CN105138783B - The design method of car body of high speed railway car end cross shock absorber damping - Google Patents
The design method of car body of high speed railway car end cross shock absorber damping Download PDFInfo
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Abstract
The present invention relates to the design method of car body of high speed railway car end cross shock absorber damping, belongs to high speed railway car suspension technical field.The present invention is by establishing the free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17, using orbital direction irregularity and horizontal irregularity as input stimulus, the minimum design object of vibration acceleration root-mean-square value of motion is sidewindered with car body, optimization design obtains the optimal damping constant of body end portion lateral damper.By designing example and SIMPACK simulating, verifyings, the damping coefficient of the available accurately and reliably body end portion lateral damper of this method, the design for car body of high speed railway car end cross shock absorber damping provides reliable design method.Using this method, the design level of high speed railway car suspension system can be not only improved, improves vehicle safety and stationarity；Meanwhile product design and testing expenses can be also reduced, strengthen the competitiveness in the international market of China's rail vehicle.
Description
Technical field
The present invention relates to high speed railway car suspension, particularly car body of high speed railway car end cross resistance of shock absorber system
Several design methods.
Background technology
Body end portion lateral damper is installed between radial type or the two adjacent car bodies for turning to posture high speed railway car,
Effectively car body can be controlled to wave, dramatically increase the globality and stability of vehicle.However, understood according to institute's inspection information, due to
Rail vehicle belongs to Mdof Vibration System, it is carried out dynamic analysis calculate it is extremely difficult, at present both at home and abroad for car
The design of body end portion lateral damper damped coefficient, the theoretical design method of system is never provided, be mostly by calculating
Machine technology, using Dynamics Simulation soft sim PACK or ADAMS/Rail, optimize by solid modelling and determine that its is big
It is small, although this method can obtain reliable simulation numerical, make vehicle that there is preferable power performance, however, with rail
The continuous improvement of road Vehicle Speed, design of the people to body end portion lateral damper damped coefficient propose higher want
Ask, the method for body end portion lateral damper damped coefficient design at present can not provide the innovation theory with directive significance, no
Development to absorber designing requirement in the case of rail vehicle constantly speed-raising can be met.Therefore, it is necessary to establish a kind of accurate, reliable
Car body of high speed railway car end cross shock absorber damping design method, meet rail vehicle constantly speed-raising in the case of
Requirement to absorber designing, the design level and product quality of high speed railway car suspension system are improved, improve vehicle traveling
Security and stationarity；Meanwhile product design and testing expenses are reduced, shorten the product design cycle, strengthen China's rail vehicle
Competitiveness in the international market.
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,
The design method of reliable car body of high speed railway car end cross shock absorber damping, its design flow diagram are as shown in Figure 1；
The free degree of high speed railway car vehicle 17 travels left view such as Fig. 2 of oscillation crosswise model, and high speed railway car 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, car body of high speed railway car end cross resistance of shock absorber system provided by the present invention
Several design method, it is characterised in that use following design procedure：
(1) free degree of the rail vehicle whole 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
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；The longitudinal register rigidity of each round pair
K1x, located lateral stiffness K1y；Every bogie one side one is the vertical equivalent stiffness K of suspension1z, vertical Equivalent damping coefficient Cd1；
The longitudinal rigidity K of every bogie central spring2x, located lateral stiffness K2y；Every system of bogie two suspends vertical equivalent firm
Spend K2z, vertical Equivalent damping coefficient Cd2；The torsional rigidity K of single anti-side rolling torsion rodθ；The damped coefficient of a pair of anti-hunting damper holders
Cs；It is for a pair two the damped coefficient C of lateral damper2；The Equivalent damping coefficient C of body end portion longitudinal shock absorber3；Car to be designed
The Equivalent damping coefficient C of body end portion lateral damperr；Vehicle wheel roll radius r, wheel tread gradient λ；Vehicle Speed v；Car
