Suspension design method and device suitable for vehicle gas turbine

Abstract

The invention discloses a suspension design method and device suitable for a vehicle gas turbine, and relates to the field of vehicle vibration isolation. The method comprises the following steps: acquiring inertial parameters of the gas turbine; determining rigidity solving constraint conditions of vibration decoupling according to the inertial parameters by adopting an elastic center method; performing modal analysis and vibration response characteristic analysis on the basis of a mechanical vibration theory, determining a dynamic differential equation, adopting standard pavement spectrum excitation of a design requirement grade, calculating a relative displacement expression between the blade and the wall surface, and determining a rigidity solving constraint condition for ensuring the structural safety; determining a safety rigidity range according to a rigidity solving constraint condition of vibration decoupling and a rigidity solving constraint condition for ensuring structural safety; and in a safe rigidity range, selecting a group of most appropriate rigidity combinations and checking, and then carrying out nonlinear rigidity matching to obtain matched data. The problem that the structure of the vehicle gas turbine is prone to being damaged under the vibration working condition is solved, and the application reliability is improved.

Description

Technical Field

[0001] The present application relates to the field of vehicle vibration isolation, and in particular to a suspension design method and device suitable for a vehicle gas turbine. Background Art

[0002] Gas turbines are used in vehicle power systems and have the following significant advantages: they can provide strong power output, respond quickly and are highly efficient, and are particularly suitable for applications that require high power density and rapid acceleration performance; they have high combustion efficiency and relatively clean emissions during operation, which helps reduce environmental pollution; in addition, gas turbines have a compact structure, are relatively easy to maintain, run smoothly, and have low noise, bringing a more comfortable and environmentally friendly driving experience to vehicles. Based on the above advantages, gas turbines are becoming a research hotspot for the engineering application of the new generation of vehicle power systems.

[0003] The rational application of vehicle gas turbines requires a mounting system. The role of the gas turbine mounting system is similar to that of the engine mounting system, which mainly plays the role of support, vibration isolation and limit. The rationality of the mounting system design is directly related to the reliability of the powertrain operation. However, due to the fundamental differences in structure and working principle between gas turbines and traditional reciprocating piston internal combustion engines, the gas turbine structure is more sophisticated, the gap between the blades and the wall is very small, and the gas turbine is prone to collision between the blades and the wall under vibration conditions. Therefore, the design focus of the gas turbine mounting system is different from that of the traditional engine mounting. The design method of the traditional reciprocating piston internal combustion engine mounting system cannot be directly applied to the design of the vehicle-mounted gas turbine mounting system, and there is little research on the mounting stiffness design method for gas turbines. The stiffness design of the mounting system has become one of the obstacles to the promotion and application of gas turbines in the vehicle field.

[0004] At present, the design methods of suspension stiffness are mostly designed for reciprocating piston internal combustion engines. The vibration isolation effect of the traditional reciprocating piston internal combustion engine suspension system is reflected in two aspects. One is to prevent the impact of uneven ground excitation on the internal combustion engine, and the other is to prevent the vibration of the internal combustion engine from being transmitted to the body. The design focus can be summarized as ensuring that the overall response displacement of the internal combustion engine powertrain to the road excitation vibration is within a safe range, and avoiding the resonance caused by the coincidence of the suspension natural frequency and the engine excitation frequency. However, the working form of the gas turbine is different from the reciprocating motion form of the piston internal combustion engine. When working, the vibration excitation generated by itself as a vibration excitation source is relatively small. After being transmitted by the suspension system, the vibration response generated by the power unit excitation of the body and the passenger compartment is very small, and the vibration and noise are at a low level. The design requirements for the powertrain as an excitation source are relatively low. In addition, unlike traditional reciprocating piston internal combustion engines, gas turbines have a more sophisticated structure and a very small gap between the blades and the wall. When designing, we must not only pay attention to the overall displacement response of the powertrain under road excitation, but also pay attention to the relative displacement between the gas turbine blades and the wall not exceeding the safe design gap. Therefore, it is crucial to solve the problem of structural damage that is prone to occur under vibration conditions of automotive gas turbines and improve application reliability. Summary of the invention

[0005] The purpose of this application is to provide a suspension design method and device suitable for a vehicle gas turbine, which can solve the problem of structural damage that is prone to occur under vibration conditions of the vehicle gas turbine and improve application reliability.

[0006] To achieve the above objectives, this application provides the following solutions:

[0007] In a first aspect, the present application provides a suspension design method applicable to a vehicle gas turbine, comprising:

[0008] Obtaining inertial parameters of the gas turbine; the inertial parameters include: the coordinates of the suspension point position and the coordinates of the gas turbine center of mass; the inertial parameters are determined based on the powertrain inertial reference coordinate system; the powertrain inertial reference coordinate system takes the intersection of the gas generator rotor axis and the gearbox rear machine and the compressor intake machine installation surface as the coordinate origin O, the airflow direction is the positive direction of the X axis, the vertical direction is the Z axis, and the X axis, Y axis and Z axis form a right-handed coordinate system;

[0009] Using the elastic center method, the stiffness solution constraint condition of the vibration decoupling is determined according to the inertia parameters;

[0010] Perform modal analysis and vibration response characteristic analysis based on mechanical vibration theory to determine the dynamic differential equation;

[0011] Based on the dynamic differential equation, the standard road surface spectrum excitation of the design requirement level is adopted to calculate the relative displacement expression between the blade and the wall surface, and the stiffness solution constraint condition to ensure the safety of the structure is determined;

[0012] According to the stiffness solution constraint conditions for vibration decoupling and the stiffness solution constraint conditions for ensuring structural safety, the safe stiffness range of each point of the gas turbine mounting is determined;

[0013] Within the safety stiffness range, a set of stiffness combinations that meet the preset requirements are selected and checked to obtain the design stiffness value;

[0014] Nonlinear stiffness matching is performed according to the designed stiffness value to obtain matching data; the matching data is parameter data for determining the suspension design of the vehicle gas turbine.

[0015] Optionally, the elastic center method is used to determine the stiffness solution constraint conditions for vibration decoupling according to the inertia parameters, specifically including:

[0016] Determine the mass center reference coordinate system according to the gas turbine mass center coordinates;

[0017] Based on the centroid reference coordinate system, the coordinates of the suspension point position are converted into the centroid coordinates of the suspension point position;

[0018] According to the principle of vibration decoupling by the elastic center method, the coordinates of the elastic center are made the same as the values ​​corresponding to the coordinates of the gas turbine center of mass, and the stiffness solution constraint conditions for vibration decoupling are obtained; the coordinates of the elastic center are determined based on the coordinates of the suspension point position.

