CN105069261B - Low speed rail vehicle two is the design method of lateral damper optimum damping coefficient - Google Patents
Low speed rail vehicle two is the design method of lateral damper optimum damping coefficient Download PDFInfo
- Publication number
- CN105069261B CN105069261B CN201510559657.6A CN201510559657A CN105069261B CN 105069261 B CN105069261 B CN 105069261B CN 201510559657 A CN201510559657 A CN 201510559657A CN 105069261 B CN105069261 B CN 105069261B
- Authority
- CN
- China
- Prior art keywords
- msub
- mrow
- mtd
- bogie
- low speed
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired - Fee Related
Links
Landscapes
- Vehicle Body Suspensions (AREA)
Abstract
The present invention relates to the design method that low speed rail vehicle two is lateral damper optimum damping coefficient, belongs to low speed rail suspension of vehicle technical field.The present invention travels the oscillation crosswise differential equation by establishing the free degree of low speed rail whole vehicle 17, utilize MATLAB/Simulink simulation softwares, construct the free degree oscillation crosswise optimization design simulation model of low speed rail whole vehicle 17, and using orbital direction irregularity and horizontal irregularity as input stimulus, the minimum design object of vibration root mean square of weighed acceleration moved with cross-car, optimization design obtain the optimum damping coefficient that low speed rail vehicle two is lateral damper.By designing example and SIMPACK simulating, verifyings, available this method two is accurately and reliably lateral damper optimum damping coefficient value, and reliable design method is provided for the design that low speed rail vehicle two is lateral damper optimum damping coefficient.Using this method, the design level and vehicle safety and stationarity of low speed rail automobile suspension system are remarkably improved.
Description
Technical field
It is lateral damper optimal damper system the present invention relates to low speed rail suspension of vehicle, particularly low speed rail vehicle two
Several design methods.
Background technology
Two be that lateral damper has important influence to the riding comfort and security of low speed rail vehicle.However,
Understood according to institute's inspection information, because low speed rail vehicle belongs to Mdof Vibration System, dynamic analysis calculating is carried out to it
It is extremely difficult, it is both at home and abroad at present the design of lateral damper damped coefficient for two, the theory for never providing system is set
Meter method, mostly it is by computer technology, using Dynamics Simulation soft sim PACK or ADAMS/Rail, passes through reality
Volume modeling optimizes and determines its size, although this method can obtain reliable simulation numerical, has vehicle preferable
Power performance, however, with the continuous development of rail vehicle industry, people are the design of lateral damper damped coefficient to two
Higher requirement is proposed, current two be that the method that lateral damper damped coefficient designs can not provide the wound with directive significance
New theory, it is impossible to meet the development to absorber designing requirement in the case of rail vehicle fast development.Therefore, it is necessary to establish one kind
Accurately, reliable low speed rail vehicle two is the design method of lateral damper optimum damping coefficient, meets that rail vehicle is quick
To the requirement of absorber designing under development, the design level and product quality of low speed rail automobile suspension system are improved, is carried
High vehicle riding comfort and security；Meanwhile product design and testing expenses are reduced, shorten the product design cycle, strengthen me
The competitiveness in the international market of state's rail vehicle.
The content of the invention
For defect present in above-mentioned prior art, the technical problems to be solved by the invention be to provide it is a kind of accurate,
Reliable low speed rail vehicle two is the design method of lateral damper optimum damping coefficient, and its design flow diagram is as shown in Figure 1；
The free degree of low speed rail whole vehicle 17 travels left view such as Fig. 2 of oscillation crosswise model, and low speed rail whole vehicle 17 is freely
The top view of degree traveling oscillation crosswise model is as shown in Figure 3.
In order to solve the above technical problems, low speed rail vehicle two provided by the present invention is lateral damper optimal damper system
Several design method, it is characterised in that use following design procedure：
(1) free degree of the low speed rail whole vehicle 17 traveling oscillation crosswise differential equation is established：
According to the quality m of the single-unit car body of rail vehicle3, rotary inertia of shaking the headSidewinder rotary inertia J3θ；Every turns
To the quality m of framework frame2, rotary inertia of shaking the headSidewinder rotary inertia J2θ；The quality m of each round pair1, rotary inertia of shaking the headEach wheel shaft weight W；The horizontal creep coefficient f of each round pair1, longitudinal creep coefficient f2；Longitudinal direction per axle-box positioning device is just
Spend K1x, lateral stiffness K1y, vertical stiffness K1z；The longitudinal rigidity K of every bogie central spring2x, located lateral stiffness K2y；Often
The vertical equivalent stiffness K of the system of platform bogie two suspension2z, vertical equivalent damping Cd2；The torsional rigidity K of single anti-side rolling torsion rodθ；Often
Platform bogie to be designed two is the Equivalent damping coefficient C of lateral damper2；Vehicle wheel roll radius r, wheel tread gradient λ；Vehicle
Travel speed v；The half b of wheel and rail contact point horizontal spacing, wheel shaft retaining spring are transversely mounted the half b of spacing1, turn
The half b of spacing is transversely mounted to frame central spring2, the half a of length between truck centers, the half a of wheel-base bogie0, axle centerline
To the height h of orbit plane0, the height h of plane on car body barycenter to central spring1, car body barycenter to two is lateral damper
Height h2, height h of the plane to framework barycenter on central spring3, the height h of bogie frame barycenter to axle centerline4, two
