CN116680804B - Critical speed design method of dual-rotor system with intermediate bearings based on strain energy distribution - Google Patents
Critical speed design method of dual-rotor system with intermediate bearings based on strain energy distribution Download PDFInfo
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Abstract
本发明公开了一种基于应变能分布的中介轴承双转子系统临界转速设计方法，涉及航空发动机转子动力学技术领域，包括：分析高低压转子及中介轴承应变能分布，结合提出的振型筛选参数，划分危险临界转速模态振型类型，进一步按照单转子弯曲振型、单转子刚体振型、双转子耦合弯曲振型的次序，有针对性地分析部件应变能分布及其变化，确定各类危险临界转速关键影响参数，从而进行结构特征参数调整以优化临界转速分布。本发明的临界转速分布设计方法，是针对中介轴承双转子系统首次提出，尤其解决了高低压转子耦合弯曲振型临界转速分布设计的问题，对引导工程实践中转子结构方案设计改进具有重要意义。
The present invention discloses a critical speed design method for an intermediate bearing dual rotor system based on strain energy distribution, which relates to the field of aero-engine rotor dynamics technology, including: analyzing the strain energy distribution of high- and low-pressure rotors and intermediate bearings, combining the proposed vibration mode screening parameters, dividing the dangerous critical speed modal vibration mode types, and further analyzing the strain energy distribution of components and their changes in the order of single rotor bending vibration mode, single rotor rigid body vibration mode, and dual rotor coupled bending vibration mode, and determining the key influencing parameters of various types of dangerous critical speeds, thereby adjusting the structural characteristic parameters to optimize the critical speed distribution. The critical speed distribution design method of the present invention is proposed for the first time for an intermediate bearing dual rotor system, and especially solves the problem of critical speed distribution design of high- and low-pressure rotor coupled bending vibration mode, which is of great significance for guiding the improvement of rotor structure design in engineering practice.
Description
Technical Field
The invention relates to the technical field of aero-engine rotor dynamics, in particular to a critical rotation speed design method of an intermediate bearing double-rotor system based on strain energy distribution.
Background
Modern advanced high thrust-weight ratio turbofan engines widely adopt an intermediate bearing double-rotor system structural layout, which consists of a high-pressure rotor and a low-pressure rotor, wherein the high-pressure rotor is supported on the low-pressure rotor through an intermediate bearing.
The requirements of high pneumatic performance and high structural efficiency lead the high-pressure rotor and the low-pressure rotor to have complex mass/rigidity distribution characteristics, and the high-pressure rotor and the low-pressure rotor are embodied to have complex modal characteristics in the aspect of mechanical characteristics. The use of intermediate bearings further couples the modes of the high-pressure rotor and the low-pressure rotor, thereby exacerbating the complexity of the mode characteristics. The load environment excited by the high-pressure rotor and the low-pressure rotor together and the complex modal characteristic of the double-rotor system enable the intermediate bearing double-rotor system to have complex critical rotation speed distribution, and the intermediate bearing double-rotor system consists of the critical rotation speed excited by the high-pressure rotor and the critical rotation speed excited by the low-pressure rotor. The coupling bending vibration mode critical rotation speed line excited by the high-voltage rotor is often located near the maximum working rotation speed, structural characteristics of the coupling bending vibration mode critical rotation speed line affect a plurality of parameters, the critical rotation speed margin is most difficult to meet and improve, the critical rotation speed distribution design problem is brought, repeated modification of a rotor structural design scheme is caused in engineering practice, and development progress is seriously affected.
The chinese patent application CN115640717a indicates that it is difficult to design an engine meeting the requirements following the critical rotation speed margin criteria, so a method for optimizing the design of the capacity mode of the intermediate bearing dual-rotor system is proposed. The method adopts three subfunctions of modal unbalance influence factor, elastic support total strain energy duty ratio and intermediate support strain energy duty ratio to form a modal capacity evaluation function as an optimal design function, and realizes the operation of a rotor system at critical rotation speed and neighborhood thereof through rotor structure optimal design. However, for an engineering practical intermediate bearing double-rotor system, the modal capacity of the coupling bending vibration mode critical rotation speed excited by the high-pressure rotor is very low, and the method is not suitable for the high-order bending vibration mode critical rotation speed with the coupling characteristic.
Therefore, a critical rotation speed distribution design method of the intermediate bearing dual-rotor system based on the current critical rotation speed margin criterion is needed, and is oriented to engineering practice, so that the influence parameters of the critical rotation speed distribution of the intermediate bearing dual-rotor system are explicitly optimized, the corresponding structural characteristic parameters are adjusted to realize the optimization of the critical rotation speed distribution, and the improvement direction of the actual rotor structural design scheme is guided.
Disclosure of Invention
The invention provides a critical rotation speed design method of an intermediate bearing double-rotor system based on strain energy distribution, which mainly aims to solve the problem of critical rotation speed division design of the intermediate bearing double-rotor system in the current engineering practice, and particularly aims at the design of coupling bending vibration mode critical rotation speed distribution excited by a high-pressure rotor.
In order to achieve the above purpose, the invention adopts the following technical scheme:
the critical rotation speed design method of the intermediate bearing double-rotor system based on strain energy distribution comprises the following steps:
Step S0, dangerous critical rotation speed determination: calculating critical rotation speed of high-pressure rotor and critical rotation speed of low-pressure rotor, drawing critical rotation speed distribution diagram of intermediate bearing double-rotor system, adding high-low pressure rotor common working line, and determining dangerous critical rotation speed near the common working rotation speed line;
step S1, dangerous critical rotation speed vibration mode classification: when the rotating speed of the non-excited rotor is 0, corresponding dangerous critical rotating speed points excited by the excited rotor are selected, strain energy distribution of the low-pressure rotor and a low-pressure first fulcrum, a low-pressure second fulcrum, a low-pressure third fulcrum, the high-pressure rotor and a high-pressure first fulcrum thereof and an intermediate bearing is analyzed, and critical rotating speed vibration mode screening parameters are utilized on the basis to divide dangerous critical rotating speed mode vibration mode types;
Step S2, rationality judgment and optimization of single-rotor bending vibration mode critical rotation speed distribution: judging the rationality of critical rotation speed distribution through the relative position relation between the exciting rotor rotation speed and a dangerous critical rotation speed curve in the maximum working state;
Step S3, rationality judgment and optimization of single-rotor rigid body vibration threshold speed distribution: judging the rationality of critical rotation speed distribution by exciting the relative position relationship between the rotation speed of a rotor and a dangerous critical rotation speed curve in a ground slow-running state;
And S4, rationality judgment and optimization of critical rotation speed distribution of the coupling bending vibration mode of the double rotors.
