CN115203898A - Gear box coupling dynamics modeling method excited by internal multi-source fault - Google Patents
Gear box coupling dynamics modeling method excited by internal multi-source fault Download PDFInfo
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Abstract
The invention discloses a gearbox coupling dynamics modeling method excited by internal multi-source faults, which comprises the following steps: s1: calculating the gear meshing time-varying rigidity, damping and dynamic friction force: calculating to obtain a functional relation of the gear meshing rigidity along with the rotation angle of the driving wheel according to the gear appearance and the local peeling fault parameters; combining a mixed elastohydrodynamic lubrication theory, and calculating dynamic friction force and meshing damping of a current gear meshing pair at each moment by inputting dynamic gear meshing force, surface roughness and entrainment speed parameters; s2: calculating the acting force of the inner ring and the outer ring of each bearing in the gearbox; s3: and (4) constructing a complete equation of the whole gearbox dynamic model and solving. The invention can accurately obtain the vibration response of the gear box under the influence of multiple factors such as meshing force, meshing friction, dynamic acting force of the inner ring and the outer ring of the bearing, elastohydrodynamic lubrication, fault excitation and the like.
Description
Technical Field
The invention relates to the technical field of health state evaluation and fault diagnosis of mechanical equipment, in particular to a gear box coupling dynamics modeling method excited by internal multi-source faults.
Background
The transmission gearbox is one of the key parts of the mechanical system, the state of the transmission gearbox directly affects the operation state of the mechanical system, and once the gearbox fails, huge economic loss is likely to be caused, and even casualties are caused. Signal processing methods such as mechanical fault feature extraction are effective ways for diagnosing faults of the gearbox, and the fault diagnosis methods usually need prior knowledge of fault features as theoretical basis. The experimental verification is one of the main methods for obtaining the prior knowledge of the fault characteristics, but the experimental verification of a specific gearbox often faces the problems of high cost, uncertain fault parameters, large system interference and the like, so that the method for analyzing the fault mechanism by obtaining the simulation signal through the dynamic modeling is very important. The internal stress condition of the gearbox is relatively complex, the failure modes are variable, and the traditional gearbox dynamic model is used for researching the failure characteristics of one part by simplifying other parts (such as bearings). The model has certain referential property for fault mechanism research but cannot accurately reflect the actual behavior of the whole gearbox and the coupling action between components, and has a larger improvement space.
Disclosure of Invention
The invention aims to provide a gearbox coupling dynamics modeling method excited by internal multi-source faults so as to solve the problems in the prior art.
In order to achieve the purpose, the invention provides the following technical scheme: a gearbox coupling dynamics modeling method excited by internal multi-source faults comprises the following steps:
s1: calculating the gear meshing time-varying rigidity, damping and dynamic friction force: calculating to obtain a functional relation of the gear meshing rigidity along with the rotation angle of the driving wheel according to the gear appearance and the local peeling fault parameters; combining a mixed elastohydrodynamic lubrication theory, and calculating dynamic friction force and meshing damping of a current gear meshing pair at each moment by inputting dynamic gear meshing force, surface roughness and entrainment speed parameters;
s2: calculating the inner and outer ring acting force of each bearing in the gearbox: inputting the relative displacement of the inner ring and the outer ring at the current moment, and obtaining the normal deformation and the tangential speed of all contact pairs in the bearing through the revolution and rotation relation of the quasi-static rolling elements; according to structural parameters and surface parameters of the bearing, the normal force of each rolling element in the bearing, which is considered to be damped, can be calculated by combining a mixed elastohydrodynamic lubrication theory and a Hertz contact theory; synthesizing the normal forces to obtain dynamic bearing inner and outer ring acting forces;
s3: and (3) constructing a complete equation of the whole gearbox dynamic model and solving: introducing a gear meshing transmission error, a driving motor, a brake and a gear box shell on the basis of the steps; establishing a complete gearbox dynamic differential equation set through Newton's second law; solving the differential equation set by using a numerical solving method to obtain the dynamic response of the system;
s4: simulation signal fault characteristic analysis: the method comprises the steps of selecting the most significant shell acceleration from a dynamic numerical solution of a gearbox system as a gearbox vibration simulation signal, analyzing the simulation signal through time domain analysis, power spectrum and envelope spectrum analysis and other basic means, and revealing the vibration behavior of the gearbox under a coupling fault condition.
Preferably, the specific step of S1 includes:
s11: local gear failure is introduced by changes in the geometric parameters. The invention introduces local faults in the gear through surface peeling; the surface exfoliation morphology is described as elliptical with a certain depth h s Width w max And length l max (ii) a Such surface flaking can affect the effective contact length L of the gears during meshing s And cross section properties at the time of stiffness calculation;
wherein x s_center Denotes the position of the center coordinate, x, of the peeled ellipse s_start Denotes the starting coordinate position, x, of the peeling ellipse s_end Represents the end coordinate position of the peeling ellipse; theta.theta. s The tilt angle representing the peeling ellipse; the numerical value of the part affected by the fault can affect the calculation of the meshing rigidity, and the whole dynamic system is affected through the meshing rigidity, so that the purpose of introducing the gear fault is achieved;
s12: the time varying stiffness of the meshing gear is calculated. The gear teeth are regarded as variable cross-section cantilever beams, and the Hertzian contact rigidity k of a certain gear tooth is obtained through energy method derivation h (iii) bending stiffness k b Shear stiffness k s Axial compression stiffness k a And angle base stiffness k f (ii) a Obtaining the comprehensive series rigidity of the two gear teeth in a meshing state:
because more than one pair of gear teeth are in contact in the gear meshing process, the two pairs of gear teeth are in contact simultaneously, the series rigidity of the two pairs of gear teeth needs to be connected in parallel, and finally, the meshing rigidity function related to the rotation angle of the driving wheel is obtained:
s13: calculating elastohydrodynamic lubrication damping and friction coefficient of gear engagement; firstly according to lubrication conditions and physicsParameter, calculating the current thickness h of lubricating oil film of the meshing pair c ：
Wherein, K Hc The surface micro-peak shape parameter is set as 1; w is a dimensionless load parameter; u is a dimensionless speed parameter; g is a dimensionless material parameter; r is the equivalent contact radius of curvature;
is the dimensionless surface roughness, V is the dimensionless hardness parameter;
the contact damping c of the gear contact pair can be obtained by approximate calculation according to the oil film thickness, the lubricant viscosity and the contact area film ：
According to the friction theory under mixed elastohydrodynamic lubrication conditions in gear meshing, the dynamic friction coefficient can be calculated by the following formula:
the intermediate parameter can be obtained by the following formula:
wherein u pi And u gi The surface speeds of the driving wheel and the driven wheel in each pair of gear contact pairs are respectively; s is the root mean square surface roughness; p is h Is Hertz contact pressure, b 1 To b 9 Is a constant value coefficient:
b 1-9 ＝-8.92,1.03,1.04,-0.35,2.81,-0.10,0.75,-0.39,0.62
due to the gear engagement processThe two pairs of teeth are meshed, so that friction force needs to be calculated on each pair of meshed gear teeth respectively and then synthesized; first according to the load distribution coefficient S ri The load on each pair of teeth is calculated:
then, a calculation formula of the friction force and the friction torque can be obtained:
wherein subscript p indicates the action on the drive wheel and subscript g indicates the action on the driven wheel;
in summary, step 1 provides a way to calculate the forces generated by the meshing parts of the gears, which fully describe the dynamic behavior of the meshing of the gears, forming part of the meshing of the gears in the global model.
