A method for designing the structure of a power turbine rotor

Abstract

This application discloses a design method for a power turbine rotor structure, belonging to the field of aero-engine technology. The method includes: arranging oil film dampers based on strain energy distribution and mode shape; optimizing the power turbine rotor structure based on the radial deformation of the connection interface; designing the segmented wall thickness of the power turbine shaft; selecting a balancing boss on the outer wall at the point of maximum displacement in each mode shape of the power turbine shaft; and designing the clearance between the central tie rod of the power turbine rotor and the power turbine disk based on the maximum vibration displacement of the power turbine disk. This application achieves good vibration reduction of the rotor by arranging oil film dampers in appropriate positions; performing deformation coordination calculations at the connection interface to optimize the connection structure and prevent damage to the connection interface during long-term operation; designing and optimizing the segmented shaft wall thickness; optimizing the critical speed margin; and considering the influence of the ultimate unbalance on the clearance to achieve good dynamic balance characteristics.

Description

Technical Field

[0001] This application relates to the field of aero-engine technology, and in particular, to a method for designing a power turbine rotor structure. Background Technology

[0002] In the field of aero-engine technology, a power turbine rotor typically consists of four bearings, two turbine disks, and a long shaft. The long shaft is hollow to reduce rotor mass and to allow for intermediate ventilation.

[0003] Currently, when designing the dynamics of multi-support long-span flexible power turbine rotors, the design experience of other similar models is generally used to lay out the oil film damper. Deformation analysis of the connection interface is usually not carried out. The long shaft generally adopts a fixed inner and outer diameter design, and the balance bosses on the long shaft are simply and uniformly distributed.

[0004] Existing power turbine rotor design methods have several drawbacks. Damping design does not consider the impact of whole-machine detachment and containment tests. Inconsistent deformation of some structures can easily lead to changes in unbalance, resulting in large rotor vibrations. It is impossible to optimize the critical speed margin and meet the 20% design margin requirement. Poor placement of the balancing boss makes it difficult to achieve dynamic balancing of multiple vibration modes of high-speed rotors. The impact of the critical speed that the detachment and containment test may reach on the boss is not considered, nor is the impact of the ultimate unbalance on the clearance. Summary of the Invention

[0005] In view of at least one of the above technical problems, this application provides a power turbine rotor structure design method, which can enable the rotor to have good vibration reduction by arranging oil film dampers in a reasonable position; perform deformation coordination calculation of the connection interface to optimize the connection structure and keep the connection interface from being damaged under long-term operation; design and optimize the segmented shaft wall thickness, optimize the design critical speed margin, and consider the influence of the ultimate unbalance on the clearance to achieve good dynamic balance characteristics.

[0006] According to one aspect of this application, a method for designing a power turbine rotor structure is provided, comprising the following steps:

[0007] S100: The strain energy distribution and mode shape of the power turbine rotor within the operating speed range are calculated by finite element software simulation, and the oil film damper is arranged according to the strain energy distribution and mode shape.

[0008] S200: The radial deformation of the connection interface between the power turbine disk and the power turbine shaft is calculated by finite element software under the working state of the power turbine rotor, and the structure of the power turbine rotor is optimized according to the radial deformation of the connection interface.

[0009] S300: The power turbine shaft is designed with segmented wall thicknesses, and the inner diameter of the power turbine shaft is fixed. The outer diameter of the power turbine shaft is divided into several segments of different sizes to meet the requirement that the critical speed of the power turbine rotor is far away from the preset value of the working speed.

[0010] S400: The first three critical vibration modes of the power turbine rotor are calculated by finite element software simulation, and a balance boss is set on the outer wall at the maximum displacement point of the power turbine shaft in each vibration mode.

[0011] S500: The dynamic response of the power turbine rotor is simulated and calculated using finite element software. The calculation conditions are selected based on the maximum imbalance caused by blade shedding, and the maximum vibration displacement of the power turbine disk is obtained. The gap between the central tie rod of the power turbine rotor and the power turbine disk is designed based on the maximum vibration displacement of the power turbine disk.

