CN111753410B - Parameter optimization method for engine torsional damper - Google Patents

Parameter optimization method for engine torsional damper Download PDF

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CN111753410B
CN111753410B CN202010526982.3A CN202010526982A CN111753410B CN 111753410 B CN111753410 B CN 111753410B CN 202010526982 A CN202010526982 A CN 202010526982A CN 111753410 B CN111753410 B CN 111753410B
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torsional vibration
tvd
vibration damper
crankshaft
inertia
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CN111753410A (en
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谭小东
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Chongqing Changan Automobile Co Ltd
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Abstract

The invention discloses a parameter optimization method of an engine torsional vibration damper, which comprises the following steps: establishing a crankshaft multi-degree-of-freedom torsion model, calculating a first-order torsional vibration modal shape and a first-order natural frequency of a crankshaft system, determining the modal inertia of the crankshaft system at the position with the maximum degree of freedom of amplitude, calculating the rotational inertia, the natural frequency and the damping coefficient of the torsional vibration damper, calculating the optimal solution of the natural frequency adjustment factor beta of the torsional vibration damper, and calculating the adjusted natural frequency omega of the torsional vibration damperoptAnd a stiffness KoptWith KoptAs the rigidity parameter for designing the torsional vibration damper, the first-order resonance at the Hub end of the crankshaft is weak and the second-order resonance is strong after the crankshaft is coupled with the TVD, so that the overhall angular acceleration curve at the Hub end of the crankshaft is approximately linearly increased along with the rotating speed, the reasonable matching of the TVD and a crankshaft system is better realized, and the rigidity parameter can be improvedNVH performance of the engine.

Description

Parameter optimization method for engine torsional damper
Technical Field
The invention belongs to the technical field of mechanical rotation and vibration noise, and particularly relates to a parameter optimization method of an engine torsional vibration damper.
Background
Torsional vibration of an engine crankshaft is a ubiquitous vibration mode, and the torsional vibration can generate additional stress on the crankshaft, cause larger noise in a resonance rotating speed range and directly influence the NVH performance of the engine and the fatigue durability of the crankshaft. As engines develop towards high power and high rotational speed, torsional vibration is receiving more and more attention from engineers.
The installation of a torsional vibration damper (i.e., TVD) at the front end of the crankshaft is the most common method for suppressing torsional resonance of the crankshaft. There are a lot of researchers to make a relatively deep study on the matching optimization problem of TVD installed on the crankshaft of the engine. The parameter design method of the TVD has various kinds, and because the principle of the TVD is the same as that of the dynamic vibration absorber and the coherent mass damper, the research method of the dynamic vibration absorber and the coherent mass damper can be referred to during design. In the optimization of the parameters of the shock absorber matched with a crankshaft system with multiple degrees of freedom, the selection of an objective function has multiple forms which can be roughly divided into: minimizing the resonance peak or maximum acceleration of one of the degrees of freedom; minimizing the weighted or sum of squares of resonance peaks for all degrees of freedom; minimizing the maximum transfer rate between some two degrees of freedom; minimizing the maximum power flow of the master oscillator system.
The TVD with a certain vibration reduction effect can be designed by using the methods, but the simplest and common method is a modal inertia method aiming at parameter optimization of the TVD of the crankshaft, namely, a crankshaft multi-degree-of-freedom system is equivalent to a single-degree-of-freedom system at the front end, and the TVD parameters are calculated by using an optimal homodyne formula. The core idea of the method is to make two resonance peaks of the equivalent single-degree-of-freedom system frequency response curve have the same height (namely, optimal coherence), the method is simple and quick, but the method has the following problems when used for optimizing the parameters of the TVD of the crankshaft:
1) the TVD has a 2-order resonance mode after being connected with a crankshaft, in engineering, a lower resonance peak is expected to be left in a conventional rotating speed interval, a higher resonance peak is expected to be pushed to a high rotating speed interval which is less reached, and the design idea of an optimal coherent formula is to obtain two resonance peaks with equal heights, so that the engineering requirements cannot be met frequently.
2) The optimal homodyne formula of the Den Hartog is obtained by derivation based on an undamped main vibration system, the crankshaft system is a weak damping system, and the influence of the damping of the crank system on an optimization result is not considered when the formula is used for calculation.