Preferably, the specific step of S2 includes:
s21: calculating the equivalent rigidity and damping of the bearing rolling body; connecting the inner contact pair and the outer contact pair of each rolling body of the bearing in series to be equivalent to a spring damping system; therefore, the rigidity and the damping under each contact pair need to be firstly solved; the hertzian contact stiffness of a bearing can be obtained by the formula:
wherein E b ，ν b And Σ ρ is the young's modulus, poisson's ratio, and raceway curvature sum, respectively; delta. For the preparation of a coating * A dimensionless deformation coefficient; according to the research related to the elastohydrodynamic lubrication, the elastohydrodynamic lubrication damping of the rolling body is generated by an inlet region of a lubricating oil film, and the inlet region lubrication damping can be calculated by the following formula:
wherein eta 0 Is the lubricating viscosity; r is x Is the effective curvature in the direction of entrainment; a is the major semi-axis length of the contact ellipse; thickness h of point contact lubricating oil film c Can be calculated from the following formula:
h c ＝2.69U 0.67 G 0.53 Q -0.067 (1-0.61e -0.73κ )R x (13)
because the bearing raceway contact pair is not completely elastic lubricated and the influence of the surface condition on the thickness of an oil film is considered, the following correction coefficients are introduced in the calculation of the film thickness:
wherein C and r are constant coefficients with values of 0.87 and 1.5, respectively; h is cT Namely the corrected film thickness;
after the rigidity and the damping of each contact pair are obtained, the equivalent series rigidity damping of the rolling body can be obtained through a complex rigidity series formula:
wherein
Linear stiffness k in formula 1,2 ＝k b δ 0.5 Linear damping c = c e (ii) a Natural frequency of free vibration of the omega contact pair;
s22: calculating the contact force of the inner ring and the outer ring of the bearing according to the contact geometry and the local fault; the local fault in the bearing raceway is described using a displacement excitation of half-sinusoidal profile, the displacement excitation equation being as follows:
wherein R is r Is the rolling element radius; l is defect Is the local fault length; phi is a unit of l Is the angle range corresponding to the fault length;
is the angle of the jth rolling element with respect to the fault position, and can be calculated by:
wherein phi d Is the starting position angle of the fault; the force between the inner ring and the outer ring of the bearing can be calculated by the following formula:
wherein lambda (delta) is a control function, when delta is a positive value, the output of the function is 1, and the output of the function is 0 under other conditions, so that the contact force is generated only when the inner ring and the outer ring are in a contact state; rolling body position angle psi j ＝2π(j-1)/N b +ψ c (ii) a And ψ c ＝0.5(1-D/D m )θ i Wherein theta i Representing the inner ring rotation angle, which is obtained by the geometrical relationship of the bearing; delta. For the preparation of a coating j The deformation amount indicating the jth rolling element position can be calculated by the following equation:
in the formula c 0 Indicating radial play.
Preferably, the S3 specifically includes:
establishing a complete gearbox dynamic model; since industrial gearboxes contain deep groove ball bearings and spur gears, the vibrations of the gearbox are mainly caused by radial excitation; all axial degrees of freedom of the gearbox are therefore neglected, and bending of the rotating shaft and wobbling motion of the shaft are not taken into account due to the relatively short shaft length of the gearbox; each shaft has freedom of movement and freedom of rotation in X and Y directions; the gearbox shell is fixedly connected with all bearing outer rings and has translational freedom degrees in two directions; the bearing inner ring is fixedly connected with the rotating shaft and rotates along with the rotating shaft; the gear box comprises four bearings, two bearings are positioned on the driving shaft, and two bearings are positioned on the driven shaft; the gearbox can be constructed into a nine-degree-of-freedom dynamic model, and the system equation is as follows:
wherein theta is L 、θ p And theta g The dynamic rotation angles of the load, the driving wheel and the driven wheel are respectively; m is p 、m g And m f The mass of the driving wheel, the driven wheel and the gear box shell is sequentially represented; i is L 、I p And I g Sequentially representing the rotational inertia of the load, the driving wheel and the driven wheel; t is Load Is the moment of force exerted on the load,
is the rotational speed of the motor. F pg Is the dynamic engagement force, and can be calculated by the following formula:
wherein k is m And c m Is the time-varying meshing stiffness and damping calculated in S1; ζ and ξ are the geometric deviations and gear transfer errors caused by tooth flank exfoliation, respectively;
in general, step 2 gives the force between the inner and outer races of the bearing; the acting force fully represents the dynamic characteristics of the bearing, and can accurately describe the dynamic action of the bearing in the gearbox system.
Preferably, the S3 specifically includes:
analyzing the fault characteristics of the simulation signals; according to the numerical result obtained by the differential equation, the vibration response of the gearbox system in a period of time can be obtained; through time domain analysis of the response, the change characteristics of the physical variables in the gearbox in the whole process can be known; meanwhile, the characteristics of the vibration of the gearbox in the time domain under the failure mode can be analyzed and obtained; converting the time domain acceleration signal into a frequency domain signal by using Fourier transform to obtain a power spectrum of the simulation signal, and obtaining frequency domain responses of the gearbox in different modes through the power spectrum; analyzing the embodiment of different fault modes on the frequency domain response of the gearbox, and summarizing to obtain the characteristics of different fault modes on a power spectrum; obtaining an envelope curve of an original simulation signal through Hilbert transform, and then performing Fourier transform on the envelope curve to obtain an envelope spectrum of the simulation signal; the characteristic frequency components of the vibration of the gearbox can be obtained more clearly through the envelope spectrum, and the frequency characteristics of the envelope spectrum of different fault modes are determined.