[0012] In some embodiments of this application, in step S100, the oil film damper is avoided from being placed on the thrust bearing of the power turbine rotor, and oil film dampers are placed at both the front and rear ends of the power turbine rotor, and the comprehensive strain energy of the power turbine rotor is selected. E 综 Arrange the supports at the locations with the highest proportion of bearing capacity.

[0013] In some embodiments of this application, the combined strain energy E 综 The calculation formula is as follows:

[0014] ;

[0015] In the formula, E 综 The proportion of comprehensive strain energy; E i For the first i Step strain energy; Q i For the first i The weight of the first strain energy, and the weight of the critical strain energy located between the ground slow speed and the design speed. Q i Take 1.0 as the critical strain energy weight located outside the range of ground slow speed to design speed. Q i Take 0.8.

[0016] In some embodiments of this application, in step S200, if the simulation calculation result shows that the interface between the power turbine disk and the power turbine shaft is deformed in a non-coordinated manner and exhibits an opening trend under the action of gyro torque and centrifugal force, then a clamping plate is added to the rear end of the power turbine rotor to coordinate the deformation of the interface.

[0017] In some embodiments of this application, in step S300, when designing the segment wall thickness of the power turbine shaft, the inner diameter of the power turbine shaft is selected according to the ventilation and torsion measurement shaft installation requirements. Based on the minimum cross-sectional bending coefficient and torsion coefficient design requirements of the power turbine shaft, the power turbine shaft is divided into three segments with different outer diameters along the axial direction to meet the requirement that the critical speed of the power turbine rotor is 20% away from the operating speed.

[0018] In some embodiments of this application, in step S400, if the positions of the maximum displacements in the power turbine shaft of each vibration mode overlap, the axial length of the balance boss at the overlapping position is extended, and the height of the balance boss must avoid rubbing against the stator.

[0019] In some embodiments of this application, when extending the axial length of the balancing boss at the overlapping position, it is necessary to ensure that the material removal quality at the overlapping position of the power turbine shaft remains unchanged.

[0020] In some embodiments of this application, in step S500, the maximum imbalance amount The calculation formula is as follows:

[0021] ;

[0022] In the formula, This is a normal imbalance quantity; The margin factor is set to 1.5. The coefficient is the roulette wheel coefficient. The support coefficient is 1.0 for simply supported wheel discs and 2.0 for cantilever wheel discs.

[0023] In some embodiments of this application, when calculating the maximum imbalance, for a single-stage turbine disk, a value 50 times the normal imbalance value is selected, i.e. Take 50; for a two-stage turbine disk, select 40 times the normal imbalance value for each disk, i.e. Take 40; for a three-stage turbine disk, select 30 times the normal imbalance value for each disk, i.e. Take 30; for a simply supported wheel with bearings at both the front and rear, select a support coefficient of 1, i.e. Take 1.0; for cantilevered wheel discs without bearings at the rear end, select a support coefficient of 2, i.e. Take 2.0.

[0024] In some embodiments of this application, in step S500, the gap between the central tie rod of the power turbine rotor and the power turbine disk is greater than the maximum vibration displacement value of the power turbine disk.

[0025] This application has the following beneficial effects:

[0026] This application's power turbine rotor structural design method comprehensively considers the impact of whole-machine detachment and containment tests on the power turbine rotor design, proposing a new power turbine rotor design method and process. Through finite element method (FEM) simulation calculations of the strain energy distribution and mode shapes at the critical speeds within the power turbine rotor's operating speed range, oil film dampers are positioned according to the strain energy distribution and mode shapes, achieving optimal damping for the rotor. The radial deformation of the connection interface between the power turbine disk and the power turbine shaft is also simulated using FEM. Based on this radial deformation, the power turbine rotor structure is optimized to avoid inconsistencies in interface deformation, ensuring a robust connection structure and preventing damage to the interface during long-term operation. This application proposes a segmented wall thickness design for the power turbine shaft. By fixing the inner diameter of the power turbine shaft first, the outer diameter is divided into several segments of different sizes, effectively meeting the requirement that the critical speed of the power turbine rotor is far from the preset operating speed. The first three critical vibration modes of the power turbine rotor are calculated using finite element method (FEM) software. Balancing bosses are then installed on the outer wall at the point of maximum displacement in each vibration mode of the power turbine shaft to achieve good dynamic balance characteristics, solving the difficult problem of dynamic balancing of multi-mode high-speed rotors. Furthermore, the number and location of the balancing bosses are designed based on the critical speed calculation results, which improves the problem of insufficient material removal surface in dynamic balancing. Simultaneously, the dynamic response of the power turbine rotor is calculated using FEM software. The maximum unbalance formed under blade shedding conditions is selected in the calculation to obtain the maximum vibration displacement of the power turbine disk. Finally, the clearance between the central tie rod and the power turbine disk is designed based on the maximum vibration displacement of the power turbine disk, considering the maximum unbalance in the clearance design. This effectively avoids problems such as rubbing of the central tie rod, ensuring the operational stability of the power turbine rotor. This design method is adapted to the long-life, low-vibration design concept of advanced civil turboshaft engines, enabling the efficient design of power turbine rotors that can successfully pass the whole-engine shedding and containment test, effectively improving the design efficiency and quality of power turbine rotors.

[0027] 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

[0028] 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:

[0029] Figure 1 This is a flowchart of a preferred embodiment of the method in this application;

[0030] Figure 2 This is a schematic diagram of the overall structure of the power turbine rotor according to a preferred embodiment of this application;

[0031] Figure 3 This is a schematic diagram of the vibration mode of the first critical speed of the power turbine rotor according to a preferred embodiment of this application;

[0032] Figure 4 This is a schematic diagram of the vibration mode of the second critical speed of the power turbine rotor according to a preferred embodiment of this application;

[0033] Figure 5 This is a schematic diagram of the vibration mode of the third critical speed of the power turbine rotor according to a preferred embodiment of this application;

[0034] Figure 6 This is a schematic diagram showing the opening trend of the disc-shaft connection interface of the power turbine rotor in a preferred embodiment of this application;

[0035] Figure 7 This is a structural schematic diagram of the segmented wall thickness of the power turbine shaft according to a preferred embodiment of this application;

[0036] Figure 8 This is a schematic diagram of the position of the balance boss under the first critical speed mode of the power turbine rotor in a preferred embodiment of this application;

[0037] Figure 9 This is a schematic diagram of the position of the balance boss under the second-order critical speed vibration mode of the power turbine rotor in a preferred embodiment of this application;

[0038] Figure 10 This is a schematic diagram of the position of the balance boss under the third critical speed mode of the power turbine rotor in a preferred embodiment of this application.

[0039] Legend: 100, power turbine shaft; 101, first section outer diameter; 102, second section outer diameter; 103, third section outer diameter. Detailed Implementation

[0040] 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.

[0041] Figure 1 This is a flowchart of a preferred embodiment of the method in this application; Figure 2 This is a schematic diagram of the overall structure of the power turbine rotor according to a preferred embodiment of this application; Figure 3This is a schematic diagram of the vibration mode of the first critical speed of the power turbine rotor according to a preferred embodiment of this application; Figure 4 This is a schematic diagram of the vibration mode of the second critical speed of the power turbine rotor according to a preferred embodiment of this application; Figure 5 This is a schematic diagram of the vibration mode of the third critical speed of the power turbine rotor according to a preferred embodiment of this application; Figure 6 This is a schematic diagram showing the opening trend of the disc-shaft connection interface of the power turbine rotor in a preferred embodiment of this application; Figure 7 This is a structural schematic diagram of the segmented wall thickness of the power turbine shaft according to a preferred embodiment of this application; Figure 8 This is a schematic diagram of the position of the balance boss under the first critical speed mode of the power turbine rotor in a preferred embodiment of this application; Figure 9 This is a schematic diagram of the position of the balance boss under the second-order critical speed vibration mode of the power turbine rotor in a preferred embodiment of this application; Figure 10 This is a schematic diagram of the position of the balance boss under the third critical speed mode of the power turbine rotor in a preferred embodiment of this application.