It is height h of the lateral damper to framework barycenter5；The barycenter O of steering framing wheel pair before respectively1ff、O1fr, trailing bogie wheel pair
Barycenter O1rf、O1rr, the barycenter O of forward and backward bogie frame2f、O2rAnd the barycenter O of car body3For the origin of coordinates；With forecarriage front-wheel
To yaw displacement y1ff, displacement of shaking the headThe yaw displacement y of forecarriage trailing wheel pair1fr, displacement of shaking the headBefore trailing bogie
The yaw displacement y of wheel pair1rf, displacement of shaking the headThe yaw displacement y of trailing bogie trailing wheel pair1rr, displacement of shaking the headForecarriage
The yaw displacement y of framework2f, displacement of shaking the headSidewinder displacement θ2f, the yaw displacement y of trailing bogie framework2r, displacement of shaking the head
Sidewinder displacement θ2r, and the yaw displacement y of car body3, displacement of shaking the headSidewinder displacement θ3For coordinate；With the forward and backward wheel of forecarriage
And the orbital direction irregularity input y at the forward and backward wheel of trailing bogiea1(t)、ya2(t)、ya3(t)、ya4And horizontal irregularity (t)
Input zθ1(t)、zθ2(t)、zθ3(t)、zθ4(t) it is input stimulus, wherein, t is time variable；Establish low speed rail whole vehicle
17 frees degree travel the oscillation crosswise differential equation, i.e.,：
1. the yaw vibration equation of forecarriage front-wheel pair：
2. the yawing equation of forecarriage front-wheel pair：
3. the yaw vibration equation of forecarriage trailing wheel pair：
4. the yawing equation of forecarriage trailing wheel pair：
5. the yaw vibration equation of trailing bogie front-wheel pair：
6. the yawing equation of trailing bogie front-wheel pair：
7. the yaw vibration equation of trailing bogie trailing wheel pair：
8. the yawing equation of trailing bogie trailing wheel pair：
9. the yaw vibration equation of forecarriage framework：
10. the rolling equation of forecarriage framework：
The yawing equation of forecarriage framework：
The yaw vibration equation of trailing bogie framework：
The rolling equation of trailing bogie framework：
The yawing equation of trailing bogie framework：
The yaw vibration equation of car body：
The rolling equation of car body：
Wherein, h=h0+h1+h3+h4；
The yawing equation of car body：
(2) free degree oscillation crosswise optimization design simulation model of low speed rail whole vehicle 17 is built：
The oscillation crosswise differential equation, profit are travelled according to the free degree of low speed rail whole vehicle 17 established in step (1)
With Matlab/Simulink simulation softwares, the free degree oscillation crosswise optimization design emulation mould of structure low speed rail whole vehicle 17
Type；
(3) the two damping optimization design object function J for being lateral damper are established：
Mould is emulated according to the free degree oscillation crosswise optimization design of low speed rail whole vehicle 17 established in step (2)
Type, using every bogie two be lateral damper Equivalent damping coefficient as design variable, with each wheel to the orbital direction at place not
It is input stimulus to smooth out stochastic inputs and horizontal irregularity stochastic inputs, utilizes the vibration of the car body weaving obtained by emulation
Frequency weighted acceleration root-mean-square valueCar body sidewinders the vibration frequency root mean square of weighed acceleration of motionAnd car body shakes
Cephalomotor vibration frequency root mean square of weighed accelerationEstablish the two damping optimization design object letters for being lateral damper
Number J, i.e.,：
In formula, vibration frequency root mean square of weighed accelerationCoefficient 1,0.63,0.2, be respectively
Car body weaving, sidewinder motion, the axle weight coefficient of yaw motion；Wherein, vibration frequency weighting accelerates at different frequencies
Spend root-mean-square valueFrequency weight values be respectively wd(fi)、we(fi)、wf(fi), i.e.,：
(4) low speed rail vehicle two is lateral damper optimum damping coefficient C optimization design：
1. according to the half a of length between truck centers, the half a of wheel-base bogie0, institute in Vehicle Speed v, and step (2)
The free degree oscillation crosswise optimization design simulation model of low speed rail whole vehicle 17 of foundation, with each wheel to the orbital direction at place not
Smooth out stochastic inputs ya1(t)、ya2(t)、ya3(t)、ya4And horizontal irregularity stochastic inputs z (t)θ1(t)、zθ2(t)、zθ3(t)、
zθ4(t) it is input stimulus, is asked using optimized algorithm in step (3) and establish the two damping optimization design objects for being lateral damper
Function J minimum value, corresponding design variable are the optimal Equivalent damping coefficient that every bogie two is lateral damper
C2；
Wherein, the relation between orbital direction irregularity stochastic inputs is： Relation between horizontal irregularity stochastic inputs is：
2. it is the installation number n of lateral damper according to every bogie two, and 1. optimization order design institute in step (4)
Obtained every bogie two is the optimal Equivalent damping coefficient C of lateral damper2, it is lateral damper that single branch two, which is calculated,
Optimum damping coefficient C, i.e.,：C=C2/n。
The present invention has the advantage that than prior art：
Because low speed rail vehicle belongs to Mdof Vibration System, it is carried out dynamic analysis calculate it is extremely difficult,
It is both at home and abroad at present the design of lateral damper damped coefficient for two, never provides the theoretical design method of system, greatly
All it is by computer technology, using Dynamics Simulation soft sim PACK or ADAMS/Rail, by solid modelling come excellent
Change and determine its size, although this method can obtain reliable simulation numerical, make vehicle that there is preferable power performance,
However, with the continuous development of rail vehicle industry, it is higher that people are that the design of lateral damper damped coefficient proposes to two
Requirement, current two be lateral damper damped coefficient design method can not provide the innovation theory with directive significance, no
The development to absorber designing requirement in the case of rail vehicle fast development can be met.