A computer device comprising a memory, a processor and a computer program stored on the memory and executable on the processor, the processor implementing the steps of any of the methods when executing the program.
A computer-readable storage medium, on which a computer program is stored which, when being executed by a processor, carries out the steps of any of the methods.
A processor for running a program, wherein the program when running performs any of the methods.
Compared with the prior art, the invention has the beneficial effects that:
1. firstly, establishing a gear box dynamic model fully considering gear engagement, bearing action and a gear box shell; the method has accurate mathematical description on local faults of a bearing, meshing geometrical characteristics of a gear and local faults of the gear in the gear box, and can accurately obtain the vibration response of the gear box under the influence of multiple factors such as meshing force, meshing friction, dynamic acting force of an inner ring and an outer ring of the bearing, elastohydrodynamic lubrication, fault excitation and the like;
2. compared with the traditional dynamic model modeling method, the invention fully considers the mathematical description method of the local faults of the bearing and the gear; by analyzing the vibration of the gearbox under different local failure modes and combining the traditional Fourier transform and envelope spectrum analysis methods, the frequency domain characteristics of the vibration of the gearbox shell under the excitation of internal multi-sources can be obtained; compared with the traditional model, the method provided by the invention can provide a transverse comparison method for responses under excitation of defects at different positions of a bearing, a gear and the like, and provides deeper mechanism analysis for fault diagnosis of the gearbox.
Drawings
The accompanying drawings, which are included to provide a further understanding of the invention and are incorporated in and constitute a part of this specification, illustrate embodiments of the invention and together with the description serve to explain the principles of the invention and not to limit the invention. In the drawings:
FIG. 1 is an internal portion of a gearbox dynamics model of the present invention;
FIG. 2 is an external portion of the gearbox dynamics model of the present invention;
FIG. 3 is a flow chart of a gearbox dynamics modeling method of internal multi-source coupled fault excitation according to the present invention;
FIG. 4 is a vibration response of a gearbox housing for a partial failure of a driven wheel and a failure of a bearing cup in accordance with the present invention;
FIG. 5 is a vibration response of a gearbox housing with a partial failure of a driven wheel and a failure of a bearing inner race according to the present invention;
FIG. 6 is a high speed shaft bearing inner race fault and driven wheel fault experimental verification of the present invention;
FIG. 7 is a high speed shaft bearing outer race fault and driven wheel fault experimental verification of the present invention;
FIG. 8 is a gear parameter table of the present invention;
FIG. 9 is a table of bearing parameters for the present invention;
FIG. 10 is a table of structural parameters for a gearbox of the present invention;
FIG. 11 is a table of the characteristic frequencies of vibration for a dynamic model of a gearbox of the present invention.
Detailed Description
In order to make the objects, technical solutions and advantages of the embodiments of the present invention more apparent, the technical solutions of the embodiments of the present invention will be described clearly and completely with reference to the accompanying drawings of the embodiments of the present invention, and it is obvious that the described embodiments are some, but not all embodiments of the present invention. All other embodiments, which can be obtained by a person skilled in the art without inventive efforts based on the embodiments of the present invention, are within the scope of protection of the present invention. Thus, the following detailed description of the embodiments of the present invention, presented in the figures, is not intended to limit the scope of the invention, as claimed, but is merely representative of selected embodiments of the invention. All other embodiments, which can be obtained by a person skilled in the art without inventive efforts based on the embodiments of the present invention, are within the scope of protection of the present invention.
Referring to fig. 1-11, in an embodiment of the present invention, a method for modeling coupling dynamics of a gearbox excited by an internal multi-source fault includes the following steps:
s1: calculating the gear meshing time-varying rigidity, damping and dynamic friction force: according to the gear appearance and the local peeling fault parameters, calculating to obtain a functional relation of the gear meshing rigidity along with the rotation angle of the driving wheel; combining a mixed elastohydrodynamic lubrication theory, and calculating dynamic friction force and meshing damping of a current gear meshing pair at each moment by inputting dynamic gear meshing force, surface roughness and entrainment speed parameters;
s2: calculating the inner and outer ring acting force of each bearing in the gearbox: inputting the relative displacement of the inner ring and the outer ring at the current moment, and obtaining the normal deformation and the tangential speed of all contact pairs in the bearing through the revolution and autorotation relationship of quasi-static rolling elements; according to structural parameters and surface parameters of the bearing, the normal force of each rolling element in the bearing, which is considered to be damped, can be calculated by combining a mixed elastohydrodynamic lubrication theory and a Hertz contact theory; synthesizing the normal forces to obtain dynamic bearing inner and outer ring acting forces;
s3: and (3) constructing a complete equation of the whole gearbox dynamic model and solving: introducing gear meshing transmission errors, a driving motor, a brake and a gear box shell on the basis of the steps; establishing a complete gearbox dynamic differential equation set through Newton's second law; solving the differential equation set by using a numerical solving method to obtain the dynamic response of the system;
s4: simulation signal fault characteristic analysis: the method comprises the steps of selecting the most significant shell acceleration from a dynamic numerical solution of a gearbox system as a gearbox vibration simulation signal, analyzing the simulation signal through time domain analysis, power spectrum and envelope spectrum analysis and other basic means, and revealing the vibration behavior of the gearbox under the coupling fault condition.
