CN113738801A - Design method of composite shock absorber and composite shock absorber - Google Patents

Design method of composite shock absorber and composite shock absorber Download PDF

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CN113738801A
CN113738801A CN202010460217.6A CN202010460217A CN113738801A CN 113738801 A CN113738801 A CN 113738801A CN 202010460217 A CN202010460217 A CN 202010460217A CN 113738801 A CN113738801 A CN 113738801A
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vibration
radial
torsional
composite
damper
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CN113738801B (en
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齐晓旭
宁明志
郭健
谢少华
王超
李冰莲
李文琪
孙雷
高恩猛
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SAIC Motor Corp Ltd
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F1/00—Springs
    • F16F1/36—Springs made of rubber or other material having high internal friction, e.g. thermoplastic elastomers
    • F16F1/38—Springs made of rubber or other material having high internal friction, e.g. thermoplastic elastomers with a sleeve of elastic material between a rigid outer sleeve and a rigid inner sleeve or pin, i.e. bushing-type
    • G—PHYSICS
    • G06—COMPUTING; CALCULATING OR COUNTING
    • G06F—ELECTRIC DIGITAL DATA PROCESSING
    • G06F30/00—Computer-aided design [CAD]
    • G06F30/10—Geometric CAD
    • G06F30/15—Vehicle, aircraft or watercraft design
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F2230/00—Purpose; Design features
    • F16F2230/32—Modular design
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F2236/00—Mode of stressing of basic spring or damper elements or devices incorporating such elements
    • F16F2236/12—Mode of stressing of basic spring or damper elements or devices incorporating such elements loaded in combined stresses
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T90/00—Enabling technologies or technologies with a potential or indirect contribution to GHG emissions mitigation

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Abstract

The invention discloses a design method of a composite shock absorber and the composite shock absorber, wherein the design method comprises the following steps: a. debugging the torsional vibration parameters of the composite damper according to the torsional resonance frequency of the transmission system when NVH complaints exist, and obtaining the range of the torsional vibration parameters; b. debugging the radial vibration parameters of the composite vibration absorber according to the radial resonance frequency of the transmission system when NVH complaints exist, and obtaining the range of the radial vibration parameters; c. obtaining a verification parameter of the composite damper when the torsional vibration and the radial vibration both meet the damping requirement according to the torsional vibration parameter range and the radial vibration parameter range, and obtaining a composite damping condition according to the verification parameter; d. and obtaining a composite vibration reduction condition according to the verification parameters, and trial-producing a sample piece according to the composite vibration reduction condition for loading verification. The parameters of the torsional vibration damper are adjusted to reach the expected range, and the purpose of simultaneously inhibiting radial vibration and torsional vibration by using one composite vibration damper is realized, so that the radial vibration damper is eliminated, and the cost is saved.

Description

Design method of composite shock absorber and composite shock absorber
Technical Field
The invention relates to the technical field of vibration reduction of an automobile transmission system, in particular to a design method of a composite vibration absorber and the composite vibration absorber.
Background
In the prior art, torsional vibration and radial vibration of a transmission system are considered separately, namely the torsional vibration is restrained by a Torsional Vibration Damper (TVD), and only the function of restraining the torsional vibration is achieved; the radial vibration Damper (Damper) is used to damp the radial vibration, and has only a function of damping the radial vibration.
Disclosure of Invention
The invention aims to solve the problems of large gravity and high cost caused by the fact that related parts are separately arranged in torsional vibration and radial vibration of a transmission system in the prior art. The method achieves the purpose of simultaneously inhibiting radial vibration and torsional vibration by using one composite vibration absorber by adjusting various parameters of the torsional vibration absorber to reach the expected range, thereby eliminating the vibration absorber of the radial vibration and saving the cost and the weight.
In order to solve the above problems, a method for designing a composite damper is provided, which comprises the following steps:
a. debugging the torsional vibration parameters of the composite vibration damper according to the torsional resonance frequency of the transmission system when NVH complaints exist, and obtaining the torsional vibration parameter range of the composite vibration damper;
b. debugging the radial vibration parameters of the composite vibration damper according to the radial resonance frequency of the transmission system when NVH complaints exist, and obtaining the radial vibration parameter range of the composite vibration damper;
c. according to the torsional vibration parameter range and the radial vibration parameter range, obtaining verification parameters of the composite vibration absorber when the torsional vibration and the radial vibration both meet the vibration reduction requirements;
d. and obtaining a composite vibration reduction condition according to the verification parameters, and trial-producing a sample piece according to the composite vibration reduction condition for loading verification.
By adopting the scheme, the approximate ranges of the torsional vibration parameter and the radial vibration parameter are determined by combining the torsional resonance frequency and the radial torsional resonance frequency of the transmission system, the conditions such as the material or the radius of the composite damper are continuously changed, the torsional vibration parameter and the radial vibration parameter are debugged, the vibration parameter range meeting the vibration reduction condition is determined, the parameter when the torsional vibration and the radial vibration meet the vibration reduction requirement is found according to the radial vibration parameter range and the torsional vibration parameter range, the composite vibration reduction condition such as the material or the radius of the composite vibration damper is obtained by taking the parameter as a verification parameter, and the composite vibration damper capable of simultaneously inhibiting the radial vibration and the torsional vibration can be obtained according to the composite vibration reduction condition. The composite damper is used for replacing two parts of a radial vibration damper and a torsional damper, the effect of suppressing vibration and noise of the whole vehicle is achieved, and the comfort of passengers in the vehicle is improved. The damper is arranged on the transmission shaft, the radial vibration damper of the transmission shaft is cancelled, the weight is reduced by about 1 Kg/vehicle, and the cost is reduced by about 25 Yuan/vehicle; the rear main reducer shock absorber is cancelled, the weight is reduced by about 1.2 Kg/car, and the cost reduction is estimated to be 80 Yuan/car.