Further, in the step S1, the critical rotation speed vibration mode screening parameters include: modal coupling factor, supporting vibration factor, rotor bending factor, the expressions of which are respectively:
In the above formulas, R c is a modal coupling factor, R sv is a supporting vibration factor, R rb is a rotor bending factor, E sum is a total modal strain energy of the dual-rotor system, E sup4 is an intermediate bearing strain energy, E sup is a supporting strain energy, and E rotor is a structural modal strain energy of the rotor;
The critical speed mode shape types are classified into: single rotor rigid body vibration mode, single rotor bending vibration mode and double rotor coupling bending vibration mode; the judgment basis of various vibration modes is as follows: if the modal coupling factor R c is smaller than 0.2 and the supporting vibration factor R sv is larger than 0.5, judging that the single-rotor rigid body vibration mode is adopted; if the modal coupling factor R c is smaller than 0.2 and the rotor bending factor R rb is larger than 0.5, judging that the single rotor bending vibration mode is adopted; if the modal coupling factor R c is more than 0.2 and the rotor bending factor R rb is more than 0.5, judging that the dual-rotor coupling bending vibration mode is adopted.
Further, in the step S2, for the critical rotation speed of the bending vibration mode of the single rotor, the rationality judgment basis is as follows: making a straight line parallel to the rotating speed coordinate axis of the exciting rotor by passing through the maximum working state point to obtain an intersection point of the straight line and a corresponding dangerous critical rotating speed line, calculating the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor, and if the margin is more than or equal to 20%, and the rotating speed corresponding to the intersection point is higher than the maximum working state rotating speed; if the critical rotation speed distribution is unreasonable, further analyzing the strain energy distribution of the rotor structure, determining critical stiffness parameters influencing the critical rotation speed, realizing the critical rotation speed distribution optimization through the adjustment of the supporting stiffness and the adjustment of the rigidity of the rotating shaft and the shaft neck, and judging the rationality of the critical rotation speed distribution again; if the critical rotation speed is reasonably distributed in the adjustable range of the supporting rigidity and the geometric configuration of the component, the critical rotation speed distribution design is completed, if the critical rotation speed distribution design is still unreasonable, the position of the low-pressure second fulcrum is redesigned on the basis of not changing the high-pressure rotor, so that the low-pressure second fulcrum is close to the low-pressure third fulcrum, and the span of the low-pressure turbine shaft is reduced; or redesign the high pressure rotor to reduce its axial length to reduce the low pressure turbine shaft span.
Further, in the step S3, for the critical rotation speed of the single-rotor rigid body vibration mode, the rationality judgment basis is as follows: the method comprises the steps of passing through a ground slow-running state point, making a straight line parallel to an excitation rotor rotating speed coordinate axis, obtaining an intersection point of the straight line and a corresponding dangerous critical rotating speed line, calculating a margin of rotating speed corresponding to the intersection point relative to the excitation rotor rotating speed, and if the margin is more than or equal to 20%, and the rotating speed corresponding to the intersection point is lower than the ground slow-running state rotating speed; if the critical rotation speed distribution is unreasonable, further analyzing the rotor fulcrum strain energy distribution, determining the critical fulcrum stiffness parameter influencing the critical rotation speed, realizing the critical rotation speed distribution optimization through the adjustment of the supporting stiffness, and judging the rationality of the critical rotation speed distribution again; if the critical rotation speed is reasonably distributed in the adjustable range of the supporting rigidity, the critical rotation speed distribution design is completed, and if the critical rotation speed distribution is still unreasonable, the high-pressure rotor is redesigned, the axial length of the high-pressure rotor is reduced, and the mass and the diameter moment of inertia of the high-pressure rotor are reduced; or redesign the fan assembly to reduce its axial or radial dimensions and reduce its mass.
Further, in the step S4, the rationality of the critical rotation speed distribution is determined according to the relative position relationship between the exciting rotor rotation speed and the critical dangerous rotation speed curve in the maximum working state, and the rationality determination basis for the critical rotation speed of the coupled bending vibration mode of the dual rotors is: making a straight line parallel to the rotating speed coordinate axis of the exciting rotor by passing through the maximum working state point to obtain an intersection point of the straight line and a corresponding dangerous critical rotating speed line, calculating the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor, and if the margin is more than or equal to 20%, and the rotating speed corresponding to the intersection point is higher than the maximum working state rotating speed; if the critical rotation speed distribution is unreasonable, further analyzing the strain energy distribution of the dual-rotor system, determining critical stiffness parameters and mass inertia parameters affecting the critical rotation speed, realizing the critical rotation speed distribution optimization through the adjustment of the supporting stiffness, the adjustment of the rigidity of the rotating shaft and the adjustment of the rotation inertia ratio of the parts, and judging the rationality of the coupling bending vibration mode critical rotation speed distribution again; if the reasonable distribution of the coupling bending vibration mode critical rotation speed can be realized in the adjustable range of the supporting rigidity and the geometric configuration of the component, the rationality of the single-rotor rigid body vibration mode critical rotation speed distribution is further judged.
Further, in the step S4, for the critical speed of the single-rotor rigid body vibration mode, the rationality judgment basis is specifically: a slow-running state point on the ground is crossed, a straight line parallel to the rotating speed coordinate axis of the exciting rotor is made, an intersection point of the straight line and a corresponding dangerous critical rotating speed line is obtained, the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor is calculated, and if the margin is more than or equal to 20%, the margin is considered reasonable; if not, the step S4 is needed to be carried out again, and if so, the dangerous critical speed distribution design is completed; if the reasonable distribution of the coupling bending vibration mode critical rotation speed is still unreasonable in the adjustable range of the supporting rigidity and the geometric configuration of the component, the position of the low-pressure second supporting point is redesigned on the basis of not changing the high-pressure rotor, so that the low-pressure second supporting point is close to the low-pressure third supporting point, and the span of the low-pressure turbine shaft is reduced; or redesigning the high pressure rotor to reduce its axial length to reduce the low pressure turbine shaft span while reducing the high pressure rotor mass and diameter moment of inertia.
Further, through analyzing strain energy distribution of the low-pressure rotor and the low-pressure first fulcrum, the low-pressure second fulcrum, the low-pressure third fulcrum, the high-pressure rotor and the high-pressure first fulcrum and the intermediate bearing, the dangerous critical rotation speed mode vibration modes are accurately classified by combining the proposed mode vibration mode screening parameters, and then the distribution of critical rotation speeds of various vibration modes is designed sequentially according to the sequence of single-rotor bending vibration mode, single-rotor rigid body vibration mode and double-rotor coupling bending vibration mode, so that iteration in critical rotation speed design of a double-rotor system of the intermediate bearing is effectively reduced;
By purposefully analyzing the strain energy distribution and the change of the component, the rigidity parameter and the mass inertia parameter which influence the critical rotation speed are determined, and various influence parameters of the critical rotation speed of the coupling bending vibration mode of the double rotors are accurately found, so that corresponding structural characteristic parameter adjustment is carried out, and the critical rotation speed distribution is optimized.
