CN114692449B - Test-based analysis and optimization method for torsional vibration of compressor - Google Patents
Test-based analysis and optimization method for torsional vibration of compressor Download PDFInfo
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Abstract
The invention relates to a method for analyzing and optimizing torsional vibration of a compressor based on a test, which adopts a finite element method to correct a concentrated parameter model to perform primary torsional vibration calculation, and then uses a primary torsional vibration test result to accurately correct the concentrated parameter model to accurately optimize torsional vibration of a shaft system.
Description
Technical Field
The invention relates to the field of torsional vibration analysis and optimization of rotating equipment, in particular to a test-based compressor torsional vibration analysis and optimization method.
Background
The compressor is used as power equipment for industrial gas compression, the operation safety and stability of the compressor are of great significance to industrial production and transportation, torsional vibration analysis is needed to be carried out on the shaft system of the compressor in the process of designing the compressor so as to ensure the safety and stability of the shaft system, but because errors exist in the process of simplifying a physical model into a mathematical model, the operation of the shaft system of the compressor is calculated by adopting a centralized parameter method or a finite element method, and a certain difference exists between the torsional vibration calculation result and the operation of the shaft system of the actual compressor. From the existing methods such as a concentrated parameter method and a finite element method, the reasons for the errors are various, such as model simplification, unit assembly, lubrication damping nonlinearity and the like, the existing method for solving the errors is not good, and the safety of the unit is ensured by adding a safety coefficient, so that whether the unit is operated in an optimal range or not is unknown after torsional vibration analysis, and the problems of overlarge unit vibration or resonance shaft breakage still occur after the torsional vibration analysis are also frequently generated on a compressor site.
The API618 standard therefore specifies that the compressor shaft natural frequency should avoid more than 5% of the excitation frequency and its multiple to ensure that the shaft will not torsional overcompensate or resonate, and if it cannot avoid, a forced vibration analysis is required to evaluate its safety. However, in the actual compressor design process, the natural frequency of the shafting always falls within the frequency multiplication range of 5% of the excitation frequency along with the increase of the frequency multiplication order of the compressor, and the compressor shafting torsional vibration is overlarge or resonates when running under specific working conditions, so that the compressor is finally unstable in running or the shafting is broken to cause huge loss.
Disclosure of Invention
The invention aims to solve the technical problems of the prior art, and provides a test-based analysis and optimization method for the torsional vibration of a compressor, which can rapidly and effectively obtain the torsional vibration value of the optimal state in the running process of the compressor and ensure the stable and safe running of a compressor shafting.
The invention is realized by the following technical scheme:
a test-based method for optimizing torsional vibration analysis of a compressor, comprising:
step 1, a three-dimensional model of a compressor shafting is established, and the three-dimensional model of the compressor shafting is converted into a plurality of concentrated parameter models;
step 2, checking by adopting a finite element analysis method, eliminating the centralized errors in each centralized parameter model, and obtaining each centralized parameter model with the centralized errors eliminated;
step 3, establishing a second-order differential equation of torsional vibration of the shafting according to each concentrated parameter model for eliminating concentrated errors, carrying out torsional vibration analysis to obtain a calculated torque response value and an optimal rotational inertia range which change along with the rotational inertia, and selecting a rotational inertia from the optimal rotational inertia range as a rotational inertia design value;
step 4, designing and manufacturing the compressor according to the rotational inertia design value to obtain the compressor;
step 5, performing torsional vibration test on the compressor to obtain a test torque response value;
step 6, comparing and analyzing the test torque response value and the calculated torque response value, and eliminating assembly errors and damping errors according to analysis results to obtain a corrected second-order differential equation of torsional vibration of the shafting;
and 7, performing torsional vibration analysis by adopting a corrected second order differential equation of the torsional vibration of the shafting to obtain a rotational inertia adjustment value, and adjusting the structure of the compressor according to the rotational inertia adjustment value.
Preferably, in step 2, a finite element analysis method is adopted for checking, so as to eliminate the concentration error in each concentration parameter model, and the specific method is as follows: and converting the compressor shafting three-dimensional model into a compressor shafting finite element model for finite element analysis, and eliminating the concentration errors in each concentration parameter model according to the Rayleigh principle.