According to another specific embodiment of the present invention, the method for designing a composite vibration absorber, which is disclosed by the embodiment of the present invention, the method for debugging the vibration parameter range obtained in the step a and the step b, comprises the following steps:
s1, establishing a radial vibration model, and obtaining radial vibration parameters according to the radial vibration model;
s2, obtaining radial vibration amplitude according to the radial vibration parameters, adjusting the radial vibration parameters to obtain a plurality of corresponding radial vibration amplitudes, and establishing a radial vibration parameter-amplitude curve according to each radial vibration amplitude;
s3, establishing a torsional vibration model, and obtaining torsional vibration parameters according to the torsional vibration model;
s4, obtaining torsional vibration amplitude according to the torsional vibration parameters, adjusting the torsional vibration parameters to obtain a plurality of corresponding torsional vibration amplitudes, and establishing a torsional vibration parameter-amplitude curve according to each torsional vibration amplitude;
and s5, obtaining a radial vibration parameter range according to the radial vibration parameter-amplitude curve, and obtaining a torsional vibration parameter range according to the torsional vibration parameter-amplitude curve.
By adopting the scheme, the debugging conditions and the association conditions required by the corresponding parameters can be rapidly and controllably determined and changed, namely, the qualitative analysis can be increased to the controllable quantitative analysis from the association relation only by adopting the scheme, the experiment times and cost are reduced, the vibration reduction effect is evaluated by the vibration amplitude, the parameter range can be rapidly and accurately determined by establishing the vibration parameter-amplitude comparison table, and therefore the verification parameters which accord with the inhibition of the double vibration can be rapidly found.
According to another embodiment of the present invention, the embodiment of the present invention discloses a method for designing a composite vibration absorber, wherein the torsional vibration parameters include a torsional natural frequency of the composite vibration absorber, a moment of inertia of an outer ring, and a torsional damping of rubber of the composite vibration absorber; the radial vibration parameters comprise the radial natural frequency of the composite vibration absorber, the mass of the outer ring and the radial damping of rubber; the torsional and radial damping of the rubber is determined by the material of the rubber.
By adopting the scheme, the effect of inhibiting the vibration is evaluated by using the amplitude, experiments show that the amplitude of the torsional vibration of the composite vibration absorber is related to the torsional natural frequency, the rotational inertia of the outer ring and the torsional damping of the rubber of the composite vibration absorber, and the effect of inhibiting the torsional vibration can be changed by changing the parameters. Meanwhile, the amplitude and the radial natural frequency of the radial vibration, the mass of the outer ring and the radial damping of the rubber can change the effect of inhibiting the radial vibration by changing the parameters. Therefore, the vibration reduction effect can be debugged according to the correlation of the parameters. Wherein, the damping is measured by a damping factor.
According to another specific embodiment of the invention, the embodiment of the invention discloses a design method of a composite damper, wherein the radial natural frequency is determined according to radial influence factors, and the radial influence factors comprise the material of rubber and the radial rigidity of the rubber; the torsional natural frequency is determined from torsional influencing factors, including the material of the rubber and the torsional stiffness of the rubber.
By adopting the scheme, experiments show that the radial natural frequency is related to the material of the rubber and the radial rigidity of the rubber, and the radial natural frequency can be changed by changing the material of the rubber and the radial rigidity of the rubber; meanwhile, the torsional natural frequency is related to the material of the rubber and the torsional rigidity of the rubber, and the torsional natural frequency can be changed by changing the material of the rubber and the torsional rigidity of the rubber. Therefore, the natural frequency can be further changed by changing the material of the rubber, the radial rigidity of the rubber and the torsional rigidity of the rubber, and the vibration damping effect can be further adjusted. The rigidity can be adjusted by arranging the structure of the allowable space design outer ring, the installation framework and the rubber, and the adjustment of the structure comprises whether holes are formed, the size of the holes, the cross section shape of the rubber and the like.
In accordance with another embodiment of the present invention, a method of designing a composite shock absorber is disclosed, wherein the radial influencing factor and the torsional influencing factor each further comprise a radius of an outer race of the composite shock absorber, the radius comprising an inner diameter and an outer diameter.
By adopting the scheme, experiments prove that the radial natural frequency and the torsional natural frequency of the composite damper can be simultaneously influenced by the radius of the outer ring of the composite damper. Therefore, the natural frequency can be further changed by changing the radius of the outer ring, and the vibration damping effect can be further debugged.
According to another embodiment of the present invention, the radial influencing factor of the design method of the composite damper further comprises an additional module, wherein the additional module comprises a pin, a metal insert or additional rubber with hardness different from that of the rubber of the composite damper, and the pin, the metal insert or the additional rubber is arranged on the composite damper.
With the above-described arrangement, it has been experimentally found that the radial natural frequency can be varied by providing additional modules, such as pins or metal inserts, to the composite damper. On the basis that the torsional natural frequency meets the vibration reduction effect, the radial natural frequency is adjusted by adding some auxiliary methods, and the better vibration reduction effect is further achieved.
According to another specific embodiment of the present invention, the embodiment of the present invention discloses a method for designing a composite damper, wherein the radial vibration amplitude and the radial vibration parameter satisfy the following relation:
Figure BDA0002510667950000041
wherein, A1-radial vibration amplitude of the composite vibration absorber;
μ1the mass ratio of the outer ring of the compound damper to the mass of the transmission system;
f1-the ratio of the radial natural frequency of the compound damper to the radial resonant frequency of the transmission system;
g1-the ratio of the radial excitation frequency of the transmission system to the radial resonance frequency of the transmission system;
ξ1-the radial damping ratio of the rubber; and the number of the first and second electrodes,
the torsional vibration amplitude and the torsional vibration parameter satisfy the relation:
Figure BDA0002510667950000042
wherein A is2-torsional vibration amplitude of the composite damper;
λ2-the ratio of the torsional excitation frequency of the transmission system to the torsional resonance frequency of the transmission system;
α2the torsional natural frequency of the composite damper is compared to the torsional resonance frequency of the transmission system;
μ2the ratio of the rotational inertia of the compound damper to the rotational inertia of the transmission system;
ξ2-torsional damping ratio of the rubber.
