CN105138786A - Collaborative optimization method for damping coefficients of high-speed rail secondary transverse damper and vehicle body end transverse damper - Google Patents

Abstract

The invention relates to a collaborative optimization method for damping coefficients of a high-speed rail secondary transverse damper and a vehicle body end transverse damper, and belongs to the technical field of suspension of high-speed rail vehicles. As a seventeen-freedom-degree transverse vibration collaborative optimization simulation model of a whole high-speed rail vehicle is set up, and irregularity in the railway direction and horizontal irregularity serve as input excitation, with the minimum root-mean-square value of vibration weighted acceleration of vehicle body transverse motion as the design target, the optimal damping coefficients of the high-speed rail secondary transverse damper and the vehicle body end transverse damper are obtained through optimization design. It can be known through design examples and SIMPACK simulation verification that the damping coefficient values of the high-speed rail secondary transverse damper and the vehicle body end transverse damper can be obtained accurately and reliably through the method, and a reliable design method is provided for designing the damping coefficients of the high-speed rail secondary transverse damper and the vehicle body end transverse damper. By means of the method, the design level of a high-speed rail vehicle suspension system can be improved, traveling safety and stability of the vehicle are improved, product design and test cost can be reduced, and the international market competitiveness of the rail vehicles in China is enhanced.

Description

High ferro two is and the cooperative optimization method of car body end cross shock absorber damping

Technical field

The present invention relates to high speed railway car suspension, particularly high ferro two is and the cooperative optimization method of car body end cross shock absorber damping.

Background technology

Two is that lateral damper and car body end cross vibration damper have important impact to the riding comfort of high ferro and security.But, known according to institute's inspection information, because high ferro belongs to Mdof Vibration System, carrying out dynamic analysis to it calculates very difficult, for high ferro two be and the design of car body end cross shock absorber damping both at home and abroad at present, never provide the theoretical design method of system, mostly be that lateral damper and car body end cross vibration damper are individually studied to two, and computer technology, utilize Dynamics Simulation soft sim PACK or ADAMS/Rail, optimize respectively by solid modelling and determine its size, although the method can obtain reliable simulation numerical, vehicle is made to have good power performance, but, because two be lateral damper and car body end cross vibration damper is a complication system intercoupled, the method that this individually modeling at present designs its shock absorber damping, be difficult to make high ferro two be and the ratio of damping of car body end cross vibration damper reaches optimum matching, and improving constantly along with high ferro travel speed, people to two are and the design of car body end cross shock absorber damping is had higher requirement, current high ferro two is and the method for car body end cross shock absorber damping design can not provide the innovation theory with directive significance, the development to absorber designing requirement in rail vehicle constantly speed-raising situation can not be met.Therefore, must set up a kind of accurately, reliably high ferro two be and the cooperative optimization method of car body end cross shock absorber damping, meet the requirement to absorber designing in rail vehicle constantly speed-raising situation, improve design level and the product quality of high ferro suspension system, improve vehicle riding comfort and security; Meanwhile, reduce product design and testing expenses, shorten the product design cycle, strengthen the competitiveness in the international market of China's rail vehicle.

Summary of the invention

For the defect existed in above-mentioned prior art, technical matters to be solved by this invention be to provide a kind of accurately, reliably high ferro two be and the cooperative optimization method of car body end cross shock absorber damping that its design flow diagram as shown in Figure 1; High ferro car load 17 degree of freedom travels the left view of transverse vibration model as Fig. 2, and high ferro car load 17 degree of freedom travels the vertical view of transverse vibration model as shown in Figure 3.

For solving the problems of the technologies described above, high ferro two provided by the present invention is and the cooperative optimization method of car body end cross shock absorber damping, it is characterized in that adopting following design procedure:

(1) set up high ferro car load 17 degree of freedom and travel the transverse vibration differential equation:

According to the quality m of the single-unit car body of high ferro ₃, moment of inertia of shaking the head sidewinder moment of inertia J _{3 θ}; The quality m of every platform bogie frame ₂, moment of inertia of shaking the head sidewinder moment of inertia J _{2 θ}; Each takes turns right quality m ₁, moment of inertia of shaking the head the heavy W of each wheel shaft; Each takes turns right horizontal creep coefficient f ₁, longitudinal creep coefficient f ₂; Each takes turns right longitudinal register stiffness K _1x, located lateral stiffness K _1y; Every platform bogie one-sided is the vertical equivalent stiffness K of suspension _1z, vertical equivalent damping C _d1; The longitudinal rigidity K of every platform bogie central spring _2x, located lateral stiffness K _2y; Every platform bogie two is the vertical equivalent stiffness K of suspension _2z, vertical equivalent damping C _d2; The torsional rigidity K of single anti-side rolling torsion rod _θ; The ratio of damping C of a pair anti-hunting damper holder _s; The Equivalent damping coefficient C of car body end longitudinal shock absorber ₃; Every platform bogie to be designed two is the Equivalent damping coefficient C of lateral damper ₂, the Equivalent damping coefficient C of car body end cross vibration damper to be designed _r; Vehicle wheel roll radius r, wheel tread gradient λ; Vehicle Speed v; The half b of wheel and rail contact point horizontal spacing, the half b of the horizontal installing space of wheel shaft retainer spring ₁, the half b of the horizontal installing space of bogie central spring ₂, the half b of the horizontal installing space of anti-hunting damper holder ₃, the half b of the horizontal installing space of car body longitudinal shock absorber ₄, the half a of length between truck centers, the half a of wheel-base bogie ₀, the half l of the longitudinal installing space of cross-car vibration damper, axle centerline is to the height h of orbit plane ₀, the height h of plane on car body barycenter to central spring ₁, car body barycenter is the height h of lateral damper to two ₂, on central spring, plane is to the height h of framework barycenter ₃, bogie frame barycenter is to the height h of axle centerline ₄, two is the height h of lateral damper to framework barycenter ₅, car body end cross vibration damper is to the height h of car body barycenter ₆; The barycenter O that respectively, steering framing wheel is right _1ff, O _1fr, trailing bogie takes turns right barycenter O _1rf, O _1rr, the barycenter O of forward and backward bogie frame _2f, O _2rand the barycenter O of car body ₃for true origin; With the yaw displacement y that forecarriage front-wheel is right _1ff, displacement of shaking the head the yaw displacement y that forecarriage trailing wheel is right _1fr, displacement of shaking the head the yaw displacement y that trailing bogie front-wheel is right _1rf, displacement of shaking the head the yaw displacement y that trailing bogie trailing wheel is right _1rr, displacement of shaking the head the yaw displacement y of forecarriage framework _2f, displacement of shaking the head sidewinder displacement θ _2f, the yaw displacement y of trailing bogie framework _2r, displacement of shaking the head sidewinder displacement θ _2r, and the yaw displacement y of car body ₃, displacement of shaking the head sidewinder displacement θ ₃for coordinate; Y is inputted with the orbital direction irregularity at the forward and backward wheel of forecarriage and the forward and backward wheel place of trailing bogie _a1(t), y _a2(t), y _a3(t), y _a4(t) and horizontal irregularity input z _{θ 1}(t), z _{θ 2}(t), z _{θ 3}(t), z _{θ 4}t () is input stimulus, wherein, t is time variable; Set up high ferro car load 17 degree of freedom and travel the transverse vibration differential equation, that is:

1. the yaw vibration equation that forecarriage front-wheel is right:

2. the yawing equation that forecarriage front-wheel is right:

3. the yaw vibration equation that forecarriage trailing wheel is right:

4. the yawing equation that forecarriage trailing wheel is right:

5. the yaw vibration equation that trailing bogie front-wheel is right:

6. the yawing equation that trailing bogie front-wheel is right:

7. the yaw vibration equation that trailing bogie trailing wheel is right:

8. the yawing equation that trailing bogie trailing wheel is right:

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=h ₀+ h ₁+ h ₃+ h ₄;

the yawing equation of car body:

(2) high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model is built:

Travel the transverse vibration differential equation according to high ferro car load 17 degree of freedom set up in step (1), utilize Matlab/Simulink simulation software, build high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model;

(3) setting up high ferro two is and the damping cooperate optimization objective function J of car body end cross vibration damper:

According to the high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model set up in step (2), be that the Equivalent damping coefficient of the Equivalent damping coefficient of lateral damper and car body end cross vibration damper is for design variable with every platform bogie two, to take turns the orbital direction irregularity stochastic inputs at place and horizontal irregularity stochastic inputs as input stimulus with each, utilize the vibration frequency root mean square of weighed acceleration of car body weaving emulating and obtain car body sidewinders the vibration frequency root mean square of weighed acceleration of motion and the vibration frequency root mean square of weighed acceleration of car body yaw motion setting up high ferro two is and the damping cooperate optimization objective function J of car body end cross vibration damper, that is:

In formula, vibration frequency root mean square of weighed acceleration coefficient 1,0.63,0.2, be respectively car body weaving, sidewinder motion, the axle weighting coefficient of yaw motion; Wherein, vibration frequency root mean square of weighed acceleration at different frequencies frequency weight values, be respectively:

$w_d\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,2\]Hz\\ 2/f_i& f_i\∈(2,80\]Hz\end{array};$

$w_e\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,1\]Hz\\ 1/f_i& f_i\∈(1,80\]Hz\end{array};$

$w_e\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,1\]Hz\\ 1/f_i& f_i\∈(1,80\]Hz\end{array};$

(4) high ferro two is lateral damper optimal damping constant C _o2and car body end cross vibration damper optimal damping constant C _oroptimal design:

1. according to the half a of length between truck centers, the half a of wheel-base bogie ₀, Vehicle Speed v, and the high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model set up in step (2), with each orbital direction irregularity stochastic inputs y taken turns place _a1(t), y _a2(t), y _a3(t), y _a4(t) and horizontal irregularity stochastic inputs z _{θ 1}(t), z _{θ 2}(t), z _{θ 3}(t), z _{θ 4}t () is input stimulus, utilize optimized algorithm to ask in step (3) to set up high ferro two be and the minimum value of damping cooperate optimization objective function J of car body end cross vibration damper that corresponding design variable is the best equivalence ratio of damping C that every platform bogie two is lateral damper ₂with the best equivalence ratio of damping C of car body end cross vibration damper _r;

Wherein, the pass between orbital direction irregularity stochastic inputs is: pass between horizontal irregularity stochastic inputs is:

$z_{\mathrm{\θ}3}\left(t\right)=z_{\mathrm{\θ}1}(t-\frac{2a}{v}),z_{\mathrm{\θ}4}\left(t\right)=z_{\mathrm{\θ}1}(t-\frac{2a+2a_0}{v});$

2. be the installation number n of lateral damper according to every platform bogie two ₁, the installation number n of car body end cross vibration damper ₂, and every platform bogie two that in step (4), 1. optimization order design obtains is the best equivalence ratio of damping C of lateral damper ₂with the best equivalence ratio of damping C of car body end cross vibration damper _r, calculate the optimal damping constant that single Zhi Gaotie bis-is lateral damper and car body end cross vibration damper, be respectively: C _o2=C ₂/ n ₁, C _or=C _r/ n ₂.

The advantage that the present invention has than prior art:

Because high ferro belongs to Mdof Vibration System, carrying out dynamic analysis to it calculates very difficult, for high ferro two be and the design of car body end cross shock absorber damping both at home and abroad at present, never provide the theoretical design method of system, mostly be that lateral damper and car body end cross vibration damper are individually studied to two, and computer technology, utilize Dynamics Simulation soft sim PACK or ADAMS/Rail, optimize respectively by solid modelling and determine its size, although the method can obtain reliable simulation numerical, vehicle is made to have good power performance, but, because two be lateral damper and car body end cross vibration damper is a complication system intercoupled, the method that this individually modeling at present designs its shock absorber damping, be difficult to make high ferro two be and the ratio of damping of car body end cross vibration damper reaches optimum matching, and improving constantly along with high ferro travel speed, people to two are and the design of car body end cross shock absorber damping is had higher requirement, current high ferro two is and the method for car body end cross shock absorber damping design can not provide the innovation theory with directive significance, the development to absorber designing requirement in rail vehicle constantly speed-raising situation can not be met.

The present invention travels the transverse vibration differential equation by setting up high ferro car load 17 degree of freedom, utilize MATLAB/Simulink simulation software, construct high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model, and with orbital direction irregularity and horizontal irregularity for input stimulus, minimum for design object with the vibration root mean square of weighed acceleration of cross-car motion, optimal design obtains the optimal damping constant that high ferro two is lateral damper and car body end cross vibration damper.By design example and SIMPACK simulating, verifying known, the method can obtain two being the damping coefficient of lateral damper and car body end cross vibration damper accurately and reliably, is and the design of car body end cross shock absorber damping provides reliable method for designing for high ferro two.Utilize the method, not only can improve design level and the product quality of high ferro suspension system, improve vehicle safety and stationarity; Meanwhile, also can reduce product design and testing expenses, shorten the product design cycle, strengthen the competitiveness in the international market of China's rail vehicle.

Accompanying drawing explanation

Be described further below in conjunction with accompanying drawing to understand the present invention better.