Further, in step 2, for a centralized parameter model, the method for eliminating centralized errors specifically includes:
2.1, acquiring an inertia matrix [ J12] and a rigidity matrix [ K12] according to the concentrated parameter model;
2.2 establishing a free vibration equationSubstitution of inertia matrix [ J12]]And stiffness matrix [ K12]]；
2.3 solving a free vibration equation to obtain a natural frequency w12 and a modal shape [ q12];
2.4, converting the three-dimensional model of the compressor shaft system into a finite element model of the compressor shaft system, and analyzing the finite element model of the compressor shaft system to obtain a natural frequency w12 'and a regular vibration mode [ q12' ];
2.5 calculating the modified stiffness matrix to be [ K12 ]']＝[q12'] T *diag[w12' 2 ]*[q12']；
2.6 let Δk= ([ K12] - [ K12 '])/[ K12' ], Δw= ([ w12] - [ w12 '])/[ w12' ], Δq= ([ q12] - [ q12 '])/[ q12' ], if Δk, Δw, Δq are greater than 1%, check if steps 2.1-2.5 are correct; if the difference is less than or equal to 1%, each torsional rigidity in the corrected rigidity matrix [ K12' ] is adopted to replace each corresponding torsional rigidity in the concentrated parameter model, and the concentrated parameter model for eliminating the concentrated error is obtained.
Further, the concentrated parameter model comprises a crankshaft concentrated parameter model, a coupling concentrated parameter model and a motor shaft concentrated parameter model, and the inertia matrix [ J12] is an inertia matrix [ J1] of a crankshaft or an inertia matrix [ J2] of a motor shaft; the rigidity matrix [ K12] is a rigidity matrix [ K1] of the crankshaft or a rigidity matrix [ K2] of the motor shaft; the natural frequency w12 is the natural frequency w1 of the crankshaft or the natural frequency w2 of the motor shaft; the mode shape [ q12] is the mode shape [ q1] of the crankshaft or the mode shape [ q2] of the motor shaft.
Preferably, in step 3, the specific steps are: establishing a second-order differential equation of torsional vibration of the shafting according to each concentrated parameter model for eliminating concentrated errorsExcitation moment substituted into calculated working condition [ F (t)]Selecting a checking point in a compressor shaft system, and calculating a torque response value T (T) of the checking point changing along with the change of moment of inertia according to a second-order differential equation of torsional vibration of a shaft system jisuan Searching an optimal range for calculating a torque response value according to a sensitivity minimum principle, obtaining an optimal moment of inertia range according to the optimal range, and selecting a moment of inertia from the optimal moment of inertia range to serve as a moment of inertia design value;
wherein [ J ] is an inertia matrix, [ K ] is a torsion matrix, and [ C ] is a damping matrix.
Further, the step 5 specifically includes: setting a strain gauge sensor at a checking point of the compressor, performing torsional vibration test, and acquiring a test torque response value T (T) through the strain gauge sensor ceshi 。
Further, in step 6, comparing the test torque response value with the calculated torque response value, and eliminating the assembly error according to the analysis result, specifically including:
6.1 true Assembly torsional stiffness K10 at the calibration Point zhenshi X times the corrected torsional stiffness K10'; the corrected torsional rigidity K10' is the corrected torsional rigidity of the correction point in the concentrated parameter model for eliminating the concentrated error;
6.2 setting the parameter for correcting the assembly error as x, and setting the value range of x, so as to correct the torsional rigidity K10 after the assembly error xiuzheng ＝K10'*x；
6.3 torsional stiffness K10 after correction of assembly errors xiuzheng Carry-in torsion matrix [ K]Based on the second order differential equation of torsional vibration of shaftingCalculating a corrected calculated torque response value T (T) as a function of the corrected assembly error parameter x jisuan_1 ；
6.4 contrast test Torque response value T (T) ceshi And correcting and calculating a torque response value T (T) jisuan_1 The amplitude error between the two is selected as x value multiplied by K10' with the minimum amplitude error to be used as the real assembly torsional rigidity K10 zhenshi The method comprises the steps of carrying out a first treatment on the surface of the True to assemble torsional stiffness K10 zhenshi Substituting the second order differential equation of the torsional vibration of the shafting to obtain the second order differential equation of the torsional vibration of the shafting, which eliminates assembly errors.