The present invention travels the oscillation crosswise differential equation by establishing the free degree of low speed rail whole vehicle 17, utilizes
MATLAB/Simulink simulation softwares, construct the free degree oscillation crosswise optimization design of low speed rail whole vehicle 17 emulation mould
Type, and using orbital direction irregularity and horizontal irregularity as input stimulus, it is equal with the vibration weighted acceleration that cross-car is moved
The minimum design object of root value, optimization design obtain the optimum damping coefficient that low speed rail vehicle two is lateral damper.It is logical
Design example and SIMPACK simulating, verifyings is crossed to understand, this method it is available accurately and reliably two be lateral damper damping system
Numerical value, reliable design method is provided for the design that low speed rail vehicle two is lateral damper damped coefficient.Utilize the party
Method, the design level and product quality of low speed rail automobile suspension system can be not only improved, improve vehicle safety peace
Stability；Meanwhile product design and testing expenses can be also reduced, shorten the product design cycle, strengthen the world of China's rail vehicle
The market competitiveness.
Brief description of the drawings
It is described further below in conjunction with the accompanying drawings for a better understanding of the present invention.
Fig. 1 is the design flow diagram that low speed rail vehicle two is lateral damper optimum damping coefficient design method；
Fig. 2 is the left view of the free degree of low speed rail whole vehicle 17 traveling oscillation crosswise model；
Fig. 3 is the top view of the free degree of low speed rail whole vehicle 17 traveling oscillation crosswise model；
Fig. 4 is the free degree oscillation crosswise optimization design simulation model of low speed rail whole vehicle 17 of embodiment；
Fig. 5 is the U.S. orbital direction irregularity random input stimuli y that embodiment is applieda1(t)；
Fig. 6 is the U.S. orbital direction irregularity random input stimuli y that embodiment is applieda2(t)；
Fig. 7 is the U.S. orbital direction irregularity random input stimuli y that embodiment is applieda3(t)；
Fig. 8 is the U.S. orbital direction irregularity random input stimuli y that embodiment is applieda4(t)；
Fig. 9 is the horizontal irregularity random input stimuli z of U.S.'s track that embodiment is appliedθ1(t)；
Figure 10 is the horizontal irregularity random input stimuli z of U.S.'s track that embodiment is appliedθ2(t)；
Figure 11 is the horizontal irregularity random input stimuli z of U.S.'s track that embodiment is appliedθ3(t)；
Figure 12 is the horizontal irregularity random input stimuli z of U.S.'s track that embodiment is appliedθ4(t)。
Specific embodiment
The present invention is described in further detail below by an embodiment.
It is lateral damper to be provided with two two on every bogie of certain low speed rail vehicle, i.e. n=2；Its single-unit car
The quality m of body3=56910kg, rotary inertia of shaking the headSidewinder rotary inertia J3θ=159300kg.m2；
The quality m of every bogie frame2=2310kg, rotary inertia of shaking the headSidewinder rotary inertia J2θ=
2080kg.m2；The quality m of each round pair1=2080kg, rotary inertia of shaking the headEach wheel shaft weight W=
160000N；The horizontal creep coefficient f of each round pair1=17250000N, longitudinal creep coefficient f2=17250000N；Determine per axle box
The longitudinal rigidity K of position device1x=17 × 106N/m, lateral stiffness K1y=1.48 × 106N/m, vertical stiffness K1z=1.48 ×
106N/m；The longitudinal rigidity K of every bogie central spring2x=0.165 × 106N/m, located lateral stiffness K2y=0.165 ×
106N/m；The vertical equivalent stiffness K of the every system of bogie two suspension2z=561.68kN/m, vertical equivalent damping Cd2=
111.39kN.s/m；The torsional rigidity K of single anti-side rolling torsion rodθ=2.5 × 106N.m/rad；Vehicle wheel roll radius r=0.43m,
Wheel tread gradient λ=0.15；The half b=0.7175m of wheel and rail contact point horizontal spacing, wheel shaft retaining spring are horizontal
Clipping room away from half b1=1.05m, bogie central spring are transversely mounted the half b of spacing2=1.2m, the one of length between truck centers
Half a=7.85m, the half a of wheel-base bogie0=1.25m, the height h of axle centerline to orbit plane0=0.43m, car body
Height h of the barycenter to plane on central spring1=0.779m, car body barycenter to two are the height h of lateral damper2=0.616m,
Height h of the plane to framework barycenter on central spring3=0.216m, the height h of bogie frame barycenter to axle centerline4=
0.075m, two be height h of the lateral damper to framework barycenter5=0.379m；Every bogie to be designed two is lateral vibration absorbing
The Equivalent damping coefficient of device is C2.The low speed rail vehicle two is the required vehicle traveling of lateral damper damped coefficient design
Speed v=100km/h, the optimum damping coefficient that the low speed rail vehicle two is lateral damper is designed.