3) The damping of the rubber (i.e. the rubber of the torsional vibration damper) of the TVD is small, and the optimal damping value (i.e. the damping coefficient reference value of the torsional vibration damper) calculated by the Den Hartog optimal homodyne formula cannot be reached.
4) The response of the front end of the engine is obtained after the response of each order is superposed, and the optimal homodyne formula is only a result obtained by derivation under the excitation of one order; when the frequency response curves of certain order excitation are in a coherent state, the frequency response curves after the overlapping of the order responses are not in the coherent state.
Disclosure of Invention
The invention aims to provide a parameter optimization method of an engine torsional damper, so as to better realize reasonable matching of a TVD and a crankshaft system and improve the NVH performance of an engine.
The invention discloses a parameter optimization method of an engine torsional vibration damper, which comprises the following steps:
s1, establishing a multi-degree-of-freedom torsional vibration model of the crankshaft system in the simulation platform, wherein the vibration equation of the torsional vibration model is as follows:
Figure BDA0002533903210000021
wherein J represents a rotational inertia matrix, C represents a damping matrix, K represents a rigidity matrix, T represents an excitation torque column vector, theta represents an angular displacement vector,
Figure BDA0002533903210000022
Represents the angular velocity vector,
Figure BDA0002533903210000023
Representing an angular acceleration vector;
s2, performing free vibration analysis on the multi-degree-of-freedom torsional vibration model of the crankshaft system, and calculating the first-order torsional vibration modal shape [ y ] of the crankshaft system1,y2,…,yN]TAnd the first order natural frequency omega of the crankshaft systemnN represents the number of degrees of freedom of the crankshaft system multi-degree-of-freedom torsional vibration model;
s3, selecting the mounting position of the torsional vibration damper at the position with the maximum freedom degree of the amplitude on the first-order torsional vibration modal shape of the crankshaft system, and determining the modal inertia of the crankshaft system at the position with the maximum freedom degree of the amplitude as follows:
Figure BDA0002533903210000024
wherein, J1Representing moment of inertia, J, of the stepped axis2Representing the moment of inertia, J, of the main journal3To JN-1Respectively representing the equivalent moment of inertia, J, of the crank-link mechanism of each cylinderNRepresenting the moment of inertia of the flywheel;
s4, calculating the moment of inertia J of the torsional dampertvdNatural frequency omega of torsional vibration dampertvdAnd damping coefficient C of the torsional vibration dampertvd；
S5 adjusting factor by natural frequency of torsional vibration damper
Figure BDA0002533903210000025
In order to optimize variables, an optimization model is established by minimizing the sum of the total angular acceleration linearity of the Hub end of the rear crankshaft (i.e. the end of the crankshaft on which the Hub is mounted, and also the front end of the crankshaft) on which the torsional vibration damper is mounted as an objective function:
Figure BDA0002533903210000026
and carrying out optimization solution in an optimization solver to obtain the optimal solution beta of the natural frequency adjustment factor beta of the torsional damperopt(ii) a Wherein, ω isadjIndicating the natural frequency adjustment value, omega, of the torsional vibration damperlbRepresenting the lowest excitation frequency, ωubRepresents the highest excitation frequency, y (omega) represents a linear function formed by connecting the lowest excitation frequency and the angular acceleration corresponding to the highest excitation frequency, minE (beta) represents an objective function, x (omega, beta) represents the frequency response of the total angular acceleration,
Figure BDA0002533903210000027
and max { } represents a maximum value, Im [, [ solution ]]The expression is taken to be imaginary, i is an imaginary unit, ω is an excitation frequency, t is time, m is the number of cylinders of the engine,
Figure BDA0002533903210000031
m、
Figure BDA0002533903210000032
2m、
Figure BDA0002533903210000033
which represents the order of ignition of the engine,
Figure BDA0002533903210000034
Xm、
Figure BDA0002533903210000035
X2m、
Figure BDA0002533903210000036
representing the angular acceleration frequency response of each order,
Figure BDA0002533903210000037
representing angular acceleration phase angles of respective orders;
s6, using the formula: omegaopt＝βoptωtvdCalculating the natural frequency omega of the tuned torsional vibration damperopt(ii) a Reuse formula: kopt＝ωopt 2JtvdAnd calculating the adjusted rigidity K of the torsional damperopt。
The moment of inertia J of the torsional vibration damper obtained by the above calculationtvdDamping coefficient C of torsional vibration dampertvdAnd the stiffness K of the tuned torsional vibration damperoptThe method is used for optimizing parameters of the engine torsional damper.