Further, in step 6, the test torque response value and the calculated torque response value are subjected to comparative error analysis, and the damping error is eliminated according to the analysis result, specifically:
set damping matrix [ C ]]With Rayleigh damping, the real damping matrix [ C] zhengshi ＝[C]* y; setting the value range of y, and correcting damping errorPoor torsional stiffness [ C] xiuzheng ＝[C]*y；
Torsional stiffness after correction of damping error [ C] xiuzheng Torsional vibration second order differential equation with replacement shaftingIn [ C ]]Calculating a corrected calculated torque response value T (T) as a function of the corrected damping error parameter y jisuan_2 ；
Comparative T (T) ceshi And T (T) jisuan_2 The amplitude error between the two is selected, and the y value with the minimum amplitude error is multiplied by [ C ]]As a real damping matrix [ C] zhenshi The real damping matrix [ C] zhenshi Substituting the second-order differential equation of the shafting torsional vibration to obtain the second-order differential equation of the shafting torsional vibration for eliminating damping errors.
Further, the step 7 specifically includes: according to the second differential equation of the corrected torsional vibration of the shaftingExcitation moment substituted into calculated working condition [ F (t)]Correction calculation torque response value T (T) of calculation check point changing along with rotation inertia change jisuan ' the torque response value T (T) will be calculated by correction jisuan ' equal to the calculated torque response value T (T) corresponding to the design value of the moment of inertia jisuan And when the corresponding moment of inertia is used as a target moment of inertia, the difference between the target moment of inertia and the moment of inertia design value is the moment of inertia adjustment value, and the compressor structure is adjusted according to the moment of inertia adjustment value.
Compared with the prior art, the invention has the following beneficial effects:
the torsional vibration optimization by adopting the centralized parameter method has the advantages of definite physical meaning, simple and quick calculation and easy correction, and has the defects that the torsional rigidity is not easy to calculate accurately and large errors are easy to exist; the advantage of optimizing the torsional vibration by adopting the finite element method is that the result of calculating a single part is accurate, the disadvantage is that the error of calculating an assembly is larger, the calculation process is complex, the time consumption is longer, and the correction is difficult; the torsional vibration is optimized through the test, the result is accurate, the error is almost avoided, the mathematical model is not needed, multiple multi-point tests are needed, the time consumption and the material cost are high, and the optimization process is complex; compared with the existing method for calculating by using a concentrated parameter method or a finite element method or optimizing the torsional vibration response of a compressor shaft system by testing, the method adopts the finite element method to correct the concentrated parameter model for primary torsional vibration calculation, and then uses a torsional vibration test result to accurately correct the concentrated parameter model to accurately optimize the torsional vibration of the shaft system.
Drawings
FIG. 1 is a diagram of a compressor shafting centralized parameter model.
FIG. 2 is a three-dimensional plot of torque response as moment of inertia increases.
Fig. 3 is a schematic diagram of a torsional vibration test.
FIG. 4 is a three-dimensional plot of torque response as the correction parameter increases.
Fig. 5 moment of inertia adjusting apparatus.
FIG. 6 is a flow chart for torsional vibration analysis and optimization.
Detailed Description
For a further understanding of the present invention, the present invention is described below in conjunction with the following examples, which are provided to further illustrate the features and advantages of the present invention and are not intended to limit the claims of the present invention.
The torsional vibration analysis optimization method based on the test comprises the steps of modeling centralized parameters, correcting centralized errors by finite elements, searching an optimal range of torsional vibration, testing the torsional vibration, eliminating assembly and damping errors by test comparison analysis, and optimizing torsional vibration of a unit by target rotational inertia.
As shown in fig. 6, taking a certain reciprocating piston compressor as an example, the following detailed description is given to the technical scheme of the present invention:
1. and (3) establishing a three-dimensional model of the compressor shafting (comprising a crankshaft three-dimensional model and a motor shaft three-dimensional model), and simplifying the three-dimensional model of the compressor shafting into a plurality of concentrated parameter models comprising a crankshaft, a coupling and a motor shaft according to a concentrated parameter method principle. As shown in fig. 1, the parameters in the crankshaft centralized parameter model include rotational inertias J1 to J9_1 (where J9_1+j9_2=j 9), torsional rigidities K1 to K8, and torsional damping C1 to C8, the parameters in the motor shaft centralized parameter model include rotational inertias J10_1 to J15 (where J10_1+j10_2=j 10), torsional rigidities K10 to K14, and torsional damping C10 to C14, the centralized parameter models of the coupling are J9_2, J10_2, K9, and C9, wherein there are centralized errors and assembly errors in the torsional rigidities (K1 to K14), and there are damping errors in the torsional damping (C1 to C14).