The half b of wheel and rail contact point horizontal spacing, wheel shaft retaining spring are transversely mounted the half b of spacing1, bogie central spring
It is transversely mounted the half b of spacing2, anti-hunting damper holder is transversely mounted the half b of spacing3, between car body longitudinal shock absorber is transversely mounted
Away from half b4, the half a of length between truck centers, the half a of wheel-base bogie0, the half of the longitudinally mounted spacing of cross-car shock absorber
L, the height h of axle centerline to orbit plane0, the height h of plane on car body barycenter to central spring1, car body barycenter to two systems
The height h of lateral damper2, height h of the plane to framework barycenter on central spring3, bogie frame barycenter to axle centerline
Height h4, two be height h of the lateral damper to framework barycenter5, height of the body end portion lateral damper to car body barycenter
h6；The barycenter O of steering framing wheel pair before respectively1ff、O1fr, the barycenter O of trailing bogie wheel pair1rf、O1rr, forward and backward bogie frame
Barycenter O2f、O2rAnd the barycenter O of car body3For the origin of coordinates；With the yaw displacement y of forecarriage front-wheel pair1ff, displacement of shaking the headThe yaw displacement y of forecarriage trailing wheel pair1fr, displacement of shaking the headThe yaw displacement y of trailing bogie front-wheel pair1rf, position of shaking the head
MoveThe yaw displacement y of trailing bogie trailing wheel pair1rr, displacement of shaking the headThe yaw displacement y of forecarriage framework2f, position of shaking the head
MoveSidewinder displacement θ2f, the yaw displacement y of trailing bogie framework2r, displacement of shaking the headSidewinder displacement θ2r, and the yaw of car body
Displacement y3, displacement of shaking the headSidewinder displacement θ3For coordinate；At the forward and backward wheel of forecarriage and the forward and backward wheel of trailing bogie
Orbital direction irregularity inputs ya1(t)、ya2(t)、ya3(t)、ya4(t) and horizontal irregularity inputs zθ1(t)、zθ2(t)、zθ3(t)、
zθ4(t) it is input stimulus, wherein, t is time variable；Establish the free degree of rail vehicle whole 17 traveling oscillation crosswise differential side
Journey, 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 rail vehicle whole 17 is built：
The oscillation crosswise differential equation is travelled according to the free degree of rail vehicle whole 17 established in step (1), utilized
Matlab/Simulink simulation softwares, build the free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17；
(3) the damping optimization design object function J of body end portion lateral damper is established：
According to the free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17 established in step (2), with
The Equivalent damping coefficient of body end portion lateral damper is design variable, the orbital direction irregularity stochastic inputs with each wheel to place
It is input stimulus with horizontal irregularity stochastic inputs, the vibration acceleration root mean square of motion is sidewindered using the car body obtained by emulation
ValueThe damping optimization design object function J of body end portion lateral damper is established, i.e.,：
(4) body end portion lateral damper optimal damping constant 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 rail vehicle whole 17 of foundation, the orbital direction irregularity with each wheel to place
Stochastic inputs ya1(t)、ya2(t)、ya3(t)、ya4And horizontal irregularity stochastic inputs z (t)θ1(t)、zθ2(t)、zθ3(t)、zθ4(t)
For input stimulus, the damping optimization design object that body end portion lateral damper is established in step (3) is sought using optimized algorithm
Function J minimum value, corresponding design variable are the best equivalence damped coefficient C of body end portion lateral damperr；
Wherein, the relation between orbital direction irregularity stochastic inputs is： Horizontal irregularity stochastic inputs it
Between relation be：
2. according to the installation number n of body end portion lateral damper, and in step (4) 1. optimization order design obtained by
The best equivalence damped coefficient C of body end portion lateral damperr, the optimal resistance of single branch body end portion lateral damper is calculated
Buddhist nun coefficient C, i.e.,：C=Cr/n。
The present invention has the advantage that than prior art：
Because rail vehicle belongs to Mdof Vibration System, it is carried out dynamic analysis calculate it is extremely difficult, at present
Both at home and abroad for the design of body end portion lateral damper damped coefficient, the theoretical design method of system is never provided, 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 improvement of rail vehicle travel speed, design of the people to body end portion lateral damper damped coefficient carries
Higher requirement is gone out, the method for body end portion lateral damper damped coefficient design at present can not be provided with directive significance
Innovation theory, it is impossible to meet development to absorber designing requirement in the case of rail vehicle constantly speed-raising.