Preferably, using the formula: j. the design is a squaretvd＝μJeqCalculating the moment of inertia J of the torsional vibration dampertvd(ii) a Wherein mu represents the inertia ratio of the torsional damper to the crankshaft system, and the value of mu is 0.25-0.3.
Preferably, using the formula: omegatvd＝fωn，
Figure BDA0002533903210000038
Calculating the natural frequency ω of the torsional vibration dampertvd。
Preferably, the formula is firstly used: ctvd'＝2ζJtvdωn，
Figure BDA0002533903210000039
Calculating to obtain a damping coefficient reference value C of the torsional vibration dampertvd'; then, the damping coefficient reference value C of the torsional vibration damper is settvd' maximum torsional damping coefficient C practically achievable with the rubber of the torsional vibration damperrubComparing; if C is presenttvd'＞CrubThen C will berubDamping coefficient C as a torsional vibration dampertvd(ii) a If C is presenttvd'≤CrubThen C will betvd' damping coefficient C as torsional dampertvd。
Preferably, the internal damping coefficient C of the crankshaft shaft section of the engine in the damping matrix C of the step S1rjWith external damping coefficient c at the cylinderojThe method comprises the following steps:
Figure BDA00025339032100000310
coj＝γJjω0,0.04≤γ≤0.08
wherein, KjDenotes the stiffness of the J-th shaft section, JjThe moment of inertia, ω, representing the jth degree of freedom0The crankshaft angular velocity is represented, η represents the adjustment coefficient of the internal damping, and γ represents the adjustment coefficient of the external damping.
The method starts from the point of view of the linearity of the angular acceleration (namely the total angular acceleration) of the Hub end of the crankshaft, minimizes the sum of the linearity of the angular acceleration of the Hub end of the crankshaft as an objective function, takes the natural frequency adjustment factor of the torsional vibration damper as an optimization variable, and then calculates and obtains the rigidity K of the adjusted torsional vibration damperoptWith KoptAs a rigidity parameter for designing the torsional vibration damper, the first-order resonance at the Hub end of the crankshaft is weak and the second-order resonance is strong after the crankshaft is coupled with the TVD, so that the overhall angular acceleration curve at the Hub end of the crankshaft is approximately linearly increased along with the rotating speed, namely, a higher acceleration resonance peak is pushed to a higher rotating speed region which is less reached, and a lower acceleration resonance peak is left in a conventional rotating speed regionCompared with the traditional modal inertia method, the method has the advantages that reasonable matching of the TVD and the crankshaft system is better realized, and the NVH performance of the engine can be improved.
Drawings
Fig. 1 is a flowchart of the optimization process of the present embodiment.
Fig. 2 is a schematic diagram of a multi-degree-of-freedom torsional vibration model of the crankshaft system established in the present embodiment.
FIG. 3 is a schematic diagram of a first-order torsional mode shape of the crankshaft system according to the present embodiment.
Fig. 4 is a model diagram of a link mechanism.
Fig. 5 is a graph of inertia ratio versus amplitude ratio.
FIG. 6 is a diagram of a model of a torsional vibration damper matched with a single degree of freedom system.
Fig. 7 is a diagram illustrating the definition of the fixed point P, Q.
Detailed Description
The present invention will be described in detail with reference to the accompanying drawings.
The theoretical principle adopted by the embodiment is as follows:
in the 2 degree-of-freedom torsional vibration system shown in FIG. 6, JeqRepresenting the equivalent inertia (also modal inertia), K, of the crankshaft systemeqRepresenting the stiffness of the crankshaft system, JtvdRepresenting the moment of inertia, K, of a TVD (i.e. torsional vibration damper)tvdDenotes the stiffness of TVD, CtvdRepresenting the damping coefficient, T, of the TVDARepresenting the magnitude of the excitation torque, ω representing the excitation frequency, t representing time, θ1Representing the angular displacement of the crankshaft system,
Figure BDA0002533903210000041
is indicative of the angular velocity of the crankshaft system,
Figure BDA0002533903210000042
representing angular acceleration, theta, of the crankshaft system2Representing the angular displacement of the TVD,
Figure BDA0002533903210000043
which represents the angular velocity of the TVD,
Figure BDA0002533903210000044
representing the angular acceleration of the TVD.