By adopting the scheme, the formula can be obtained through modeling analysis and experiments. Thereby obtaining the torsional vibration parameters influencing the torsional vibration amplitude and the relation between the radial vibration parameters influencing the radial vibration amplitude. And then make the debugging process more quick, accurate and controllable, improved developer's design efficiency, reduced development cost.
According to another embodiment of the present invention, a method for designing a composite vibration absorber is disclosed, wherein the natural torsional frequency is determined according to the following relation:
Figure BDA0002510667950000051
wherein f isTorsion-torsional natural frequency of the composite damper;
Ktorsion-the torsional stiffness of the outer ring about the axial direction of the composite damper;
rho is the material density of the outer ring of the composite shock absorber;
h is the width of the outer ring of the composite damper;
r1-an inner diameter of an outer race of the compound damper;
r2-an outer diameter of an outer race of the composite damper;
the radial natural frequency is determined according to the following relation:
Figure BDA0002510667950000052
wherein, KRadial directionThe deformation rigidity of the outer ring of the composite damper in the radial direction;
fradial directionRadial natural frequency of the composite damper.
By adopting the scheme, the formula can be obtained through modeling analysis and experiments. Resulting in various conditions affecting the torsional natural frequency and various conditions affecting the radial natural frequency. And then make the debugging process more quick, accurate and controllable, improved developer's design efficiency, reduced development cost.
According to another specific embodiment of the present invention, the method for designing a composite vibration absorber according to the embodiment of the present invention, wherein the method for obtaining the composite vibration absorbing condition according to the verification parameter in step d comprises:
d-1, obtaining the required radius range of the rotational inertia of the outer ring according to the mass range of the outer ring;
d-2, determining a rubber material according to the torsional damping and the radial damping;
d-3, determining the structure of the rubber of the composite damper according to the radial natural frequency and the torsional natural frequency;
d-4, adding an additional module according to the radial natural frequency on the basis of meeting the torsional natural frequency.
By adopting the scheme, after the verification parameters which can simultaneously play two vibration reduction effects are found through debugging, the verification parameters need to be further converted into the verification parameters which are represented by the verification parameters and accord with the vibration reducer materials, the modes, the additional parts and the like, namely, the verification parameters are converted into the composite vibration reduction conditions. The radius of the composite shock absorber can be obtained through the mass and the rotational inertia of the outer ring; determining a rubber material of the composite shock absorber according to the damping, and further determining a rubber formula according to the rubber material; determining the radial stiffness and the torsional stiffness of the rubber of the composite damper, for example, the rubber, based on the radial natural frequency and the torsional natural frequency; and the required accessory modules such as an insert plate, a pin and the like can be determined according to the radial natural frequency.
The composite shock absorber comprises an installation framework, rubber and an outer ring, wherein the rubber is arranged on the installation framework, the outer ring is sleeved on the rubber, and the composite shock absorber is designed according to the design method of the composite shock absorber.
The invention has the beneficial effects that:
the method achieves the purpose of simultaneously inhibiting radial vibration and torsional vibration by using one composite vibration absorber by adjusting various parameters of the composite vibration absorber to reach the expected range, thereby eliminating the vibration absorber with radial vibration and saving cost and weight.
Drawings
FIG. 1 is a flow chart of a method of designing a compound shock absorber according to embodiment 1 of the present invention;
FIG. 2 is a simplified radial vibration model of embodiment 1 of the present invention;
FIG. 3 is a simplified model diagram of torsional vibration in accordance with embodiment 1 of the present invention;
FIG. 4 is a radial amplitude frequency response graph of example 1 of the present invention;
FIG. 5 is a radial amplitude-frequency response curve under the influence of different outer ring qualities in embodiment 1 of the present invention;
FIG. 6 is a radial amplitude-frequency response curve under the influence of different damping factors of embodiment 1 of the present invention;
FIG. 7 is a graph of the torsional amplitude frequency response of example 1 of the present invention;
fig. 8 is a schematic structural view of a composite vibration damper according to embodiment 2 of the present invention.
Description of reference numerals:
1: an outer ring; 2: rubber; 3: installing a framework; 4: a transmission system; 5: a damper.
Detailed Description
The following description of the embodiments of the present invention is provided for illustrative purposes, and other advantages and effects of the present invention will become apparent to those skilled in the art from the present disclosure. While the invention will be described in conjunction with the preferred embodiments, it is not intended that features of the invention be limited to these embodiments. On the contrary, the invention is described in connection with the embodiments for the purpose of covering alternatives or modifications that may be extended based on the claims of the present invention. In the following description, numerous specific details are set forth in order to provide a thorough understanding of the present invention. The invention may be practiced without these particulars. Moreover, some of the specific details have been left out of the description in order to avoid obscuring or obscuring the focus of the present invention. It should be noted that the embodiments and features of the embodiments may be combined with each other without conflict.
It should be noted that in this specification, like reference numerals and letters refer to like items in the following drawings, and thus, once an item is defined in one drawing, it need not be further defined and explained in subsequent drawings.
In the description of the present embodiment, it should be noted that the terms "upper", "lower", "inner", "bottom", and the like indicate orientations or positional relationships based on the orientations or positional relationships shown in the drawings or orientations or positional relationships that the products of the present invention usually place when used, and are only used for convenience of description and description, but do not indicate or imply that the devices or elements indicated must have specific orientations, be constructed in specific orientations, and operate, and thus, should not be construed as limiting the present invention.
The terms "first," "second," and the like are used solely to distinguish one from another and are not to be construed as indicating or implying relative importance.