[0019] Optionally, the stiffness solving constraint condition of vibration decoupling includes: solving the constraint equation of the vertical stiffness of the rubber shock absorbing block;

[0020] The expression for solving the constraint equation of the vertical stiffness of the rubber damping block is:

[0021]

[0022] The coordinates of the elastic center are given by:

[0023]

[0024] Among them, x s is the horizontal coordinate of the elastic center; y s is the ordinate of the elastic center; k ri is the radial stiffness of the rubber shock absorber at the ith suspension point; x si y is the horizontal coordinate of the center of mass of the suspension point; si x is the vertical coordinate of the center of mass of the suspension point; c is the horizontal coordinate of the gas turbine mass center; c is the vertical coordinate of the gas turbine center of mass.

[0025] Optionally, modal analysis and vibration response characteristic analysis are performed based on mechanical vibration theory to determine the dynamic differential equation, specifically including:

[0026] Based on the mechanical vibration theory, the motion information data of the suspension under vertical vibration is determined with the coordinates of the gas turbine mass center as the reference; the motion information data includes: inertia parameters and kinetic energy;

[0027] Determine a characteristic root equation according to the motion information;

[0028] Determining modal parameters based on the characteristic root equation; the modal parameters include: natural frequency and vibration mode;

[0029] According to the dynamic response parameters and the modal parameters, based on the dual-mass vibration model, the vibration relative displacement is determined; the dual-mass vibration model is a physical model constructed based on the suspension mass, suspension stiffness, suspension damping, the mass of the blade and the stiffness of the shaft; the dynamic response parameters include: suspension equivalent stiffness and equivalent linear damping;

[0030] The dynamic differential equation is determined according to the vibration relative displacement.

[0031] Optionally, the expression of the dynamic differential equation is:

[0032]

[0033] Among them, m1 is the suspended mass; m2 is the mass of the blade; is the vertical acceleration of the suspension; is the vertical acceleration of the axis; z1 is the vertical displacement of the suspension; z2 is the vertical displacement of the axis; c1 is the suspension damping; is the vertical speed of the suspension; is the first-order derivative of the road roughness function; k1 is the suspension stiffness; k2 is the stiffness of the shaft; q is the road roughness function.

[0034] Optionally, determining a characteristic root equation according to the motion information specifically includes:

[0035] determining a mass matrix based on the kinetic energy;

[0036] determining elastic potential energy based on the inertia parameter;

[0037] determining a stiffness matrix based on the elastic potential energy;

[0038] determining a dynamic stiffness matrix according to the stiffness matrix;

[0039] The characteristic root equation is determined according to the dynamic stiffness matrix and the mass matrix.

[0040] Optionally, the expression of the elastic potential energy is:

[0041]

[0042] Where U is the elastic potential energy; k ti is the tangential stiffness of the rubber shock absorber at the ith suspension point; x i is the horizontal coordinate corresponding to the inertia parameter of the i-th suspension point; k ri is the radial stiffness of the rubber shock absorber at the i-th suspension point; i is the ordinate corresponding to the inertia parameter of the i-th suspension point; z i is the vertical coordinate corresponding to the inertia parameter of the i-th suspension point.

[0043] Optionally, the expression of the dynamic stiffness matrix is:

[0044]

[0045] The expression of the characteristic root equation is:

[0046]

[0047] Among them, K d is the dynamic stiffness matrix; n d is the vibration reduction parameter; K is the stiffness matrix; k ti is the tangential stiffness of the rubber shock absorber at the i-th suspension point; k ri is the radial stiffness of the rubber shock absorber at the i-th suspension point; M is the mass matrix; ω n is the natural frequency; det() is the determinant of the matrix.

[0048] Optionally, the suspension stiffness solution constraint conditions specifically include:

[0049]

[0050] Wherein, P1 is the vibration response amplitude of the gas turbine disk assembly; P2 is the vibration response amplitude of the power turbine disk assembly; P 壳 is the vibration response amplitude of the powertrain housing; 11 is the vertical radial clearance between the blade tip of the gas turbine disk assembly and the gas blade shroud; z 22 It is the vertical radial clearance between the blade tip of the power turbine disk assembly and the power turbine guide vane.

[0051] In a second aspect, the present application provides a suspension design device suitable for a vehicle gas turbine, comprising:

[0052] An information acquisition module is used to acquire inertia parameters of the gas turbine; the inertia parameters include: the coordinates of the suspension point position and the coordinates of the gas turbine center of mass; the inertia parameters are determined based on the powertrain inertia reference coordinate system; the powertrain inertia reference coordinate system takes the intersection of the gas generator rotor axis and the gearbox rear machine and the compressor intake machine installation surface as the coordinate origin O, the airflow direction is the positive direction of the X axis, the vertical direction is the Z axis, and the X axis, Y axis and Z axis form a right-handed coordinate system;

[0053] A constraint condition determination module, used to determine the stiffness solution constraint condition of vibration decoupling according to the inertia parameters by using the elastic center method;

[0054] The dynamic differential equation determination module is used to perform modal analysis and vibration response characteristic analysis based on mechanical vibration theory to determine the dynamic differential equation;

[0055] A suspension stiffness solution constraint determination module is used to calculate the relative displacement expression between the blade and the wall surface based on the dynamic differential equation and adopt the standard road surface spectrum excitation of the design requirement level, and determine the stiffness solution constraint condition to ensure the safety of the structure;

[0056] A stiffness range determination module is used to determine the safe stiffness range of each point of the gas turbine suspension according to the stiffness solution constraint conditions of vibration decoupling and the stiffness solution constraint conditions of ensuring structural safety;

[0057] A design stiffness value determination module is used to select a set of stiffness combinations that meet preset requirements within the safety stiffness range and perform verification to obtain a design stiffness value;

[0058] A matching module is used to perform nonlinear stiffness matching according to the designed stiffness value to obtain matching data; the matching data is parameter data for determining the suspension design of the vehicle gas turbine.

[0059] According to the specific embodiments provided in this application, this application discloses the following technical effects:

[0060] The present application provides a suspension design method and device suitable for vehicle gas turbines. According to the elastic center method, the stiffness solution constraint conditions required for vibration decoupling are designed; according to modal analysis and vibration response characteristic analysis, the standard road spectrum excitation of the design requirement level is used to design the relative displacement between the blade and the wall caused by the control vibration, and the stiffness solution constraint conditions to ensure the safety of the structure are determined; based on the above two aspects of stiffness solution constraint conditions, the safe stiffness range of each point of the suspension system is obtained. Within the safe stiffness range, a group of stiffness combinations that meet the preset requirements are selected and checked to obtain the design stiffness value, and then nonlinear stiffness matching is performed. That is, based on the iteration method, the nonlinear elastic characteristic curve of the rubber suspension is optimized and designed, and the suspension device is subsequently developed and designed using high temperature resistant damping rubber materials to ensure the normal operation of the vehicle-mounted gas turbine, thereby solving the technical problem that the vehicle gas turbine is prone to structural damage under vibration conditions, removing a major technical obstacle for its application and promotion, and allowing the gas turbine to give full play to the huge market application prospects and development potential in the automotive industry. Therefore, the present application can solve the problem that the vehicle gas turbine is prone to structural damage under vibration conditions and improve application reliability. BRIEF DESCRIPTION OF THE DRAWINGS

[0061] In order to more clearly illustrate the embodiments of the present application or the technical solutions in the prior art, the drawings required for use in the embodiments will be briefly introduced below. Obviously, the drawings described below are only some embodiments of the present application. For ordinary technicians in this field, other drawings can be obtained based on these drawings without paying creative work.