The vibration equation for the degree of freedom torsional vibration system is then:
Figure BDA0002533903210000045
the amplitude of the steady state solution of the crankshaft system can be obtained as follows:
Figure BDA0002533903210000046
for the purpose of theoretical analysis, the following non-dimensional symbols are introduced, as shown in Table 1.
TABLE 1 symbols of parameters and physical meanings
Figure BDA0002533903210000051
The ratio of amplitude to static deformation (amplitude ratio for short) of a crankshaft system can be expressed as:
Figure BDA0002533903210000052
when the rigidity and the rotational inertia of the TVD are unchanged, no matter how the damping coefficient of the TVD is changed, the frequency response curve of the crankshaft system passes through P, Q two fixed points. The fixed points P and Q can be obtained by intersecting two curves in formula (3) by letting ξ ═ 0 and ξ ∞ respectively, as shown in fig. 7. The TVD has the best damping effect when the two fixed points are equal in height and just at the two formant peaks, respectively.
When ζ is 0, the ratio of amplitude to static deformation of the crankshaft system is:
Figure BDA0002533903210000053
when ζ ∞, the ratio of the amplitude of the crankshaft system to the static deformation is:
Figure BDA0002533903210000054
the two formulas are combined to obtain:
(2+μ)g4-2(1+f2+μf2)g2+2f2＝0 (6)
according to the Weddar theorem, P, Q two points satisfy the functional relationship:
Figure BDA0002533903210000061
when P, Q has two points equal in height, both of these two points are on the amplitude ratio curve ζ ∞, which can be obtained from equation (5):
Figure BDA0002533903210000062
simplifying to obtain:
Figure BDA0002533903210000063
formula (7) and formula (9) are combined to obtain:
Figure BDA0002533903210000064
the frequency ratio g between P, Q points can be obtained by substituting equation (10) for equation (6)PAnd gQIn the formula (3), the damping coefficient is adjusted to make the derivative of the amplitude magnification curve zero when the amplitude magnification curve passes through P, Q, so that P, Q two points are ensured as the maximum point of the amplitude magnification curve.
Through complicated calculation, the optimal damping ratio is finally obtained as follows:
Figure BDA0002533903210000065
let the left hand setpoint P be lower than the right hand setpoint Q, equation (8) becomes:
Figure BDA0002533903210000066
simplifying to obtain:
Figure BDA0002533903210000067
combining formula (13) with formula (7):
Figure BDA0002533903210000068
therefore, under the condition that the moment of inertia and the damping coefficient of the TVD are kept unchanged, the natural frequency of the TVD is reduced, and the left peak value and the right peak value of the crankshaft system are reduced and increased; conversely, increasing the natural frequency of the TVD, the left peak of the crankshaft system increases and the right peak decreases.
As shown in fig. 1, in the present embodiment, the engine is a four-cylinder engine, and the parameter optimization method of the engine torsional damper includes the following steps:
s1, establishing a crankshaft system multi-degree-of-freedom torsional vibration model in Amesim software (see figure 2), wherein the vibration equation of the crankshaft system multi-degree-of-freedom torsional vibration model is as follows:
Figure BDA0002533903210000071
wherein J represents a rotational inertia matrix, C represents a damping matrix, K represents a rigidity matrix, T represents an excitation torque column vector, theta represents an angular displacement vector,
Figure BDA0002533903210000072
Represents the angular velocity vector,
Figure BDA0002533903210000073
Representing an angular acceleration vector.
The main principles to be followed in modeling are the following two points:
(1) each degree of freedom is regarded as a rigid rotating body only with inertia, the degrees of freedom are connected with a damper through a spring without mass, and the excitation torque of an engine is completely applied to the corresponding degree of freedom.
(2) The portions with greater or concentrated moment of inertia are considered degrees of freedom and the portions with less or dispersed inertia are considered elastic segments. The moment of inertia of the shaft section is equally distributed to the two degrees of freedom connected to it.