[0042] A method for designing a power turbine rotor structure includes the following steps:

[0043] S100: The strain energy distribution and mode shape of the power turbine rotor within the operating speed range are calculated by finite element software simulation, and the oil film damper is arranged according to the strain energy distribution and mode shape.

[0044] S200: The radial deformation of the connection interface between the power turbine disk and the power turbine shaft 100 under the working state of the power turbine rotor is calculated by finite element software simulation, and the structure of the power turbine rotor is optimized according to the radial deformation of the connection interface.

[0045] S300: The power turbine shaft 100 is designed with segmented wall thickness, the inner diameter of the power turbine shaft 100 is fixed, and the outer diameter of the power turbine shaft 100 is divided into several segments of different sizes to meet the requirement that the critical speed of the power turbine rotor is far away from the preset value of the working speed.

[0046] S400: The first three critical vibration modes of the power turbine rotor are calculated by finite element software simulation, and a balance boss is set on the outer wall at the maximum displacement of the power turbine shaft 100 in each vibration mode.

[0047] S500: The dynamic response of the power turbine rotor is simulated and calculated using finite element software. The calculation conditions are selected based on the maximum imbalance caused by blade shedding, and the maximum vibration displacement of the power turbine disk is obtained. The gap between the central tie rod of the power turbine rotor and the power turbine disk is designed based on the maximum vibration displacement of the power turbine disk.

[0048] The calculation of the critical speed within the operating speed range of the power turbine rotor here needs to consider the influence of the first three critical speeds (the influence of the first three critical speeds needs to be considered when considering the whole machine detachment and containment test). However, conventional power turbine rotors only need to consider the influence of the first or at most two critical speeds, without considering the influence of the whole machine detachment and containment test, and without considering the critical speed that the detachment and containment test may reach. Therefore, the damping design is usually simple.

[0049] The term "whole-engine detachment containment test" refers to a special test item necessitated by the structural characteristics of a turboshaft engine. Considering a failure mode where the transmission shaft breaks, the power turbine rotor loses its load, but airflow excitation remains, causing the turbine rotor to spin excessively. The powerful centrifugal load leads to the turbine disk rupture, resulting in hazardous consequences. To prevent disk rupture, a design concept is adopted where the blades detach before the disk. For example, the blades detach at 150% of the design speed, causing the disk to stop spinning due to the loss of airflow excitation. During the test, the power turbine rotor may reach 150% to 200% of the design speed. Special design is required for the power turbine rotor to ensure it successfully reaches these speeds and verifies containment.

[0050] This application's power turbine rotor structural design method generally considers the impact of whole-machine detachment and containment tests on the power turbine rotor design, proposing a new power turbine rotor design method and process. The strain energy distribution and mode shapes at the critical speed within the power turbine rotor's operating speed range are calculated using finite element method (FEM) simulation. Based on this strain energy distribution and mode shapes, oil film dampers are positioned to achieve optimal vibration reduction. The radial deformation of the connection interface between the power turbine disk and the power turbine shaft 100 under operating conditions is also calculated using FEM simulation. Based on this radial deformation, the power turbine rotor structure is optimized to avoid inconsistencies in interface deformation, ensuring a robust connection structure and preventing damage to the interface during long-term operation. This application proposes a segmented wall thickness design for the power turbine shaft 100. By fixing the inner diameter of the power turbine shaft 100 and dividing its outer diameter into several segments of different sizes, the requirement that the power turbine rotor's critical speed be far from the preset operating speed value is effectively met. The first three critical vibration modes of the power turbine rotor are calculated using finite element method (FEM) software. Balancing bosses are then installed on the outer wall of the power turbine shaft 100 at the point of maximum displacement in each vibration mode to achieve good dynamic balance characteristics of the power turbine shaft 100, solving the difficult problem of dynamic balancing of multi-mode high-speed rotors. Furthermore, the number and location of the balancing bosses are designed based on the critical speed calculation results, which improves the problem of insufficient material removal surface in dynamic balancing. Simultaneously, the dynamic response of the power turbine rotor is calculated using FEM software. The maximum unbalance formed under blade shedding conditions is selected in the calculation to obtain the maximum vibration displacement of the power turbine disk. Finally, the clearance between the central tie rod and the power turbine disk is designed based on the maximum vibration displacement of the power turbine disk, realizing clearance design considering the maximum unbalance and effectively avoiding problems such as rubbing of the central tie rod, ensuring the operational stability of the power turbine rotor. This design method is adapted to the long-life, low-vibration design concept of advanced civil turboshaft engines, achieving efficient design of power turbine rotors that can successfully pass the whole-engine shedding and containment test, effectively improving the design efficiency and quality of power turbine rotors.