The present invention travels the oscillation crosswise differential equation by establishing the free degree of rail vehicle whole 17, utilizes MATLAB/
Simulink simulation softwares, the free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17 is constructed, and with track
Direction irregularity and horizontal irregularity are input stimulus, and the minimum design of vibration acceleration root-mean-square value of motion is sidewindered with car body
Target, optimization design obtain the optimal damping constant of body end portion lateral damper.Tested by designing example and SIMPACK emulation
Card understands that the damping coefficient of the available accurately and reliably body end portion lateral damper of this method, is high speed railway car car
The design of body end portion lateral damper damped coefficient provides reliable design method.Using this method, can not only improve at a high speed
The design level and product quality of rail vehicle suspension system, improve vehicle safety and stationarity；Meanwhile it can also reduce
Product design and testing expenses, shorten the product design cycle, strengthen the competitiveness in the international market of China's rail vehicle.
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 of car body of high speed railway car end cross shock absorber damping design method；
Fig. 2 is the left view of the free degree of high speed railway car vehicle 17 traveling oscillation crosswise model；
Fig. 3 is the top view of the free degree of high speed railway car vehicle 17 traveling oscillation crosswise model；
Fig. 4 is the free degree oscillation crosswise optimization design simulation model figure of rail vehicle whole 17 of embodiment；
Fig. 5 is the German orbital direction irregularity random input stimuli y that embodiment is applieda1(t)；
Fig. 6 is the German orbital direction irregularity random input stimuli y that embodiment is applieda2(t)；
Fig. 7 is the German orbital direction irregularity random input stimuli y that embodiment is applieda3(t)；
Fig. 8 is the German orbital direction irregularity random input stimuli y that embodiment is applieda4(t)；
Fig. 9 is the horizontal irregularity random input stimuli z of German track that embodiment is appliedθ1(t)；
Figure 10 is the horizontal irregularity random input stimuli z of German track that embodiment is appliedθ2(t)；
Figure 11 is the horizontal irregularity random input stimuli z of German track that embodiment is appliedθ3(t)；
Figure 12 is the horizontal irregularity random input stimuli z of German track that embodiment is appliedθ4(t)。
Embodiment
The present invention is described in further detail below by an embodiment.
Four body end portion longitudinal shock absorbers and a body end are installed between two adjacent car bodies of certain high speed railway car
Portion's lateral damper, i.e. n=1, the quality m of its single-unit car body3=63966kg, rotary inertia of shaking the head
Sidewinder rotary inertia J3θ=77200kg.m2；The quality m of every bogie frame2=2758kg, rotary inertia of shaking the headSidewinder rotary inertia J2θ=2212kg.m2；The quality m of each round pair1=1721kg, rotary inertia of shaking the headEach wheel shaft weight W=150000N；The horizontal creep coefficient f of each round pair1=16990000N, longitudinal creep
Coefficient f2=16990000N；The longitudinal register stiffness K of each round pair1x=13.739 × 106N/m, located lateral stiffness K1y=
4.892×106N/m；Every bogie one side one is the vertical equivalent stiffness K of suspension1z=2.74 × 106N/m, vertical equivalent resistance
Buddhist nun's coefficient Cd1=28.3kN.s/m；The longitudinal rigidity K of every bogie central spring2x=0.18 × 106N/m, located lateral are firm
Spend K2y=0.18 × 106N/m；The vertical equivalent stiffness K of the every system of bogie two suspension2z=1.1368 × 106N/m, vertical etc.
Imitate damped coefficient Cd2=118.7kN.s/m；The torsional rigidity K of single anti-side rolling torsion rodθ=2.5 × 106N.m/rad；One confrontation
The damped coefficient C of hunting Dampers=1027kN.s/m；It is for a pair two the damped coefficient C of lateral damper2=90kN.s/m；Car
The Equivalent damping coefficient C of body end portion longitudinal shock absorber3=2897.6kN.s/m；Vehicle wheel roll radius r=0.445m, wheel tread
Gradient λ=0.15；The half b=0.7465m of wheel and rail contact point horizontal spacing, wheel shaft retaining spring are transversely mounted spacing
Half b1=1.15m, bogie central spring are transversely mounted the half b of spacing2=1.3m, anti-hunting damper holder are transversely mounted
The half b of spacing3=1.4m, car body longitudinal shock absorber are transversely mounted the half b of spacing4=1.2m, the half a=of length between truck centers
9.5m, the half a of wheel-base bogie0=1.35m, the half l=13.3m of the longitudinally mounted spacing of cross-car shock absorber, axletree
Height h of the center line to orbit plane0=0.347m, the height h of plane on car body barycenter to central spring1=0.8m, car body matter
The heart is to the height h that two be lateral damper2=0.616m, height h of the plane to framework barycenter on central spring3=0.416m, turn
To framework frame barycenter to the height h of axle centerline4=0.137m, two be height h of the lateral damper to framework barycenter5=
0.6m, the height h of body end portion lateral damper to car body barycenter6=0.5m；Body end portion lateral damper to be designed etc.