The cylinder pressure curve is obtained by test, the relevant damping coefficient takes an empirical value, and the internal damping coefficient C of the engine crankshaft shaft section in the damping matrix C is usedrjWith external damping coefficient c at the cylinderojThe method comprises the following steps:
Figure BDA0002533903210000074
coj＝γJjω0,0.04≤γ≤0.08
wherein, KjDenotes the stiffness of the J-th shaft section, JjThe moment of inertia, ω, representing the jth degree of freedom0The crankshaft angular velocity is represented, η represents the adjustment coefficient of the internal damping, and γ represents the adjustment coefficient of the external damping.
S2, performing free vibration analysis on the multi-degree-of-freedom torsional vibration model of the crankshaft system, and calculating the first-order torsional vibration modal shape [ y ] of the crankshaft system1,y2,…,y7]T(see FIG. 3) and first order natural frequency ω of the crankshaft systemn。
S3, selecting the installation position of the torsional vibration damper at the position with the maximum degree of freedom of amplitude on the first-order torsional vibration mode of the crankshaft system, and determining the modal inertia of the crankshaft system at the position with the maximum degree of freedom of amplitude as follows:
Figure BDA0002533903210000075
wherein, J1Representing moment of inertia, J, of the stepped axis2Representing the moment of inertia, J, of the main journal3To J6Respectively representing equivalent inertia moments, J, of 4 crank-link mechanisms7Representing the moment of inertia of the flywheel.
The installation position of the torsional damper is selected at the position with the maximum freedom degree of amplitude on the first-order torsional vibration mode of the crankshaft system, and the purpose is to absorb more energy. The maximum freedom degree of amplitude on the first-order torsional vibration modal shape of the crankshaft system is positioned at the front end of the crankshaft, and the modal inertia JeqWith a first degree of freedom J1The equivalent single-degree-of-freedom system has the same energy as the original multi-degree-of-freedom system, and the modal inertia J can be obtained according to the energy conservation principleeqThe calculation formula of (c).
The equivalent moment of inertia of the crank-link mechanism being the sum of the crank inertia and the equivalent inertia of the link mechanism, e.g. J3＝Jw+JconWherein, JwIs the inertia of the crank, JconIs the equivalent inertia of the linkage.
The equivalent inertia of the linkage is calculated as:
Figure BDA0002533903210000081
wherein m isBRepresenting the equivalent mass of the big end of the connecting rod, mARepresenting the mass of the small end of the connecting rod, mpRepresenting the mass of the piston assembly and r is the crank radius. In the link mechanism shown in fig. 4, it is assumed that point c is the center of mass of the link and the distance from the center of the left side to point c is l1The distance from the center of the circle on the right side is l2，mBAnd mAThe calculation method of (c) is as follows:
Figure BDA0002533903210000082
s4, calculating the moment of inertia J of the torsional dampertvdNatural frequency omega of torsional vibration dampertvdAnd damping coefficient C of the torsional vibration dampertvd。
Using the formula: j. the design is a squaretvd＝μJeqCalculating the moment of inertia J of the torsional vibration dampertvd(ii) a Wherein mu represents the inertia ratio of the torsional damper to the crankshaft system, and the value of mu is 0.25-0.3. Maximum value of the ratio between amplitude and static deformation of the crankshaft system (theta) in the optimum coherent regime1/δst) Is composed of
Figure BDA0002533903210000083
The relationship with μ is shown in fig. 5. The larger the moment of inertia of the TVD, the better the damping effect, but the larger the moment of inertia is, the contrary to the light weight design concept, and in order to achieve a balance between the light weight design and the damping effect, the ratio of the moment of inertia to the modal inertia of the TVD is selected to be about 0.3.
In practical application, the TVD is divided into a secondary stage and a primary stage, the Ring of the secondary TVD is arranged on the outer Ring, the belt is connected with the Ring, the Ring of the primary TVD is arranged on the inner Ring, and the belt is connected with the Hub. Because the Ring turning radius of the primary TVD is smaller, the Ring mass is too large if the inertia design is larger, and the Ring turning radius of the secondary TVD is larger than that of the primary TVD, so that the inertia of the Ring can be properly increased to improve the damping effect. It is proposed to select the inertia ratio of the secondary TVD to be 0.3 and the inertia ratio of the primary TVD to be 0.25.