Preferably, the specific step of S1 includes:
s11: local gear failure is introduced by changes in the geometric parameters. The present invention introduces localized faults in the gear through the form of surface spalling; the surface exfoliation morphology is described as elliptical with a certain depth h s Width w max And length l max (ii) a Such surface flaking can affect the effective contact length L during gear meshing s And cross section properties at the time of stiffness calculation;
wherein x s_center Denotes the center coordinate position, x, of the peeled ellipse s_start Denotes the starting coordinate position, x, of the peeling ellipse s_end Represents the end coordinate position of the peeling ellipse; theta s The inclination angle of the peeling ellipse; the numerical value of the part affected by the fault can affect the calculation of the meshing rigidity, and the whole dynamic system is affected through the meshing rigidity, so that the purpose of introducing the gear fault is achieved;
s12: calculating the time-varying stiffness of the meshing gear. The gear teeth are regarded as variable cross-section cantilever beams, and the Hertzian contact rigidity k of a certain gear tooth is obtained through energy method derivation h Bending stiffness k b Shear stiffness k s Axial compression stiffness k a And angle base stiffness k f (ii) a To obtain twoComprehensive series stiffness of gear teeth in a meshing state:
because more than one pair of gear teeth are in contact in the gear meshing process, the two pairs of gear teeth are in contact simultaneously, the series rigidity of the two pairs of gear teeth needs to be connected in parallel, and finally, the meshing rigidity function related to the rotation angle of the driving wheel is obtained:
s13: calculating elastohydrodynamic lubrication damping and friction coefficient of gear engagement; firstly, calculating the thickness h of the current lubricating oil film of the meshing pair according to the lubricating condition and the physical parameter c ：
Wherein, K Hc The surface micro-peak shape parameter is set as 1; w is a dimensionless load parameter; u is a dimensionless speed parameter; g is a dimensionless material parameter; r is the equivalent contact radius of curvature;
is a dimensionless surface roughness, V is a dimensionless hardness parameter;
the contact damping c of the gear contact pair can be obtained by approximate calculation according to the oil film thickness, the lubricant viscosity and the contact area film ：
According to the friction theory under mixed elastohydrodynamic lubrication conditions in gear meshing, the dynamic friction coefficient can be calculated by the following formula:
the intermediate parameter can be obtained by the following formula:
wherein u is pi And u gi The surface speeds of the driving wheel and the driven wheel in each pair of gear contact pairs are respectively; s is the root mean square surface roughness; p h Is Hertz contact pressure, b 1 To b 9 Is a constant value coefficient:
b 1-9 ＝-8.92,1.03,1.04,-0.35,2.81,-0.10,0.75,-0.39,0.62
because two pairs of teeth may be meshed in the gear meshing process, friction needs to be calculated on each pair of meshed gear teeth respectively and then synthesized; first according to the load distribution coefficient S ri The load on each pair of teeth is calculated:
then, a calculation formula of the friction force and the friction torque can be obtained:
where the subscript p indicates the action on the primary wheels and the subscript g indicates the action on the secondary wheels.
In general, S1 provides a way to calculate the forces generated by the meshing portions of the gears, which substantially describe the dynamic behavior of the gear mesh and form part of the gear mesh in the integral model.
Preferably, the specific step of S2 includes:
s21: calculating the equivalent stiffness and damping of the bearing rolling body; the inner contact pair and the outer contact pair of each rolling body of the bearing are connected in series to be equivalent to a spring damping system; therefore, the rigidity and the damping under each contact pair need to be solved firstly; the hertzian contact stiffness of a bearing can be obtained by the formula:
wherein E b ，ν b And Σ ρ is the sum of young's modulus, poisson's ratio, and raceway curvature, respectively; delta * A dimensionless deformation coefficient; according to the research related to the elastohydrodynamic lubrication, the elastohydrodynamic lubrication damping of the rolling body is generated by an inlet region of a lubricating oil film, and the inlet region lubrication damping can be calculated by the following formula:
wherein eta 0 Is the lubricating viscosity; r x Is the effective curvature in the direction of entrainment; a is the major semi-axis length of the contact ellipse; thickness h of point contact lubricating oil film c Can be calculated from the following formula:
h c ＝2.69U 0.67 G 0.53 Q -0.067 (1-0.61e -0.73κ )R x (13)
because the bearing raceway contact pair is not completely elastic lubricated, and the influence of the surface condition on the thickness of an oil film is considered, a correction coefficient is introduced into the calculation of the film thickness as follows:
wherein C and r are constant coefficients with values of 0.87 and 1.5, respectively; h is cT Namely the corrected film thickness;
after the rigidity and the damping of each contact pair are obtained, the equivalent series rigidity damping of the rolling element can be obtained through a complex rigidity series formula:
wherein
Linear stiffness k in formula 1,2 ＝k b δ 0.5 Linear damping c = c e (ii) a Natural frequency of free vibration of the omega contact pair;
s22: calculating the contact force of the inner ring and the outer ring of the bearing according to the contact geometry and the local fault; the local fault in the bearing raceway is described using a displacement excitation of half-sinusoidal profile, the displacement excitation equation being as follows:
wherein R is r Is the rolling element radius; l is defect Is the local fault length; phi is a l Is the angle range corresponding to the fault length;
is the angle of the jth rolling element relative to the fault location, which can be calculated by:
wherein phi d Is the starting position angle of the fault; the force acting between the inner ring and the outer ring of the bearing can be calculated by the following formula:
wherein lambda (delta) is a control function, when delta is a positive value, the output of the function is 1, and the output of the function is 0 under other conditions, so that the contact force is generated only when the inner ring and the outer ring are in a contact state; rolling body position angle psi j ＝2π(j-1)/N b +ψ c (ii) a And ψ c ＝0.5(1-D/D m )θ i Wherein theta i Representing the inner ring rotation angle, which is obtained by the geometrical relationship of the bearing; delta j The deformation amount indicating the jth rolling element position can be calculated by the following equation:
in the formula c 0 Indicating radial play.
In general, S2 gives the force between the inner and outer races of the bearing. The acting force fully reflects the dynamic characteristics of the bearing, and can accurately describe the dynamic action of the bearing in the gearbox system.
Preferably, the S3 specifically includes:
establishing a complete gearbox dynamic model; since industrial gearboxes contain deep groove ball bearings and spur gears, the vibrations of the gearbox are mainly caused by radial excitation; all axial degrees of freedom of the gearbox are therefore ignored, and bending of the rotating shaft and wobbling movement of the shaft are not taken into account due to the relatively short shaft length of the gearbox; each axis has freedom of movement and freedom of rotation in X and Y directions; the gear box shell is fixedly connected with all the bearing outer rings and has translational freedom degrees in two directions; the bearing inner ring is fixedly connected with the rotating shaft and rotates along with the rotating shaft; the gear box comprises four bearings, two bearings are positioned on the driving shaft, and two bearings are positioned on the driven shaft; the gearbox can be constructed into a nine-degree-of-freedom dynamic model, and the system equation is as follows:
wherein theta is L 、θ p And theta g The dynamic rotation angles of the load, the driving wheel and the driven wheel are respectively; m is a unit of p 、m g And m f The mass of the driving wheel, the driven wheel and the gear box shell is sequentially represented; I.C. A L 、I p And I g Sequentially representing the rotational inertia of the load, the driving wheel and the driven wheel; t is Load Is the moment exerted on the load，
Is the rotational speed of the motor. F pg Is the dynamic engagement force, and can be calculated by the following formula:
wherein k is m And c m Time-varying meshing stiffness and damping calculated in S1; ζ and ξ are the geometric deviation and the gear transmission error caused by the tooth surface exfoliation, respectively.