[0051] Preferably, in step S100, the oil film damper is avoided from being placed on the thrust bearing of the power turbine rotor, and oil film dampers are placed at both the front and rear ends of the power turbine rotor, and the comprehensive strain energy of the power turbine rotor is selected. E 综 The support positions, which account for the highest proportion, are arranged accordingly. This achieves optimal structural layout and vibration reduction, resulting in smoother operation of the power turbine rotor within its variable speed range.

[0052] In this preferred embodiment, the combined strain energy E 综 The calculation formula is as follows:

[0053] ;

[0054] In the formula, E 综 The proportion of comprehensive strain energy; E i For the first i Step strain energy; Q i For the first i The weight of the first strain energy, and the weight of the critical strain energy located between the ground slow speed and the design speed. Q i Take 1.0 as the critical strain energy weight located outside the range of ground slow speed to design speed. Q i Take 0.8.

[0055] It should be noted that, considering prioritizing vibration reduction near the operating speed, the critical strain energy weight between the ground slow speed and the design speed is... Q i Take 1.0 as the critical strain energy weight located outside the range of ground slow speed to design speed. Q i Take 0.8.

[0056] In some embodiments, please refer to Figure 1 The power turbine rotor includes four bearings as support points, namely support point 1, support point 2, support point 5, and support point 6. The strain energy distribution is shown in Table 1.

[0057] Table 1 Strain Energy Distribution

[0058] ;

[0059] In Table 1, the first and second critical speeds are below the ground slowdown speed, while the third critical speed is above the design speed; therefore, all weights are taken as 0.8. The oil film damper should not be installed on the thrust bearing (bearing #1); instead, the oil film damper should be installed at bearing #2, based on the comprehensive strain energy. E 综 Based on the calculation results, the oil film damper is selected at the 6# support point. This allows the oil film damper to be placed in a reasonable position, thus enabling the rotor to have good vibration reduction.

[0060] Preferably, in step S200, if the simulation calculation result shows that the interface between the power turbine disk and the power turbine shaft is deformed in a non-coordinated manner and exhibits an opening trend under the action of gyro torque and centrifugal force, then a clamping plate is added at the rear end of the power turbine rotor to coordinate the deformation of the interface.

[0061] Understandably, considering the influence of gyroscopic torque and centrifugal force under rotor operating conditions and performing deformation coordination calculations at the connection interface can effectively avoid deformation inconsistencies at the connection interface.

[0062] It should be noted that when performing simulation calculations using finite element method (FEM) software, the radial deformation of the connection interface under the working conditions is calculated based on the applied temperature, gyroscopic torque, centrifugal force, and preload. If the difference in radial deformation between the two connected surfaces increases with rotational speed, the structure is modified by applying a clamping plate to harmonize the deformation of the connection interface. Please refer to [reference needed]. Figure 5 As shown, under the action of gyroscopic torque and centrifugal force, the interface between the power turbine disk and the power turbine shaft tends to open. Calculations show that the difference in radial deformation between the two connected surfaces increases with rotational speed. By applying a clamping plate to the rear end of the power turbine rotor, the disk-shaft connection surface is pressed together. After clamping, the connection structure is robust, achieving coordinated structural deformation, which helps to ensure that the connection interface remains undamaged during long-term operation.