Using the formula: omegatvd＝fωn，
Figure BDA0002533903210000084
Calculating the natural frequency ω of the torsional vibration dampertvd。
Firstly, using a formula: ctvd'＝2ζJtvdωn，
Figure BDA0002533903210000085
Calculating to obtain a damping coefficient reference value C of the torsional vibration dampertvd'; then, the damping coefficient reference value C of the torsional vibration damper is settvd' maximum torsional damping coefficient C practically achievable with the rubber of the torsional vibration damperrubComparing; if C is presenttvd'＞CrubThe maximum torsional damping coefficient C that can be actually achieved by the rubber of the torsional vibration damperrubAs torsional vibration dampersDamping coefficient C oftvd(ii) a If C is presenttvd'≤CrubThen reference value C of damping coefficient of the torsional vibration dampertvd' damping coefficient C as torsional dampertvd. Normally, the maximum torsion damping coefficient actually achieved by the rubber of the TVD cannot exceed 0.4Nm/rad/s, and the optimal damping coefficient required by the TVD is far higher than the maximum torsion damping coefficient, so that the damping coefficient C of the torsion damper is reducedtvdThe rate was taken to be 0.4 Nm/rad/s.
S5, connecting matlab software with Amesim software, calling fmincon optimization solver in matlab, and adjusting factors by natural frequency of the torsional vibration damper
Figure BDA0002533903210000091
In order to optimize variables, the sum of the total angular acceleration linearity of the rear crankshaft Hub end provided with the torsional damper is minimized to be an objective function, and an optimization model is established:
Figure BDA0002533903210000092
obtaining the optimal solution beta of the natural frequency adjustment factor beta of the torsional damper by optimal solutionopt(ii) a Wherein, ω isadjIndicating the natural frequency adjustment value, omega, of the torsional vibration damperlbRepresenting the lowest excitation frequency, ωubRepresents the highest excitation frequency, y (omega) represents a linear function formed by connecting the lowest excitation frequency and the angular acceleration corresponding to the highest excitation frequency, minE (beta) represents an objective function, x (omega, beta) represents the frequency response of the total angular acceleration (namely, the overhall angular acceleration),
Figure BDA0002533903210000093
and max { } denotes taking the maximum value, Im [, [ 2 ]]The expression takes the imaginary part, i denotes the unit of imaginary number, ω denotes the excitation frequency, t denotes the time, X2、X4、X6、X8、X10Representing the angular acceleration frequency response of each order,
Figure BDA0002533903210000094
indicating the angular acceleration phase angle of each order.
S6, using the formula: omegaopt＝βoptωtvdCalculating the natural frequency omega of the tuned torsional vibration damperopt(ii) a Reuse formula: kopt＝ωopt 2JtvdAnd calculating the adjusted rigidity K of the torsional damperopt. Stiffness K with tuned torsional vibration damperoptAs the rigidity used in the design of the TVD, the first-order resonance of the crankshaft Hub end after the crankshaft system is coupled with the TVD is weak, and the second-order resonance is strong, so that the overhall angular acceleration curve of the crankshaft Hub end is increased approximately linearly along with the rotating speed.