Compared with the prior art, the invention has the advantages that:
1) Through strain energy distribution analysis of the low-pressure rotor and the pivot thereof (comprising a low-pressure first pivot, a low-pressure second pivot and a low-pressure third pivot in general), the high-pressure rotor and the pivot thereof (comprising a high-pressure first pivot in general) and the intermediate bearing, the dangerous critical speed mode vibration mode can be accurately classified by combining the proposed mode vibration mode screening parameters, and then the distribution of critical speed of various vibration modes is designed sequentially according to the sequence of single-rotor bending vibration mode, single-rotor rigid body vibration mode and double-rotor coupling bending vibration mode, so that iteration in critical speed design of the intermediate bearing double-rotor system is effectively reduced.
2) By purposefully analyzing the strain energy distribution and the change of the component, the rigidity parameter and the mass inertia parameter which influence the critical rotation speed can be determined, and particularly, various influence parameters of the critical rotation speed of the coupling bending vibration mode of the double rotors can be accurately found, so that the corresponding structural characteristic parameter adjustment is carried out, the critical rotation speed distribution is optimized, a clear direction is provided for the structural improvement of the engineering practice rotor, and the method has important significance for guiding the improvement of the overall structural design scheme of the rotor.
Drawings
FIG. 1 is a schematic diagram of an intermediate bearing dual rotor system according to an embodiment of the present invention;
FIG. 2 is a schematic diagram of components of an intermediate bearing dual rotor system in accordance with an embodiment of the present invention;
FIG. 3 is a flow chart of a critical rotational speed design method of an intermediate bearing dual-rotor system based on strain energy distribution according to the present invention;
FIG. 4 is a flow chart of the threshold speed distribution optimization according to the present invention;
FIG. 5 is a graph of critical rotational speed profile of an intermediate bearing dual rotor system in accordance with an embodiment of the present invention;
FIG. 6 is a graph of the strain energy distribution of critical speed at risk for each stage in accordance with an embodiment of the present invention;
FIG. 7 is a graph of a single rotor bending mode dangerous critical rotational speed strain energy profile in accordance with an embodiment of the present invention;
FIG. 8 is a chart of critical speed strain energy for a single rotor rigid body vibration mode risk in accordance with an embodiment of the present invention;
FIG. 9a is a graph of critical rotational speed strain energy for a dual rotor coupled flexural mode at 0 rpm for a non-excited rotor in accordance with an embodiment of the present invention;
FIG. 9b is a graph of bi-rotor coupled bending mode threshold rotational speed strain energy near maximum operating speed in accordance with an embodiment of the present invention;
FIG. 10 is a graph illustrating critical rotational speed profiles of an intermediate bearing dual rotor system in accordance with an embodiment of the present invention.
Detailed Description
It should be understood that the specific embodiments described herein are for purposes of illustration only and are not intended to limit the scope of the invention. The technical means used in the examples are conventional means well known to those skilled in the art unless otherwise indicated.
As shown in fig. 1, the intermediate bearing dual-rotor system of the present invention is composed of a low-pressure rotor 1 and a high-pressure rotor 2, the high-pressure rotor 2 is supported on the low-pressure rotor 1 through an intermediate bearing 3, and the high-pressure rotor 2 and the low-pressure rotor 1 are reversed.
As shown in fig. 2, the low-pressure rotor 1 mainly includes: a fan front journal 5, a fan assembly 6, a fan rear journal 7, a low pressure turbine shaft 9, a low pressure turbine assembly 11, a low pressure turbine journal 10, a low pressure first fulcrum 4, a low pressure second fulcrum 8, and a low pressure third fulcrum 12. The high-pressure rotor 2 mainly includes: a high pressure compressor front journal 14, a high pressure compressor assembly 15, a high pressure compressor rear cone shell 16, a high pressure drum shaft 17, a high pressure turbine assembly 18, a high pressure turbine rear journal 19, and a high pressure first fulcrum 13.
As shown in fig. 3, the method for designing critical rotation speed of the intermediate bearing dual-rotor system based on strain energy distribution according to the embodiment of the invention comprises the following steps:
Step S0, dangerous critical rotation speed determination: calculating critical rotation speed of the high-pressure rotor 2 and critical rotation speed of the low-pressure rotor 1, drawing critical rotation speed distribution diagram of the intermediate bearing double-rotor system, adding a common working line of the high-pressure rotor and the low-pressure rotor, and determining dangerous critical rotation speed near the common working rotation speed line.
Step S1, dangerous critical rotation speed vibration mode classification: when the rotation speed of the non-excited rotor is 0, corresponding dangerous critical rotation speed points excited by the excited rotor are selected, strain energy distribution of the low-pressure rotor 1, the low-pressure first supporting point 4, the low-pressure second supporting point 8, the low-pressure third supporting point 12, the high-pressure rotor 2, the high-pressure first supporting point 13 and the intermediate bearing is analyzed, and critical rotation speed vibration mode screening parameters are utilized on the basis to divide dangerous critical rotation speed mode vibration mode types.
The critical rotation speed vibration mode screening parameters comprise: modal coupling factor, supporting vibration factor, rotor bending factor, the expressions of which are respectively:
In the above formulas, R c is a modal coupling factor, R sv is a supporting vibration factor, R rb is a rotor bending factor, E sum is a total modal strain energy of the dual-rotor system, E sup4 is an intermediate bearing strain energy, E sup is a supporting strain energy, and E rotor is a structural modal strain energy of the rotor;
The critical speed mode shape types are classified into: single rotor rigid body vibration mode, single rotor bending vibration mode and double rotor coupling bending vibration mode. The judgment basis of various vibration modes is as follows: if the modal coupling factor R c is smaller than 0.2 and the supporting vibration factor R sv is larger than 0.5, judging that the single-rotor rigid body vibration mode is adopted; if the modal coupling factor R c is smaller than 0.2 and the rotor bending factor R rb is larger than 0.5, judging that the single rotor bending vibration mode is adopted; if the modal coupling factor R c is more than 0.2 and the rotor bending factor R rb is more than 0.5, judging that the dual-rotor coupling bending vibration mode is adopted.