In the description of the present embodiment, it should be further noted that, unless explicitly stated or limited otherwise, the terms "disposed," "connected," and "connected" are to be interpreted broadly, e.g., as a fixed connection, a detachable connection, or an integral connection; can be mechanically or electrically connected; they may be connected directly or indirectly through intervening media, or they may be interconnected between two elements. The specific meanings of the above terms in the present embodiment can be understood in specific cases by those of ordinary skill in the art.
In order to make the objects, technical solutions and advantages of the present invention more apparent, embodiments of the present invention will be described in detail with reference to the accompanying drawings.
Example 1
In order to solve the above problems, a method for designing a composite vibration absorber is provided, as shown in fig. 1, including the steps of:
a. debugging the torsional vibration parameters of the composite vibration damper according to the torsional resonance frequency of the transmission system when NVH complaints exist, and obtaining the torsional vibration parameter range of the composite vibration damper;
b. debugging the radial vibration parameters of the composite vibration damper according to the radial resonance frequency of the transmission system when NVH complaints exist, and obtaining the radial vibration parameter range of the composite vibration damper;
c. according to the torsional vibration parameter range and the radial vibration parameter range, obtaining verification parameters of the composite vibration absorber when the torsional vibration and the radial vibration both meet the vibration reduction requirements;
d. and obtaining a composite vibration reduction condition according to the verification parameters, and trial-producing a sample piece according to the composite vibration reduction condition for loading verification.
Specifically, the radial frequency in this specification refers to a vibration frequency in the radial direction; the torsional frequency refers to the vibration frequency in the circumferential rotation direction. The torsional and radial resonant frequencies of the driveline are the frequencies of NVH that are complained about due to resonance.
NVH is an english abbreviation for Noise, Vibration and Harshness (Noise, Vibration, Harshness), which is a comprehensive index in the field for measuring automobile manufacturing quality. The natural frequency of the transmission system is divided into a radial natural frequency and a torsional natural frequency, and the radial natural frequency and the torsional natural frequency of the transmission system are the natural properties of the transmission system. There are two categories of driveline complaints (i.e., the frequency of excitation to which the system is subjected is close to the natural frequency), radial (bending) resonance complaints, and torsional resonance complaints, with radial (bending) resonance complaints occurring when the driveline is subjected to radial excitation near the radial natural frequency, and torsional resonance complaints occurring when the driveline is subjected to torsional excitation near the torsional natural frequency.
It should be understood that, for the sake of convenience of distinction, the torsional natural frequency in the following description refers to the inherent property of the torsional vibration of the component or the transmission system, and when the torsional frequency reaches the torsional natural frequency, torsional resonance occurs and the transmission system is unstable; the radial natural frequency, the natural property of the radial vibration of the component or the drive train, up to which the radial resonance occurs and the drive train destabilizes.
More specifically, the structure of the composite vibration damper in the present embodiment is the same as that of the composite vibration damper provided in embodiment 2, and as shown in fig. 6, includes an outer ring 1, a mounting frame 3, and a rubber 2 provided between the outer ring 1 and the mounting frame 3. The drive train, which refers to the location or vicinity where resonance occurs, may be mounted on or at the end of the drive shaft. The compound vibration damper is arranged on a transmission shaft of the transmission system.
The vibration parameters are changed by changing the conditions of the outer ring and the rubber affecting the vibration damping effect, and the conditions can be the mass, the radius, the rubber material and the like of the outer ring. For example, modeling in three-dimensional software, adjusting the size and measuring the mass and inertia in software are carried out according to the calculated outer ring size, inertia and mass, and the inertia and mass are adjusted to meet the required range.
More specifically, the specific conditions for changing the vibration parameters of the compound vibration absorber can be determined through multiple experiments or by establishing a vibration reduction model for analysis and the like. The method comprises the steps of determining the approximate range of vibration parameters by continuously changing conditions influencing the vibration damping effect of the composite vibration damper and combining the torsional resonance frequency and the radial resonance frequency of a transmission system, further debugging the torsional vibration parameters and the radial vibration parameters, finding out parameters when the torsional vibration and the radial vibration meet vibration damping requirements, obtaining composite vibration damping conditions such as materials or radius of the composite vibration damper according to the parameters as verification parameters, and obtaining the composite vibration damper capable of simultaneously inhibiting the radial vibration and the torsional vibration according to the composite vibration damping conditions.
For example, for radial vibration, radial frequency, radial damping, and mass are measured; and measuring the torsional frequency, the torsional damping and the moment of inertia of the torsional vibration. And after the torsional vibration parameter range and the radial vibration parameter range are obtained, drawing a digital model in three-dimensional software, measuring the rotational inertia of the digital model in the software, and adjusting the size of the outer ring to enable the rotational inertia to meet the requirement. Then, in the CAE software, the stiffness of the rubber in the torsional direction and the radial direction is adjusted by analyzing the rubber hardness and the like, the torsional frequency can be calculated from the stiffness in the torsional direction and the moment of inertia, and the radial frequency can be calculated from the stiffness in the radial direction and the mass of the outer ring.
It is to be understood that in this specification, in addition to the adjustment of the rubber material, the rigidity (including the radial rigidity and the torsional rigidity) may be adjusted by designing the structure of the outer ring, the mounting frame and the rubber with a space allowed by the arrangement, and the adjustment of the structure includes whether to open the hole, the size of the hole, the shape of the cross section of the rubber, and the like.
The torsional vibration parameters and the radial vibration parameters which influence the vibration effect can be obtained by finding out influence factors influencing and correlating according to modeling analysis or experiments. For example, the torsional vibration parameter may include a torsional natural frequency obtained from a torsional rigidity, an outer ring radius, and a rubber material, a moment of inertia of the outer ring, and a torsional damping of the rubber determined by the rubber material, and may also be a radius, a mass, a density, or a width of the outer ring. Also the radial vibration parameters may be the radial natural frequency, the mass of the outer ring and the radial damping of the rubber, etc.