[0062] Figure 1 A flow chart of a method for designing a suspension for a vehicle gas turbine for the present application;

[0063] Figure 2 Schematic diagram of the gas turbine mount stiffness design process;

[0064] Figure 3 Construct a schematic for a right-handed coordinate system;

[0065] Figure 4 It is a schematic diagram of the physical model of the double-mass vibration system;

[0066] Figure 5 A simplified physical model schematic diagram of a power turbine working blade acting as a concentrated load at the center of mass of a rotating assembly;

[0067] Figure 6 A simplified physical model schematic diagram showing the mass of the turbine disk acting as a concentrated load at the center of mass of the rotating assembly;

[0068] Figure 7 A simplified diagram of the vibration isolation of road excitation force;

[0069] Figure 8 It is a schematic diagram of the root mean square value curve of vertical vibration acceleration of the powertrain under different levels of road excitation;

[0070] Fig. 9 A schematic diagram of the mounting device connecting the gas turbine to the vehicle frame. DETAILED DESCRIPTION

[0071] The following will be combined with the drawings in the embodiments of the present application to clearly and completely describe the technical solutions in the embodiments of the present application. Obviously, the described embodiments are only part of the embodiments of the present application, not all of the embodiments. Based on the embodiments in the present application, all other embodiments obtained by ordinary technicians in this field without creative work are within the scope of protection of this application.

[0072] In order to make the above-mentioned objects, features and advantages of the present application more obvious and easy to understand, the present application is further described in detail below with reference to the accompanying drawings and specific implementation methods.

[0073] The suspension stiffness design of the gas turbine should focus on the vibration of the gas turbine caused by ground unevenness excitation, and focus on the relative displacement between the blades and the wall caused by vibration. This application aims to isolate the vibration caused by road excitation when using a vehicle gas turbine, and designs a suspension design method suitable for a vehicle gas turbine. This method meets the basic requirements of the suspension design for vibration decoupling in all directions of the gas turbine powertrain, focuses on the collision damage between the gas turbine blades and the wall caused by vibration, ensures that the powertrain can work safely under certain road unevenness excitation, improves the application reliability of vehicle gas turbines, and opens up a feasible way for the application and promotion of vehicle gas turbines.

[0074] First, the technical terms appearing in this application are explained:

[0075] Suspension refers to a type of device that isolates the powertrain from the vehicle body to reduce the transmission of vibration and noise.

[0076] A gas turbine is a high-speed rotary power machine that converts the chemical energy or thermal energy of fuel into mechanical energy. It achieves energy conversion by driving the impeller to rotate at high speed through a continuous flow of gas (mainly air and gas).

[0077] Road roughness excitation refers to the dynamic impact of the ups and downs of the road surface on the vehicle's driving process. This impact is caused by the deviation between the road surface and the ideal plane (usually a horizontal and smooth plane). Road roughness excitation is the main source of vibration of the vehicle system.

[0078] Modal analysis: Mode is the natural vibration characteristic of mechanical structure. Each mode has a specific natural frequency, damping ratio and modal vibration shape. These modal parameters can be obtained through calculation or experimental analysis. This analysis process is called modal analysis.

[0079] Vibration decoupling refers to designing the suspension system so that the vibrations of the powertrain in multiple degrees of freedom are independent or separated from each other, thereby reducing the mutual influence between vibrations and improving the vibration isolation effect.

[0080] like Figure 1 As shown, the embodiment of the present application provides a suspension design method applicable to a vehicle gas turbine, the method comprising:

[0081] Step 100: Obtain the inertial parameters of the gas turbine. The inertial parameters include: the coordinates of the mounting point position and the coordinates of the gas turbine center of mass; the inertial parameters are determined based on the powertrain inertial reference coordinate system; the powertrain inertial reference coordinate system uses the intersection of the gas generator rotor axis and the gearbox rear machine and the compressor intake machine installation surface as the coordinate origin O, the airflow direction is the positive direction of the X axis, the vertical direction is the Z axis, and the X axis, Y axis and Z axis form a right-handed coordinate system.

[0082] Step 200: Using the elastic center method, determine the stiffness solution constraint conditions for vibration decoupling according to the inertia parameters.

[0083] Step 300: Perform modal analysis and vibration response characteristic analysis based on mechanical vibration theory to determine the dynamic differential equation.

[0084] Step 400: Based on the dynamic differential equation, the standard road surface spectrum excitation of the design requirement level is used to calculate the relative displacement expression between the blade and the wall, and determine the stiffness solution constraint conditions to ensure structural safety.

[0085] Step 500: Determine the safe stiffness range of each gas turbine mounting point according to the stiffness solution constraint conditions for vibration decoupling and the stiffness solution constraint conditions for ensuring structural safety.

[0086] Step 600: Within the safety stiffness range, a set of stiffness combinations that meet preset requirements is selected and verified to obtain a design stiffness value.

[0087] Step 700: Perform nonlinear stiffness matching according to the designed stiffness value to obtain matching data. The matching data is parameter data used to determine the mounting design of the vehicle gas turbine.

[0088] In one embodiment, the elastic center method is used to determine the stiffness solution constraint conditions for vibration decoupling based on inertia parameters, specifically including:

[0089] The center of mass reference coordinate system is determined according to the center of mass coordinates of the gas turbine.

[0090] Based on the centroid reference coordinate system, the coordinates of the suspension point position are converted into the centroid coordinates of the suspension point position.

[0091] According to the principle of vibration decoupling by the elastic center method, the coordinates of the elastic center are made the same as the values ​​corresponding to the coordinates of the gas turbine center of mass, and the stiffness solution constraint conditions for vibration decoupling are obtained; the coordinates of the elastic center are determined based on the coordinates of the suspension point position.

[0092] The stiffness solution constraint conditions for vibration decoupling include: solving the vertical stiffness constraint equation of the rubber shock-absorbing block.