2. And (5) respectively calculating the natural frequencies and the mode shapes of the crankshaft and the motor shaft in the shafting.
Taking a crankshaft as an example, the natural frequency w1 and the mode shape [ q1] are calculated as follows:
2.1, according to a crankshaft centralized parameter model, taking rotational inertia J1-J9_1 of a crankshaft to form an inertia matrix [ J1], and torsional rigidity K1-K8 to form a rigidity matrix [ K1];
2.2 establishing a free vibration equation for the undamped concentration parameters of the crankshaftMatrix of inertia [ J1]]Stiffness matrix [ K1]]Carrying out a free vibration equation of undamped centralized parameters of the crankshaft;
2.3, calculating to obtain the natural frequency of the free vibration equation as w1 and the modal regular vibration mode as q 1;
the same mode of steps 2.1-2.3 is adopted, only the moment of inertia J10_1-J15 of the motor shaft is taken to form an inertia matrix [ J2], the torsional rigidity K10-K14 is taken to form a rigidity matrix [ K2], and the inherent frequency w2 and the modal regular vibration type [ q2] of the motor shaft are calculated.
3. And converting the compressor shafting three-dimensional model into a compressor shafting finite element model for finite element analysis, and correcting and eliminating the concentrated errors in the concentrated parameter model according to the Rayleigh principle to obtain the concentrated parameter model for eliminating the concentrated errors.
Taking a crankshaft as an example, converting a crankshaft three-dimensional model into a crankshaft finite element model for finite element analysis, and carrying out model correction elimination on concentration errors in a crankshaft concentration parameter model according to a Rayleigh principle to obtain the crankshaft concentration parameter model for eliminating the concentration errors, wherein the method comprises the following specific steps of:
3.1, analyzing a finite element model of the crankshaft to obtain a natural frequency w1 'and a regular vibration mode [ q1' ];
3.2 calculating the modified stiffness matrix to be [ K1 ]']＝[q1'] T *diag[w1' 2 ]*[q1']；
3.3 let Δk= ([ K1] - [ K1 '])/[ K1' ], Δw= ([ w1] - [ w1 '])/[ w1' ], Δq= ([ q1] - [ q1 '])/[ q1' ], if Δk, Δw, Δq are greater than 1%, the process of calculation is simplified by checking the centralized parameters and finite elements, and determining if the process is correct;
3.4 if Δk, Δw and Δq are all less than or equal to 1%, then [ K1'] is considered as a stiffness matrix for successful correction, and each corrected torsional stiffness (K1' -K8 ') in the stiffness matrix [ K1' ] is adopted to replace each corresponding torsional stiffness in the crankshaft centralized parameter model, so that the crankshaft centralized parameter model for eliminating the centralized error is obtained.
4. And (3) correcting the concentrated errors of the rigidity matrix in the motor shaft concentrated parameter model in the same manner as in the steps 3.1-3.3 to obtain a corrected rigidity matrix [ K2'], and replacing the corresponding torsional rigidities in the motor shaft concentrated parameter model with the corrected torsional rigidities (K10' -K14 ') in the rigidity matrix [ K2' ], so as to obtain the motor shaft concentrated parameter model with the concentrated errors eliminated.
5. According to the crankshaft concentrated parameter model for eliminating concentrated errors and the motor shaft concentrated parameter model for eliminating concentrated errors, a second-order differential equation of torsional vibration of a shafting is establishedExcitation moment substituted into calculated working condition [ F (t)]Selecting a certain position (near the assembly error point) in the shafting as a checking point (K10 in figure 1), and calculating a calculated torque response value T (T) of the checking point along with the change of moment of inertia (J9 in figure 1) according to a second-order differential equation of shafting torsional vibration jisuan Searching the optimal range ((i.e. T (T)) of the calculated torque response value of the checking point of the compressor according to the sensitivity minimum principle jisuan The variation of (a) is smaller along with the variation of the moment of inertia, and the distance between the left and right resonance peaks is longer, as shown in figure 2), and the optimal moment of inertia can be judged to be 1400-160 Kg.M 2 . Wherein the inertia matrix [ J]Consists of moment of inertia J1-J15, torsional matrix [ K ]]From [ K1 ]']、[K2']And K9, damping matrix [ C ]]Consists of C1-C14.