The invention uses Runge Kutta algorithm to solve the dynamic differential equation set, the time step length is set to 10 mus, the equation initial value is set as follows: x is a radical of a fluorine atom p ＝-3.5×10 -6 m，y p ＝-8.4×10 -6 m，x g ＝3.5×10 -6 m and y g ＝8.4×10 -6 m; the rotational speed is set to 1000RPM and the load torque is set to 16Nm.
Preferably, the S3 specifically includes:
analyzing the fault characteristics of the simulation signals; according to the numerical result obtained by the differential equation, the vibration response of the gearbox system in a period of time can be obtained; through time domain analysis of the response, the change characteristics of the physical variables in the gearbox in the whole process can be known; meanwhile, the characteristics of the vibration of the gearbox in the time domain under the obstructed fault mode can be analyzed and obtained; converting the time domain acceleration signal into a frequency domain signal by using Fourier transform to obtain a power spectrum of the simulation signal, and obtaining frequency domain responses of the gearbox in different modes through the power spectrum; analyzing the embodiment of different fault modes on the frequency domain response of the gearbox, and summarizing to obtain the characteristics of different fault modes on the power spectrum; obtaining an envelope curve of an original simulation signal through Hilbert transform, and then performing Fourier transform on the envelope curve to obtain an envelope spectrum of the simulation signal; the characteristic frequency components of the vibration of the gearbox can be obtained more clearly through the envelope spectrum, and the frequency characteristics of the envelope spectrum of different fault modes are clarified.
Through numerical solution and analysis, the coupled vibration response of the bearing outer ring fault and the gear driven wheel fault is shown in fig. 4. The bearing failure comprises two conditions, namely a high-speed shaft bearing failure and a low-speed shaft bearing failure. Each fault condition is analyzed by three means of time domain signals, power spectrum and envelope spectrum. It can be seen from the time domain signals that bearing outer race faults and driven wheel faults can produce significant impact on the gearbox housing. The fault condition inside the gearbox can be judged by the impact in the time domain signal, but the specific fault mode cannot be known from the time domain only. The dominant frequency components in the power spectrum are the mesh frequency and its harmonics, with the largest frequency being 4gmf, which is caused by the modulation of the natural frequency of the gearbox housing, which is determined by the gearbox parameters. The fault of the bearing outer ring excites an obvious side frequency band with the characteristic frequency of the outer ring as an interval in a power spectrum, and the interval of the side frequency band is different between the high-speed shaft bearing and the low-speed shaft bearing due to different rotating speeds. These side bands are distributed mainly around 4gmf, which means that the bearing characteristic frequency, like the meshing frequency, is amplitude modulated by the natural frequency of the gearbox housing. Similarly, the side bands excited by the failure of the driven wheel at intervals of the driven wheel speed have the same distribution characteristics in the power spectrum, which indicates that these frequencies are also subjected to natural frequency modulation. Because the edge frequency bands excited by the gear faults are relatively dense in interval and relatively low in amplitude, and are distributed in the same way as the edge frequency bands of the bearing faults, the edge frequency bands are easily covered by the bearing fault and are difficult to find. These easily covered bands are identified in the figure by red dashed lines. In the envelope spectrum, the characteristic frequencies occur mainly in the form of fundamental frequencies and lower order harmonic frequencies. The characteristic frequency of the bearing outer ring exceeds the meshing frequency at this time to become a main frequency component, but the meshing frequency is still visible in the envelope spectrum. The frequency conversion component caused by gear failure is mainly concentrated in the low-frequency area of the envelope spectrum due to low frequency and small amplitude, and is marked by a red dotted line. The bearing fault characteristic frequency and the gear fault characteristic frequency in the envelope spectrum and the power spectrum are in a superposed state, and no obvious mutual modulation phenomenon is found.
The failure of the bearing inner race and the failure of the driven wheel are shown in fig. 5. In line with the outer ring failure, the inner ring failure is also considered from both the high speed shaft bearing and the low speed shaft bearing. Under the inner ring fault, the fault position rotates along with the inner ring, the stress of the fault position fluctuates between the maximum value and zero, so that the time domain impact of the bearing is smaller than that of the outer ring, but the impact caused by the bearing fault can still be found. The impact caused by gear failure is consistent with that of the outer ring. The frequency domain response of the inner race fault and the driven wheel fault are also shown in fig. 5. The sidebands due to gear failure, identified by the red dashed lines, are the same as in the case of bearing outer race failure, indicating that the mode of bearing failure does not affect the frequency domain response of the gear failure to gearbox housing vibrations. The characteristic frequency caused by the failure of the inner ring of the shaft is more complex, and the characteristic frequency in the power spectrum and the envelope spectrum is modulated by the frequency conversion of the bearing except for the side frequency band taking the characteristic frequency of the inner ring as an interval. This is consistent with the conclusions drawn from the dynamic model of a single bearing, which also demonstrates the correctness of the model established by the invention. And the characteristic frequency of the bearing inner ring fault is consistent with that of the outer ring fault, the characteristic frequency of the bearing inner ring fault is not modulated with the characteristic frequency of the gear fault, and the characteristic frequencies are mutually overlapped.
To further verify the correctness of the invention, a single stage spur gear drive gearbox was used driven by an ac motor. The load is applied by a magnetic particle brake mounted at the end of the output shaft. The vibration signal is acquired by a sensor mounted on the gearbox housing. A gear local failure is placed on the driven wheel and a bearing local failure is placed in the high speed shaft bearing. In order to eliminate noise interference of the experimental signals, the experimental signals are all subjected to denoising processing. And selecting frequency bands of [9216Hz and 10240Hz ] to filter the simulation signals and the experiment signals. Fig. 6 shows experimental and simulation results comparing inner ring bearing failure and driven wheel failure. Fig. 7 shows experimental and simulation results comparing the failure of the outer ring bearing and the failure of the driven wheel. The green dots in the graph identify the gear failure characteristic frequency, i.e., the driven wheel rotation frequency, and the red dots identify the bearing failure characteristic frequency and the meshing frequency. From the experimental comparison graph, it can be known that although the frequency amplitude has partial errors, the frequency characteristics are basically consistent, which can prove that the experiment can prove the accuracy of the industrial gearbox dynamics modeling method excited by the internal multisource coupling fault provided by the invention.