[0093] The expression for solving the constraint equation of the vertical stiffness of the rubber shock absorber is:

[0094]

[0095] The coordinates of the elastic center are given by:

[0096]

[0097] Among them, x s is the horizontal coordinate of the elastic center; y s is the ordinate of the elastic center; k ri is the radial stiffness of the rubber shock absorber at the ith suspension point; x si y is the horizontal coordinate of the center of mass of the suspension point; si x is the vertical coordinate of the center of mass of the suspension point; c is the horizontal coordinate of the gas turbine mass center; c is the vertical coordinate of the gas turbine center of mass.

[0098] In one embodiment, modal analysis and vibration response characteristic analysis are performed based on mechanical vibration theory to determine the dynamic differential equation, which specifically includes:

[0099] Based on the mechanical vibration theory and taking the coordinates of the center of mass of the gas turbine as a reference, the motion information data of the suspension under vertical vibration is determined; the motion information data includes: inertia parameters and kinetic energy.

[0100] The characteristic root equation is determined according to the motion information; the modal parameters are determined based on the characteristic root equation; the modal parameters include: natural frequency and vibration mode.

[0101] According to the dynamic response parameters and modal parameters, the vibration relative displacement is determined based on the double-mass vibration model; the double-mass vibration model is a physical model constructed based on the suspension mass, suspension stiffness, suspension damping, the mass of the blade and the stiffness of the shaft; the dynamic response parameters include: suspension equivalent stiffness and equivalent linear damping.

[0102] Determine the dynamic differential equation based on the vibration relative displacement.

[0103] The expression of the dynamic differential equation is:

[0104]

[0105] Among them, m1 is the suspended mass; m2 is the mass of the blade; is the vertical acceleration of the suspension; is the vertical acceleration of the axis; z1 is the vertical displacement of the suspension; z2 is the vertical displacement of the axis; c1 is the suspension damping; is the vertical speed of the suspension; is the first-order derivative of the road roughness function; k1 is the suspension stiffness; k2 is the stiffness of the shaft; q is the road roughness function.

[0106] The characteristic root equation is determined according to the motion information, specifically including: determining the mass matrix according to the kinetic energy; determining the elastic potential energy according to the inertia parameter; determining the stiffness matrix according to the elastic potential energy; determining the dynamic stiffness matrix according to the stiffness matrix; determining the characteristic root equation according to the dynamic stiffness matrix and the mass matrix.

[0107] The expression of elastic potential energy is:

[0108]

[0109] Where U is the elastic potential energy; k ti is the tangential stiffness of the rubber shock absorber at the ith suspension point; x i is the horizontal coordinate corresponding to the inertia parameter of the i-th suspension point; k ri is the radial stiffness of the rubber shock absorber at the i-th suspension point; i is the ordinate corresponding to the inertia parameter of the i-th suspension point; z i is the vertical coordinate corresponding to the inertia parameter of the i-th suspension point.

[0110] The expression of the dynamic stiffness matrix is:

[0111]

[0112] The expression of the characteristic root equation is:

[0113]

[0114] Among them, K d is the dynamic stiffness matrix; n d is the vibration reduction parameter; K is the stiffness matrix; k ti is the tangential stiffness of the rubber shock absorber at the i-th suspension point; k ri is the radial stiffness of the rubber shock absorber at the i-th suspension point; M is the mass matrix; ω n is the natural frequency; det() is the determinant of the matrix, which is a representation method in linear algebra.

[0115] The constraints for solving the suspension stiffness include:

[0116]

[0117] Wherein, P1 is the vibration response amplitude of the gas turbine disk assembly; P2 is the vibration response amplitude of the power turbine disk assembly; P 壳 is the vibration response amplitude of the powertrain housing; 11 is the vertical radial clearance between the blade tip of the gas turbine disk assembly and the gas blade shroud; z 22 It is the vertical radial clearance between the blade tip of the power turbine disk assembly and the power turbine guide vane.

[0118] In practical applications, the design method of the present application includes five steps: vibration decoupling based on the elastic center method, suspension modal analysis, gas turbine modal analysis, suspension stiffness solution and verification. The suspension device includes three parts: a gas turbine conversion joint, a rubber shock absorber block and a system bracket.

[0119] Specifically, the gas turbine suspension stiffness design process described in this application is as follows: Figure 2 shown.

[0120] The position of the mounting point is determined according to the actual reserved mounting position of the gas turbine. The mounting position is generally symmetrically distributed left and right and evenly distributed front and back. In the powertrain inertial reference coordinate system (the coordinate system is generally defined as the intersection of the gas generator rotor axis and the gearbox rear machine and the compressor intake machine installation surface as the coordinate origin O, the airflow direction is the positive direction of the X axis, the vertical direction is the Z axis, and the upward direction is positive. The X, Y, and Z axes form a right-handed coordinate system, such as Figure 3 The coordinates of the mounting point and the center of mass of the gas turbine are obtained. The mass and center of mass coordinates of the gas turbine are shown in Table 1. The mounting arrangement position is shown in Table 2.

[0121] Table 1 Gas turbine mass and center of mass coordinates

[0122]

[0123] Table 2 Suspension layout position table

[0124] Location x-coordinate y coordinate z coordinate 1 -496.079 96.4325 -282.038 2 -496.129 -79.2678 -291.793 3 -44.499 208.45 -187.69 4 -44.499 -208.45 -187.69 5 420.101 0 -180

[0125] The position coordinates of the mounting point in the powertrain inertial reference coordinate system are converted into the position coordinates (x si ,y si ,z si ). The suspension arrangement position table with the center of mass as the origin is shown in Table 3.

[0126] Table 3 Suspension layout position table with the center of mass as the origin

[0127] Location x-coordinate y coordinate z coordinate 1 -500.990 92.499 -223.688 2 -501.040 -83.201 -233.443 3 -49.410 204.517 -129.340 4 -49.410 -212.383 -129.340 5 415.190 -3.933 -121.650

[0128] The design basis of the powertrain suspension system is vibration decoupling to eliminate vibration coupling in different directions. Usually, the six natural vibration modes of the engine suspension system are coupled in multiple degrees of freedom. Excitation in a certain degree of freedom will produce coupled vibration, which greatly widens the range of resonance frequency and increases the chance of resonance. If you want to achieve a better vibration isolation effect, you need to use softer suspension elements, which will cause a larger relative displacement between the engine powertrain and the surrounding parts, causing collisions with the surrounding parts and destroying the ride comfort of the whole vehicle. At the same time, the large displacement of the suspension element will increase the strain of the suspension element and affect its service life. Therefore, the design of modern automobile engine suspension is developing in the direction of complete decoupling or partial decoupling. Due to the difficulty of complete decoupling, the usual practice is to decouple several vibration modes. This application mainly considers acceleration and vertical displacement in three directions. Therefore, the elastic center method is used to decouple the vibration of the powertrain in some directions. The specific method is to make the coordinates of the center of mass of the gas turbine in the x and y directions coincide with the coordinates of the elastic center of the suspension in the x and y directions. The elastic center is determined by the stiffness and geometric arrangement of the elastic element and has nothing to do with the mass of the supported object. When the elastic center of the suspension is located at the center of mass of the engine suspension system, elastic coupling can be eliminated. The coordinates of the elastic center are expressed as (x s ,y s ), then according to the definition of elastic center:

[0129]

[0130] Among them, k ri is the radial stiffness of the rubber shock absorber at the i-th suspension point; i=1, 2, 3, 4, 5.

