According to geometric similarity requirements, the geometric parameters of the vibration model mass are determined. The ratio of length to width is 21. According to the requirements of the test bench and the power requirements of the power plant, the mass of the vibration model mass is determined to be 100 kg. If the vibration model mass is set Length is l, width is b, height is h1, and spring stiffness is k. According to the requirements of similarity theorem, if the model is similar to prototype dynamics, it must satisfy the model that the natural frequency or period of vibration is equal to the prototype. According to the gravity requirements of the model, It is lbh1=100kg where the density of the model material is 7800kgm-3. The natural period of vibration of the model is: Tz=2m2k, and its value is determined between 0.51.0s. The geometric dimensions of the model are as follows: l=0.8m, b=0.4m, h1=0.06m, m=100kg, k=4531Nm-1 , Tz = 0.66s.
The inherent period of vertical vibration is within the prototype vibration cycle and meets the requirements. Design Principle Because the cam can obtain any desired continuous or discontinuous movement during movement, it can simulate the vehicle's driving speed, sinusoidal road surface with different wavelengths and elevations, and obtain the excitation that meets the test requirements. Therefore, it can be used as an excitation input. The most typical road surface profile can be summarized as the superposition of various amplitudes and the harmonic path type at the circular frequency, as shown by the dirt roads tracked or tracked by wheeled vehicles. That is, y=iYisin2Lix+i(4) where: y is the unevenness of the ground, mm; Yi is the maximum elevation of each harmonic, mm; L is the wavelength of the ground contour, mm; x is the travel distance relative to the origin vehicle, Mm;i is the phase angle of each harmonic.
Because the tracked vehicle's ground during actual driving is relatively complex, only the most typical non-planar ground is selected as the input of the interference. Assume that the road surface frequency and elevation passed by the vehicle are constant values, and ignore the influence of the track. The obtained excitation (road) function is f=hsin2Lx(5). According to the relative relativity of the movement, when the road wheel function f=hsin2Lx, the weighted wheel with radius R is moving at a constant speed, the motion trajectory at the axial center is f= hsin2Lx, and the outer edge of the contour line = R-hsin (R is the radius of the base circle in the theoretical contour line) when the cam is moving at the same speed in the horizontal plane, the axial trajectory is also f=hsin2Lx.
The damping coefficient is determined when the excitation force is less than the force required to cause relative sliding of the damping force, ie, over damping, the damper is in a locked state, causing it to become a rigid connector; and when the damper provides a damping force relative to If the excitation force is too small, the effect of vibration reduction will not be achieved. Therefore, determining the required damping force range is a key to designing the shock absorber of the test bench. Of course, in the case of overdamping and critical damping, the system has no vibration, and the optimal value of the damping coefficient of the test bench under under-damped conditions is mainly discussed.
Due to the presence of damping, the transmission rate T continuously varies with the frequency ratio, and regardless of the magnitude of the damping ratio, the respective transmission rate curves at the different damping ratios shown in FIG. When the damping ratio increases after 2, the value of transmissibility is increased. In this case, the principle of geometric similarity and dynamic similarity is discussed in this paper. The design principle of the vibration test rig is discussed and the parameters are calculated. The input to its stimulus was analyzed and calculated from the principle to the example. Through the dynamic analysis, the damping coefficient was determined.
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