The economic benefit lies in that the device is arranged on a transmission shaft, a radial vibration damper of the transmission shaft is cancelled, the weight is reduced by about 1 Kg/vehicle, and the cost is reduced by about 25 Yuan/vehicle; the rear main reducer Damper (RDM Damper) is removed, the weight is reduced by about 1.2 Kg/vehicle, and the cost is reduced by about 80 yuan/vehicle.
In a preferred embodiment, the method for debugging the range of the vibration parameter obtained in the steps a and b comprises the following steps:
s1, establishing a radial vibration model, and obtaining radial vibration parameters according to the radial vibration model;
s2, obtaining radial vibration amplitude according to the radial vibration parameters, adjusting the radial vibration parameters to obtain a plurality of corresponding radial vibration amplitudes, and establishing a radial vibration parameter-amplitude curve according to each radial vibration amplitude;
s3, establishing a torsional vibration model, and obtaining torsional vibration parameters according to the torsional vibration model;
s4, obtaining torsional vibration amplitude according to the torsional vibration parameters, adjusting the torsional vibration parameters to obtain a plurality of corresponding torsional vibration amplitudes, and establishing a torsional vibration parameter-amplitude curve according to each torsional vibration amplitude;
and s5, obtaining a radial vibration parameter range according to the radial vibration parameter-amplitude curve, and obtaining a torsional vibration parameter range according to the torsional vibration parameter-amplitude curve.
Specifically, the radial vibration model (system simplified mathematical model) is established in s1 as shown in fig. 2, according to the radial natural frequency calculation formula:
Figure BDA0002510667950000101
a part of a transmission system (such as a transmission shaft and a rear axle input shaft) is simplified into two parameters of radial (bending) rigidity and mass (damping is ignored by metal parts), the specific structural shape of the parts is ignored, and the composite shock absorber is also simplified into three parameters of radial rigidity, mass and rubber radial damping, and the specific structural shape of the composite shock absorber is ignored. Wherein, the transmission system 4 is a part which generates radial resonance; k1 is the radial stiffness of the part where radial resonance occurs; m1 is the equivalent mass of the transmission system, and k2, c2 and m2 in the shock absorber 5 respectively represent the radial rigidity, the radial damping and the outer ring mass of the composite shock absorber. In addition, x1(t) refers to the variation of the radial vibration displacement of the transmission system with time, x2(t) refers to the variation of the radial vibration displacement of the composite vibration damper with time, and only the vibration displacement of the transmission system is concerned in the model, and the vibration displacement of the composite vibration damper is not discussed.
The torsional vibration model (simplified mathematical model of the system) established in s3 is as shown in fig. 3, namely simplified into torsional rigidity and moment of inertia, and the corresponding torsional natural frequency formula is as follows regardless of the structural shape:
Figure BDA0002510667950000102
the action of the composite damper in the torsional direction is simplified into torsional rigidity, rotational inertia and rubber torsional damping. Wherein K1 is the torsional stiffness of the part in which torsional resonance occurs; j1 is the moment of inertia of the part where torsional resonance occurs; kd, C and Jd respectively represent torsional rigidity, torsional damping and rotational inertia of the composite damper.
More specifically, the effect of suppressing vibration is evaluated by amplitude, and debugging conditions and related conditions required by corresponding parameters can be rapidly and controllably determined and changed after a model is established, irrelevant influence factors are eliminated, and the experiment times are reduced.
For example, it can be analyzed by modeling that the most effective torsional vibration parameters can be the torsional natural frequency derived from the torsional stiffness, the outer ring radius and the rubber material, the moment of inertia of the outer ring and the torsional damping of the rubber determined by the rubber material, etc.; the radial vibration parameters may be the radial natural frequency, the mass of the outer ring and the radial damping of the rubber, etc.
And after determining each influence factor and each influence parameter, the functional relationship among the parameters can be confirmed, and the parameter range can be rapidly and accurately determined by establishing a vibration parameter-amplitude comparison table, so that the verification parameter which is in line with the inhibition of the double vibration can be rapidly found.
In the embodiment, after the radial vibration model is established, parameters such as the radial frequency of the vibration absorber, the outer ring quality, the rubber damping and the like can be determined and adjusted, and the radial vibration amplitude of the part of the transmission system can be reduced; and parameters such as the torsional frequency of the shock absorber, the rotational inertia of the outer ring, the rubber damping and the like are adjusted, so that the torsional vibration amplitude of the part of the transmission system is reduced.
Taking the established models shown in fig. 2 and 3 as an example, after determining the radial vibration parameter and the torsional vibration parameter, establishing a radial vibration parameter-amplitude curve shown in fig. 4-6 and a torsional vibration parameter-amplitude curve shown in fig. 7, specifically:
the radial vibration parameter-amplitude curves are radial amplitude-frequency response curves shown in fig. 4, fig. 5 and fig. 6, wherein fig. 4 is the amplitude-frequency response curve under the influence of the radial natural frequency ratio of different compound vibration dampers to the resonance frequency of the transmission system, f1, f2 and f3 are natural-resonance frequency ratios respectively representing the frequency ratio of the radial natural frequency of different compound vibration dampers to the radial resonance frequency of the transmission system, and f0Is in the original state; FIG. 5 is an amplitude-frequency response curve under the influence of different outer ring masses, wherein m, 2m and 3m respectively represent different outer ring masses; fig. 6 is an amplitude-frequency response curve under the influence of different damping factors, ξ a, ξ b, ξ c respectively represent the amplitude-frequency response curve under the influence of different damping factors, and ξ a ═ 2, ξ b ═ 0.2, ξ c ═ 0.1. The abscissas of FIGS. 4-6 are the excited radial vibration frequency of the drive train and the radial resonance frequency of the drive train(radial frequency of occurrence of complaints).