In order to solve the problem that the traditional dynamic model is incomplete and cannot provide a theoretical basis for the vibration of the gearbox under the internal multi-source fault mechanism, the invention constructs a brand-new industrial gearbox dynamic modeling method under the excitation of the internal multi-source coupling fault by starting from the traditional gear meshing model and fully considering the aspects of dynamic acting force of a bearing, a gearbox shell, elastohydrodynamic lubrication, local fault description and the like. And starting from the dynamic model, the vibration response of the gearbox shell under the internal multi-source coupling fault mechanism is researched. The method has the advantages that the vibration characteristics of the gearbox under different fault modes are analyzed by using a classical Fourier transform and envelope spectrum analysis method, the superposition action mechanism of the local faults of the bearing and the gear on the shell of the gearbox is disclosed, and a theoretical basis is provided for signal processing methods such as fault characteristic extraction and the like. Finally, the experimental signal measured by the gearbox test bed is compared with the simulation signal obtained in the invention, so that the dynamic modeling method of the industrial gearbox excited by the internal multi-source coupling fault has certain effectiveness.
The invention is described in detail below with reference to simulated signal analysis and experimental verification.
Schematic diagrams of gearbox dynamics models as shown in fig. 1 and 2 below, the inner and outer models interact with each other through bearing forces to form a complete gearbox. Gearbox parameters used in the present invention are shown in figures 8-10:
the S4 comprises the following steps:
analyzing the fault characteristics of the simulation signals; according to the numerical result obtained by the differential equation, the vibration response of the gearbox system in a period of time can be obtained; through time domain analysis of the response, the change characteristics of the physical variables in the gearbox in the whole process can be known; meanwhile, the characteristics of the vibration of the gearbox in the time domain under the failure mode can be analyzed and obtained; converting the time domain acceleration signal into a frequency domain signal by using Fourier transform to obtain a power spectrum of the simulation signal, and obtaining frequency domain responses of the gearbox in different modes through the power spectrum; analyzing the embodiment of different fault modes on the frequency domain response of the gearbox, and summarizing to obtain the characteristics of different fault modes on a power spectrum; obtaining an envelope curve of an original simulation signal through Hilbert transform, and then performing Fourier transform on the envelope curve to obtain an envelope spectrum of the simulation signal; the characteristic frequency components of the vibration of the gearbox can be obtained more clearly through the envelope spectrum, and the frequency characteristics of the envelope spectrum of different fault modes are determined;
through numerical solution and analysis, the coupled vibration response of the bearing outer ring fault and the gear driven wheel fault is shown in fig. 4; the bearing fault comprises two conditions, namely a high-speed shaft bearing fault and a low-speed shaft bearing fault; each fault condition is analyzed by three means of time domain signals, power spectrum and envelope spectrum; obvious impact can be generated on the shell of the gearbox due to the failure of the outer ring of the bearing and the failure of the driven wheel from the time domain signal; the fault condition inside the gearbox can be judged through the impact in the time domain signal, but the specific fault mode cannot be known only from the time domain; the frequency components in the power spectrum which are dominant are the meshing frequency and harmonic frequencies thereof, wherein the maximum frequency is 4gmf, which is caused by the modulation of the natural frequency of the gearbox shell, and the natural frequency is determined by the parameters of the gearbox; the fault of the bearing outer ring excites an obvious side frequency band with the characteristic frequency of the outer ring as an interval in a power spectrum, and the intervals of the side frequency bands of the high-speed shaft bearing and the low-speed shaft bearing are different due to different rotating speeds; these side bands are mainly distributed around 4gmf, which means that the characteristic frequency of the bearing is the same as the meshing frequency and is subjected to amplitude modulation of the natural frequency of the gearbox shell; similarly, the side frequency bands excited by the driven wheel fault and spaced by the rotating speed of the driven wheel have the same distribution characteristics in the power spectrum, which shows that the frequencies are also modulated by the natural frequency; because the edge frequency bands excited by the gear fault have dense intervals and low amplitude and are distributed the same as the edge frequency bands of the bearing fault, the edge frequency bands are easily covered by the bearing fault and are not easy to find; these easily covered bands are identified in the figure by red dashed lines; in the envelope spectrum, the characteristic frequencies mainly appear in the form of fundamental frequencies and low-order harmonic frequencies; the characteristic frequency of the bearing outer ring exceeds the meshing frequency at the moment to become a main frequency component, but the meshing frequency is still visible in an envelope spectrum; the frequency conversion component caused by gear failure is mainly concentrated in a low-frequency area of an envelope spectrum due to lower frequency and smaller amplitude, and is marked by a red dotted line; bearing fault characteristic frequency and gear fault characteristic frequency in the envelope spectrum and the power spectrum are in a superposed state, and no obvious mutual modulation phenomenon is found;
the failure of the bearing inner ring and the failure of the driven wheel are shown in figure 5; the fault of the inner ring is considered from two conditions of a high-speed shaft bearing and a low-speed shaft bearing, and is consistent with the fault of the outer ring; under the condition of inner ring failure, the failure position rotates along with the inner ring, and the stress of the failure position fluctuates between the maximum value and zero, so that the time domain impact of the bearing is smaller than that of the outer ring, but the impact caused by the bearing failure can still be found; the impact caused by the gear fault is consistent with that of the outer ring; the frequency domain response of the inner race fault and the driven wheel fault are also shown in fig. 5; the sidebands caused by gear failure are identified by red dashed lines, which are the same as in the case of bearing outer ring failure, indicating that the mode of bearing failure does not affect the frequency domain response of gear failure to gearbox housing vibrations; the characteristic frequency caused by the failure of the inner ring of the shaft is more complex, and the characteristic frequency in the power spectrum and the envelope spectrum is modulated by the frequency conversion of the bearing except for the edge frequency band taking the characteristic frequency of the inner ring as an interval; this is consistent with the conclusions drawn from the dynamic model of a single bearing, which also proves the correctness of the model established by the invention; the characteristic frequency of the bearing inner ring fault is not modulated with the characteristic frequency of the gear fault and is in a superposed state;
in order to further verify the correctness of the invention, a gearbox test bed is used for experimental verification; the single-stage straight-tooth gear transmission gear box is driven by an alternating current motor; the load is applied through a magnetic powder brake arranged at the tail end of the output shaft; the vibration signal is obtained through a sensor arranged on a gearbox body; a gear local fault is arranged on the driven wheel, and a bearing local fault is arranged in the high-speed shaft bearing; in order to eliminate noise interference of experimental signals, the experimental signals are subjected to denoising treatment; selecting frequency bands of [9216Hz,10240Hz ] to filter the simulation signals and the experiment signals; FIG. 6 shows experimental and simulation results comparing inner race bearing failure and driven wheel failure; FIG. 7 shows experimental and simulation results comparing failure of an outer race bearing and failure of a driven wheel; the green dots in the graph identify the gear fault characteristic frequency, namely the rotation frequency of the driven wheel, and the red dots identify the bearing fault characteristic frequency and the meshing frequency; from the experimental comparison graph, the frequency characteristics are basically consistent although the frequency amplitude has partial errors, which can show that the experiment can prove the accuracy of the internal multi-source coupling fault excitation industrial gearbox dynamics modeling method provided by the invention;
in order to solve the problem that the traditional dynamic model is incomplete and cannot provide a theoretical basis for the vibration of the gearbox subjected to the internal multi-source fault mechanism, the invention constructs a brand-new dynamic modeling method of the industrial gearbox excited by the internal multi-source coupling fault by starting from the traditional gear meshing model and fully considering the aspects of dynamic acting force of a bearing, a gearbox shell, elastohydrodynamic lubrication, local fault description and the like; based on the dynamic model, the vibration response of the gearbox shell under an internal multi-source coupling fault mechanism is researched; the method has the advantages that the vibration characteristics of the gearbox under different fault modes are analyzed by using a classical Fourier transform and envelope spectrum analysis method, the superposition action mechanism of the local faults of the bearing and the gear on the shell of the gearbox is disclosed, and a theoretical basis is provided for signal processing methods such as fault characteristic extraction and the like; finally, the experimental signal measured by the gearbox test bed is compared with the simulation signal obtained in the invention, so that the dynamic modeling method for the industrial gearbox excited by the internal multi-source coupling fault has certain effectiveness.
