance between the central tie rod 103 and the power turbine shaft based on the distribution calculation results of the dual-rotor coupled vibration modes of the gas generator rotor and the power turbine rotor include: S301, when the power turbine rotor passes the second critical speed, it is necessary to avoid the midpoint of the power turbine shaft from rubbing against the central tie rod 103; S302, in the power turbine blade shedding test, when the power turbine rotor overclocks past the third critical speed, the clearance between the balance boss of the power turbine shaft and the central tie rod 103 should be designed to be greater than the sum of the maximum value of the third critical response at the balance boss at the rear section of the power turbine shaft under the maximum unbalance and the calculated maximum value of the response of the central tie rod 103 under the allowable maximum unbalance.

[0090] Understandably, it is necessary to combine the coupled vibration analysis of the gas generator rotor and the power turbine rotor, and consider the bending vibration mode of the power turbine rotor over-rotating past the third critical speed, to optimize the gap between the power turbine balance boss and the central tie rod 103 of the gas generator rotor, and to fully consider the rotor gap design ideas and methods for special working conditions.

[0091] In this preferred embodiment, when both the central tie rod 103 and the power turbine shaft are hollow shafts with uniform axial thickness, the clearance value calculated during the power turbine rotor's over-rotation past the third critical speed in the power turbine blade shedding test in S302 is selected. When the central tie rod 103 or the power turbine shaft is a hollow shaft with different axial thicknesses, the clearance calculated twice—in S301 when the power turbine rotor passes the second critical speed and in S320 when the power turbine rotor over-rotates past the third critical speed—is combined to ensure that each segment of the power turbine shaft does not rub against the central tie rod 103.

[0092] Please refer to Figure 10 When the power turbine rotor passes the second critical speed, the maximum allowable unbalance of the rotor is applied when calculating the rotor response. This yields the second critical response mode shape of the power turbine and its minimum clearance with the first critical response of the ignition / generator rotor, and the minimum clearance is not negative. Please refer to [reference needed]. Figure 11 When the power turbine rotor overclocks past the third critical speed, twenty times the maximum allowable unbalance of the rotor is applied when calculating the rotor response to obtain the third critical response mode of the power turbine and its minimum gap with the second critical response of the ignition rotor. The minimum gap is not negative.

[0093] Understandably, when the power turbine rotor passes the second-order critical speed (the power turbine's operating speed is between the second and third-order critical speeds), it approaches the first-order critical speed of the gas generator rotor. Therefore, it's necessary to avoid rubbing between the midpoint of the power turbine shaft and the central tie rod 103 of the gas generator. At this point, the rotor imbalance is relatively small. When calculating the rotor response, the maximum allowable rotor imbalance is applied, and the bending mode is as follows: Figure 10 As shown. In the power turbine blade shedding test, when the power turbine rotor overclocks past the third critical speed, the gas generator rotor speed continuously decreases from its maximum speed (passing through the second critical speed). This can lead to rubbing between the rear section of the power turbine shaft and the central tie rod 103 of the gas generator. Severe rubbing can cause the central tie rod 103 to break. At this point, the rotor imbalance is relatively large, far exceeding the maximum allowable rotor imbalance applied in the previous rotor response calculation. The maximum imbalance in this step of the response calculation increases by 20 times, and the bending mode is as follows: Figure 11 As shown.

[0094] Therefore, when considering the clearance between the power turbine shaft and the central tie rod 103 of the combustion rotor, if both the central tie rod 103 and the power turbine shaft are hollow shafts with uniform axial thickness, the clearance calculated in the second calculation (i.e., the clearance value calculated when the power turbine rotor overclocks past the third critical speed) should be used as the standard; if the central tie rod 103 or the power turbine shaft can be machined into hollow shafts with different axial thicknesses, the clearances calculated in the two calculations should be combined to ensure that each segment in the axial direction meets the clearance requirements.