To be high ferro two be Fig. 1 and the design flow diagram of car body end cross shock absorber damping cooperative optimization method;

Fig. 2 is the left view that high ferro car load 17 degree of freedom travels transverse vibration model;

Fig. 3 is the vertical view that high ferro car load 17 degree of freedom travels transverse vibration model;

Fig. 4 is the high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model of embodiment;

Fig. 5 is the German orbital direction irregularity random input stimuli y that embodiment applies _a1(t);

Fig. 6 is the German orbital direction irregularity random input stimuli y that embodiment applies _a2(t);

Fig. 7 is the German orbital direction irregularity random input stimuli y that embodiment applies _a3(t);

Fig. 8 is the German orbital direction irregularity random input stimuli y that embodiment applies _a4(t);

Fig. 9 is the German track horizontal irregularity random input stimuli z that embodiment applies _{θ 1}(t);

Figure 10 is the German track horizontal irregularity random input stimuli z that embodiment applies _{θ 2}(t);

Figure 11 is the German track horizontal irregularity random input stimuli z that embodiment applies _{θ 3}(t);

Figure 12 is the German track horizontal irregularity random input stimuli z that embodiment applies _{θ 4}(t).

Specific embodiments

Below by an embodiment, the present invention is described in further detail.

It is lateral damper that every platform bogie of certain high ferro is provided with two two, is provided with four car body end longitudinal shock absorbers and a car body end cross vibration damper, i.e. n between two adjacent car bodies ₁=2, n ₂=1; The quality m of its single-unit car body ₃=63966kg, moment of inertia of shaking the head sidewinder moment of inertia J _{3 θ}=77200kg.m ²; The quality m of every platform bogie frame ₂=2758kg, moment of inertia of shaking the head sidewinder moment of inertia J _{2 θ}=2212kg.m ²; Each takes turns right quality m ₁=1721kg, moment of inertia of shaking the head the heavy W=150000N of each wheel shaft; Each takes turns right horizontal creep coefficient f ₁=16990000N, longitudinal creep coefficient f ₂=16990000N; Each takes turns right longitudinal register stiffness K _1x=13.739 × 10 ⁶n/m, located lateral stiffness K _1y=4.892 × 10 ⁶n/m; Every platform bogie one-sided is the vertical equivalent stiffness K of suspension _1z=2.74 × 10 ⁶n/m, vertical equivalent damping C _d1=28.3kN.s/m; The longitudinal rigidity K of every platform bogie central spring _2x=0.18 × 10 ⁶n/m, located lateral stiffness K _2y=0.18 × 10 ⁶n/m; Every platform bogie two is the vertical equivalent stiffness K of suspension _2z=1.1368 × 10 ⁶n/m, vertical equivalent damping C _d2=118.7kN.s/m; The torsional rigidity K of single anti-side rolling torsion rod _θ=2.5 × 10 ⁶n.m/rad; The ratio of damping C of a pair anti-hunting damper holder _s=1027kN.s/m; The Equivalent damping coefficient C of car body end longitudinal shock absorber ₃=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, the half b of the horizontal installing space of wheel shaft retainer spring ₁=1.15m, the half b of the horizontal installing space of bogie central spring ₂=1.3m, the half b of the horizontal installing space of anti-hunting damper holder ₃=1.4m, the half b of the horizontal installing space of car body longitudinal shock absorber ₄=1.2m, the half a=9.5m of length between truck centers, the half a of wheel-base bogie ₀=1.35m, the half l=13.3m of the longitudinal installing space of cross-car vibration damper, axle centerline is to the height h of orbit plane ₀=0.347m, the height h of plane on car body barycenter to central spring ₁=0.8m, car body barycenter is the height h of lateral damper to two ₂=0.616m, on central spring, plane is to the height h of framework barycenter ₃=0.416m, bogie frame barycenter is to the height h of axle centerline ₄=0.137m, two is the height h of lateral damper to framework barycenter ₅=0.6m, car body end cross vibration damper is to the height h of car body barycenter ₆=0.5m; Every platform bogie to be designed two is the Equivalent damping coefficient of lateral damper is C ₂, the Equivalent damping coefficient of car body end cross vibration damper to be designed is C _r.This high ferro two be and car body end cross shock absorber damping design required by Vehicle Speed v=300km/h, be that the ratio of damping of lateral damper and the ratio of damping of car body end cross vibration damper design to this high ferro two.