Effect damped coefficient is Cr.The required vehicle traveling speed of car body of high speed railway car end cross shock absorber damping design
V=300km/h is spent, the optimal damping constant of the car body of high speed railway car end cross shock absorber is designed.
The design method for the car body of high speed railway car end cross shock absorber damping that present example is provided, its
Design flow diagram is high as shown in figure 1, left view such as Fig. 2 of the free degree of high speed railway car 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 rail vehicle whole 17 traveling oscillation crosswise differential equation is established：
According to the quality m of the single-unit car body of rail vehicle3=63966kg, rotary inertia of shaking the head
Sidewinder rotary inertia J3θ=77200kg.m2；The quality m of every bogie frame2=2758kg, rotary inertia of shaking the headSidewinder rotary inertia J2θ=2212kg.m2；The quality m of each round pair1=1721kg, rotary inertia of shaking the headEach wheel shaft weight W=150000N；The horizontal creep coefficient f of each round pair1=16990000N, longitudinal creep
Coefficient f2=16990000N；The longitudinal register stiffness K of each round pair1x=13.739 × 106N/m, located lateral stiffness K1y=
4.892×106N/m；Every bogie one side one is the vertical equivalent stiffness K of suspension1z=2.74 × 106N/m, vertical equivalent resistance
Buddhist nun's coefficient Cd1=28.3kN.s/m；The longitudinal rigidity K of every bogie central spring2x=0.18 × 106N/m, located lateral are firm
Spend K2y=0.18 × 106N/m；The vertical equivalent stiffness K of the every system of bogie two suspension2z=1.1368 × 106N/m, vertical etc.
Imitate damped coefficient Cd2=118.7kN.s/m；The torsional rigidity K of single anti-side rolling torsion rodθ=2.5 × 106N.m/rad；One confrontation
The damped coefficient C of hunting Dampers=1027kN.s/m；It is for a pair two the damped coefficient C of lateral damper2=90kN.s/m；Car
The Equivalent damping coefficient C of body end portion longitudinal shock absorber3=2897.6kN.s/m；Body end portion lateral damper to be designed it is equivalent
Damped coefficient Cr；Vehicle wheel roll radius r=0.445m, wheel tread gradient λ=0.15；Vehicle Speed v=300km/h；
The half b=0.7465m of wheel and rail contact point horizontal spacing, wheel shaft retaining spring are transversely mounted the half b of spacing1=
1.15m, bogie central spring are transversely mounted the half b of spacing2=1.3m, anti-hunting damper holder are transversely mounted the half of spacing
b3=1.4m, car body longitudinal shock absorber are transversely mounted the half b of spacing4=1.2m, the half a=9.5m of length between truck centers, bogie
The half a of wheelbase0=1.35m, the half l=13.3m of the longitudinally mounted spacing of cross-car shock absorber, axle centerline to track
The height h of plane0=0.347m, the height h of plane on car body barycenter to central spring1=0.8m, car body barycenter are horizontal to two systems
The height h of shock absorber2=0.616m, height h of the plane to framework barycenter on central spring3=0.416m, bogie frame barycenter
To the height h of axle centerline4=0.137m, two be height h of the lateral damper to framework barycenter5=0.6m, body end portion are horizontal
To shock absorber to the height h of car body barycenter6=0.5m；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；Before forecarriage
The yaw displacement y of wheel pair1ff, displacement of shaking the headThe yaw displacement y of forecarriage trailing wheel pair1fr, displacement of shaking the headTrailing bogie
The yaw displacement y of front-wheel pair1rf, displacement of shaking the headThe yaw displacement y of trailing bogie trailing wheel pair1rr, displacement of shaking the headFront steering