[0095] In summary, this application presents a dynamic design method for the turbine engine gas generator rotor structure. It proposes criteria for support layout design, central tie rod clearance design, end tooth radial height design, and strain energy distribution design. These criteria have been validated through rotor dynamics simulation analysis and experimental verification, providing a basis for rotor structure dynamics design. This application proposes the overall approach to the dynamic design of the gas generator rotor structure; it presents an analysis method and optimization path for rotor strain energy distribution; it proposes an analysis method and optimization path for the stiffness loss of the end tooth connection structure; it proposes structural forms and methods for optimizing rotor assembly relationships; and it proposes a design approach and method for rotor clearance considering special operating conditions. Using this method, the dynamic design of a certain type of engine gas generator rotor structure was completed, and it passed the maximum unbalance airworthiness test, the turbine blade shedding containment test, and the vibration compliance analysis.

[0096] It should be understood that in the foregoing description of the embodiments in this specification, various features are combined in a single embodiment, drawing, or description for the purpose of simplifying the description and aiding in the understanding of a feature. However, this does not mean that the combination of these features is necessary, and those skilled in the art may readily identify some of the devices as separate embodiments when reading this specification. That is, the embodiments in this specification can also be understood as an integration of multiple secondary embodiments. It is also valid when each secondary embodiment contains fewer than all the features of a single foregoing disclosed embodiment.

[0097] Finally, it should be understood that the embodiments disclosed herein are illustrative of the principles of the embodiments described in this specification. Other modified embodiments are also within the scope of this specification. Therefore, the embodiments disclosed in this specification are merely examples and not limitations. Those skilled in the art can implement the applications described in this specification by adopting alternative configurations based on the embodiments in this specification. Therefore, the embodiments in this specification are not limited to the embodiments precisely described in the application. For those skilled in the art, several improvements, modifications, or changes can be made without departing from the principles of this application, and the above technical features can also be combined in an appropriate manner; these improvements, modifications, changes, or combinations, or the direct application of the inventive concept and technical solution to other occasions without modification, should all be considered as protection of this application.

Claims

1. A method for dynamic design of a gas generator rotor structure, characterized in that, It includes the following steps: S100: Finite element modeling of the rotor system (100) of the gas generator; S200: The finite element model is modified based on the end tooth connection structure of the rotor system (100) of the gas generator, and then the critical speed, bending mode shape and strain energy distribution of the rotor are calculated. When modifying the finite element model based on the end tooth connection structure of the rotor system (100) of the gas generator, the end tooth connection position is modeled separately. The end tooth thickness is set to h, and the modification area is 3h. The formula for calculating the elastic modulus of the connection part of the two parts in the modification area is as follows: 1 In the formula, This represents the corrected elastic modulus; This represents the original elastic modulus of the connection structure; This is a correction factor; For contact stiffness loss; This is due to assembly stiffness loss; This is for bending stiffness loss; Contact stiffness loss The calculation formula is as follows: 2 In the formula, This represents the area of ​​the contact surface between the teeth. Let be the cross-sectional area of ​​the power turbine shaft. The angle between the end tooth contact surface and the center line of the power turbine shaft; Assembly stiffness loss The calculation formula is as follows: 3 In the formula, The average contact stress of the pre-set end tooth contact surface; The average contact stress on the end tooth contact surface during operation; Bending stiffness loss The calculation formula is as follows: 4 In the formula, The angle between the position of the lower tooth under bending deformation and its tangent; S300: Based on the calculation results of the bending vibration mode and the strain energy distribution of the end tooth connection structure, the radial height of the end tooth is designed; based on the calculation results of the strain energy distribution of the pivot bearing of the rotor system (100) and the first-stage rotor disk (102), the assembly structure of the first-stage rotor disk front journal (1021), the second-stage rotor disk, the third-stage rotor disk and the central tie rod (103) is designed; based on the distribution calculation results of the dual rotor coupled vibration mode, the radial height of the central tie rod (103) and the clearance between the central tie rod (103) and the power turbine shaft are designed.