Fig. 7 is a torsional vibration parameter versus amplitude curve, and fig. 7 is an amplitude frequency response curve under the influence of different torsional natural frequencies. Similarly, the effect of the other parameters affecting the vibration of the torsional vibration parameters can be similar to the torsional vibration parameter-amplitude curve established with reference to fig. 5 or fig. 6, which is not illustrated in the present embodiment. The abscissa of FIG. 7 is the ratio of the excited torsional vibration frequency of the drive train to the torsional resonance frequency of the drive train (the torsional frequency at which complaints occur).
The ordinate is the amplitude amplification factor, and when the transmission system is subjected to a constant excitation force, the transmission system has a stable deformation amplitude, and when the excitation force changes according to a certain frequency, the deformation amplitude of the transmission system can be amplified, and the amplitude amplification factor refers to the ratio of the amplified amplitude to the static deformation amplitude. Wherein, the ordinate of fig. 4-6 is the radial vibration amplitude amplification factor of the composite vibration absorber, and the ordinate of fig. 7 is the torsional vibration amplitude amplification factor of the composite vibration absorber.
It should also be understood that the excited radial frequency refers to the external excitation force to which the drive train is subjected, not a constant value, but a value that varies according to a certain frequency, which is the excited radial frequency of the drive train. The excited torsional frequency works the same.
More specifically, the damping effect is measured in terms of amplitude, and according to fig. 4-7, the lower the various amplitudes, indicating that the vibrations are smaller, the goal is to reduce the amplitude of the original state, the lower the effect the better.
Taking fig. 4 as an example, fig. 4 shows the damping effect when tuning the radial natural frequency of the damper, and it can be seen that under the influence of different f0, f1, f2 and f3, a curve of the corresponding relationship between the natural-resonance frequency ratio and the amplitude amplification factor can be obtained. From the simple model established in fig. 2, it can be known that the amplitude and the radial natural frequency have a correlation, and a radial vibration parameter range conforming to the radial vibration damping effect can be found according to fig. 4. Similarly, the determination method of the torsional vibration parameter is similar.
In a preferred embodiment, the torsional vibration parameters include the torsional natural frequency of the composite damper, the moment of inertia of the outer race, and the torsional damping of the rubber of the composite damper; the radial vibration parameters comprise the radial natural frequency of the composite vibration absorber, the mass of the outer ring and the radial damping of rubber; the torsional and radial damping of the rubber is determined by the material of the rubber.
Specifically, through experiments or modeling analysis, the amplitude of the torsional vibration of the composite damper is related to the torsional natural frequency, the rotational inertia of the outer ring and the torsional damping of the rubber of the composite damper, and the effect of suppressing the torsional vibration can be changed by changing the above parameters. Meanwhile, the amplitude and the radial natural frequency of the radial vibration, the mass of the outer ring and the radial damping of the rubber can change the effect of inhibiting the radial vibration by changing the parameters. Therefore, the vibration reduction effect can be debugged according to the correlation of the parameters.
Further, determining the radial natural frequency according to radial influence factors, wherein the radial influence factors comprise the material of the rubber and the radial rigidity of the rubber; the torsional natural frequency is determined from torsional influencing factors, including the material of the rubber and the torsional stiffness of the rubber.
Experiments show that the radial natural frequency is related to the material of the rubber and the radial rigidity of the rubber, and the radial natural frequency can be changed by changing the material of the rubber and the radial rigidity of the rubber; meanwhile, the torsional natural frequency is related to the material of the rubber and the torsional rigidity of the rubber, and the torsional natural frequency can be changed by changing the material of the rubber and the torsional rigidity of the rubber. Therefore, the natural frequency can be further changed by changing the material of the rubber, the radial rigidity of the rubber and the torsional rigidity of the rubber, and the vibration damping effect can be further adjusted.
Further, both the radial and torsional influencing factors also include the radius of the outer race of the composite damper, including the inner and outer diameters.
By adopting the scheme, experiments prove that the radial natural frequency and the torsional natural frequency of the composite damper can be simultaneously influenced by the radius of the outer ring of the composite damper. Therefore, the natural frequency can be further changed by changing the radius of the outer ring, and the vibration damping effect can be further debugged.
Further, the radial influencing factor further comprises an additional module, wherein the additional module comprises a pin, a metal insert or additional rubber with hardness different from that of the rubber of the composite damper, and the pin, the metal insert or the additional rubber is arranged on the composite damper.
Specifically, it has been experimentally found that the radial natural frequency can be changed by providing an additional module such as a pin or a metal insert in the composite damper. On the basis that the torsional natural frequency meets the vibration reduction effect, the radial natural frequency is adjusted by adding some auxiliary methods, and the better vibration reduction effect is further achieved.
More specifically, a pin or metal insert may be installed to increase its radial mode to change the radial natural frequency. When the radial mode needs to be reduced, the pin or the metal insert is cancelled, the radial mode is reduced, or rubber with hardness different from that of the rubber of the shock absorber is adopted to replace the pin, so that the radial mode is adjusted. The radial mode refers to a mode frequency, namely a mode frequency, and is a natural frequency calculated by CAE software modal analysis.
Further, the radial vibration amplitude and the radial vibration parameter satisfy the relation:
Figure BDA0002510667950000131
wherein, A1 is the radial vibration amplitude amplification coefficient of the composite vibration damper;
μ1the mass ratio of the outer ring of the compound damper to the mass of the transmission system;
f1-the ratio of the radial natural frequency of the compound damper to the radial resonant frequency of the transmission system;
g1-the ratio of the radial excitation frequency of the transmission system to the radial resonance frequency of the transmission system;
ξ1-the radial damping ratio of the rubber; and the number of the first and second electrodes,
the torsional vibration amplitude and the torsional vibration parameter satisfy the relation:
Figure BDA0002510667950000141
wherein A is2-the torsional vibration amplitude amplification factor of the composite damper;
λ2-the ratio of the torsional excitation frequency of the transmission system to the torsional resonance frequency of the transmission system;
α2the torsional natural frequency of the composite damper is compared to the torsional resonance frequency of the transmission system;
μ2the ratio of the rotational inertia of the compound damper to the rotational inertia of the transmission system;
ξ2-torsional damping ratio of the rubber.