6. According to the optimal rotational inertia ranges of the compressors under all the calculated working conditions, the median value of 1500Kg.M in the J9 optimal rotational inertia ranges is preferably selected 2 And as the rotational inertia design value, designing a compressor shaft system according to the rotational inertia design value, and reserving the installation position of the inertia adjusting device so as to adjust the rotational inertia of the shaft system.
7. After the design and manufacture of the compressor shafting are completed, a strain gauge sensor (at K10 in FIG. 1) is stuck on the checking point, and a test torque response value T (T) of the shafting is tested ceshi The test principle is shown in fig. 3.
8. Will test the torque response value T (T) ceshi And calculating a torque response value T (T) jisuan Performing contrast analysis, and correcting assembly errors according to a contrast analysis result, wherein the principle is as follows:
8.1 true Assembly torsional stiffness K10 at shafting verification Point zhenshi X times the corrected torsional stiffness K10';
8.2 setting the parameter of the corrected assembly error as x, and setting the value range of x (for example, the value range of x is 0.1-3 in the present example), the torsional rigidity K10 after the corrected assembly error xiuzheng ＝K10'*x；
8.3 to K10 xiuzheng Carry-in torsion matrix [ K]Based on the second order differential equation of torsional vibration of shaftingCalculating a corrected calculated torque response value T (T) as a function of the corrected assembly error parameter x jisuan_1 As shown in fig. 4;
8.4 comparative T (T) ceshi And T (T) jisuan_1 Amplitude error between the two is selectedThe x value with the smallest amplitude error multiplied by K10' is used as the real assembly torsional rigidity K10 zhenshi ；
8.5 setting damping matrix [ C ]]Rayleigh damping is adopted, and a shafting real damping matrix [ C ] is adopted] zhengshi ＝[C]*y；
8.6 obtaining a shafting real damping matrix [ C ] according to the method of the steps 8.2-8.4] zhenshi The method comprises the steps of carrying out a first treatment on the surface of the The method comprises the following steps:
setting the parameter of the corrected damping error as y and setting the value range of y, and correcting the torsional rigidity [ C ] after the damping error] xiuzheng ＝[C]*y；
Will [ C ]] xiuzheng Torsional vibration second order differential equation with replacement shaftingIn [ C ]]Calculating a corrected calculated torque response value T (T) as a function of the corrected damping error parameter y jisuan_2 ；
Comparative T (T) ceshi And T (T) jisuan_2 The amplitude error between the two is selected, and the y value with the minimum amplitude error is multiplied by [ C ]]As a real damping matrix [ C] zhenshi 。
9. True to assemble torsional stiffness K10 zhenshi Substitution into a torsion matrix [ K]Let damping matrix [ C ]]Is a real damping matrix [ C] zhenshi Updating second-order differential equation of torsional vibration of shaftingExcitation moment substituted into calculated working condition [ F (t)]Calculating a corrected calculated torque response value T (T) in which the correction point changes with the change in the moment of inertia of J9 jisuan ' the torque response value T (T) will be calculated by correction jisuan ' equal to the calculated torque response value T (T) corresponding to the design value of the moment of inertia jisuan The moment of inertia corresponding to the moment of inertia is taken as a target moment of inertia, and the difference between the target moment of inertia of J9 and the moment of inertia design value is the moment of inertia which needs to be increased or decreased, namely the moment of inertia adjustment value. The inertia adjusting apparatus scheme is shown in fig. 5. The inertia adjusting device consists of two semicircular rings 1, two sets of bolts 2 and nuts 3, and an inner hole of each semicircular ring is tightly arranged on a reserved circular shaft of the compressorThe upper shaft system rotates together, the two sets of bolt fastening devices tightly connect the two semicircular rings 1 together, the inner diameter of the two semicircular rings 1 is determined by the outer diameter of the reserved circular shaft, the outer diameter of the two semicircular rings 1 is determined by the moment of inertia needing to be increased, and if the moment of inertia needing to be increased is larger, the outer diameter of the two semicircular rings 1 is larger.