Step S2, rationality judgment and optimization of single-rotor bending vibration mode critical rotation speed distribution, and specifically comprises three steps of SJ1, SA1 and SO 1:
SJ1: judging the rationality of critical rotation speed distribution through the relative position relation between the exciting rotor rotation speed and a dangerous critical rotation speed curve in the maximum working state;
SA1: analyzing strain energy distribution of a single rotor, and determining key rigidity parameters affecting critical rotation speed;
SO1: in the adjustable range of the supporting rigidity and the geometric configuration of the component, the critical rotation speed distribution is optimized through the adjustment of the supporting rigidity and the adjustment of the rigidity of the rotating shaft and the shaft neck, and then the step SJ1 is carried out to judge the rationality of the critical rotation speed distribution.
Step S3, rationality judgment and optimization of single-rotor rigid body vibration mode critical rotation speed distribution, which specifically comprises three steps of SJ2, SA2 and SO 2:
SJ2: judging the rationality of critical rotation speed distribution by exciting the relative position relationship between the rotation speed of a rotor and a dangerous critical rotation speed curve in a ground slow-running state;
SA2: analyzing the strain energy distribution of the rotor fulcrum, and determining the rigidity of the key fulcrum affecting the critical rotation speed;
SO2: in the supporting rigidity adjustable range, the critical rotation speed distribution is optimized through the adjustment of the supporting rigidity, and then the step SJ2 is carried out to judge the rationality of the critical rotation speed distribution.
Step S4, rationality judgment and optimization of critical rotation speed distribution of the double-rotor coupling bending vibration mode specifically comprises four steps of SJ3, SA3, SO3 and SJ 4:
SJ3: judging the rationality of critical rotation speed distribution through the relative position relation between the exciting rotor rotation speed and a dangerous critical rotation speed curve in the maximum working state;
SA3: analyzing strain energy distribution and change of the dual-rotor system, and determining key rigidity parameters and mass inertia parameters affecting the critical rotation speed;
SO3: and in the adjustable range of the supporting rigidity and the geometric configuration of the component, the critical rotation speed distribution is optimized through the adjustment of the supporting rigidity, the adjustment of the rigidity of the rotating shaft and the adjustment of the rotation inertia ratio of the component, and then the step SJ3 is carried out to judge the rationality of the critical rotation speed distribution.
SJ4: and judging the rationality of critical rotation speed distribution through the relative position relation between the exciting rotor rotation speed and a dangerous critical rotation speed curve in the ground slow-running state, wherein the judging basis is different from that of the SJ2 step.
As shown in fig. 4, which is a specific flowchart of optimizing the critical rotation speed distribution according to an embodiment of the present invention, the critical rotation speed distribution optimizing step SOi refers to a specific operation performed by steps SO1, SO2, SO3, and includes the following steps:
SM1: judging whether the supporting rigidity can be continuously increased or reduced according to the engineering actual supporting rigidity adjustable range;
SM2: increasing or decreasing the support stiffness of the corresponding critical pivot point; the key pivot points refer to pivot points with the highest strain energy ratio except the intermediate pivot points;
SM3: judging whether the rigidity of the journal can be continuously increased or reduced according to the strength requirement of the engineering actual journal, the mass limit of the journal, the modifiable range of the geometric configuration of the journal and the like;
SM4: changing the geometric configuration of the journal to realize the increase or decrease of the rigidity of the journal;
SM5: judging whether the rigidity of the rotating shaft can be continuously increased according to the engineering actual rotating shaft strength requirement, the rotating shaft quality limit, the rotating shaft geometric configuration modifiable range and the like;
SM6: the geometric configuration of the rotating shaft is changed, the rigidity of the rotating shaft is increased;
SM7: judging whether the axial dimension of the compression part can be continuously shortened according to the minimum axial dimension limit of the pneumatic design of the compression part in engineering practice;
SM8: the overall axial dimension of the compression component is shortened to realize the reduction of polar moment of inertia;
SM9: if the SM7 step is performed, the moment of inertia of the compression part is judged to be not adjustable, the fact that the critical rotation speed distribution is still unreasonable at the moment is indicated, and the rotor scheme needs to be redesigned.
The steps in the critical rotation speed design method of the intermediate bearing dual-rotor system based on the strain energy distribution will be described in further detail by specific implementation examples.
And S0, determining critical dangerous rotating speed.
Establishing a finite element model according to an engineering actual intermediate bearing double-rotor system, wherein a motion differential equation can be expressed as follows:
Wherein M, K, C is rotor system mass matrix, damping matrix, rigidity matrix, G l、Gh is gyro effect matrix of low-pressure rotor and high-pressure rotor, omega l、ωh is rotation speed of low-pressure rotor and high-pressure rotor, r, The generalized displacement vector, the generalized velocity vector and the generalized acceleration vector of the rotor system are respectively, and f is a generalized force vector.
Solving the critical rotation speed of the intermediate bearing double-rotor system, namely solving the characteristic value of the formula (1), wherein the characteristic equation of the formula (1) can be expressed as follows:
f(ωl,ωh,Ωl,Ωh)＝0 (2)
in the above formula, Ω l、Ωh is the precession speed of the low pressure rotor and the high pressure rotor, respectively.
The critical rotation speed of the intermediate bearing double-rotor system comprises the critical rotation speed of the low-pressure rotor and the critical rotation speed of the high-pressure rotor. When solving the critical rotation speed of low-pressure excitation, the following relation exists:
Ω＝Ωl＝Ωh＝ωl (3)
substituting the above formula into the characteristic equation (2) to obtain:
f(Ω,ωh)＝0 (4)
and sequentially giving high-pressure rotating speeds, solving the characteristic equation, obtaining critical rotating speeds of the low-pressure rotor excited at the corresponding high-pressure rotating speeds, and connecting all critical rotating speed points at different high-pressure rotating speeds according to the corresponding orders to obtain critical rotating speed lines of the low-pressure rotor excited.
When solving the critical rotation speed of high-pressure excitation, the following relation exists:
Ω＝Ωl＝Ωh＝ωh (5)
substituting the above formula into the characteristic equation (2) to obtain:
f(Ω,ωl)＝0 (6)
And sequentially giving low-pressure rotating speeds, solving the characteristic equation, obtaining critical rotating speeds of the high-pressure rotor under the corresponding low-pressure rotating speeds, and connecting all critical rotating speed points under different low-pressure rotating speeds according to corresponding orders to obtain critical rotating speed lines of the high-pressure rotor under the corresponding low-pressure rotating speeds.
And drawing a critical rotation speed line excited by the low-pressure rotor and a critical rotation speed line excited by the high-pressure rotor in the same graph to obtain a critical rotation speed distribution diagram of the intermediate bearing double-rotor system shown in fig. 5.