The design method for the system's lateral damper optimum damping coefficient of low speed rail vehicle two that present example is provided, its
Design flow diagram is low as shown in figure 1, left view such as Fig. 2 of the free degree of low speed rail whole vehicle 17 traveling oscillation crosswise model
The top view of the fast free degree of rail vehicle whole 17 traveling oscillation crosswise model is as shown in figure 3, comprise the following steps that：
(1) free degree of the low speed rail whole vehicle 17 traveling oscillation crosswise differential equation is established：
According to the quality m of the single-unit car body of rail vehicle3=56910kg, rotary inertia of shaking the head
Sidewinder rotary inertia J3θ=159300kg.m2；The quality m of every bogie frame2=2310kg, rotary inertia of shaking the headSidewinder rotary inertia J2θ=2080kg.m2；The quality m of each round pair1=2080kg, rotary inertia of shaking the headEach wheel shaft weight W=160000N；The horizontal creep coefficient f of each round pair1=17250000N, longitudinal direction are compacted
Sliding coefficient f2=17250000N；Longitudinal rigidity K per axle-box positioning device1x=17 × 106N/m, lateral stiffness K1y=1.48 ×
106N/m, vertical stiffness K1z=1.48 × 106N/m；The longitudinal rigidity K of every bogie central spring2x=0.165 × 106N/
M, located lateral stiffness K2y=0.165 × 106N/m；The vertical equivalent stiffness K of the every system of bogie two suspension2z=561.68kN/
M, vertical equivalent damping Cd2=111.39kN.s/m；The torsional rigidity K of single anti-side rolling torsion rodθ=2.5 × 106N.m/rad；Often
Platform bogie to be designed two is the Equivalent damping coefficient C of lateral damper2；Vehicle wheel roll radius r=0.43m, wheel tread are oblique
Spend λ=0.15；Vehicle Speed v=100km/h；The half b=0.7175m of wheel and rail contact point horizontal spacing, wheel
Axle retaining spring is transversely mounted the half b of spacing1=1.05m, bogie central spring are transversely mounted the half b of spacing2=
1.2m, the half a=7.85m of length between truck centers, the half a of wheel-base bogie0=1.25m, axle centerline to orbit plane
Height h0=0.43m, the height h of plane on car body barycenter to central spring1=0.779m, car body barycenter to two are lateral vibration absorbing
The height h of device2=0.616m, height h of the plane to framework barycenter on central spring3=0.216m, bogie frame barycenter to car
The height h of shaft centre line4=0.075m, two be height h of the lateral damper to framework barycenter5=0.379m；Respectively with front steering
The barycenter O of frame wheel pair1ff、O1fr, the barycenter O of trailing bogie wheel pair1rf、O1rr, the barycenter O of forward and backward bogie frame2f、O2rAnd car
The barycenter O of body3For the origin of coordinates；With the yaw displacement y of forecarriage front-wheel pair1ff, displacement of shaking the headForecarriage trailing wheel pair
Yaw displacement y1fr, displacement of shaking the headThe yaw displacement y of trailing bogie front-wheel pair1rf, displacement of shaking the headTrailing bogie trailing wheel
To yaw displacement y1rr, displacement of shaking the headThe yaw displacement y of forecarriage framework2f, displacement of shaking the headSidewinder displacement θ2f, after
The yaw displacement y of bogie frame2r, displacement of shaking the headSidewinder displacement θ2r, and the yaw displacement y of car body3, displacement of shaking the headSide
Roll displacement θ3For coordinate；Y is inputted with the orbital direction irregularity at the forward and backward wheel of forecarriage and the forward and backward wheel of trailing bogiea1
(t)、ya2(t)、ya3(t)、ya4(t) and horizontal irregularity inputs zθ1(t)、zθ2(t)、zθ3(t)、zθ4(t) it is input stimulus, its
In, t is time variable；The free degree of the low speed rail whole vehicle 17 traveling oscillation crosswise differential equation is established, i.e.,：
1. the yaw vibration equation of forecarriage front-wheel pair：
2. the yawing equation of forecarriage front-wheel pair：
3. the yaw vibration equation of forecarriage trailing wheel pair：
4. the yawing equation of forecarriage trailing wheel pair：
5. the yaw vibration equation of trailing bogie front-wheel pair：
6. the yawing equation of trailing bogie front-wheel pair：
7. the yaw vibration equation of trailing bogie trailing wheel pair：
8. the yawing equation of trailing bogie trailing wheel pair：
9. the yaw vibration equation of forecarriage framework：
10. the rolling equation of forecarriage framework：
The yawing equation of forecarriage framework：
The yaw vibration equation of trailing bogie framework：
The rolling equation of trailing bogie framework：
The yawing equation of trailing bogie framework：
The yaw vibration equation of car body：
The rolling equation of car body：
Wherein, h=h0+h1+h3+h4；
The yawing equation of car body：
(2) free degree oscillation crosswise optimization design simulation model of low speed rail whole vehicle 17 is built：
The oscillation crosswise differential equation, profit are travelled according to the free degree of low speed rail whole vehicle 17 established in step (1)
With Matlab/Simulink simulation softwares, the free degree oscillation crosswise optimization design emulation mould of structure low speed rail whole vehicle 17
Type, as shown in Figure 4；