The invention combines the advantages of various methods such as a centralized parameter method, a finite element method, a torsional vibration test and the like, obtains an accurate shafting torsional vibration model of the compressor, optimizes the torsional vibration of the compressor shafting, and reduces the risk of shafting faults. The method has definite physical meaning, rapid and convenient calculation and high robustness, and greatly improves the running stability and safety of the compressor shafting.
Claims (3)
1. A method for testing-based torsional vibration analysis optimization of a compressor, comprising:
step 1, a three-dimensional model of a compressor shafting is established, and the three-dimensional model of the compressor shafting is converted into a plurality of concentrated parameter models;
step 2, checking by adopting a finite element analysis method, eliminating the centralized errors in each centralized parameter model, and obtaining each centralized parameter model with the centralized errors eliminated;
step 3, establishing a second-order differential equation of torsional vibration of the shafting according to each concentrated parameter model for eliminating concentrated errors, carrying out torsional vibration analysis to obtain a calculated torque response value and an optimal rotational inertia range which change along with the rotational inertia, and selecting a rotational inertia from the optimal rotational inertia range as a rotational inertia design value;
step 4, designing and manufacturing the compressor according to the rotational inertia design value to obtain the compressor;
step 5, performing torsional vibration test on the compressor to obtain a test torque response value;
step 6, comparing and analyzing the test torque response value and the calculated torque response value, and eliminating assembly errors and damping errors according to analysis results to obtain a corrected second-order differential equation of torsional vibration of the shafting;
step 7, adopting a corrected second order differential equation of the torsional vibration of the shafting to carry out torsional vibration analysis to obtain a rotational inertia adjustment value, and adjusting the structure of the compressor according to the rotational inertia adjustment value;
in the step 2, checking is carried out by adopting a finite element analysis method, and the concentration error in each concentration parameter model is eliminated, wherein the specific method comprises the following steps: converting the compressor shafting three-dimensional model into a compressor shafting finite element model for finite element analysis, and eliminating the concentration error in each concentration parameter model according to the Rayleigh principle;
in step 2, for a centralized parameter model, the method for eliminating centralized errors specifically includes:
2.1, acquiring an inertia matrix [ J12] and a rigidity matrix [ K12] according to the concentrated parameter model;
2.2 establishing a free vibration equationSubstitution of inertia matrix [ J12]]And stiffness matrix [ K12]]；
2.3 solving a free vibration equation to obtain a natural frequency w12 and a modal shape [ q12];
2.4, converting the three-dimensional model of the compressor shaft system into a finite element model of the compressor shaft system, and analyzing the finite element model of the compressor shaft system to obtain a natural frequency w12 'and a regular vibration mode [ q12' ];
2.5 calculating the modified stiffness matrix to be [ K12 ]']＝[q12'] T *diag[w12' 2 ]*[q12']；
2.6 let Δk= ([ K12] - [ K12 '])/[ K12' ], Δw= ([ w12] - [ w12 '])/[ w12' ], Δq= ([ q12] - [ q12 '])/[ q12' ], if Δk, Δw, Δq are greater than 1%, check if steps 2.1-2.5 are correct; if the difference is less than or equal to 1%, replacing the corresponding torsional rigidity in the concentrated parameter model by adopting the corrected torsional rigidity in the corrected rigidity matrix [ K12' ], and obtaining a concentrated parameter model for eliminating concentrated errors;
in step 3, specifically: establishing a second-order differential equation of torsional vibration of the shafting according to each concentrated parameter model for eliminating concentrated errorsExcitation moment substituted into calculated working condition [ F (t)]Selecting a checking point in a compressor shaft system, and calculating a torque response value T (T) of the checking point changing along with the change of moment of inertia according to a second-order differential equation of torsional vibration of a shaft system jisuan Searching an optimal range for calculating a torque response value according to a sensitivity minimum principle, obtaining an optimal moment of inertia range according to the optimal range, and selecting a moment of inertia from the optimal moment of inertia range to serve as a moment of inertia design value;