The yaw displacement y of framework frame2f, 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；It is forward and backward with forecarriage
Orbital direction irregularity input y at wheel and the forward and backward wheel of trailing bogiea1(t)、ya2(t)、ya3(t)、ya4(t) with level not
Smooth-going input zθ1(t)、zθ2(t)、zθ3(t)、zθ4(t) it is input stimulus, wherein, t is time variable；Establish rail vehicle whole
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 rail vehicle whole 17 is built：
The oscillation crosswise differential equation is travelled according to the free degree of rail vehicle whole 17 established in step (1), utilized
Matlab/Simulink simulation softwares, the free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17 is built, is such as schemed
Shown in 4；
(3) the damping optimization design object function J of body end portion lateral damper is established：
According to the free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17 established in step (2), with
The Equivalent damping coefficient of body end portion lateral damper is design variable, the orbital direction irregularity stochastic inputs with each wheel to place
It is input stimulus with horizontal irregularity stochastic inputs, the vibration acceleration root mean square of motion is sidewindered using the car body obtained by emulation
ValueThe damping optimization design object function J of body end portion lateral damper is established, i.e.,：
(4) body end portion lateral damper optimal damping constant C optimization design：
1. according to the half a=9.5m of length between truck centers, the half a of wheel-base bogie0=1.35m, Vehicle Speed v=
The free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17 established in 300km/h, and step (2), with each
Take turns the orbital direction irregularity stochastic inputs y to placea1(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 and establish body end portion in step (3) and laterally subtract
Shake device damping optimization design object function J minimum value, optimization design obtains the best equivalence of body end portion lateral damper
Damped coefficient Cr=165kN.s/m；
Wherein, the relation between orbital direction irregularity stochastic inputs is：ya2(t)=ya1(t-0.0324s), ya3(t)=
ya1(t-0.228s), ya4(t)=ya1(t-0.2604s)；Relation between horizontal irregularity stochastic inputs is：zθ2(t)=zθ1
(t-0.0324s), zθ3(t)=zθ1(t-0.228s), zθ4(t)=zθ1(t-0.2604s)；Vehicle Speed v=300km/h
When, it is each to take turns the German orbital direction irregularity random input stimuli applied to place, respectively as shown in Fig. 5, Fig. 6, Fig. 7, Fig. 8；
The horizontal irregularity random input stimuli of German track applied, respectively as shown in Fig. 9, Figure 10, Figure 11, Figure 12；
2. according to the installation number n=1 of body end portion lateral damper, and 1. optimization order design gained in step (4)
The best equivalence damped coefficient C of the body end portion lateral damper arrivedr=165.3kN.s/m, single branch body end portion is calculated
The optimal damping constant C of lateral damper, i.e.,：C=Cr/ n=165.3kN.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 optimal damping constant of the car body of high speed railway car end cross shock absorber is C=165.5kN.s/m；Can
Know, using the optimal damping constant C=165.3kN.s/m of the body end portion lateral damper obtained by Optimization Design, with
Optimal damping constant C=165.5kN.s/m obtained by SIMPACK simulating, verifyings matches, and both deviations are only 0.2kN.s/
M, relative deviation are only 0.18%, show car body of high speed railway car end cross shock absorber damping provided by the present invention
Design method be correct.