Finally, it should be noted that: although the present invention has been described in detail with reference to the foregoing embodiments, it will be apparent to those skilled in the art that modifications may be made to the embodiments described above, or equivalents may be substituted for elements thereof. Any modification, equivalent replacement, or improvement made within the spirit and principle of the present invention should be included in the protection scope of the present invention.
Claims (8)
1. A gearbox coupling dynamics modeling method excited by internal multisource faults is characterized by comprising the following steps of: the method comprises the following steps:
s1: calculating the gear meshing time-varying rigidity, damping and dynamic friction force: according to the gear appearance and the local peeling fault parameters, calculating to obtain a functional relation of the gear meshing rigidity along with the rotation angle of the driving wheel; combining a mixed elastohydrodynamic lubrication theory, and calculating dynamic friction force and meshing damping of a current gear meshing pair at each moment by inputting dynamic gear meshing force, surface roughness and entrainment speed parameters;
s2: calculating the acting force of the inner ring and the outer ring of each bearing in the gearbox: inputting the relative displacement of the inner ring and the outer ring at the current moment, and obtaining the normal deformation and the tangential speed of all contact pairs in the bearing through the revolution and autorotation relationship of quasi-static rolling elements; according to structural parameters and surface parameters of the bearing, the normal force of each rolling element in the bearing, which is considered to be damped, can be calculated by combining a mixed elastohydrodynamic lubrication theory and a Hertz contact theory; synthesizing the normal forces to obtain dynamic bearing inner and outer ring acting forces;
s3: and (3) constructing a complete equation of the whole gearbox dynamic model and solving: introducing a gear meshing transmission error, a driving motor, a brake and a gear box shell on the basis of the steps; establishing a complete gearbox dynamic differential equation set through Newton's second law; solving the differential equation set by using a numerical solving method to obtain the dynamic response of the system;
s4: simulation signal fault characteristic analysis: the method comprises the steps of selecting the most significant shell acceleration from a dynamic numerical solution of a gearbox system as a gearbox vibration simulation signal, analyzing the simulation signal through time domain analysis, power spectrum and envelope spectrum analysis and other basic means, and revealing the vibration behavior of the gearbox under a coupling fault condition.
2. The method of modeling gearbox coupling dynamics excited by an internal multisource fault of claim 1, where: the specific steps of S1 include:
s11: local gear failure is introduced by changes in the geometric parameters. The invention introduces local faults in the gear through surface peeling; the surface exfoliation morphology is described as elliptical with a certain depth h s Width w max And length l max (ii) a Such surface flaking can affect the effective contact length L of the gears during meshing s And cross section properties at the time of stiffness calculation;
wherein x s_center Denotes the position of the center coordinate, x, of the peeled ellipse s_start Denotes the starting coordinate position, x, of the peeling ellipse s_end Represents the end coordinate position of the peeling ellipse; theta.theta. s The tilt angle representing the peeling ellipse; the numerical value of the part affected by the fault can affect the calculation of the meshing rigidity, and the whole dynamic system is affected through the meshing rigidity, so that the purpose of introducing the gear fault is achieved;
s12: calculating the time-varying stiffness of the meshing gear. The gear teeth are regarded as variable cross-section cantilever beams, and the Hertzian contact rigidity k of a certain gear tooth is obtained through energy method derivation h (iii) bending stiffness k b Shear stiffness k s Axial compression stiffness k a And angle base stiffness k f (ii) a Obtaining the comprehensive series rigidity of the two gear teeth in a meshing state:
because more than one pair of gear teeth are in contact in the gear meshing process, the two pairs of gear teeth are in contact simultaneously, the series rigidity of the two pairs of gear teeth needs to be connected in parallel, and finally, the meshing rigidity function related to the rotation angle of the driving wheel is obtained:
s13: calculating elastohydrodynamic lubrication damping and friction coefficient of gear engagement; firstly, calculating the thickness h of the current lubricating oil film of the meshing pair according to the lubricating condition and the physical parameter c ：
Wherein, K Hc Surface micro-peak morphology parameters are set as 1; w is a dimensionless load parameter; u is a dimensionless speed parameter; g is a dimensionless material parameter; r is the equivalent contact radius of curvature;
is a dimensionless surface roughness, V is a dimensionless hardness parameter;
the contact damping c of the gear contact pair can be obtained by approximate calculation according to the oil film thickness, the lubricant viscosity and the contact area film ：
According to the friction theory under mixed elastohydrodynamic lubrication conditions in gear meshing, the dynamic friction coefficient can be calculated by the following formula:
the intermediate parameter can be obtained by the following formula:
whereinu pi And u gi The surface speeds of the driving wheel and the driven wheel in each pair of gear contact pairs are respectively; s is the root mean square surface roughness; p is h Is Hertz contact pressure, b 1 To b 9 Is a constant value coefficient:
b 1-9 ＝-8.92,1.03,1.04,-0.35,2.81,-0.10,0.75,-0.39,0.62
because two pairs of teeth may be meshed in the gear meshing process, friction needs to be calculated on each pair of meshed gear teeth and then synthesized; first according to the load distribution coefficient S ri The load on each pair of teeth is calculated:
then, a calculation formula of the friction force and the friction torque can be obtained:
where the subscript p indicates the action on the primary wheels and the subscript g indicates the action on the secondary wheels.