Specifically, the above formula can be obtained through experiments and modeling analysis. The relation between the parameters such as the radial frequency, the outer ring mass, the rubber damping and the like of the shock absorber and the radial vibration amplitude of the part of the transmission system, and the relation between the parameters such as the torsional frequency, the outer ring rotational inertia, the rubber damping and the like of the shock absorber and the torsional vibration amplitude of the part of the transmission system are quantitatively embodied in the formula.
More specifically, the torsional natural frequency is an inherent property of the composite damper, the torsional resonance frequency of the driveline is the NVH frequency complained of due to resonance, and the radial natural frequency and the radial resonance frequency are the same. The damping ratio refers to the ratio of actual damping to critical damping, and both the actual damping and the critical damping refer to the actual damping and the critical damping of the rubber of the composite shock absorber. The rubber damping comprises radial damping and torsional damping, and is measured through experiments. Further, the moment of inertia of the composite damper refers to the moment of inertia of the outer race. Wherein, the radial natural frequency is determined by rubber damping and rubber radial rigidity; the torsional natural frequency is determined by rubber damping and rubber torsional rigidity; the rubber damping is determined by the rubber material, which is determined by the rubber formulation.
More specifically, as previously mentioned, the amplitude magnification factor refers to the ratio of the amplified amplitude to the amplitude of the static deformation. Taking the model established in fig. 2 as an example, x1(t) in fig. 2 indicates the variation of radial vibration displacement of the transmission system with time, and a1 is the maximum value of x1 (t).
It is to be understood that, in the present embodiment, the above formula is obtained based on the models established in fig. 2 and 3, wherein g1 and a1 in the formula correspond to the abscissa (frequency ratio) and the ordinate (amplitude magnification factor) in fig. 4 to 6, respectively, and λ 2 and a2 correspond to the abscissa (frequency ratio) and the ordinate (amplitude magnification factor) in fig. 7, respectively.
Further, in a preferred embodiment, the torsional natural frequency is determined according to the following relationship:
Figure BDA0002510667950000151
wherein f isTorsion-torsional natural frequency of the composite damper;
Ktorsion-the torsional stiffness of the outer ring about the axial direction of the composite damper;
rho is the material density of the outer ring of the composite shock absorber;
h is the width of the outer ring of the composite damper;
r1-an inner diameter of an outer race of the compound damper;
r2-an outer diameter of an outer race of the composite damper;
the radial natural frequency is determined according to the following relation:
Figure BDA0002510667950000152
wherein, KRadial directionThe deformation rigidity of the outer ring of the composite damper in the radial direction;
fradial directionRadial natural frequency of the composite damper.
The formula can be obtained through modeling analysis and experiments.
More specifically, from the above formula, it can be understood that the torsional rigidity K of the damper rubber is adjustedTorsionAnd radial stiffness KRadial directionThe torsional and radial frequencies, the damper radii r1 and r2, and the torsional frequency can be variedFrequency and radial frequency. Also, the proportional relationship between the torsional natural frequency and the radial natural frequency can be determined according to the above formula. Namely, it is
Figure BDA0002510667950000153
The vibration damping effect can be conveniently calculated according to the proportional relation.
More specifically, KTorsion、KRadial directionAnd fTorsion、fRadial directionThe vector direction of (2) is as shown in fig. 8 in embodiment 2.
In a preferred embodiment, the method for obtaining the composite damping condition according to the verification parameter in step d comprises the following steps:
d-1, obtaining the required radius range of the rotational inertia of the outer ring according to the mass range of the outer ring;
d-2, determining a rubber material according to the torsional damping and the radial damping;
d-3, determining the structure of the rubber of the composite damper according to the radial natural frequency and the torsional natural frequency;
d-4, adding an additional module according to the radial natural frequency on the basis of meeting the torsional natural frequency.
Specifically, since the material of the outer ring is determined, for example, iron, the density is determined, and a desired radius range of the moment of inertia of the outer ring can be obtained according to the range of the mass of the outer ring. The rubber structure in d-3 means a cross-sectional structure, whether or not holes are formed in the entire circumference, and the like, in order to adjust the rigidity of the rubber. After the parameters and the verification parameters which can simultaneously play two vibration damping effects are found through debugging, the verification parameters need to be further converted into the verification parameters which are represented by the verification parameters and accord with the vibration damper materials, modes, additional parts and the like, namely, the verification parameters are converted into composite vibration damping conditions. The radius of the composite shock absorber can be obtained through the mass and the rotational inertia of the outer ring; determining a rubber material of the composite shock absorber according to the damping, and further determining a rubber formula according to the rubber material; determining the radial stiffness and the torsional stiffness of the rubber of the composite damper, for example, the rubber, based on the radial natural frequency and the torsional natural frequency; and the required accessory modules such as an insert plate, a pin and the like can be determined according to the radial natural frequency.
Example 2
A composite damper is shown in fig. 8 and comprises a mounting framework 3, rubber 2 and an outer ring 1, wherein the rubber 2 is arranged on the mounting framework 3, the outer ring 1 is sleeved on the rubber 2, and the composite damper is designed according to the design method of the composite damper in embodiment 1.
While the invention has been shown and described with reference to certain preferred embodiments thereof, it will be understood by those skilled in the art that the foregoing is a more detailed description of the invention, taken in conjunction with the specific embodiments thereof, and that no limitation of the invention is intended thereby. Various changes in form and detail, including simple deductions or substitutions, may be made by those skilled in the art without departing from the spirit and scope of the invention.