(3) the two damping optimization design object function J for being lateral damper are established：
Mould is emulated according to the free degree oscillation crosswise optimization design of low speed rail whole vehicle 17 established in step (2)
Type, using every bogie two be lateral damper Equivalent damping coefficient as design variable, with each wheel to the orbital direction at place not
It is input stimulus to smooth out stochastic inputs and horizontal irregularity stochastic inputs, utilizes the vibration of the car body weaving obtained by emulation
Frequency weighted acceleration root-mean-square valueCar body sidewinders the vibration frequency root mean square of weighed acceleration of motionAnd car body shakes
Cephalomotor vibration frequency root mean square of weighed accelerationEstablish the two damping optimization design object letters for being lateral damper
Number J, i.e.,：
In formula, vibration frequency root mean square of weighed accelerationCoefficient 1,0.63,0.2, be respectively
Car body weaving, sidewinder motion, the axle weight coefficient of yaw motion；Wherein, vibration frequency weighting accelerates at different frequencies
Spend root-mean-square valueFrequency weight values be respectively wd(fi)、we(fi)、wf(fi), i.e.,：
(4) low speed rail vehicle two is lateral damper optimum damping coefficient C optimization design：
1. according to the half a=7.85m of length between truck centers, the half a of wheel-base bogie0=1.25m, Vehicle Speed v
The free degree oscillation crosswise optimization design of the low speed rail whole vehicle 17 emulation mould established in=100km/h, and step (2)
Type, the orbital direction irregularity stochastic inputs y with each wheel to placea1(t)、ya2(t)、ya3(t)、ya4(t) and horizontal irregularity is random
Input zθ1(t)、zθ2(t)、zθ3(t)、zθ4(t) it is input stimulus, is asked using optimized algorithm and two systems transverse direction is established in step (3)
The damping optimization design object function J of shock absorber minimum value, it is lateral damper that optimization design, which obtains every bogie two,
Optimal Equivalent damping coefficient C2=86.3kN.s/m；
Wherein, the relation between orbital direction irregularity stochastic inputs is：ya2(t)=ya1(t-0.09s), ya3(t)=ya1
(t-0.5652s), ya4(t)=ya1(t-0.6552s)；Relation between horizontal irregularity random input stimuli is：zθ2(t)=
zθ1(t-0.09s), zθ3(t)=zθ1(t-0.5652s), zθ4(t)=zθ1(t-0.6552s)；Vehicle Speed v=100km/
It is each to take turns the U.S.'s orbital direction irregularity random input stimuli applied to place during h, respectively as shown in Fig. 5, Fig. 6, Fig. 7, Fig. 8；
The horizontal irregularity random input stimuli of U.S.'s track applied, respectively as shown in Fig. 9, Figure 10, Figure 11, Figure 12；
2. it is the installation number n=2 of lateral damper according to every bogie two, and 1. optimization order is set in step (4)
Every bogie two obtained by meter is the optimal Equivalent damping coefficient C of lateral damper2=86.3kN.s/m, is calculated list
Branch two is the optimum damping coefficient C of lateral damper, i.e.,：C=C2/ n=43.15kN.s/m.
The vehicle parameter provided according to embodiment, using rail vehicle special-purpose software SIMPACK, imitated by solid modelling
True checking can obtain, and the low speed rail vehicle two is that the optimum damping coefficient of lateral damper is C=43.17kN.s/m；Understand, profit
It is the optimum damping coefficient C=43.15kN.s/m of lateral damper with the low speed rail vehicle two obtained by Optimization Design,
Matched with the optimum damping coefficient C=43.17kN.s/m obtained by SIMPACK simulating, verifyings, both are only at deviation
0.02kN.s/m, relative deviation are only 0.046%, show low speed rail vehicle two provided by the present invention be lateral damper most
The design method of excellent damped coefficient is correct.
Claims (1)
1. low speed rail vehicle two is the design method of lateral damper optimum damping coefficient, its specific design step is as follows：
(1) free degree of the low speed rail whole vehicle 17 traveling oscillation crosswise differential equation is established：
According to the quality m of the single-unit car body of rail vehicle3, rotary inertia of shaking the headSidewinder rotary inertia J3θ；Every steering structure
The quality m of frame2, rotary inertia of shaking the headSidewinder rotary inertia J2θ；The quality m of each round pair1, rotary inertia of shaking the headOften
One wheel shaft weight W；The horizontal creep coefficient f of each round pair1, longitudinal creep coefficient f2；Longitudinal rigidity K per axle-box positioning device1x、
Lateral stiffness K1y, vertical stiffness K1z；The longitudinal rigidity K of every bogie central spring2x, located lateral stiffness K2y；Every steering
The vertical equivalent stiffness K of the system of frame two suspension2z, vertical equivalent damping Cd2；The torsional rigidity K of single anti-side rolling torsion rodθ；Every steering
Frame to be designed two is the Equivalent damping coefficient C of lateral damper2；Vehicle wheel roll radius r, wheel tread gradient λ；Vehicle traveling speed
Spend v；The half b of wheel and rail contact point horizontal spacing, wheel shaft retaining spring are transversely mounted the half b of spacing1, in bogie