The high ferro two that example of the present invention provides is and the cooperative optimization method of car body end cross shock absorber damping, its design flow diagram as shown in Figure 1, high ferro car load 17 degree of freedom travels the left view of transverse vibration model as Fig. 2, high ferro car load 17 degree of freedom travels the vertical view of transverse vibration model as shown in Figure 3, and concrete steps are as follows:

(1) set up high ferro car load 17 degree of freedom and travel the transverse vibration differential equation:

According to the quality m of the single-unit car body of high ferro ₃=63966kg, moment of inertia of shaking the head sidewinder moment of inertia J _{3 θ}=77200kg.m ²; The quality m of every platform bogie frame ₂=2758kg, moment of inertia of shaking the head sidewinder moment of inertia J _{2 θ}=2212kg.m ²; Each takes turns right quality m ₁=1721kg, moment of inertia of shaking the head the heavy W=150000N of each wheel shaft; Each takes turns right horizontal creep coefficient f ₁=16990000N, longitudinal creep coefficient f ₂=16990000N; Each takes turns right longitudinal register stiffness K _1x=13.739 × 10 ⁶n/m, located lateral stiffness K _1y=4.892 × 10 ⁶n/m; Every platform bogie one-sided is the vertical equivalent stiffness K of suspension _1z=2.74 × 10 ⁶n/m, vertical equivalent damping C _d1=28.3kN.s/m; The longitudinal rigidity K of every platform bogie central spring _2x=0.18 × 10 ⁶n/m, located lateral stiffness K _2y=0.18 × 10 ⁶n/m; Every platform bogie two is the vertical equivalent stiffness K of suspension _2z=1.1368 × 10 ⁶n/m, vertical equivalent damping C _d2=118.7kN.s/m; The torsional rigidity K of single anti-side rolling torsion rod _θ=2.5 × 10 ⁶n.m/rad; The ratio of damping C of a pair anti-hunting damper holder _s=1027kN.s/m; The Equivalent damping coefficient C of car body end longitudinal shock absorber ₃=2897.6kN.s/m; Every platform bogie to be designed two is the Equivalent damping coefficient C of lateral damper ₂, the Equivalent damping coefficient C of car body end cross vibration damper to be designed _r; 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, the half b of the horizontal installing space of wheel shaft retainer spring ₁=1.15m, the half b of the horizontal installing space of bogie central spring ₂=1.3m, the half b of the horizontal installing space of anti-hunting damper holder ₃=1.4m, the half b of the horizontal installing space of car body longitudinal shock absorber ₄=1.2m, the half a=9.5m of length between truck centers, the half a of wheel-base bogie ₀=1.35m, the half l=13.3m of the longitudinal installing space of cross-car vibration damper, axle centerline is to the height h of orbit plane ₀=0.347m, the height h of plane on car body barycenter to central spring ₁=0.8m, car body barycenter is the height h of lateral damper to two ₂=0.616m, on central spring, plane is to the height h of framework barycenter ₃=0.416m, bogie frame barycenter is to the height h of axle centerline ₄=0.137m, two is the height h of lateral damper to framework barycenter ₅=0.6m, car body end cross vibration damper is to the height h of car body barycenter ₆=0.5m; The barycenter O that respectively, steering framing wheel is right _1ff, O _1fr, trailing bogie takes turns right barycenter O _1rf, O _1rr, the barycenter O of forward and backward bogie frame _2f, O _2rand the barycenter O of car body ₃for true origin; With the yaw displacement y that forecarriage front-wheel is right _1ff, displacement of shaking the head the yaw displacement y that forecarriage trailing wheel is right _1fr, displacement of shaking the head the yaw displacement y that trailing bogie front-wheel is right _1rf, displacement of shaking the head the yaw displacement y that trailing bogie trailing wheel is right _1rr, displacement of shaking the head the yaw displacement y of forecarriage framework _2f, displacement of shaking the head sidewinder displacement θ _2f, the yaw displacement y of trailing bogie framework _2r, displacement of shaking the head sidewinder displacement θ _2r, and the yaw displacement y of car body ₃, displacement of shaking the head sidewinder displacement θ ₃for coordinate; Y is inputted with the orbital direction irregularity at the forward and backward wheel of forecarriage and the forward and backward wheel place of trailing bogie _a1(t), y _a2(t), y _a3(t), y _a4(t) and horizontal irregularity input z _{θ 1}(t), z _{θ 2}(t), z _{θ 3}(t), z _{θ 4}t () is input stimulus, wherein, t is time variable; Set up high ferro car load 17 degree of freedom and travel the transverse vibration differential equation, that is:

1. the yaw vibration equation that forecarriage front-wheel is right:

2. the yawing equation that forecarriage front-wheel is right:

3. the yaw vibration equation that forecarriage trailing wheel is right:

4. the yawing equation that forecarriage trailing wheel is right:

5. the yaw vibration equation that trailing bogie front-wheel is right:

6. the yawing equation that trailing bogie front-wheel is right:

7. the yaw vibration equation that trailing bogie trailing wheel is right:

8. the yawing equation that trailing bogie trailing wheel is right:

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=h ₀+ h ₁+ h ₃+ h ₄;

the yawing equation of car body:

(2) high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model is built:

Travel the transverse vibration differential equation according to high ferro car load 17 degree of freedom set up in step (1), utilize Matlab/Simulink simulation software, build high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model, as shown in Figure 4;

(3) setting up high ferro two is and the damping cooperate optimization objective function J of car body end cross vibration damper:

According to the high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model set up in step (2), be that the Equivalent damping coefficient of the Equivalent damping coefficient of lateral damper and car body end cross vibration damper is for design variable with every platform bogie two, to take turns the orbital direction irregularity stochastic inputs at place and horizontal irregularity stochastic inputs as input stimulus with each, utilize the vibration frequency root mean square of weighed acceleration of car body weaving emulating and obtain car body sidewinders the vibration frequency root mean square of weighed acceleration of motion and the vibration frequency root mean square of weighed acceleration of car body yaw motion setting up high ferro two is and the damping cooperate optimization objective function J of car body end cross vibration damper, that is:

In formula, vibration frequency root mean square of weighed acceleration coefficient 1,0.63,0.2, be respectively car body weaving, sidewinder motion, the axle weighting coefficient of yaw motion; Wherein, vibration frequency root mean square of weighed acceleration at different frequencies frequency weight values, be respectively:

$w_d\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,2\]Hz\\ 2/f_i& f_i\∈(2,80\]Hz\end{array};$

$w_e\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,1\]Hz\\ 1/f_i& f_i\∈(1,80\]Hz\end{array};$

$w_e\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,1\]Hz\\ 1/f_i& f_i\∈(1,80\]Hz\end{array};$

(4) high ferro two is lateral damper optimal damping constant C _o2and car body end cross vibration damper optimal damping constant C _oroptimal design:

1. according to the half a=9.5m of length between truck centers, the half a of wheel-base bogie ₀=1.35m, Vehicle Speed v=300km/h, and the high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model set up in step (2), with each orbital direction irregularity stochastic inputs y taken turns place _a1(t), y _a2(t), y _a3(t), y _a4(t) and horizontal irregularity stochastic inputs z _{θ 1}(t), z _{θ 2}(t), z _{θ 3}(t), z _{θ 4}t () is input stimulus, utilize optimized algorithm to ask in step (3) to set up high ferro two be and the minimum value of damping cooperate optimization objective function J of car body end cross vibration damper that optimal design obtains the best equivalence ratio of damping C that every platform bogie two is lateral damper ₂=106.8kN.s/m, the best equivalence ratio of damping C of car body end cross vibration damper _r=176.3kN.s/m;

Wherein, the pass between orbital direction irregularity stochastic inputs is: y _a2(t)=y _a1(t-0.0324s), y _a3(t)=y _a1(t-0.228s), y _a4(t)=y _a1(t-0.2604s); Pass between horizontal irregularity stochastic inputs is: z _{θ 2}(t)=z _{θ 1}(t-0.0324s), z _{θ 3}(t)=z _{θ 1}(t-0.228s), z _{θ 4}(t)=z _{θ 1}(t-0.2604s); During Vehicle Speed v=300km/h, the German orbital direction irregularity random input stimuli that each wheel applies place, respectively as shown in Fig. 5, Fig. 6, Fig. 7, Fig. 8; The horizontal irregularity random input stimuli of the German track applied, respectively as shown in Fig. 9, Figure 10, Figure 11, Figure 12;

2. be the installation number n of lateral damper according to every platform bogie two ₁=2, the installation number n of car body end cross vibration damper ₂=1, and every platform bogie two that in step (4), 1. optimization order design obtains is the best equivalence ratio of damping C of lateral damper ₂the best equivalence ratio of damping C of=106.8kN.s/m and car body end cross vibration damper _r=176.3kN.s/m, calculates the optimal damping constant that single Zhi Gaotie bis-is lateral damper and car body end cross vibration damper, is respectively: C _o2=C ₂/ n ₁=53.4kN.s/m, C _or=C _r/ n ₂=176.3kN.s/m.