Claims (5)

1. A method of optimizing parameters of an engine torsional vibration damper, comprising:
s1, establishing a multi-degree-of-freedom torsional vibration model of the crankshaft system in the simulation platform, wherein the vibration equation of the torsional vibration model is as follows:
Figure FDA0002533903200000011
wherein J represents a rotational inertia matrix, C represents a damping matrix, K represents a rigidity matrix, T represents an excitation torque column vector, theta represents an angular displacement vector,
Figure FDA0002533903200000012
Represents the angular velocity vector,
Figure FDA0002533903200000013
Representing an angular acceleration vector;
s2, performing free vibration analysis on the multi-degree-of-freedom torsional vibration model of the crankshaft system, and calculating the first-order torsional vibration modal shape [ y ] of the crankshaft system1,y2,…,yN]TAnd the first order natural frequency omega of the crankshaft systemnN represents the number of degrees of freedom of the crankshaft system multi-degree-of-freedom torsional vibration model;
s3, selecting the mounting position of the torsional vibration damper at the position with the maximum freedom degree of the amplitude on the first-order torsional vibration modal shape of the crankshaft system, and determining the modal inertia of the crankshaft system at the position with the maximum freedom degree of the amplitude as follows:
Figure FDA0002533903200000014
wherein, J1Representing moment of inertia, J, of the stepped axis2Representing the moment of inertia, J, of the main journal3To JN-1Respectively representing the equivalent moment of inertia, J, of the crank-link mechanism of each cylinderNRepresenting the moment of inertia of the flywheel;
s4, calculating the moment of inertia J of the torsional dampertvdNatural frequency omega of torsional vibration dampertvdAnd damping coefficient C of the torsional vibration dampertvd；
S5 adjusting factor by natural frequency of torsional vibration damper
Figure FDA0002533903200000015
In order to optimize variables, the sum of the total angular acceleration linearity of the rear crankshaft Hub end provided with the torsional damper is minimized to be an objective function, and an optimization model is established:
Figure FDA0002533903200000016
and carrying out optimization solution in an optimization solver to obtain the optimal solution beta of the natural frequency adjustment factor beta of the torsional damperopt(ii) a Wherein, ω isadjIndicating the natural frequency adjustment value, omega, of the torsional vibration damperlbRepresenting the lowest excitation frequency, ωubRepresents the highest excitation frequency, y (omega) represents a linear function formed by connecting the lowest excitation frequency and the angular acceleration corresponding to the highest excitation frequency, minE (beta) represents an objective function, x (omega, beta) represents the frequency response of the total angular acceleration,
Figure FDA0002533903200000017
and max { } denotes taking the maximum value, Im [, [ 2 ]]The expression is taken to be imaginary, i is an imaginary unit, ω is an excitation frequency, t is time, m is the number of cylinders of the engine,
Figure FDA0002533903200000018
m、
Figure FDA0002533903200000019
2m、
Figure FDA00025339032000000110
which represents the order of ignition of the engine,
Figure FDA00025339032000000111
Xm、
Figure FDA00025339032000000112
X2m、
Figure FDA00025339032000000113
representing the angular acceleration frequency response of each order,
Figure FDA00025339032000000114
representing angular acceleration phase angles of respective orders;
s6, using the formula: omegaopt＝βoptωtvdCalculating the natural frequency omega of the tuned torsional vibration damperopt(ii) a Reuse formula: kopt＝ωopt 2JtvdAnd calculating the adjusted rigidity K of the torsional damperopt。
2. The method for optimizing parameters of an engine shock absorber according to claim 1, characterized in that the formula is used: j. the design is a squaretvd＝μJeqCalculating the moment of inertia J of the torsional vibration dampertvd(ii) a Wherein mu represents the inertia ratio of the torsional damper to the crankshaft system, and the value of mu is 0.25-0.3.
3. The method for optimizing parameters of an engine torsional vibration damper of claim 2, characterized in that the formula is utilized: omegatvd＝fωn，
Figure FDA0002533903200000021
Calculating the natural frequency ω of the torsional vibration dampertvd。
4. A method for optimizing parameters of an engine torsional vibration damper according to claim 2 or 3, characterized in that the formula is used: ctvd'＝2ζJtvdωn，
Figure FDA0002533903200000022
Calculating to obtain a damping coefficient reference value C of the torsional vibration dampertvd'; then, the damping coefficient reference value C of the torsional vibration damper is settvd' maximum torsional damping coefficient C practically achievable with the rubber of the torsional vibration damperrubComparing; if C is presenttvd'＞CrubThen C will berubDamping coefficient C as a torsional vibration dampertvd(ii) a If C is presenttvd'≤CrubThen C will betvd' damping coefficient C as torsional dampertvd。
5. The method for optimizing parameters of an engine torsional vibration damper of claim 4, wherein the internal damping coefficient C of the engine crankshaft section in the damping matrix C of step S1rjWith external damping coefficient c at the cylinderojThe method comprises the following steps:
Figure FDA0002533903200000023
coj＝γJjω0,0.04≤γ≤0.08
wherein, KjDenotes the stiffness of the J-th shaft section, JjThe moment of inertia, ω, representing the jth degree of freedom0The crankshaft angular velocity is represented, η represents the adjustment coefficient of the internal damping, and γ represents the adjustment coefficient of the external damping.
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