Centre spring is transversely mounted the half b of spacing2, the half a of length between truck centers, the half a of wheel-base bogie0, axle centerline to track
The height h of plane0, the height h of plane on car body barycenter to central spring1, car body barycenter to two is the height of lateral damper
h2, height h of the plane to framework barycenter on central spring3, the height h of bogie frame barycenter to axle centerline4, two systems horizontal stroke
To shock absorber to the height h of framework barycenter5；The barycenter O of steering framing wheel pair before respectively1ff、O1fr, the barycenter of trailing bogie wheel pair
O1rf、O1rr, the barycenter O of forward and backward bogie frame2f、O2rAnd the barycenter O of car body3For the origin of coordinates；With forecarriage front-wheel pair
Yaw displacement y1ff, displacement of shaking the headThe yaw displacement y of forecarriage trailing wheel pair1fr, displacement of shaking the headTrailing bogie front-wheel pair
Yaw displacement y1rf, displacement of shaking the headThe yaw displacement y of trailing bogie trailing wheel pair1rr, displacement of shaking the headForecarriage framework
Yaw displacement y2f, displacement of shaking the headSidewinder displacement θ2f, the yaw displacement y of trailing bogie framework2r, displacement of shaking the headSidewinder
Displacement θ2r, and the yaw displacement y of car body3, displacement of shaking the headSidewinder displacement θ3For coordinate；With the forward and backward wheel of forecarriage and after
Orbital direction irregularity input y at the forward and backward wheel of bogiea1(t)、ya2(t)、ya3(t)、ya4(t) inputted with horizontal irregularity
zθ1(t)、zθ2(t)、zθ3(t)、zθ4(t) it is input stimulus, wherein, t is time variable；Establish low speed rail whole vehicle 17 certainly
The oscillation crosswise differential equation is travelled by degree, i.e.,：
1. the yaw vibration equation of forecarriage front-wheel pair：
2. the yawing equation of forecarriage front-wheel pair：
3. the yaw vibration equation of forecarriage trailing wheel pair：
4. the yawing equation of forecarriage trailing wheel pair：
5. the yaw vibration equation of trailing bogie front-wheel pair：
6. the yawing equation of trailing bogie front-wheel pair：
7. the yaw vibration equation of trailing bogie trailing wheel pair：
8. the yawing equation of trailing bogie trailing wheel pair：
9. the yaw vibration equation of forecarriage framework：
10. the rolling equation of forecarriage framework：
The yawing equation of forecarriage framework：
The yaw vibration equation of trailing bogie framework：
The rolling equation of trailing bogie framework：
The yawing equation of trailing bogie framework：
The yaw vibration equation of car body：
The rolling equation of car body：
Wherein, h=h0+h1+h3+h4；
The yawing equation of car body：
(2) free degree oscillation crosswise optimization design simulation model of low speed rail whole vehicle 17 is built：
The oscillation crosswise differential equation is travelled according to the free degree of low speed rail whole vehicle 17 established in step (1), utilized
Matlab/Simulink simulation softwares, the free degree oscillation crosswise optimization design simulation model of structure low speed rail whole vehicle 17；
(3) the two damping optimization design object function J for being lateral damper are established：
According to the free degree oscillation crosswise optimization design simulation model of low speed rail whole vehicle 17 established in step (2), with
Every bogie two is that the Equivalent damping coefficient of lateral damper is design variable, with each wheel to the orbital direction irregularity at place with
Machine inputs and horizontal irregularity stochastic inputs are input stimulus, and the vibration frequency using the car body weaving obtained by emulation adds
Weigh acceleration root-mean-square valueCar body sidewinders the vibration frequency root mean square of weighed acceleration of motionAnd car body is shaken the head fortune
Dynamic vibration frequency root mean square of weighed accelerationThe two damping optimization design object function J for being lateral damper are established,
I.e.：
In formula, vibration frequency root mean square of weighed accelerationCoefficient 1,0.63,0.2, respectively car body
Weaving, sidewinder motion, the axle weight coefficient of yaw motion；Wherein, vibration frequency weighted acceleration is equal at different frequencies
Root valueFrequency weight values be respectively wd(fi)、we(fi)、wf(fi), i.e.,：
<mrow>
<msub>
<mi>w</mi>
<mi>d</mi>
</msub>
<mrow>
<mo>(</mo>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>)</mo>
</mrow>
<mo>=</mo>
<mfenced open = "{" close = "">
<mtable>
<mtr>
<mtd>
<mn>1</mn>
</mtd>
<mtd>
<mrow>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>&Element;</mo>
<mo>&lsqb;</mo>
<mn>0.5</mn>
<mo>,</mo>
<mn>2</mn>
<mo>&rsqb;</mo>
<mi>H</mi>
<mi>z</mi>
</mrow>
</mtd>
</mtr>
<mtr>
<mtd>
<mrow>
<mn>2</mn>
<mo>/</mo>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
</mrow>
</mtd>
<mtd>
<mrow>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>&Element;</mo>
<mo>(</mo>
<mn>2</mn>
<mo>,</mo>
<mn>80</mn>
<mo>&rsqb;</mo>
<mi>H</mi>
<mi>z</mi>
</mrow>
</mtd>
</mtr>
</mtable>
</mfenced>
<mo>;</mo>
</mrow>
<mrow>
<msub>
<mi>w</mi>
<mi>e</mi>
</msub>
<mrow>