2. The method for dynamic design of a gas generator rotor structure according to claim 1, characterized in that, In step S200, when calculating the rotor critical speed, bending mode shape, and strain energy distribution, the strain energy of different structural positions in the rotor system (100) is distinguished, and the calculation formula is as follows: 5 In the formula, Indicates the proportion of strain energy; The different structural positions are indicated, including the blade disk, the central tie rod (103), the end tooth connection structure, and the pivot bearing; The strain energy of each structure under different bending modes; Let be the strain energy of the entire rotor system (100) under this bending mode.

3. The method for dynamic design of a gas generator rotor structure according to claim 1, characterized in that, In steps S200 and S300, when calculating the rotor critical speed and bending mode shape, the critical speed margin for the bending mode shape should satisfy the following formula: 6 In the formula, This indicates the critical rotational speed corresponding to the bending mode. This indicates the maximum operating speed of the rotor; if the requirement is not met, the radial height of the end teeth is increased until Formula 6 is satisfied.

4. The method for dynamic design of a gas generator rotor structure according to claim 1, characterized in that, In step S300, based on the strain energy distribution calculation results of the pivot bearing and the first-stage rotor disk (102) of the rotor system (100), the length of the front journal (1021) of the first-stage rotor disk is increased, and the front pivot bearing (101) of the rotor is supported on the outer periphery of the front journal (1021) of the first-stage rotor disk. The central tie rod (103) is connected to the inner hole of the first-stage rotor disk (102) by thread, and the connection position is located on the side of the front pivot bearing (101) facing the second-stage rotor disk. The second-stage rotor disk and the third-stage rotor disk are designed as an integrated machine, and the end tooth connection structure between the second and third-stage disks is eliminated.

5. The method for dynamic design of a gas generator rotor structure according to claim 4, characterized in that, Step S300 also includes optimizing the assembly structure of the clamping member (106) at the front of the central tie rod (103) and the centrifugal impeller (104) and the gas turbine (105): a transition plate (107) is provided between the clamping member (106) and the centrifugal impeller (104). The clamping member (106) is threadedly connected to the central tie rod (103) and, after being tightened, abuts against the rear end of the centrifugal impeller (104) through the transition plate (107). The clamping member (106) is used to lock the central tie rod (103) and the centrifugal impeller (104). The rear journal of the centrifugal impeller (104) is connected to the front journal of the gas turbine (105) through end teeth, and the centrifugal impeller (104) and the gas turbine (105) are clamped through the central tie rod (103).

6. The method for dynamic design of a gas generator rotor structure according to claim 1, characterized in that, In step S300, based on the distribution calculation results of the dual rotor coupled vibration modes of the gas generator rotor and the power turbine rotor, the conditions for designing the radial height of the central tie rod (103) and the clearance between the central tie rod (103) and the power turbine shaft include: S301, when the power turbine rotor passes the second critical speed, it is necessary to avoid the midpoint of the power turbine shaft from rubbing against the central tie rod (103); S302, in the power turbine blade shedding test, when the power turbine rotor overclocks past the third critical speed, the clearance between the balance boss of the power turbine shaft and the central tie rod (103) should be designed to be greater than the sum of the maximum value of the third critical response at the balance boss at the rear section of the power turbine shaft under the maximum unbalance and the calculated maximum value of the response of the central tie rod (103) under the allowable maximum unbalance.

7. The method for dynamic design of a gas generator rotor structure according to claim 6, characterized in that, When both the central tie rod (103) and the power turbine shaft are hollow shaft structures with uniform axial thickness, the gap value is obtained by calculation when the power turbine rotor over-rotates past the third critical stage in the power turbine blade shedding test in S302.

8. The method for dynamic design of a gas generator rotor structure according to claim 6, characterized in that, When the central tie rod (103) or the power turbine shaft is a hollow shaft with different thicknesses in the axial direction, the gap calculated twice in S301 when the power turbine rotor passes the second critical speed and in S320 when the power turbine rotor over-rotates past the third critical speed is combined to ensure that each segment of the power turbine shaft in the axial direction does not rub against the central tie rod (103).

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