[0063] Preferably, in step S300, when designing the segmented wall thickness of the power turbine shaft 100, the inner diameter of the power turbine shaft 100 is selected according to the ventilation and torsion measurement shaft installation requirements. Based on the minimum cross-sectional bending coefficient and torsion coefficient design requirements of the power turbine shaft 100, the power turbine shaft 100 is divided into three segments with different outer diameters along the axial direction to meet the requirement that the critical speed of the power turbine rotor is 20% away from the working speed.

[0064] It should be noted that, since the inner diameter of the power turbine shaft 100 is difficult to machine and not conducive to segmented design, the inner diameter of the power turbine shaft 100 is fixed first, and then the outer diameter of the power turbine shaft 100 is designed with different wall thicknesses in segments. When designing, the requirements for ventilation, torsion measurement shaft installation, strength requirements, and space requirements are taken into account as boundary conditions, and the critical speed is comprehensively optimized.

[0065] In some embodiments, the power turbine shaft 100 is divided axially into three segments with different outer diameters: a first segment with an outer diameter of 101, a second segment with an outer diameter of 102, and a third segment with an outer diameter of 103. Since the support position of the power turbine is basically fixed, the inner diameter of the power turbine shaft 100 is selected (based on the requirements for ventilation and torsion measurement shaft installation). The minimum cross-sectional bending coefficient and torsional coefficient of the power turbine shaft 100 are subject to product requirements. Therefore, the parameter that mainly affects the critical speed is the wall thickness of the power turbine shaft 100. However, conventional designs often fail to meet the requirement that the critical speed be 20% away from the operating speed. This application proposes dividing the outer diameter of the power turbine shaft 100 into three segments: left, middle, and right. Each segment is assigned a different outer diameter, and optimized design is carried out to effectively achieve a critical speed margin of over 20%.

[0066] Preferably, in step S400, if the positions of the maximum displacements in the power turbine shaft 100 of each vibration mode overlap, the axial length of the balance boss at the overlapping position is extended, and the height of the balance boss must avoid rubbing against the stator.

[0067] Understandably, the balancing bosses on the outer wall of the power turbine shaft 100 are limited by structural space and cannot be designed with a large outer diameter (as this would make them prone to rubbing against the combustion rotor). This application designs multiple balancing bosses based on the critical speed calculation results; please refer to [reference needed]. Figure 8 , Figure 9 , Figure 10 Furthermore, considering that the power turbine rotor needs to pass through the second or even third critical speed, this design effectively improves the problem of insufficient material removal surface in dynamic balance. Insufficient material removal surface in dynamic balance refers to the failure to fully remove material from the rotor surface where material needs to be removed to achieve a balanced state during dynamic balance correction.

[0068] Preferably, when extending the axial length of the balancing boss at the overlapping position, it is necessary to ensure that the material removal quality of the power turbine shaft 100 at the overlapping position remains unchanged.

[0069] Understandably, to avoid insufficient dynamic balancing surface when designing balancing bosses, if the design positions of the balancing bosses overlap, the length of the balancing bosses should be extended to ensure that the material removal quality remains unchanged and to ensure dynamic balance is achieved.

[0070] Preferably, in step S500, the maximum imbalance amount The calculation formula is as follows:

[0071] ;

[0072] In the formula, This is a normal imbalance quantity; The margin factor is set to 1.5. The coefficient is the roulette wheel coefficient. The support coefficient is 1.0 for simply supported wheel discs and 2.0 for cantilever wheel discs.

[0073] In this preferred embodiment, for a single-stage turbine disk, a value 50 times the normal imbalance value is selected, i.e. Take 50; for a two-stage turbine disk, select 40 times the normal imbalance value for each disk, i.e. Take 40; for a three-stage turbine disk, select 30 times the normal imbalance value for each disk, i.e. Take 30; for a simply supported wheel with bearings at both the front and rear, select a support coefficient of 1, i.e. Take 1.0; for cantilevered wheel discs without bearings at the rear end, select a support coefficient of 2, i.e. Take 2.0.