[0131] According to the principle of vibration decoupling by the elastic center method, the elastic center of the suspension system should be the same as the coordinate value of the gas turbine mass center, thereby obtaining the vertical stiffness of the rubber shock absorber to solve the constraint equation:

[0132]

[0133] The equation can also be expressed as:

[0134]

[0135] Modal analysis mainly analyzes the linear natural frequency and formation of the system, which is an important reference factor for the stiffness design of the suspension system, and is also the basis for the static and dynamic response analysis of the system. The modal analysis of the suspension system must first establish the dynamic differential equation of the suspension system, with the center of mass C as the origin. The coordinates describe the vertical vibration of the center of mass, with the positive direction pointing upwards, and the x c The coordinate is in the direction of airflow as the positive direction, x c coordinate, coordinate, The coordinates form a right-handed coordinate system.

[0136] The kinetic energy T can be expressed as:

[0137]

[0138] m c is the gas turbine mass; For x c Vibration speed on the coordinate axis; for Vibration speed on the coordinate axis; for Vibration speed on the coordinate axis.

[0139] The variable X is defined as:

[0140] X=[x c y c z c ] T (4)

[0141] The mass matrix M can be expressed as,

[0142]

[0143] Among them, ξ i ,ξ j Can represent One of them.

[0144] Assume that the axial deformation of each wire mesh shock absorber relative to the static equilibrium position is z i ,but

[0145] z i =z c (6)

[0146] The elastic potential energy U can be expressed as:

[0147]

[0148] Therefore, the stiffness matrix K can be expressed as:

[0149]

[0150] The ratio of dynamic elastic modulus Ed to static elastic modulus Es, nd = Ed / Es, is an important parameter. In vibration reduction design, nd is generally taken as 1 to 2. Therefore, the relationship between dynamic stiffness and static stiffness is:

[0151] k id =n d k i (9)

[0152] Among them, k id represents the dynamic stiffness of the rubber damping block at the ith suspension point, k i It represents the vertical static stiffness of the rubber shock absorber at the ith suspension point. d is the vibration reduction parameter.

[0153] According to equations (7), (8) and (9), the dynamic stiffness matrix K can be obtained: d :

[0154]

[0155] Characteristic root equation:

[0156]

[0157] Among them, ω n Indicates the natural frequency.

[0158] By solving the characteristic root equation, the natural frequency and vibration mode can be calculated. According to the mechanical vibration theory, to prevent the powertrain suspension system from amplifying the road excitation, the natural frequency of the suspension vertical vibration should be ensured to avoid the natural frequency of the wheel vertical vibration (10Hz~15Hz) and the first-order natural frequency of the vehicle suspension (0.9Hz~1.2Hz).

[0159] When designing the stiffness of the gas turbine suspension system, the static and dynamic response characteristics of the powertrain vibration should be analyzed. Among them, the static response characteristic analysis only considers the force of the suspension when the powertrain is subjected to a steady-state load. The static response excitation of the powertrain is its own gravity. Under the action of its own gravity acceleration g, the vertical displacement of the center of mass is z c , then the equilibrium equation of the system can be written as:

[0160]

[0161] m c g is the gravity on the gas turbine. Since the response under constant excitation is generally composed of transient response and steady-state response, but the transient response decays very quickly, it is generally ignored, and only the steady-state response is considered here. Then under constant excitation, when the system reaches a steady state, the first-order and second-order derivatives of the variable X mentioned above are zero, then:

[0162]

[0163] The design index constraint of gas turbine displacement response under gravity is set as z max ,Right now:

[0164] z c ≤z max (14)

[0165] When designing the suspension stiffness, analyzing the dynamic response of the gas turbine can evaluate its dynamic performance, such as vibration amplitude, vibration frequency, etc., to determine whether the structural design meets the design requirements. In the dynamic response analysis, the basic parameters required are the equivalent stiffness and damping characteristics of the suspension. The specific solution adopts appropriate simplification, and the process is as follows.

[0166] Since the stiffness of the rubber damping block is nonlinear, an equivalent linearization must be performed before the vibration mode and natural frequency analysis, where the elastic force F of the i-th rubber damping block is si and deformation δ i The relationship between can be expressed as:

[0167] F si =f si (δ i )(15)

[0168] Among them, f si is a nonlinear function.

[0169] According to the equivalent linearization principle, k i 、b i The square integral of the error should be minimized, that is:

[0170]

[0171] Where: i represents the error, (δ ai ,δ bi ) is the allowable working range of the ith rubber shock absorber.

[0172]

[0173] b i is the intercept of the linear stiffness function.

[0174] Ignoring the difference in stiffness between compression and recovery of the rubber shock absorber, the equivalent stiffness of the rubber shock absorber can be calculated.

[0175] Since the damping of the rubber shock absorber is hysteresis damping, an equivalent linearization must be performed before linear analysis, where the damping force F of the i-th rubber shock absorber is ci and deformation speed The relationship between can be expressed as:

[0176]

[0177] Since the damping is in the form of a loop, the equivalent linearization result is related to the deformation frequency and deformation amplitude. Therefore, the deformation circular frequency is selected as the natural circular frequency ω of the vertical vibration of the rigid body connected z , the amplitude X is selected as half of the allowable deformation, then the equivalent linear damping c i for:

[0178]

[0179] For ordinary nitrile rubber, the damping ratio of the rubber support is usually 0.05 to 0.15.

[0180] After obtaining the basic parameters for dynamic response calculation, the dynamic response analysis begins. The dynamic response analysis of the gas turbine suspension stiffness design focuses on the vibration relative displacement of the gas turbine power turbine blades, centrifugal impeller blades and gas turbine casing (including blade guard ring). The vibration of the power turbine shaft, centrifugal impeller shaft and power assembly casing (including blade guard ring) inside the power assembly can be equivalent to dual-mass vibration. The physical model of the dual-mass vibration system is as follows: Figure 4 As shown, the equivalent dual-mass model consists of suspension mass m1, suspension stiffness k1, suspension damping c1, blade mass m2, shaft stiffness k2, and the damping of the shaft and casing is neglected.