And drawing a joint working rotating speed line of the high-voltage rotor and the low-voltage rotor of the intermediate bearing double-rotor system, wherein the joint working rotating speed line comprises 4 main working states, the 1 st working state corresponds to a ground slow-running state, and the 4 th working state corresponds to a maximum working state.
In the embodiment of the invention, the judgment basis for the critical dangerous rotating speed is as follows: the common working line between the critical rotation speed line and the ground slow-running state to the maximum working state has an intersection point, or the rotation speed point on the critical rotation speed line, which is closest to the ground slow-running state and the maximum working state, has a margin of less than 20% with the ground slow-running state and the maximum working state.
As can be seen from fig. 5, the LPX-2 nd and LPX-3 rd orders, which are closest to the ground slow vehicle state, are critical dangerous rotational speeds; LPX-4 th, HPX-6 th and HPX-7 th, nearest to the maximum working state, are critical speeds of risk. In fig. 5 HPX represents high pressure rotor excitation and LPX represents low pressure rotor excitation.
And S1, classifying dangerous critical rotation speed vibration modes.
The critical dangerous rotational speed determined based on step S1 includes: LPX-2 nd order, LPX-3 rd order, LPX-4 th order, HPX-6 th order, HPX-7 th order. And for the critical rotation speed, when the rotation speed of the non-excited rotor is 0, analyzing strain energy distribution of the low-pressure rotor and a fulcrum thereof, the high-pressure rotor and a fulcrum thereof and an intermediate bearing at a corresponding dangerous critical rotation speed point excited by the excited rotor, namely a corresponding A, B, C, D, E point in fig. 5. The modal strain energy calculation expression of each structure is as follows:
Wherein E com is the modal strain energy of the corresponding structure, and u re、uim is the real modal displacement vector and the virtual modal displacement vector of the corresponding structure, respectively. Further, the ratio of the modal strain energy of the corresponding structure to the modal strain energy of the whole dual-rotor system is calculated, and the calculation expression is as follows:
Wherein, R e represents the strain energy ratio of the corresponding structure, E com、Esum represents the modal strain energy of the corresponding structure and the modal strain energy of the intermediate bearing dual-rotor system, and the modal strain energy of the dual-rotor system is the sum of all the component strain energies. The strain energy duty ratio of the corresponding structure is plotted in the same graph, and the strain energy distribution of the critical dangerous rotating speed of each step is shown in fig. 6.
As can be seen from analysis of fig. 6, in the LPX-2 nd order, the ratio of the high-pressure rotor and its fulcrum strain energy is close to 1, so it is a critical speed for the high-pressure rotor vibration; in the LPX-3 rd, LPX-4 th and HPX-7 th steps, the ratio of the low-pressure rotor and the fulcrum strain energy thereof is close to 1, so that the low-pressure rotor is a dangerous critical rotation speed taking the vibration of the low-pressure rotor as a vibration; in the HPX-6 th order, the strain energy ratio of the low-pressure rotor, the high-pressure rotor and the intermediate bearing is equivalent, and the preliminary judgment is the critical dangerous rotating speed of the high-low pressure coupling mode vibration.
Further, the modal coupling factor, the supporting vibration factor and the rotor bending factor of each dangerous critical rotation speed are calculated to accurately judge the critical rotation speed vibration mode as shown in table 2.
TABLE 2 dangerous critical speed mode shape determination
S2, rationality judgment and optimization of single-rotor bending vibration mode critical rotation speed distribution:
in the embodiment of the invention, for the critical dangerous rotating speed of the bending vibration mode of the single rotor, the rationality judgment basis is as follows: and (3) making a straight line parallel to the rotating speed coordinate axis of the exciting rotor by passing through the maximum working state point to obtain an intersection point of the straight line and the corresponding dangerous critical rotating speed line, calculating the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor, and if the margin is more than or equal to 20% and the rotating speed corresponding to the intersection point is higher than the maximum working state rotating speed, considering that the critical rotating speed distribution is reasonable. If not, the critical rotation speed distribution is considered unreasonable.
The critical risk speed margin calculation expression is as follows:
In the above formula, M cr is a critical speed margin, n er is an excitation rotor speed, and n cr is a critical speed.
As shown in fig. 5, for the LPX-4 th order in the single rotor bending vibration mode critical rotation speed, a line parallel to the low pressure rotor rotation speed coordinate axis is made to cross the LPX-4 th order critical rotation speed line at the W4 point, and the margin of the relative maximum working state rotation speed point is 11%, so that the LPX-4 th order critical rotation speed distribution is not reasonable.
Aiming at HPX-7 th order in the single rotor bending vibration mode critical rotating speed, a straight line parallel to a high-pressure rotor rotating speed coordinate axis is made through a maximum state rotating speed point, and the straight line intersects with the HPX-7 th order critical rotating speed line at a W5 point, and the margin of the relative maximum working state rotating speed point is 12%, so that the judgment of the HPX-7 th order critical rotating speed distribution is unreasonable.
Further, considering the dangerous critical rotation speed of the LPX-4 th and HPX-7 th order low pressure rotor vibration, the modal strain energy distribution of each part of the low pressure rotor and the fulcrum thereof at the W4 point and the W5 point is analyzed as shown in figure 7, and the critical rigidity parameter influencing the critical rotation speed is determined.
Further, as shown in fig. 7, the sum of the strain energy ratios of the low-pressure turbine shaft, the low-pressure first fulcrum, the low-pressure second fulcrum, and the low-pressure turbine journal reaches 91% for the LPX-4 th-order critical rotational speed, and the strain energy ratios thereof are 46%, 15%, 18%, and 12%, respectively. The rigidity of the structure is a key rigidity parameter, and has potential important influence on the critical speed of the LPX-4 th order danger.
Further, as shown in fig. 7, the sum of the strain energy ratios of the low-pressure turbine shaft, the front and rear fan journals, and the low-pressure turbine journal reaches 95% for the HPX-7 th-order critical speed, and the strain energy ratios thereof are 64%, 19%, and 12%, respectively. The rigidity of the structure is a key rigidity parameter, and has potential important influence on the HPX-7 th-order dangerous critical rotating speed.
Further, the LPX-4 th and HPX-7 th order critical speed distributions are optimized.
Firstly, the adjustability of the supporting rigidity of the low-pressure first supporting point and the low-pressure second supporting point is judged, and in the embodiment of the invention, the supporting rigidity adjustable range of each supporting point is considered to be 1 multiplied by 10 7～1×108 N/m. The supporting rigidity of the low-pressure first supporting point and the low-pressure second supporting point are respectively 2 multiplied by 10 7N/m、4×107 N/m, so that the supporting rigidity can be further increased. After the supporting rigidity of the two is improved to the upper limit of the adjustable range, the critical rotation speed margin of the LPX-4 th order and the HPX-7 th order is respectively improved to 15% and 12%, and the critical rotation speed distribution is still unreasonable.