[0074] It should be noted that during the whole-machine detachment containment test, the impact of blade detachment on the rotor will cause huge deformation of the wheel disk and shaft, resulting in a large imbalance. If the clearance between the two rotors is designed too small under the condition of a large imbalance, the two rotors may rub against each other, which may lead to rotor breakage and harmful consequences. Here, the two rotors are the gas turbine rotor and the power turbine rotor. Conventional clearance design only considers the influence of vibration amplitude under normal imbalance (about 20 g·mm); it does not consider the influence of vibration amplitude under the condition of a large imbalance (about 1000 g·mm to 2000 g·mm).

[0075] Preferably, in step S500, the gap between the central tie rod of the power turbine rotor and the power turbine disk is greater than the maximum vibration displacement value of the power turbine disk.

[0076] Understandably, considering the maximum imbalance in the clearance design can avoid the impact of the extreme imbalance on the clearance between the center tie rod and the power turbine disk, ensuring that the center tie rod will not rub against the power turbine disk.

[0077] In summary, the power turbine rotor design method of this application considers the influence of the whole-machine detachment containment test in the damping layout. It considers not only the critical speed range within the operating speed range but also the critical speed that the detachment containment test may reach. Oil film dampers are placed in reasonable positions to ensure good vibration reduction of the rotor. This application determines the connection interface structure design through radial deformation coordination, avoiding problems such as uneven deformation of some structures, which can easily damage and deform the connection structure under long-term operation, causing changes in imbalance and leading to large rotor vibration. Based on the critical speed optimization design requirements, this application proposes a fixed 100mm inner diameter for the power turbine shaft and a multi-segment structure with different thicknesses for the outer diameter, effectively meeting the critical speed design requirements. Simultaneously, this application considers the influence of the critical speed that the detachment containment test may reach on the boss, setting a balancing boss in the optimal position to solve the problem of difficult dynamic balancing of multi-vibration modes in high-speed rotors. Finally, this application considers the large imbalance and designs the gap between the center tie rod and the power turbine disk of the power turbine rotor, improving the rationality and accuracy of the gap design. This application adapts to the long-life, low-vibration design concept of advanced civil turboshaft engines, achieving a power turbine rotor design that can successfully pass the whole-machine detachment containment test.

[0078] It should be noted that, in this document, the terms “comprising,” “including,” or any other variations thereof are intended to cover non-exclusive inclusion, such that a process, method, article, or apparatus that comprises a list of elements includes not only those elements but also other elements not expressly listed, or elements inherent to such process, method, article, or apparatus.

[0079] 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.

[0080] Each patent, patent application, publication of the patent application, and other materials such as articles, books, specifications, publications, documents, articles, etc., cited herein may be incorporated by reference. The entire contents used for all purposes, except for any history of prosecution documents associated with it, that may be inconsistent with or conflict with this document, or that may have a limiting effect on the widest extent of the claims, are now or hereafter associated with this document. For example, in the event of any inconsistency or conflict between the description, definition, and / or use of terms associated with any of the included materials and the terms, description, definition, and / or used in connection with this document, the terms used herein shall prevail.

[0081] 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 designing a power turbine rotor structure, characterized in that, It includes the following steps: S100: The strain energy distribution and mode shape of the power turbine rotor within the operating speed range are calculated by finite element software simulation, and the oil film damper is arranged according to the strain energy distribution and mode shape. S200: The radial deformation of the connection interface between the power turbine disk and the power turbine shaft (100) under the working state of the power turbine rotor is calculated by finite element software simulation, and the structure of the power turbine rotor is optimized according to the radial deformation of the connection interface. S300: Design the segmented wall thickness of the power turbine shaft (100), fix the inner diameter of the power turbine shaft (100), and divide the outer diameter of the power turbine shaft (100) into several segments of different sizes to meet the requirement that the critical speed of the power turbine rotor is far away from the preset value of the working speed. S400: The first three critical vibration modes of the power turbine rotor are calculated by finite element software simulation, and a balance boss is set on the outer wall at the maximum displacement of the power turbine shaft (100) in each vibration mode. S500: The dynamic response of the power turbine rotor is simulated and calculated using finite element software. The calculation conditions are selected based on the maximum imbalance caused by blade shedding, and the maximum vibration displacement of the power turbine disk is obtained. The gap between the central tie rod of the power turbine rotor and the power turbine disk is designed based on the maximum vibration displacement of the power turbine disk.