<mo>(</mo>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>)</mo>
</mrow>
<mo>=</mo>
<mfenced open = "{" close = "">
<mtable>
<mtr>
<mtd>
<mn>1</mn>
</mtd>
<mtd>
<mrow>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>&Element;</mo>
<mo>&lsqb;</mo>
<mn>0.5</mn>
<mo>,</mo>
<mn>1</mn>
<mo>&rsqb;</mo>
<mi>H</mi>
<mi>z</mi>
</mrow>
</mtd>
</mtr>
<mtr>
<mtd>
<mrow>
<mn>1</mn>
<mo>/</mo>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
</mrow>
</mtd>
<mtd>
<mrow>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>&Element;</mo>
<mo>(</mo>
<mn>1</mn>
<mo>,</mo>
<mn>80</mn>
<mo>&rsqb;</mo>
<mi>H</mi>
<mi>z</mi>
</mrow>
</mtd>
</mtr>
</mtable>
</mfenced>
<mo>;</mo>
</mrow>
<mrow>
<msub>
<mi>w</mi>
<mi>f</mi>
</msub>
<mrow>
<mo>(</mo>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>)</mo>
</mrow>
<mo>=</mo>
<mfenced open = "{" close = "">
<mtable>
<mtr>
<mtd>
<mn>1</mn>
</mtd>
<mtd>
<mrow>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>&Element;</mo>
<mo>&lsqb;</mo>
<mn>0.5</mn>
<mo>,</mo>
<mn>1</mn>
<mo>&rsqb;</mo>
<mi>H</mi>
<mi>z</mi>
</mrow>
</mtd>
</mtr>
<mtr>
<mtd>
<mrow>
<mn>1</mn>
<mo>/</mo>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
</mrow>
</mtd>
<mtd>
<mrow>
<msub>
<mi>f</mi>
<mi>i</mi>
</msub>
<mo>&Element;</mo>
<mo>(</mo>
<mn>1</mn>
<mo>,</mo>
<mn>80</mn>
<mo>&rsqb;</mo>
<mi>H</mi>
<mi>z</mi>
</mrow>
</mtd>
</mtr>
</mtable>
</mfenced>
<mo>;</mo>
</mrow>
(4) low speed rail vehicle two is lateral damper optimum damping coefficient C optimization design：
1. according to the half a of length between truck centers, the half a of wheel-base bogie0, established in Vehicle Speed v, and step (2)
The free degree oscillation crosswise optimization design simulation model of low speed rail whole vehicle 17, with each wheel to the orbital direction irregularity at place with
Machine inputs ya1(t)、ya2(t)、ya3(t)、ya4And horizontal irregularity stochastic inputs z (t)θ1(t)、zθ2(t)、zθ3(t)、zθ4(t) it is
Input stimulus, asked using optimized algorithm in step (3) and establish the two damping optimization design object function J for being lateral damper
Minimum value, corresponding design variable are the optimal Equivalent damping coefficient C that every bogie two is lateral damper2；
Wherein, the relation between orbital direction irregularity stochastic inputs is： Relation between horizontal irregularity stochastic inputs is：
2. be the installation number n of lateral damper according to every bogie two, and in step (4) 1. optimization order design obtained by
Every bogie two be lateral damper optimal Equivalent damping coefficient C2, be calculated single branch two be lateral damper most
Excellent damped coefficient C, i.e.,：C=C2/n。
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CN201510559657.6A CN105069261B (en) | 2015-09-06 | 2015-09-06 | Low speed rail vehicle two is the design method of lateral damper optimum damping coefficient |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CN201510559657.6A CN105069261B (en) | 2015-09-06 | 2015-09-06 | Low speed rail vehicle two is the design method of lateral damper optimum damping coefficient |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| CN105069261A CN105069261A (en) | 2015-11-18 |
| CN105069261B true CN105069261B (en) | 2017-12-12 |
Family
ID=54498627
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CN201510559657.6A Expired - Fee Related CN105069261B (en) | 2015-09-06 | 2015-09-06 | Low speed rail vehicle two is the design method of lateral damper optimum damping coefficient |
Country Status (1)
| Country | Link |
|---|---|
| CN (1) | CN105069261B (en) |
Families Citing this family (10)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CN106777481B (en) * | 2016-11-18 | 2019-06-14 | 中铁二院工程集团有限责任公司 | Deck Arch Bridges lateral stiffness control method |
| CN109766635B (en) * | 2019-01-11 | 2023-02-03 | 中国铁路总公司 | Optimized layout method for state perception sensor of mechanical part of locomotive |
| CN110502726B (en) * | 2019-08-28 | 2021-04-27 | 西南交通大学 | Method and device for predicting noise in railway vehicle |
| CN111222087B (en) * | 2020-01-15 | 2021-09-14 | 西南交通大学 | Method for calculating rolling frequency and swing center position of railway passenger car body |
| CN114279659B (en) * | 2020-09-27 | 2023-06-23 | 中车株洲电力机车研究所有限公司 | Hinge system test method, system, medium and equipment |
| CN113239449B (en) * | 2020-12-29 | 2022-07-01 | 西南交通大学 | Method for analyzing snaking motion of flexible bogie of railway vehicle |
| CN112765728B (en) * | 2021-01-11 | 2022-03-29 | 四川轻化工大学 | Shimmy optimization control method for single-axle bogie straddle type monorail vehicle |
| CN217260056U (en) * | 2021-10-15 | 2022-08-23 | 合肥工业大学 | Rail vehicle interconnected secondary suspension transverse vibration damper |
| CN113859392B (en) * | 2021-10-20 | 2022-12-27 | 东风汽车集团股份有限公司 | Vehicle shock absorber adjusting method and system |
| CN114248813B (en) * | 2021-11-29 | 2023-06-02 | 同济大学 | Variable damping dynamic vibration absorber for railway vehicle |
Citations (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US6732033B2 (en) * | 2002-01-11 | 2004-05-04 | Active Shock, Inc. | Semi-active shock absorber control system |
| DE102012015974A1 (en) * | 2012-08-11 | 2013-02-28 | Daimler Ag | Suspension assembly for supporting vibrating unit relative to supporting structure in body assembly of motor vehicle, comprises a damping device that damps the relative movement of suspension assembly along a specific direction |