3. The method of modeling gearbox coupling dynamics excited by an internal multisource fault of claim 2, where: the specific steps of S2 include:
s21: calculating the equivalent rigidity and damping of the bearing rolling body; connecting the inner contact pair and the outer contact pair of each rolling body of the bearing in series to be equivalent to a spring damping system; therefore, the rigidity and the damping under each contact pair need to be solved firstly; the hertzian contact stiffness of a bearing can be obtained by the formula:
wherein E b ，ν b And Σ ρ is the sum of young's modulus, poisson's ratio, and raceway curvature, respectively; delta * A dimensionless deformation coefficient; according to the research related to the elastohydrodynamic lubrication, the elastohydrodynamic lubrication damping of the rolling body is generated in an inlet area of a lubricating oil film, and the inlet area lubrication damping can be calculated by the following formula:
wherein eta 0 Is the lubricating viscosity; r is x Is the effective curvature in the direction of entrainment; a is the major semi-axis length of the contact ellipse; thickness h of point contact lubricating oil film c Can be calculated from the following formula:
h c ＝2.69U 0.67 G 0.53 Q -0.067 (1-0.61e -0.73κ )R x (13)
because the bearing raceway contact pair is not completely elastic lubricated, and the influence of the surface condition on the thickness of an oil film is considered, a correction coefficient is introduced into the calculation of the film thickness as follows:
wherein C and r are constant coefficients with values of 0.87 and 1.5, respectively; h is cT Namely the corrected film thickness;
after the rigidity and the damping of each contact pair are obtained, the equivalent series rigidity damping of the rolling element can be obtained through a complex rigidity series formula:
wherein
Linear stiffness k in formula 1,2 ＝k b δ 0.5 Linear damping c = c e ；ωThe natural frequency of free vibration of the contact pair;
s22: calculating the contact force of the inner ring and the outer ring of the bearing according to the contact geometry and the local fault; the local fault in the bearing raceway is described using a displacement excitation of half-sinusoidal profile, the displacement excitation equation being as follows:
wherein R is r Is the rolling element radius; l is defect Is the local fault length; phi l is the angle range corresponding to the fault length;
is the angle of the jth rolling element relative to the fault location, which can be calculated by:
wherein phi d Is the starting position angle of the fault; the force acting between the inner ring and the outer ring of the bearing can be calculated by the following formula:
wherein lambda (delta) is a control function, when delta is a positive value, the output of the function is 1, and the output of the function is 0 under other conditions, so that the contact force is generated only when the inner ring and the outer ring are in a contact state; rolling body position angle psi j ＝2π(j-1)/N b +ψ c (ii) a And ψ c ＝0.5(1-D/D m )θ i Wherein θ i Representing the inner ring rotation angle, which is derived from the bearing geometry; delta. For the preparation of a coating j The deformation amount indicating the jth rolling element position can be calculated by the following equation:
in the formula c 0 Indicating radial play.
4. The method of modeling gearbox coupling dynamics excited by an internal multi-source fault according to claim 3, wherein: the S3 specifically includes:
establishing a complete gearbox dynamic model; since industrial gearboxes contain deep groove ball bearings and spur gears, the vibrations of the gearbox are mainly caused by radial excitation; all axial degrees of freedom of the gearbox are therefore ignored, and bending of the rotating shaft and wobbling movement of the shaft are not taken into account due to the relatively short shaft length of the gearbox; each axis has freedom of movement and freedom of rotation in X and Y directions; the gearbox shell is fixedly connected with all bearing outer rings and has translational freedom degrees in two directions; the bearing inner ring is fixedly connected with the rotating shaft and rotates along with the rotating shaft; the gear box comprises four bearings, two bearings are positioned on the driving shaft, and two bearings are positioned on the driven shaft; the gearbox can be constructed into a nine-degree-of-freedom dynamic model, and the system equation is as follows:
wherein theta is L 、θ p And theta g The dynamic rotation angles of the load, the driving wheel and the driven wheel are respectively; m is p 、m g And m f The mass of the driving wheel, the driven wheel and the gear box shell is sequentially represented; i is L 、I p And I g Sequentially representing the rotational inertia of the load, the driving wheel and the driven wheel; t is a unit of Load Is the moment of force exerted on the load,
is the rotational speed of the motor. F pg Is the dynamic engagement force, which can be calculated from the following formula:
wherein k is m And c m Time-varying meshing stiffness and damping calculated in S1; ζ and ξ are the geometric deviations and gear transfer errors caused by tooth flank peeling, respectively.
5. The method of modeling gearbox coupling dynamics excited by an internal multisource fault of claim 4, where: the S3 specifically includes:
analyzing the fault characteristics of the simulation signals; according to the numerical result obtained by the differential equation, the vibration response of the gearbox system in a period of time can be obtained; through time domain analysis of the response, the change characteristics of the physical variables in the gearbox in the whole process can be known; meanwhile, the characteristics of the vibration of the gearbox in the time domain under the obstructed fault mode can be analyzed and obtained; converting the time domain acceleration signal into a frequency domain signal by using Fourier transform to obtain a power spectrum of the simulation signal, and obtaining frequency domain responses of the gearbox in different modes through the power spectrum; analyzing the embodiment of different fault modes on the frequency domain response of the gearbox, and summarizing to obtain the characteristics of different fault modes on a power spectrum; obtaining an envelope curve of an original simulation signal through Hilbert transform, and then performing Fourier transform on the envelope curve to obtain an envelope spectrum of the simulation signal; the characteristic frequency components of the vibration of the gearbox can be obtained more clearly through the envelope spectrum, and the frequency characteristics of the envelope spectrum of different fault modes are determined.
6. A computer device comprising a memory, a processor and a computer program stored on the memory and executable on the processor, characterized in that the steps of the method of any of claims 1 to 5 are implemented when the program is executed by the processor.
7. A computer-readable storage medium, on which a computer program is stored which, when being executed by a processor, carries out the steps of the method of any one of claims 1 to 5.
8. A processor, characterized in that the processor is configured to run a program, wherein the program when running performs the method of any of claims 1 to 5.
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