Judging the adjustability of the supporting rigidity of the low-pressure first supporting point and the low-pressure second supporting point again, at this time, the supporting rigidity of the two components is increased to the upper limit of the supporting rigidity adjusting range, and the two components cannot be further adjusted. Therefore, judging the rigidity adjustability of the low-pressure turbine journal, the rigidity of the low-pressure turbine journal can not be further improved through geometric configuration design and local thickening at present, and the low-pressure turbine journal can not be further adjusted. Therefore, the rigidity adjustability of the low-pressure turbine shaft is judged, and the current low-pressure turbine shaft can be further increased in outer diameter and reduced in thickness on the basis of meeting the strength requirement and the size limitation of the front and rear journals of high pressure. After the rigidity of the low-pressure turbine shaft is improved by adopting the measures, the critical rotation speed margin of the LPX-4 th order and the HPX-7 th order is improved to 23% and 22%, and the critical rotation speed distribution is reasonable at the moment, so that the design of the single-rotor bending vibration mode critical rotation speed is completed.
Step S3, rationality judgment and optimization of single-rotor rigid body vibration threshold speed distribution:
In the embodiment of the invention, for the critical dangerous rotating speed of the single-rotor rigid body vibration mode, the rationality judgment basis is as follows: and (3) a passing ground slow-running state point is subjected to a straight line parallel to the rotating speed coordinate axis of the exciting rotor, so as to obtain an intersection point of the straight line and a corresponding dangerous critical rotating speed line, calculating the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor, and if the margin is more than or equal to 20%, judging that the critical rotating speed is distributed reasonably. If not, the critical rotation speed distribution is considered unreasonable.
As shown in fig. 5, for the LPX-2 nd order, ground passing slow-running state rotation speed point in the single-rotor rigid body vibration mode critical rotation speed, a straight line parallel to the low-pressure rotor rotation speed coordinate axis is made, and intersects with the LPX-2 nd order critical rotation speed line at a point W1, and the margin of the LPX-2 nd order critical rotation speed point relative to the ground slow-running state rotation speed point is 1%, so that the LPX-2 nd order critical rotation speed distribution is unreasonable.
Aiming at the LPX-3 rd order in the single-rotor rigid body vibration mode critical rotating speed, a straight line parallel to the rotating speed coordinate axis of the low-voltage rotor is made for the rotating speed point of the ground passing slow vehicle state, the straight line intersects with the LPX-3 rd order critical rotating speed line at a W2 point, and after the step 3, the margin of the rotating speed point of the low-speed state relative to the ground is 25%, so that the LPX-3 rd order critical rotating speed distribution is judged to be reasonable.
Further, considering the dangerous critical rotation speed of the LPX-2 nd order high pressure rotor mainly, the modal strain energy distribution of each part of the W1 point high pressure rotor and the supporting point thereof is analyzed, and as shown in fig. 8, the critical rigidity parameter influencing the critical rotation speed is determined.
Further, as shown in fig. 8, for the LPX-2 nd order critical rotational speed, the strain energy ratio of the high-pressure first fulcrum reaches 83%, and the strain energy ratio of the rest structure is very low, so that the supporting rigidity is a critical rigidity parameter, and has a potential important influence on the LPX-2 nd order critical rotational speed.
Further, the LPX-2 nd order critical speed distribution is optimized.
First, the adjustability of the supporting rigidity of the high-pressure first supporting point is judged, and in the embodiment of the invention, the supporting rigidity of the supporting point is considered to be 1X 10 7～1×108 N/m. The current high pressure first fulcrum has a supporting stiffness of 3 x 10 7 N/m, so it can be further reduced to make the LPX-2 nd order critical speed line to be located on the right side of the ground slow speed point with sufficient margin. After the supporting rigidity is reduced to 1.5X10 7 N/m, the critical rotation speed margin of LPX-2 nd order is increased to 21%, and the critical rotation speed distribution is reasonable at this time, so that the design of the critical rotation speed of the single-rotor rigid body vibration mode is completed.
And S4, rationality judgment and optimization of critical rotation speed distribution of the double-rotor coupling bending vibration mode.
In the embodiment of the invention, for the critical rotation speed of the coupling bending vibration mode of the double rotors, the rationality judgment basis is as follows: and (3) making a straight line parallel to the rotating speed coordinate axis of the exciting rotor by passing through the maximum working state point to obtain an intersection point of the straight line and the corresponding dangerous critical rotating speed line, calculating the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor, and if the margin is more than or equal to 20% and the rotating speed corresponding to the intersection point is higher than the maximum working state rotating speed, considering that the critical rotating speed distribution is reasonable. If not, the critical rotation speed distribution is considered unreasonable.
As shown in FIG. 5, for HPX-6 th order in the critical rotation speed of single rotor bending vibration mode, a line parallel to the rotation speed coordinate axis of low pressure rotor is made through the maximum state rotation speed point, and intersects with HPX-6 th order critical rotation speed line at W3 point, the margin of relative maximum working state rotation speed point is 4%, and is lower than the maximum working state rotation speed, so that it is unreasonable to judge HPX-6 th order critical rotation speed distribution.
Further, considering that the HPX-6 th order is the critical speed of the coupling vibration danger of the double rotors, the modal strain energy distribution of each component of the double rotor system of the W3 point and the D point is analyzed, and the critical rigidity parameter and the mass inertia parameter affecting the critical speed are determined as shown in fig. 9a and 9b respectively.
Further, as shown in fig. 9a and 9b, the sum of the strain energy ratios of the intermediate fulcrum, the low pressure turbine shaft, the high pressure drum shaft, and the fan shaft is 86% for the HPX-6 th-order critical speed, and the strain energy ratios are 31%, 27%, 10%, 9%, and 9%, respectively. The rigidity of the structure is a key rigidity parameter, and has potential important influence on the HPX-6 th-order critical speed.
Further, as can be seen from comparative analysis of strain energy distribution changes of the W3 point and the D point, for the HPX-6 th-order dangerous critical rotation speed, the sum of the strain energy ratio of the high-pressure drum shaft and the high-pressure compressor is reduced from 43% to 10%, namely, under the gyroscopic effect of the high-pressure compressor and the high-pressure turbine, the bending deformation of the high-pressure rotor is obviously reduced, the gyroscopic effect obviously increases the equivalent rigidity of the high-pressure rotor, and the HPX-6 th-order dangerous critical rotation speed is improved. The strength of the gyroscopic effect depends on the ratio of the polar moment of inertia to the diameter moment of inertia of the large-mass inertia component, and for an intermediate bearing double-rotor system adopted by an advanced high thrust-weight ratio turbofan engine, a single-stage high-pressure turbine design is adopted in most cases, the moment of inertia ratio is difficult to adjust, so that the ratio of the polar moment of inertia to the diameter moment of inertia of the high-pressure compressor is a key mass inertia parameter, and has potential important influence on the HPX-6 th-order dangerous critical rotating speed.