2. The power turbine rotor structure design method according to claim 1, characterized in that, In step S100, the oil film dampers are avoided from being placed on the thrust bearing of the power turbine rotor, and oil film dampers are placed at both the front and rear ends of the power turbine rotor. The comprehensive strain energy of the power turbine rotor is selected. E 综 Arrange the supports at the locations with the highest proportion of bearing capacity.

3. The power turbine rotor structure design method according to claim 2, characterized in that, The formula for calculating the comprehensive strain energy is as follows: ； In the formula, E 综 The proportion of comprehensive strain energy; E i For the first i Step strain energy; Q i For the first i The weight of the first strain energy, and the weight of the critical strain energy located between the ground slow speed and the design speed. Q i Take 1.0 as the critical strain energy weight located outside the range of ground slow speed to design speed. Q i Take 0.

8.

4. The power turbine rotor structure design method according to claim 1, characterized in that, In step S200, if the simulation calculation result shows that under the action of gyro torque and centrifugal force, the deformation of the connection interface between the power turbine disk and the power turbine shaft is inconsistent and tends to open, then a clamping plate is added to the rear end of the power turbine rotor to make the deformation of the connection interface consistent.

5. The power turbine rotor structure design method according to claim 1, characterized in that, In step S300, when designing the segmented wall thickness of the power turbine shaft (100), the inner diameter of the power turbine shaft (100) is selected according to the ventilation and torsion measurement shaft installation requirements. Based on the minimum cross-sectional bending coefficient and torsion coefficient design requirements of the power turbine shaft (100), the power turbine shaft (100) is divided into three segments with different outer diameters along the axial direction to meet the requirement that the critical speed of the power turbine rotor is 20% away from the working speed.

6. The power turbine rotor structure design method according to claim 1, characterized in that, In step S400, if the positions of the maximum displacements in the dynamic turbine shaft (100) of each vibration mode overlap, the axial length of the balance boss at the overlapping position is extended, and the height of the balance boss must avoid rubbing against the stator.

7. The power turbine rotor structure design method according to claim 6, characterized in that, When extending the axial length of the balance boss at the overlapping position, it is necessary to ensure that the material removal quality of the power turbine shaft (100) at the overlapping position remains unchanged.

8. The power turbine rotor structure design method according to claim 1, characterized in that, In step S500, the maximum imbalance amount The calculation formula is as follows: ； In the formula, This is a normal imbalance quantity; The margin factor is set to 1.

5. The coefficient is the roulette wheel coefficient. The support coefficient is 1.0 for simply supported wheel discs and 2.0 for cantilever wheel discs.

9. The power turbine rotor structure design method according to claim 8, characterized in that, When calculating the maximum imbalance, for a single-stage turbine disk, a value 50 times the normal imbalance value is selected. Take 50; for a two-stage turbine disk, select 40 times the normal imbalance value for each disk, i.e. Take 40; for a three-stage turbine disk, select 30 times the normal imbalance value for each disk, i.e. Take 30; for a simply supported wheel with bearings at both the front and rear, select a support coefficient of 1, i.e. Take 1.0; for cantilevered wheel discs without bearings at the rear end, select a support coefficient of 2, i.e. Take 2.

0.

10. The power turbine rotor structure design method according to claim 1, characterized in that, In step S500, the gap between the central tie rod of the power turbine rotor and the power turbine disk is greater than the maximum vibration displacement value of the power turbine disk.

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