wherein [ J ] is an inertia matrix, [ K ] is a torsion matrix, and [ C ] is a damping matrix;
in step 6, comparing the test torque response value with the calculated torque response value, and eliminating the assembly error according to the analysis result, specifically including:
6.1 true Assembly torsional stiffness K10 at the calibration Point zhenshi X times the corrected torsional stiffness K10'; the corrected torsional rigidity K10' is the corrected torsional rigidity of the correction point in the concentrated parameter model for eliminating the concentrated error;
6.2 setting the parameter for correcting the assembly error as x, and setting the value range of x, so as to correct the torsional rigidity K10 after the assembly error xiuzheng ＝K10'*x；
6.3 torsional stiffness K10 after correction of assembly errors xiuzheng Carry-in torsion matrix [ K]Based on the second order differential equation of torsional vibration of shaftingCalculating a corrected calculated torque response value T (T) as a function of the corrected assembly error parameter x jisuan_1 ；
6.4 contrast test Torque response value T (T) ceshi And correcting and calculating a torque response value T (T) jisuan_1 The amplitude error between the two is selected as x value multiplied by K10' with the minimum amplitude error to be used as the real assembly torsional rigidity K10 zhenshi The method comprises the steps of carrying out a first treatment on the surface of the True to assemble torsional stiffness K10 zhenshi Substituting the second-order differential equation of the torsional vibration of the shafting to obtain the second-order differential equation of the torsional vibration of the shafting, which eliminates assembly errors;
in step 6, comparing the test torque response value with the calculated torque response value, and eliminating the damping error according to the analysis result, specifically:
set damping matrix [ C ]]With Rayleigh damping, the real damping matrix [ C] zhengshi ＝[C]* y; setting the value range of y, the torsional rigidity [ C ] after the damping error is corrected] xiuzheng ＝[C]*y；
Torsional stiffness after correction of damping error [ C] xiuzheng Torsional vibration second order differential equation with replacement shaftingIn [ C ]]Calculating a corrected calculated torque response value T (T) as a function of the corrected damping error parameter y jisuan_2 ；
Comparative T (T) ceshi And T (T) jisuan_2 The amplitude error between the two is selected, and the y value with the minimum amplitude error is multiplied by [ C ]]As a real damping matrix [ C] zhenshi The real damping matrix [ C] zhenshi Substituting the second-order differential equation of the torsional vibration of the shafting to obtain the second-order differential equation of the torsional vibration of the shafting for eliminating damping errors;
the step 7 is specifically as follows: according to the second differential equation of the corrected torsional vibration of the shaftingExcitation moment substituted into calculated working condition [ F (t)]Correction calculation torque response value T (T) of calculation check point changing along with rotation inertia change jisuan ' the torque response value T (T) will be calculated by correction jisuan ' equal to the calculated torque response value T (T) corresponding to the design value of the moment of inertia jisuan And when the corresponding moment of inertia is used as a target moment of inertia, the difference between the target moment of inertia and the moment of inertia design value is the moment of inertia adjustment value, and the compressor structure is adjusted according to the moment of inertia adjustment value.
2. The test-based compressor torsional vibration analysis optimization method according to claim 1, wherein the concentrated parameter model comprises a crankshaft concentrated parameter model, a coupling concentrated parameter model and a motor shaft concentrated parameter model, and the inertia matrix [ J12] is an inertia matrix [ J1] of a crankshaft or an inertia matrix [ J2] of a motor shaft; the rigidity matrix [ K12] is a rigidity matrix [ K1] of the crankshaft or a rigidity matrix [ K2] of the motor shaft; the natural frequency w12 is the natural frequency w1 of the crankshaft or the natural frequency w2 of the motor shaft; the mode shape [ q12] is the mode shape [ q1] of the crankshaft or the mode shape [ q2] of the motor shaft.
3. The method for optimizing torsional vibration analysis of a compressor based on a test of claim 1, wherein step 5 is specifically: setting a strain gauge sensor at a checking point of the compressor, performing torsional vibration test, and acquiring a test torque response value T (T) through the strain gauge sensor ceshi 。
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CN202210290500.8A CN114692449B (en) | 2022-03-23 | 2022-03-23 | Test-based analysis and optimization method for torsional vibration of compressor |
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