Claims (1)
1. the design method of car body of high speed railway car end cross shock absorber damping, its specific design step is as follows：
(1) free degree of the rail vehicle whole 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 headIt is each
Wheel shaft weight W；The horizontal creep coefficient f of each round pair1, longitudinal creep coefficient f2；The longitudinal register stiffness K of each round pair1x, laterally
Locating stiffness K1y；Every bogie one side one is the vertical equivalent stiffness K of suspension1z, vertical Equivalent damping coefficient Cd1；Every turns
To the longitudinal rigidity K of frame central spring2x, located lateral stiffness K2y；The vertical equivalent stiffness K of the every system of bogie two suspension2z、
Vertical Equivalent damping coefficient Cd2；The torsional rigidity K of single anti-side rolling torsion rodθ；The damped coefficient C of a pair of anti-hunting damper holderss；One
It is the damped coefficient C of lateral damper to two2；The Equivalent damping coefficient C of body end portion longitudinal shock absorber3；Body end portion to be designed
The Equivalent damping coefficient C of lateral damperr；Vehicle wheel roll radius r, wheel tread gradient λ；Vehicle Speed v；Wheel and steel
The half b of rail contact point horizontal spacing, wheel shaft retaining spring are transversely mounted the half b of spacing1, bogie central spring transverse direction peace
Fill the half b of spacing2, anti-hunting damper holder is transversely mounted the half b of spacing3, car body longitudinal shock absorber is transversely mounted the one of spacing
Half b4, the half a of length between truck centers, the half a of wheel-base bogie0, the half l of the longitudinally mounted spacing of cross-car shock absorber, axletree
Height h of the center line to orbit plane0, the height h of plane on car body barycenter to central spring1, car body barycenter laterally subtracts to two systems
Shake the height h of device2, height h of the plane to framework barycenter on central spring3, the height of bogie frame barycenter to axle centerline
h4, two be height h of the lateral damper to framework barycenter5, the height h of body end portion lateral damper to car body barycenter6；Respectively
The barycenter O of steering framing wheel pair in the past1ff、O1fr, the barycenter O of trailing bogie wheel pair1rf、O1rr, the barycenter of forward and backward bogie frame
O2f、O2rAnd the barycenter O of car body3For the origin of coordinates；With the yaw displacement y of forecarriage front-wheel pair1ff, displacement of shaking the headFront steering
The yaw displacement y of frame trailing wheel pair1fr, displacement of shaking the headThe yaw displacement y of trailing bogie front-wheel pair1rf, displacement of shaking the headAfter turn
Yaw displacement y to frame trailing wheel pair1rr, displacement of shaking the headThe yaw displacement y of forecarriage framework2f, 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, shake
Head displacementSidewinder displacement θ3For coordinate；With the orbital direction at the forward and backward wheel of forecarriage and the forward and backward wheel of trailing bogie not
Smooth-going input ya1(t)、ya2(t)、ya3(t)、ya4(t) and horizontal irregularity inputs zθ1(t)、zθ2(t)、zθ3(t)、zθ4(t) to be defeated
Enter excitation, wherein, t is time variable；The free degree of the rail vehicle whole 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 rail vehicle whole 17 is built：
The oscillation crosswise differential equation is travelled according to the free degree of rail vehicle whole 17 established in step (1), utilizes Matlab/
Simulink simulation softwares, build the free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17；
(3) the damping optimization design object function J of body end portion lateral damper is established：
According to the free degree oscillation crosswise optimization design simulation model of rail vehicle whole 17 established in step (2), with car body
The Equivalent damping coefficient of end cross shock absorber is design variable, orbital direction irregularity stochastic inputs and water with each wheel to place
Flat irregularity stochastic inputs are input stimulus, and the vibration acceleration root-mean-square value of motion is sidewindered using the car body obtained by emulationThe damping optimization design object function J of body end portion lateral damper is established, i.e.,：
<mrow>
<mi>J</mi>
<mo>=</mo>
<msub>
<mi>&sigma;</mi>
<msub>
<mover>
<mi>&theta;</mi>
<mo>&CenterDot;&CenterDot;</mo>
</mover>
<mn>3</mn>
</msub>
</msub>
<mo>;</mo>
</mrow>
(4) body end portion lateral damper optimal damping constant 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 rail vehicle whole 17, it is defeated at random to the orbital direction irregularity at place with each wheel
Enter ya1(t)、ya2(t)、ya3(t)、ya4And horizontal irregularity stochastic inputs z (t)θ1(t)、zθ2(t)、zθ3(t)、zθ4(t) it is input
Excitation, ask the middle damping optimization design object function J's for establishing body end portion lateral damper of step (3) using optimized algorithm
Minimum value, corresponding design variable are the best equivalence damped coefficient C of body end portion lateral damperr；
Wherein, the relation between orbital direction irregularity stochastic inputs is： Relation between horizontal irregularity stochastic inputs is：
2. according to the installation number n of body end portion lateral damper, and the 1. resulting car body of optimization order design in step (4)
The best equivalence damped coefficient C of end cross shock absorberr, the optimal damping system of single branch body end portion lateral damper is calculated
Number C, i.e.,：C=Cr/n。
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| Simulation and analysis on handling stability of tractor-semitrailer;Ou Chuanjin;《Transportation Information and Safety (ICTIS), 2015 International Conference on》;20150628;第94-101页 * |
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