[0181] The input road roughness function is q(t), denoted as q, the vertical displacement of the suspension is z1(t), denoted as z1, the vertical displacement of the shaft is z2(t), denoted as z2, and the vertical velocity of the suspension is Recorded as The vertical speed of the axis is Recorded as The vertical acceleration of the suspension is Recorded as The vertical acceleration of the axis is Recorded as The differential equation of motion of the system is:

[0182]

[0183] Perform Fourier transform on both sides of equation (20), and we get:

[0184]

[0185] Let A1=k2,A2=-ω 2 m2+k2,A3=-ω 2 m1+k2+k1, ω is the generalized frequency in Fourier transform; from formula (21), the frequency response function of z2~z1 is:

[0186]

[0187] The frequency response function of z1~q is:

[0188]

[0189] The method for solving the stiffness of the shaft and the natural frequency of the blade vibration on the shaft is as follows. According to the general structure of the gas turbine, the suspension stiffness design focuses on the vibration of the gas turbine rotor assembly on the centrifugal impeller shaft and the vibration of the power turbine rotor assembly on the power turbine shaft. The centrifugal impeller shaft can be simplified as a cantilever beam, and the gas turbine rotor assembly is applied to the centrifugal impeller shaft as a uniformly distributed load; the power turbine shaft can be simplified as a simply supported beam, and the power turbine shaft mass load is evenly distributed. The mass of the power turbine working blades and turbine disks is applied as a concentrated load at the center of mass of the rotating assembly. The simplified physical model is as follows Figure 5 and Figure 6 shown.

[0190] in, Figure 5 The l1 in it represents the length of the centrifugal impeller shaft, and q' represents that the mass of the centrifugal impeller is evenly distributed on the centrifugal impeller shaft.

[0191] Figure 6 Here, l2 represents the length of the power turbine shaft, a represents the distance from the power turbine to the right fulcrum, b represents the distance from the power turbine to the left fulcrum, and F represents the gravity of the power turbine.

[0192] According to the material mechanics bending deformation calculation formula, the deflection w1 of the centrifugal impeller shaft at the center of mass of the gas turbine rotor assembly and the deflection w2 of the power turbine shaft at the center of mass of the power turbine rotor assembly can be calculated. According to the stiffness calculation formula The stiffness K1 of the centrifugal impeller shaft at the mass center of the gas turbine rotor assembly and the stiffness K2 of the power turbine shaft at the mass center of the power turbine rotor assembly are obtained. Where, F is force; x is displacement.

[0193] according to The natural frequency w of the gas turbine rotor assembly vibration is obtained n1 and the natural frequency w of the power turbine rotor assembly vibration n2 The natural frequency of the gas turbine casing (i.e. modal analysis) is obtained by using the finite element method. The modal analysis of the blade peripheral casing is performed using Ansys Workbench software to obtain the natural frequency of the vertical vibration of the casing w n壳 . m is the mass.

[0194] The road excitation of dynamic response analysis should select the corresponding standard road spectrum according to the vehicle design index. To ensure the design margin, the suspension natural frequency 1Hz is used as the excitation frequency, and the body displacement response is equivalent to a sine function with an amplitude of P0. According to equations (24)-(26), (27) and (28), the vibration response amplitude of the gas turbine disk assembly P1, the vibration response amplitude of the power turbine disk assembly P2, and the vibration response amplitude of the powertrain housing P 壳 . P0 is the displacement of the vehicle body; k 悬置 is the suspension stiffness; c 悬置 is the suspension damping ratio; m 动力叶片 is the mass of the power blade; m 动力 is the power mass; m 悬置 is the suspended mass; m 燃气轴 is the mass of the gas shaft; m 燃气叶片 is the mass of the gas blade; k 壳 is the shell stiffness.

[0195]

[0196] To ensure the safety margin, the suspension stiffness is designed to solve the constraint equation:

[0197]

[0198]

[0199] Specific clearance requirements are shown in Table 4.

[0200] Table 4 Requirements for clearance between blades and guard rings

[0201]

[0202] The above is the analysis of the design conditions of the gas turbine mount stiffness. Based on the vibration decoupling constraint of formula (2) and the blade clearance constraint of formula (27) and formula (28), the safe stiffness range of each point of the gas turbine mount can be calculated. Within the safe stiffness range, a genetic algorithm is used to select a group of stiffness combinations with better comprehensive characteristics, and the values ​​are shown in Table 5.

[0203] Table 5 Suspension design plan

[0204] Serial number Z-axis stiffness (N / mm) 1 227 2 100 3 100 4 148 5 424

[0205] After obtaining the suspension stiffness values ​​of each point in the design, the energy decoupling rate of each degree of freedom can be checked by using Adams multi-body dynamics simulation software. If the energy decoupling rate is above 80%, it can be considered that the vibration decoupling design effect is good. The static response characteristics of the system can be checked according to equations (12) to (14) to see whether they meet the design requirements, especially whether the static displacement under gravity is within the design index range. The radial clearance between the blade tip of the gas turbine disk assembly and the gas blade guard ring, and the radial clearance between the blade tip of the power turbine disk assembly and the power turbine guide can be checked according to equations (24) to (28) to see whether they meet the design requirements. In the dynamic response analysis verification stage, it is also necessary to check whether the root mean square value of the vertical vibration acceleration of the powertrain meets the requirements under different levels of road excitation. The verification calculation process is as follows.

[0206] Since the suspension system is much smaller than the vehicle body and is arranged at the center of the vehicle, the pitch vibration and roll vibration of the vehicle body can be ignored. Since the vibrations in the three coordinate axes of the suspension system have been decoupled, only the influence of vertical vibration on the system is considered. For the convenience of calculation, the vehicle body is simplified to a single-degree-of-freedom vibration system. Assuming that the damping ratio of the suspension system is 0.2, the natural frequency of the system ω n =2πf n =2π(0.9~1.2), the damping coefficient is about c=4.11×10 3 , the road roughness is A to D, the vehicle speed is v, then the transfer function of the suspension is H s (jω) can be written as:

[0207]

[0208] Where r is the frequency ratio, r = ω / ω n ζ s is the damping ratio of the suspension; r s is the frequency ratio of the suspension.

[0209] Pavement Road Spectrum G q (ω) can be expressed as:

[0210]

[0211] G q (n0) is the power spectral density function of the road surface roughness at the reference spatial frequency n0, which is called the road surface roughness coefficient; n0 is the reference spatial frequency.

[0212] Then the excitation spectrum of the vehicle body to the suspension support system can be expressed as:

[0213]

[0214] in, H sThe conjugate of (jω).

[0215] The transfer function of the suspension can be written as:

[0216]

[0217] ζ t is the damping ratio of the suspension; r t is the suspension frequency ratio.