Further, the HPX-6 th order critical speed distribution is optimized.
First, the adjustability of the stiffness of the intermediate fulcrum is judged. The intermediate fulcrum cannot adopt an elastic supporting structure, the rigidity of which depends on the structural characteristics of the intermediate bearing itself, so that the intermediate fulcrum is not adjustable.
Further, the rigidity adjustability of the low-pressure turbine journal is judged, and the rigidity of the low-pressure turbine journal cannot be further improved through geometric configuration design and local thickening at present, so that the low-pressure turbine journal cannot be further adjusted.
Further, the adjustability of the rigidity of the high-pressure drum shaft is judged, the tangential speed of the current drum shaft can be further improved, the axial size of the combustion chamber can be further shortened, and the rigidity of the high-pressure drum shaft is adjustable. The measure of shortening the axial dimension of the drum shaft and increasing the outer diameter is adopted, after the rigidity is improved, the LHPX th-6 th order critical rotating speed margin is improved to 9%, and the critical rotating speed distribution is still unreasonable.
In addition, the axial dimension of the high-pressure drum shaft is shortened, the single-rotor bending critical rotation speed margin is improved, and the LPX-4 th order critical rotation speed margin and the HPX-7 th order critical rotation speed margin are respectively improved to 24% and 25%.
Further, the adjustability of the rigidity of the low-pressure turbine shaft is judged, the outer diameter of the middle section of the turbine shaft can be further increased and the thickness of the middle section of the turbine shaft can be reduced on the basis that the current low-pressure turbine shaft meets the strength requirement and the size limitation of the front shaft neck and the rear shaft neck of the high pressure, after the rigidity of the low-pressure turbine shaft is improved by adopting the measures, the margin of the HPX-6 th order critical rotation speed is improved to 14%, and at the moment, the critical rotation speed distribution is still unreasonable.
Further, the adjustability of the rotational inertia of the compression part is judged, the radial dimension of the high-pressure compressor is not greatly modified under the current design scheme, and the axial dimension of the high-pressure compressor can be further shortened, so that the diameter rotational inertia of the high-pressure compressor can be reduced, and the ratio of the polar rotational inertia to the diameter rotational inertia of the high-pressure compressor is improved. After the measures are adopted, the rotational inertia of the high-pressure compressor is increased to 1.2 compared with that of the high-pressure compressor from 0.8 in the original scheme, so that the margin of the HPX-6 th order critical rotation speed is increased to 21%, and the critical rotation speed is reasonably distributed at the moment, so that the design of the critical rotation speed of the double-rotor coupling bending vibration mode is completed.
In addition, the axial dimension of the high-pressure drum shaft is shortened, the single-rotor bending critical speed margin is improved, and the LPX-4 th order critical speed margin and the HPX-7 th order critical speed margin are respectively improved to 27% and 26%.
Thus, the critical rotation speed distribution design of the intermediate bearing dual-rotor system in the embodiment is completed, and the critical rotation speed distribution is shown in fig. 10.
According to the embodiment, the critical rotation speed design method of the intermediate bearing double-rotor system based on the strain energy distribution, disclosed by the invention, can be used for accurately classifying the critical rotation speed mode vibration modes by analyzing the critical rotation speed mode strain energy distribution, and further sequentially designing the distribution of various vibration modes according to the sequence of single-rotor bending vibration mode, single-rotor rigid body vibration mode and double-rotor coupling bending vibration mode. In particular, various influencing parameters of the critical rotation speed of the coupling bending vibration mode of the double rotors can be accurately found, so that corresponding structural characteristic parameter adjustment is carried out, critical rotation speed distribution is optimized, a clear direction is provided for structural improvement of engineering practice rotors, and the method has important significance for guiding the improvement of the overall structural design scheme of the rotors.
The above embodiments are only illustrative of the preferred embodiments of the present invention and are not intended to limit the scope of the present invention, and various modifications, variations, alterations, substitutions made by those skilled in the art to the technical solution of the present invention should fall within the protection scope defined by the claims of the present invention without departing from the spirit of the design of the present invention.
Claims (6)
1. The critical rotation speed design method of the intermediate bearing double-rotor system based on strain energy distribution is characterized by comprising the following steps of:
Step S0, dangerous critical rotation speed determination: calculating critical rotation speed of high-pressure rotor and critical rotation speed of low-pressure rotor, drawing critical rotation speed distribution diagram of intermediate bearing double-rotor system, adding high-low pressure rotor common working line, and determining dangerous critical rotation speed near the common working rotation speed line;
step S1, dangerous critical rotation speed vibration mode classification: when the rotating speed of the non-excited rotor is 0, corresponding dangerous critical rotating speed points excited by the excited rotor are selected, strain energy distribution of the low-pressure rotor and a low-pressure first fulcrum, a low-pressure second fulcrum, a low-pressure third fulcrum, the high-pressure rotor and a high-pressure first fulcrum thereof and an intermediate bearing is analyzed, and critical rotating speed vibration mode screening parameters are utilized on the basis to divide dangerous critical rotating speed mode vibration mode types;
the critical rotation speed vibration mode screening parameters comprise: modal coupling factor, supporting vibration factor, rotor bending factor, the expressions of which are respectively:
，
，
，
in the above-mentioned formulae, the first and second light-emitting elements, As a mode-coupling factor,In order to support the vibration factor,As a result of the rotor bending factor,Is the total modal strain energy of the dual-rotor system,In order to mediate the strain energy of the bearing,In order to support the strain energy it is desirable,The modal strain energy of the rotor structure;
the critical speed mode shape types are classified into: single rotor rigid body vibration mode, single rotor bending vibration mode and double rotor coupling bending vibration mode; the judgment basis of various vibration modes is as follows: if the mode coupling factor And supports the vibration factorJudging the single-rotor rigid body vibration mode; if the mode coupling factorAnd rotor bending factorJudging the bending vibration mode of the single rotor; if the mode coupling factorAnd rotor bending factorJudging the coupled bending vibration mode of the double rotors;
Step S2, rationality judgment and optimization of single-rotor bending vibration mode critical rotation speed distribution: judging the rationality of critical rotation speed distribution through the relative position relation between the exciting rotor rotation speed and a dangerous critical rotation speed curve in the maximum working state;
Step S3, rationality judgment and optimization of single-rotor rigid body vibration threshold speed distribution: judging the rationality of critical rotation speed distribution by exciting the relative position relationship between the rotation speed of a rotor and a dangerous critical rotation speed curve in a ground slow-running state;
And S4, rationality judgment and optimization of critical rotation speed distribution of single-rotor rigid body vibration mode or double-rotor coupling bending vibration mode.