According to the vehicle parameter that embodiment provides, utilize rail vehicle special software SIMPACK, can be obtained by solid modelling simulating, verifying, this high ferro two is the optimal damping constant of lateral damper is C _o2=53.7kN.s/m, the optimal damping constant of car body end cross vibration damper is C _or=176.1kN.s/m; Known, the high ferro two utilizing cooperative optimization method to obtain is the optimal damping constant C of lateral damper _o2=53.4kN.s/m, the optimal damping constant C of car body end cross vibration damper _or=176.3kN.s/m, obtain with SIMPACK simulating, verifying two is the optimal damping constant C of lateral damper _o2=53.7kN.s/m, the optimal damping constant C of car body end cross vibration damper _or=176.3kN.s/m matches, both are respectively 0.3kN.s/m, 0.2kN.s/m at deviation, relative deviation is respectively 0.56%, 0.11%, shows that high ferro two provided by the present invention is and the cooperative optimization method of car body end cross shock absorber damping is correct.

Claims (1)

1. high ferro two is and the cooperative optimization method of car body end cross shock absorber damping, and its specific design step is as follows:

(1) set up high ferro car load 17 degree of freedom and travel the transverse vibration differential equation:

According to the quality m of the single-unit car body of high ferro ₃, moment of inertia of shaking the head sidewinder moment of inertia J _{3 θ}; The quality m of every platform bogie frame ₂, moment of inertia of shaking the head sidewinder moment of inertia J _{2 θ}; Each takes turns right quality m ₁, moment of inertia of shaking the head the heavy W of each wheel shaft; Each takes turns right horizontal creep coefficient f ₁, longitudinal creep coefficient f ₂; Each takes turns right longitudinal register stiffness K _1x, located lateral stiffness K _1y; Every platform bogie one-sided is the vertical equivalent stiffness K of suspension _1z, vertical equivalent damping C _d1; The longitudinal rigidity K of every platform bogie central spring _2x, located lateral stiffness K _2y; Every platform bogie two is the vertical equivalent stiffness K of suspension _2z, vertical equivalent damping C _d2; The torsional rigidity K of single anti-side rolling torsion rod _θ; The ratio of damping C of a pair anti-hunting damper holder _s; The Equivalent damping coefficient C of car body end longitudinal shock absorber ₃; Every platform bogie to be designed two is the Equivalent damping coefficient C of lateral damper ₂, the Equivalent damping coefficient C of car body end cross vibration damper to be designed _r; Vehicle wheel roll radius r, wheel tread gradient λ; Vehicle Speed v; The half b of wheel and rail contact point horizontal spacing, the half b of the horizontal installing space of wheel shaft retainer spring ₁, the half b of the horizontal installing space of bogie central spring ₂, the half b of the horizontal installing space of anti-hunting damper holder ₃, the half b of the horizontal installing space of car body longitudinal shock absorber ₄, the half a of length between truck centers, the half a of wheel-base bogie ₀, the half l of the longitudinal installing space of cross-car vibration damper, axle centerline is to the height h of orbit plane ₀, the height h of plane on car body barycenter to central spring ₁, car body barycenter is the height h of lateral damper to two ₂, on central spring, plane is to the height h of framework barycenter ₃, bogie frame barycenter is to the height h of axle centerline ₄, two is the height h of lateral damper to framework barycenter ₅, car body end cross vibration damper is to the height h of car body barycenter ₆; The barycenter O that respectively, steering framing wheel is right _1ff, O _1fr, trailing bogie takes turns right barycenter O _1rf, O _1rr, the barycenter O of forward and backward bogie frame _2f, O _2rand the barycenter O of car body ₃for true origin; With the yaw displacement y that forecarriage front-wheel is right _1ff, displacement of shaking the head the yaw displacement y that forecarriage trailing wheel is right _1fr, displacement of shaking the head the yaw displacement y that trailing bogie front-wheel is right _1rf, displacement of shaking the head the yaw displacement y that trailing bogie trailing wheel is right _1rr, displacement of shaking the head the yaw displacement y of forecarriage framework _2f, displacement of shaking the head sidewinder displacement θ _2f, the yaw displacement y of trailing bogie framework _2r, displacement of shaking the head sidewinder displacement θ _2r, and the yaw displacement y of car body ₃, displacement of shaking the head sidewinder displacement θ ₃for coordinate; Y is inputted with the orbital direction irregularity at the forward and backward wheel of forecarriage and the forward and backward wheel place of trailing bogie _a1(t), y _a2(t), y _a3(t), y _a4(t) and horizontal irregularity input z _{θ 1}(t), z _{θ 2}(t), z _{θ 3}(t), z _{θ 4}t () is input stimulus, wherein, t is time variable; Set up high ferro car load 17 degree of freedom and travel the transverse vibration differential equation, that is:

1. the yaw vibration equation that forecarriage front-wheel is right:

2. the yawing equation that forecarriage front-wheel is right:

3. the yaw vibration equation that forecarriage trailing wheel is right:

4. the yawing equation that forecarriage trailing wheel is right:

5. the yaw vibration equation that trailing bogie front-wheel is right:

6. the yawing equation that trailing bogie front-wheel is right:

7. the yaw vibration equation that trailing bogie trailing wheel is right:

8. the yawing equation that trailing bogie trailing wheel is right:

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=h ₀+ h ₁+ h ₃+ h ₄;

the yawing equation of car body:

(2) high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model is built:

Travel the transverse vibration differential equation according to high ferro car load 17 degree of freedom set up in step (1), utilize Matlab/Simulink simulation software, build high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model;

(3) setting up high ferro two is and the damping cooperate optimization objective function J of car body end cross vibration damper:

According to the high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model set up in step (2), be that the Equivalent damping coefficient of the Equivalent damping coefficient of lateral damper and car body end cross vibration damper is for design variable with every platform bogie two, to take turns the orbital direction irregularity stochastic inputs at place and horizontal irregularity stochastic inputs as input stimulus with each, utilize the vibration frequency root mean square of weighed acceleration of car body weaving emulating and obtain car body sidewinders the vibration frequency root mean square of weighed acceleration of motion and the vibration frequency root mean square of weighed acceleration of car body yaw motion setting up high ferro two is and the damping cooperate optimization objective function J of car body end cross vibration damper, that is:

In formula, vibration frequency root mean square of weighed acceleration coefficient 1,0.63,0.2, be respectively car body weaving, sidewinder motion, the axle weighting coefficient of yaw motion; Wherein, vibration frequency root mean square of weighed acceleration at different frequencies frequency weight values, be respectively:

$\begin{array}{c}{w}_d\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,2\]Hz\\ 2/f_i& f_i\∈(2,80\]Hz\end{array};\\ w_e\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,1\]Hz\\ 1/f_i& f_i\∈(1,80\]Hz\end{array};\\ w_e\left(f_i\right)=\{\begin{array}{cc}1& f_i\∈\[0.5,1\]Hz\\ 1/f_i& f_i\∈(1,80\]Hz\end{array};\end{array}$

(4) high ferro two is lateral damper optimal damping constant C _o2and car body end cross vibration damper optimal damping constant C _oroptimal design:

1. according to the half a of length between truck centers, the half a of wheel-base bogie ₀, Vehicle Speed v, and the high ferro car load 17 degree of freedom transverse vibration cooperate optimization realistic model set up in step (2), with each orbital direction irregularity stochastic inputs y taken turns place _a1(t), y _a2(t), y _a3(t), y _a4(t) and horizontal irregularity stochastic inputs z _{θ 1}(t), z _{θ 2}(t), z _{θ 3}(t), z _{θ 4}t () is input stimulus, utilize optimized algorithm to ask in step (3) to set up high ferro two be and the minimum value of damping cooperate optimization objective function J of car body end cross vibration damper that corresponding design variable is the best equivalence ratio of damping C that every platform bogie two is lateral damper ₂with the best equivalence ratio of damping C of car body end cross vibration damper _r;

Wherein, the pass between orbital direction irregularity stochastic inputs is: $y_{a2}\left(t\right)=y_{a1}(t-\frac{2a_0}{v}),y_{a3}\left(t\right)=y_{a1}(t-\frac{2a}{v}),$ $y_{a4}\left(t\right)=y_{a1}(t-\frac{2a+2a_0}{v});$ Pass between horizontal irregularity stochastic inputs is: $z_{\mathrm{\θ}2}\left(t\right)=z_{\mathrm{\θ}1}(t-\frac{2a_0}{v}),$ $z_{\mathrm{\θ}3}\left(t\right)=z_{\mathrm{\θ}1}(t-\frac{2a}{v}),z_{\mathrm{\θ}4}\left(t\right)=z_{\mathrm{\θ}1}(t-\frac{2a+2a_0}{v});$

2. be the installation number n of lateral damper according to every platform bogie two ₁, the installation number n of car body end cross vibration damper ₂, and every platform bogie two that in step (4), 1. optimization order design obtains is the best equivalence ratio of damping C of lateral damper ₂with the best equivalence ratio of damping C of car body end cross vibration damper _r, calculate the optimal damping constant that single Zhi Gaotie bis-is lateral damper and car body end cross vibration damper, be respectively: C _o2=C ₂/ n ₁, C _or=C _r/ n ₂.