[0218] The response spectrum of the suspension support system can be expressed as:

[0219]

[0220] The mean square value of acceleration is:

[0221]

[0222] See the schematic diagram of road excitation force isolation for Figure 7 According to equations (31) to (37), the root mean square value of the vertical vibration acceleration of the powertrain under different levels of road excitation can be obtained as follows: Figure 8 As shown, it can be judged that the stiffness design meets the performance requirements. is the square of the vibration transfer function; H t It is the abbreviation of suspension vibration transfer function; H s is the abbreviation of the suspension vibration transfer function; v(n0) is the vehicle speed. t (jω) is the suspension vibration transfer function; is the transpose of the suspension vibration transfer function; H s (jω) is the suspension vibration transfer function; is the transpose of the suspension vibration transfer function.

[0223] After the suspension stiffness design is verified, the nonlinear stiffness matching is performed by using an iterative method based on the designed stiffness value. Thus, a suspension device connecting the gas turbine and the frame can be designed. The schematic diagram of the device is shown in the figure. Fig. 9 shown.

[0224] Fig. 9 In the figure, 1 is a gas turbine conversion joint, 2 is a suspension bracket, and 3 is a rubber shock-absorbing block.

[0225] The suspension system is designed with the rubber shock absorber block 3 as the core. The rubber shock absorber block 3 is cylindrical in shape with a through hole in the middle. On the one hand, it provides space for rubber compression and is also a bolt connection hole. The rubber shock absorber block 3 and the suspension bracket 2 are connected by bolts, and the other end of the suspension bracket 2 is connected to the frame plane by four bolts. Because the rubber shock absorber block 3 has a pre-compression requirement when it is installed to ensure that the stiffness is within the design range, the suspension bracket 2 can only be connected to it on one side of the rubber shock absorber block 3, and a fork-shaped structure cannot be used to connect it to both sides of the rubber shock absorber block 3. The suspension bracket 2 has a boss at the part connected to the rubber shock absorber block 3 to ensure the connection strength; the suspension bracket 2 is similar to a cantilever design as a whole, and the position connected to the suspension is subjected to a large force displacement. In order to ensure the reliable strength design of the part, a rib is designed for the weak part of the bracket. In addition, the static analysis based on the finite element method verifies the stiffness of the part. On the other side of the rubber shock absorber block 3 is the gas turbine conversion section 1, which is a transition piece between the gas turbine and the suspension system. An interference fit is adopted between the gas turbine conversion section 1 and the rubber shock-absorbing block 3, which ensures the reliability of the connection on the one hand, and the pre-compression amount of the suspension installation on the other hand to meet the stiffness design requirements; the other side of the gas turbine conversion section 1 is connected to the suspension installation interface reserved for the gas turbine through four screws.

[0226] Based on the same inventive concept, the embodiment of the present application also provides a suspension design device for vehicle gas turbines for implementing the above-mentioned suspension design method for vehicle gas turbines. The implementation scheme for solving the problem provided by the device is similar to the implementation scheme recorded in the above-mentioned method, so the specific limitations in one or more embodiments of the suspension design device for vehicle gas turbines provided below can refer to the limitations of the suspension design method for vehicle gas turbines above, and will not be repeated here.

[0227] A suspension design device applicable to a vehicle gas turbine, the suspension design device applicable to a vehicle gas turbine comprising:

[0228] The information acquisition module is used to acquire the inertia parameters of the gas turbine; the inertia parameters include: the coordinates of the suspension point position and the coordinates of the gas turbine center of mass; the inertia parameters are determined based on the powertrain inertia reference coordinate system; the powertrain inertia reference coordinate system takes the intersection of the gas generator rotor axis and the gear box rear machine and the compressor intake machine installation surface as the coordinate origin O, the air flow direction is the positive direction of the X axis, the vertical direction is the Z axis, and the X axis, Y axis and Z axis form a right-handed coordinate system.

[0229] The constraint condition determination module is used to determine the stiffness solution constraint condition of vibration decoupling according to the inertia parameters by using the elastic center method.

[0230] The dynamic differential equation determination module is used to perform modal analysis and vibration response characteristic analysis based on mechanical vibration theory to determine the dynamic differential equation.

[0231] The suspension stiffness solution constraint determination module is used to calculate the relative displacement expression between the blade and the wall based on the dynamic differential equation and adopt the standard road surface spectrum excitation of the design requirement level, and determine the stiffness solution constraint condition to ensure the safety of the structure.

[0232] The stiffness range determination module is used to determine the safe stiffness range of each point of the gas turbine suspension according to the stiffness solution constraint conditions of vibration decoupling and the stiffness solution constraint conditions of ensuring structural safety.

[0233] The design stiffness value determination module is used to select a set of stiffness combinations that meet preset requirements within the safety stiffness range and perform verification to obtain the design stiffness value.

[0234] A matching module is used to perform nonlinear stiffness matching according to the designed stiffness value to obtain matching data; the matching data is parameter data for determining the suspension design of the vehicle gas turbine.

[0235] The technical features of the above embodiments may be combined arbitrarily. To make the description concise, not all possible combinations of the technical features in the above embodiments are described. However, as long as there is no contradiction in the combination of these technical features, they should be considered to be within the scope of this specification.

[0236] This article uses specific examples to illustrate the principles and implementation methods of this application. The description of the above embodiments is only used to help understand the method and core ideas of this application. At the same time, for those skilled in the art, according to the ideas of this application, there will be changes in the specific implementation methods and application scope. In summary, the content of this specification should not be understood as limiting this application.

Claims

1. A method for designing a mounting structure for a vehicle gas turbine, characterized in that: The suspension design method applicable to a vehicle gas turbine includes: Obtaining inertial parameters of the gas turbine; the inertial parameters include: the coordinates of the suspension point position and the coordinates of the gas turbine center of mass; the inertial parameters are determined based on the powertrain inertial reference coordinate system; the powertrain inertial reference coordinate system takes the intersection of the gas generator rotor axis and the gearbox rear machine and the compressor intake machine installation surface as the coordinate origin O, the airflow direction is the positive direction of the X axis, the vertical direction is the Z axis, and the X axis, Y axis and Z axis form a right-handed coordinate system; Using the elastic center method, the stiffness solution constraint condition of the vibration decoupling is determined according to the inertia parameters; Perform modal analysis and vibration response characteristic analysis based on mechanical vibration theory to determine the dynamic differential equation; Based on the dynamic differential equation, the standard road surface spectrum excitation of the design requirement level is adopted to calculate the relative displacement expression between the blade and the wall surface, and the stiffness solution constraint condition to ensure the safety of the structure is determined; According to the stiffness solution constraint conditions for vibration decoupling and the stiffness solution constraint conditions for ensuring structural safety, the safe stiffness range of each point of the gas turbine mounting is determined; Within the safety stiffness range, a set of stiffness combinations that meet the preset requirements are selected and checked to obtain the design stiffness value; Nonlinear stiffness matching is performed according to the designed stiffness value to obtain matching data; the matching data is parameter data for determining the suspension design of the vehicle gas turbine.