Claims (10)

1. A method of designing a composite shock absorber comprising the steps of:
a. debugging the torsional vibration parameters of the composite vibration damper according to the torsional resonance frequency of the transmission system when NVH complaints exist, and obtaining the torsional vibration parameter range of the composite vibration damper;
b. debugging the radial vibration parameter of the composite vibration damper according to the radial resonance frequency of the transmission system when NVH complaints exist, and obtaining the radial vibration parameter range of the composite vibration damper;
c. according to the torsional vibration parameter range and the radial vibration parameter range, obtaining verification parameters of the composite vibration absorber when both the torsional vibration and the radial vibration meet the vibration reduction requirements;
d. and obtaining a composite vibration reduction condition according to the verification parameters, and trial-producing a sample piece according to the composite vibration reduction condition for loading verification.
2. The method for designing a composite vibration absorber according to claim 1, wherein the tuning method for obtaining the torsional vibration parameter range and the radial vibration parameter range in the steps a and b comprises the steps of:
s1, establishing a radial vibration model, and obtaining the radial vibration parameters according to the radial vibration model;
s2, obtaining radial vibration amplitude according to the radial vibration parameters, adjusting the radial vibration parameters to obtain a plurality of corresponding radial vibration amplitudes, and establishing a radial vibration parameter-amplitude curve according to each radial vibration amplitude;
s3, establishing a torsional vibration model, and obtaining the torsional vibration parameters according to the torsional vibration model;
s4, obtaining torsional vibration amplitude according to the torsional vibration parameters, adjusting the torsional vibration parameters to obtain a plurality of corresponding torsional vibration amplitudes, and establishing a torsional vibration parameter-amplitude curve according to each torsional vibration amplitude;
s5, obtaining the radial vibration parameter range according to the radial vibration parameter-amplitude curve, and obtaining the torsional vibration parameter range according to the torsional vibration parameter-amplitude curve.
3. The method of designing a composite vibration absorber according to claim 2, wherein the torsional vibration parameters include a torsional natural frequency of the composite vibration absorber, a rotational inertia of an outer race, and a torsional damping of rubber of the composite vibration absorber;
the radial vibration parameters comprise the radial natural frequency of the composite vibration damper, the mass of the outer ring and the radial damping of the rubber;
the torsional damping and the radial damping of the rubber are determined depending on the material of the rubber.
4. The method of designing a compound vibration damper according to claim 3, wherein the radial natural frequency is determined based on a radial influence factor, the radial influence factor including a material of the rubber and a radial stiffness of the rubber;
the torsional natural frequency is determined from torsional influencing factors including the material of the rubber and the torsional stiffness of the rubber.
5. The method of designing a composite shock absorber according to claim 4, wherein the radial influencing factor and the torsional influencing factor each further comprise a radius of an outer race of the composite shock absorber, the radius comprising an inner diameter and an outer diameter.
6. The method of designing a composite vibration damper according to claim 5, wherein said radial influencing factors further comprise an additional module comprising a pin, a metal insert, or an additional rubber of a different hardness than the rubber of said composite vibration damper disposed on said composite vibration damper.
7. The method of designing a compound vibration damper according to claim 6, wherein the radial vibration amplitude and the radial vibration parameter satisfy the relation:
Figure FDA0002510667940000021
wherein, A1 is the radial vibration amplitude amplification coefficient of the composite vibration damper;
μ1the mass ratio of the outer ring of the compound damper to the mass of the transmission system;
f1-the ratio of the radial natural frequency of the compound damper to the radial resonant frequency of the transmission system;
g1-the ratio of the radial excitation frequency of the transmission system to the radial resonance frequency of the transmission system;
ξ1-the radial damping ratio of the rubber; and the number of the first and second electrodes,
the torsional vibration amplitude and the torsional vibration parameter satisfy the relation:
Figure FDA0002510667940000022
wherein A is2-the torsional vibration amplitude amplification factor of the composite damper;
λ2-the ratio of the torsional excitation frequency of the transmission system to the torsional resonance frequency of the transmission system;
α2the torsional natural frequency of the composite damper is compared to the torsional resonance frequency of the transmission system;
μ2the ratio of the rotational inertia of the compound damper to the rotational inertia of the transmission system;
ξ2-torsional damping ratio of the rubber.
8. The method of designing a composite vibration damper according to claim 7, wherein the torsional natural frequency is determined according to the following relation:
Figure FDA0002510667940000031
wherein f isTorsion-torsional natural frequency of the composite damper;
Ktorsion-the torsional stiffness of the outer ring about the axial direction of the composite damper;
rho is the material density of the outer ring of the composite shock absorber;
h is the width of the outer ring of the composite damper;
r1-an inner diameter of an outer race of the compound damper;
r2-an outer diameter of an outer race of the composite damper;
the radial natural frequency is determined according to the following relation:
Figure FDA0002510667940000032
wherein, KRadial directionThe deformation rigidity of the outer ring of the composite damper in the radial direction;
fradial directionRadial natural frequency of the composite damper.
9. The method of designing a compound vibration damper according to claim 8, wherein the method of obtaining the compound vibration damping condition based on the verification parameter in step d includes:
d-1, obtaining the radius range of the required rotational inertia of the outer ring according to the mass range of the outer ring;
d-2, determining a rubber material according to the torsional damping and the radial damping;
d-3, determining the structure of the rubber of the composite damper according to the radial natural frequency and the torsional natural frequency;
d-4, adding the additional module according to the radial natural frequency on the basis of meeting the torsional natural frequency.
10. A composite vibration absorber comprising a mounting frame, rubber and an outer ring, wherein the rubber is arranged on the mounting frame, and the outer ring is sleeved on the rubber, characterized in that the composite vibration absorber is designed according to the design method of the composite vibration absorber as claimed in any one of claims 1 to 9.
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