2. The critical rotation speed design method of the intermediate bearing dual-rotor system based on strain energy distribution according to claim 1, wherein in the step S2, for the critical rotation speed of the single-rotor bending vibration mode, the rationality judgment basis is: making a straight line parallel to the rotating speed coordinate axis of the exciting rotor by passing through the maximum working state point to obtain an intersection point of the straight line and a corresponding dangerous critical rotating speed line, calculating the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor, and if the margin is more than or equal to 20%, and the rotating speed corresponding to the intersection point is higher than the maximum working state rotating speed; if the critical rotation speed distribution is unreasonable, further analyzing the strain energy distribution of the rotor structure, determining critical stiffness parameters influencing the critical rotation speed, realizing the critical rotation speed distribution optimization through the adjustment of the supporting stiffness and the adjustment of the rigidity of the rotating shaft and the shaft neck, and judging the rationality of the critical rotation speed distribution again; if the critical rotation speed is reasonably distributed in the adjustable range of the supporting rigidity and the geometric configuration of the component, the critical rotation speed distribution design is completed, if the critical rotation speed distribution design is still unreasonable, the position of the low-pressure second fulcrum is redesigned on the basis of not changing the high-pressure rotor, so that the low-pressure second fulcrum is close to the low-pressure third fulcrum, and the span of the low-pressure turbine shaft is reduced; or redesign the high pressure rotor to reduce its axial length to reduce the low pressure turbine shaft span.
3. The critical rotation speed design method of the intermediate bearing dual-rotor system based on strain energy distribution according to claim 2, wherein in the step S3, for the critical rotation speed of the single-rotor rigid body vibration mode, the rationality judgment basis is: the method comprises the steps of passing through a ground slow-running state point, making a straight line parallel to an excitation rotor rotating speed coordinate axis, obtaining an intersection point of the straight line and a corresponding dangerous critical rotating speed line, calculating a margin of rotating speed corresponding to the intersection point relative to the excitation rotor rotating speed, and if the margin is more than or equal to 20%, and the rotating speed corresponding to the intersection point is lower than the ground slow-running state rotating speed; if the critical rotation speed distribution is unreasonable, further analyzing the rotor fulcrum strain energy distribution, determining the critical fulcrum stiffness parameter influencing the critical rotation speed, realizing the critical rotation speed distribution optimization through the adjustment of the supporting stiffness, and judging the rationality of the critical rotation speed distribution again; if the critical rotation speed is reasonably distributed in the adjustable range of the supporting rigidity, the critical rotation speed distribution design is completed, and if the critical rotation speed distribution is still unreasonable, the high-pressure rotor is redesigned, the axial length of the high-pressure rotor is reduced, and the mass and the diameter moment of inertia of the high-pressure rotor are reduced; or redesign the fan assembly to reduce its axial or radial dimensions and reduce its mass.
4. The critical rotational speed design method of the intermediate bearing dual-rotor system based on strain energy distribution according to claim 3, wherein in the step S4, the rationality judgment basis for the critical rotational speed of the dual-rotor coupling bending vibration mode is: making a straight line parallel to the rotating speed coordinate axis of the exciting rotor by passing through the maximum working state point to obtain an intersection point of the straight line and a corresponding dangerous critical rotating speed line, calculating the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor, and if the margin is more than or equal to 20%, and the rotating speed corresponding to the intersection point is higher than the maximum working state rotating speed; if the critical rotation speed distribution is unreasonable, further analyzing the strain energy distribution of the dual-rotor system, determining critical stiffness parameters and mass inertia parameters affecting the critical rotation speed, realizing the critical rotation speed distribution optimization through the adjustment of the supporting stiffness, the adjustment of the rigidity of the rotating shaft and the adjustment of the rotation inertia ratio of the parts, and judging the rationality of the coupling bending vibration mode critical rotation speed distribution again; if the reasonable distribution of the coupling bending vibration mode critical rotation speed can be realized in the adjustable range of the supporting rigidity and the geometric configuration of the component, the rationality of the single-rotor rigid body vibration mode critical rotation speed distribution is further judged.
5. The critical rotation speed design method of the intermediate bearing dual-rotor system based on strain energy distribution according to claim 3, wherein in the step S4, for the critical rotation speed of the single-rotor rigid body vibration mode, the rationality judgment is specifically as follows: a slow-running state point on the ground is crossed, a straight line parallel to the rotating speed coordinate axis of the exciting rotor is made, an intersection point of the straight line and a corresponding dangerous critical rotating speed line is obtained, the margin of the rotating speed corresponding to the intersection point relative to the rotating speed of the exciting rotor is calculated, and if the margin is more than or equal to 20%, the margin is considered reasonable; if not, the step S4 is needed to be carried out again, and if so, the dangerous critical speed distribution design is completed; if the reasonable distribution of the coupling bending vibration mode critical rotation speed is still unreasonable in the adjustable range of the supporting rigidity and the geometric configuration of the component, the position of the low-pressure second supporting point is redesigned on the basis of not changing the high-pressure rotor, so that the low-pressure second supporting point is close to the low-pressure third supporting point, and the span of the low-pressure turbine shaft is reduced; or redesigning the high pressure rotor to reduce its axial length to reduce the low pressure turbine shaft span while reducing the high pressure rotor mass and diameter moment of inertia.
6. The critical rotational speed design method of the intermediate bearing double-rotor system based on strain energy distribution according to claim 1, wherein the critical rotational speed design method is characterized in that the strain energy distribution of a low-pressure rotor, a low-pressure first fulcrum, a low-pressure second fulcrum, a low-pressure third fulcrum, a high-pressure rotor, a high-pressure first fulcrum and an intermediate bearing is analyzed, and the critical rotational speed mode is accurately classified by combining the proposed mode vibration mode screening parameters, so that the critical rotational speed distribution of various vibration modes is designed sequentially according to the sequence of a single-rotor bending vibration mode, a single-rotor rigid body vibration mode and a double-rotor coupling bending vibration mode, and iteration in the critical rotational speed design of the intermediate bearing double-rotor system is effectively reduced;
By purposefully analyzing the strain energy distribution and the change of the component, the rigidity parameter and the mass inertia parameter which influence the critical rotation speed are determined, and various influence parameters of the critical rotation speed of the coupling bending vibration mode of the double rotors are accurately found, so that corresponding structural characteristic parameter adjustment is carried out, and the critical rotation speed distribution is optimized.
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