| CN104455157A (en) * | 2014-10-29 | 2015-03-25 | 山东理工大学 | Obtaining method of car seat suspension hydraulic buffer nonlinear speed characteristic parameter |
-
2015
- 2015-09-06 CN CN201510559657.6A patent/CN105069261B/en not_active Expired - Fee Related
Patent Citations (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US6732033B2 (en) * | 2002-01-11 | 2004-05-04 | Active Shock, Inc. | Semi-active shock absorber control system |
| DE102012015974A1 (en) * | 2012-08-11 | 2013-02-28 | Daimler Ag | Suspension assembly for supporting vibrating unit relative to supporting structure in body assembly of motor vehicle, comprises a damping device that damps the relative movement of suspension assembly along a specific direction |
| CN104455157A (en) * | 2014-10-29 | 2015-03-25 | 山东理工大学 | Obtaining method of car seat suspension hydraulic buffer nonlinear speed characteristic parameter |
Non-Patent Citations (5)
| Title |
|---|
| B型城市轨道交通车辆转向架低动力作用仿真研究;赵雷;《万方学术论文数据库》;20101231 * |
| Electromagnetic Energy-Harvesting Shock Absorbers: Design, Modeling, and Road Tests;Zhongjie Li 等;《IEEE TRANSACTIONS ON VEHICULAR TECHNOLOGY》;20130331;第62卷(第3期);第1065-1074页 * |
| 二系横向减振器阻尼系数对车辆横向振动影响的仿真研究;池毓敢 等;《铁道车辆 》;20140430;第52卷(第4期);第15-16页 * |
| 基于虚拟样机的自行火炮行驶平顺性仿真研究;吴大林;《兵工学报》;20061130;第27卷(第6期);第970-973页 * |
| 采用MATLAB的车辆悬挂半主动控制的仿真研究;王月明 等;《铁道车辆》;20020930;第40卷(第9期);第1-4页 * |
Also Published As
| Publication number | Publication date |
|---|---|
| CN105069261A (en) | 2015-11-18 |
Similar Documents
| Publication | Publication Date | Title |
|---|---|---|
| CN105069261B (en) | Low speed rail vehicle two is the design method of lateral damper optimum damping coefficient | |
| CN101847270B (en) | Virtual prototype-based suspension system control simulation method of four-axle heavy truck | |
| CN105117556B (en) | One system of high ferro and the cooperative optimization method of two systems and end shock absorber damping | |
| CN105138783B (en) | The design method of car body of high speed railway car end cross shock absorber damping | |
| CN105160105B (en) | High ferro two is vertical and the cooperative optimization method of end longitudinal shock absorber damped coefficient | |
| CN105069260B (en) | High speed railway car two is the Optimization Design of vertical suspension Optimal damping ratio | |
| CN105160103B (en) | The system of high speed railway car one and two be vertical suspension damping ratio cooperative optimization method | |
| CN208673333U (en) | A kind of suspension type monorail vehicle Coupled Dynamics analogue system | |
| CN105117554B (en) | High speed railway car one is the design method of vertical suspension Optimal damping ratio | |
| CN105183982B (en) | The design method of car body of high speed railway car end longitudinal shock absorber damped coefficient | |
| CN105183979B (en) | High ferro is vertical and the cooperative optimization method of body end portion longitudinal shock absorber damped coefficient | |
| CN105138786A (en) | Collaborative optimization method for damping coefficients of high-speed rail secondary transverse damper and vehicle body end transverse damper | |
| CN105224718B (en) | The system of high speed railway car two laterally suspends the Optimization Design of Optimal damping ratio | |
| CN105069259B (en) | Low speed rail vehicle two is the Analytic Calculation Method of vertical suspension optimum damping ratio | |
| CN105160179B (en) | The system of high speed railway car two laterally suspends the Analytic Calculation Method of Optimal damping ratio | |
| Shahriar et al. | Simulation and analysis of half-car passive suspension system | |
| CN105160180B (en) | High speed railway car two is the Analytic Calculation Method of vertical suspension Optimal damping ratio | |
| CN105138784B (en) | The Analytic Calculation Method of high speed railway car seat suspension optimum damping ratio | |
| Dumitriu | Numerical analysis of the influence of lateral suspension parameters on the ride quality of railway vehicles | |
| CN105183984B (en) | Low speed rail vehicle two is the Optimization Design of vertical suspension optimum damping ratio | |
| Hu | Road profile recovery using vertical acceleration data | |
| CN110949436A (en) | Optimal damping regulation and control method for variable-track-pitch power bogie secondary system and motor transverse shock absorber | |
| CN105069263B (en) | High speed railway car seat and two be vertical suspension damping ratio cooperative optimization method | |
| CN105138785B (en) | High-speed rail seat and a system and two be vertical suspension damping ratio cooperative optimization method | |
| Chen | The research of vehicle’s ride comfort in the nonlinear suspension system |
Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| C06 | Publication | ||
| PB01 | Publication | ||
| C10 | Entry into substantive examination | ||
| SE01 | Entry into force of request for substantive examination | ||
| GR01 | Patent grant | ||
| GR01 | Patent grant | ||
| CF01 | Termination of patent right due to non-payment of annual fee |
Granted publication date: 20171212 Termination date: 20200906 |
|
| CF01 | Termination of patent right due to non-payment of annual fee |