2. The method for designing a mounting structure for a vehicle gas turbine according to claim 1, characterized in that: The elastic center method is used to determine the stiffness solution constraint conditions for vibration decoupling according to the inertia parameters, specifically including: Determine the mass center reference coordinate system according to the gas turbine mass center coordinates; Based on the centroid reference coordinate system, the coordinates of the suspension point position are converted into the centroid coordinates of the suspension point position; According to the principle of vibration decoupling by the elastic center method, the coordinates of the elastic center are made the same as the values ​​corresponding to the coordinates of the gas turbine center of mass, and the stiffness solution constraint conditions for vibration decoupling are obtained; the coordinates of the elastic center are determined based on the coordinates of the suspension point position.

3. The method for designing a mounting structure for a vehicle gas turbine according to claim 2, characterized in that: The stiffness solution constraint conditions for vibration decoupling include: solving the constraint equation of the vertical stiffness of the rubber shock-absorbing block; The expression for solving the constraint equation of the vertical stiffness of the rubber damping block is: The coordinates of the elastic center are given by: Among them, x s is the horizontal coordinate of the elastic center; y s is the ordinate of the elastic center; k ri is the radial stiffness of the rubber shock absorber at the ith suspension point; x si y is the horizontal coordinate of the center of mass of the suspension point; si x is the vertical coordinate of the center of mass of the suspension point; c is the horizontal coordinate of the gas turbine mass center; c is the vertical coordinate of the gas turbine center of mass.

4. The method for designing a mounting structure for a vehicle gas turbine according to claim 1, characterized in that: Based on the mechanical vibration theory, modal analysis and vibration response characteristics analysis are carried out to determine the dynamic differential equation, including: Based on the mechanical vibration theory, the motion information data of the suspension under vertical vibration is determined with the coordinates of the gas turbine mass center as the reference; the motion information data includes: inertia parameters and kinetic energy; Determine a characteristic root equation according to the motion information; Determining modal parameters based on the characteristic root equation; the modal parameters include: natural frequency and vibration mode; According to the dynamic response parameters and the modal parameters, based on the dual-mass vibration model, the vibration relative displacement is determined; the dual-mass vibration model is a physical model constructed based on the suspension mass, suspension stiffness, suspension damping, the mass of the blade and the stiffness of the shaft; the dynamic response parameters include: suspension equivalent stiffness and equivalent linear damping; The dynamic differential equation is determined according to the vibration relative displacement.

5. The method for designing a mounting structure for a vehicle gas turbine according to claim 4, characterized in that: The expression of the dynamic differential equation is: Among them, m1 is the suspended mass; m2 is the mass of the blade; is the vertical acceleration of the suspension; is the vertical acceleration of the axis; z1 is the vertical displacement of the suspension; z2 is the vertical displacement of the axis; c1 is the suspension damping; is the vertical speed of the suspension; is the first-order derivative of the road roughness function; k1 is the suspension stiffness; k2 is the stiffness of the shaft; q is the road roughness function.

6. The method for designing a mounting structure for a vehicle gas turbine according to claim 4, characterized in that: Determining the characteristic root equation according to the motion information specifically includes: determining a mass matrix based on the kinetic energy; determining elastic potential energy based on the inertia parameter; determining a stiffness matrix based on the elastic potential energy; determining a dynamic stiffness matrix according to the stiffness matrix; The characteristic root equation is determined according to the dynamic stiffness matrix and the mass matrix.

7. The method for designing a mounting structure for a vehicle gas turbine according to claim 5, characterized in that: The expression of the elastic potential energy is: Where U is the elastic potential energy; k ti is the tangential stiffness of the rubber shock absorber at the ith suspension point; x i is the horizontal coordinate corresponding to the inertia parameter of the i-th suspension point; k ri is the radial stiffness of the rubber shock absorber at the i-th suspension point; i is the ordinate corresponding to the inertia parameter of the i-th suspension point; z i is the vertical coordinate corresponding to the inertia parameter of the i-th suspension point.

8. The method for designing a mounting structure for a vehicle gas turbine according to claim 5, characterized in that: The expression of the dynamic stiffness matrix is: The expression of the characteristic root equation is: Among them, K d is the dynamic stiffness matrix; n d is the vibration reduction parameter; K is the stiffness matrix; k ti is the tangential stiffness of the rubber shock absorber at the i-th suspension point; k ri is the radial stiffness of the rubber shock absorber at the i-th suspension point; M is the mass matrix; ω n is the natural frequency; det() is the determinant of the matrix.

9. The method for designing a mounting structure for a vehicle gas turbine according to claim 1, characterized in that: The constraints for solving the suspension stiffness include: Wherein, P1 is the vibration response amplitude of the gas turbine disk assembly; P2 is the vibration response amplitude of the power turbine disk assembly; P 壳 is the vibration response amplitude of the powertrain housing; 11 is the vertical radial clearance between the blade tip of the gas turbine disk assembly and the gas blade shroud; z 22 It is the vertical radial clearance between the blade tip of the power turbine disk assembly and the power turbine guide vane.

10. A suspension design device suitable for a vehicle gas turbine, characterized in that: The suspension design device suitable for a vehicle gas turbine comprises: An information acquisition module is used to acquire inertia parameters of the gas turbine; the inertia parameters include: the coordinates of the suspension point position and the coordinates of the gas turbine center of mass; the inertia parameters are determined based on the powertrain inertia reference coordinate system; the powertrain inertia reference coordinate system takes the intersection of the gas generator rotor axis and the gearbox rear machine and the compressor intake machine installation surface as the coordinate origin O, the airflow direction is the positive direction of the X axis, the vertical direction is the Z axis, and the X axis, Y axis and Z axis form a right-handed coordinate system; A constraint condition determination module, used to determine the stiffness solution constraint condition of vibration decoupling according to the inertia parameters by using the elastic center method; The dynamic differential equation determination module is used to perform modal analysis and vibration response characteristic analysis based on mechanical vibration theory to determine the dynamic differential equation; A suspension stiffness solution constraint determination module is used to calculate the relative displacement expression between the blade and the wall surface based on the dynamic differential equation and adopt the standard road surface spectrum excitation of the design requirement level, and determine the stiffness solution constraint condition to ensure the safety of the structure; A stiffness range determination module is used to determine the safe stiffness range of each point of the gas turbine suspension according to the stiffness solution constraint conditions of vibration decoupling and the stiffness solution constraint conditions of ensuring structural safety; A design stiffness value determination module is used to select a set of stiffness combinations that meet preset requirements within the safety stiffness range and perform verification to obtain a design stiffness value; A matching module is used to perform nonlinear stiffness matching according to the designed stiffness value to obtain matching data; the matching data is parameter data for determining the suspension design of the vehicle gas turbine.

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