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CN112733265A - Design calculation and optimization method for electric vehicle power assembly suspension system - Google Patents

Design calculation and optimization method for electric vehicle power assembly suspension system Download PDF

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CN112733265A
CN112733265A CN202011640091.7A CN202011640091A CN112733265A CN 112733265 A CN112733265 A CN 112733265A CN 202011640091 A CN202011640091 A CN 202011640091A CN 112733265 A CN112733265 A CN 112733265A
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康英姿
肖兵
上官文斌
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South China University of Technology SCUT
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Abstract

The invention discloses a design calculation and optimization method of an electric automobile power assembly suspension system, which comprises the following steps: firstly, establishing a power assembly mass center coordinate system, and acquiring parameters such as an inertia parameter, a mass center position, a mounting position of a rubber suspension, a static stiffness curve and the like of the power assembly on the basis; further establishing a dynamic model of the power assembly suspension system, and deriving a motion differential equation; establishing a target function according to an energy decoupling theory and transient response characteristics; the method is characterized by comprising the steps of selecting the purposes that the suspension system is highest in energy decoupling rate, the power assembly mass center longitudinal acceleration under transient response and the impact amplitude are minimum, taking the suspension linear section rigidity as a design variable, reasonably distributing natural frequency and changing range of the suspension rigidity as constraint conditions, optimizing by adopting a multi-island genetic algorithm, and finally verifying the feasibility of the method through an example.

Description

Design calculation and optimization method for electric vehicle power assembly suspension system
Technical Field
The invention belongs to the field of optimization design of an automobile power assembly suspension system, and particularly relates to a design calculation and optimization method of an electric automobile power assembly suspension system.
Background
The power assembly serves as a main excitation source of the automobile, and vibration generated by the power assembly is transmitted to the auxiliary frame through the suspension system and then transmitted to the automobile body, so that vibration of the automobile body is caused. The suspension system with good performance can not only reduce the vibration transmitted to the frame by the power assembly and improve the riding comfort, but also can better prolong the service life of the power assembly and other parts, so the design of the suspension system becomes more important. Different from the traditional fuel vehicle, the electric vehicle driving motor has the advantages of quick output response and large output torque. The output torque of the driving motor can rise to hundreds of Nm within a few milliseconds usually, and the driving motor has a relatively obvious impact characteristic, which puts new requirements on the design of an electric automobile suspension system.
In the current research on the suspension system of the power assembly, reasonable distribution of natural frequency of each order and maximum energy decoupling rate are generally optimized. The vibration isolation performance of the system under a specific working condition is neglected to a certain extent by purely considering the inherent characteristics of the power assembly suspension system. Particularly in the suspension system of the electric automobile, the excitation of the electric drive assembly has a relatively obvious impact characteristic, and if only the inherent characteristics of the suspension system of the power assembly are considered, the influence of the transient impact on the system cannot be evaluated. At present, no patent of invention relating to the aspect of the optimization design of the suspension system of the electric automobile under the impact working condition exists.
In a chinese granted patent "a design optimization method and optimization device of a powertrain suspension system" (CN102609551B), "the constructed powertrain suspension system model is a six-degree-of-freedom model, and the influence of a subframe widely used in an existing automobile on system vibration is not considered, and only the intrinsic characteristics of the powertrain suspension system are analyzed and optimized, that is, only static optimization is performed, and transient operating conditions frequently occurring in an electric automobile are not analyzed, so that the defect of such treatment is that the intrinsic characteristics may reach the optimization target, but the problem of poor vibration still exists in the actual operating conditions.
Disclosure of Invention
In order to solve the problems in the prior art, the invention provides a design calculation and optimization method of an electric automobile power assembly suspension system, which introduces two indexes of shock degree and longitudinal acceleration while considering the natural frequency and the energy decoupling rate of the power assembly suspension system, fully considers the natural characteristic and the transient response of the system, takes the reasonable distribution of the natural frequency, the highest energy decoupling rate and the optimal transient response index as optimization targets, and provides guidance for the design research and development of the electric automobile power assembly suspension system.
The object of the invention is achieved by at least one of the following solutions.
A design calculation and optimization method for an electric automobile power assembly suspension system comprises the following steps:
(1) establishing a reference coordinate system and acquiring inertia parameters of the power assembly and the auxiliary frame;
(2) acquiring the mounting position and the mounting angle of the suspension;
(3) establishing an undamped motion differential equation of the power assembly suspension system;
(4) establishing a parameterized dynamic model of the powertrain suspension system, calculating the mass center longitudinal acceleration and the impact degree of the powertrain by taking a step function as excitation, and taking the suspension rigidity and the installation position as model input parameters, wherein the mass center longitudinal acceleration (A)x) And the degree of impact (J) as output of the model;
(5) the method comprises the steps of establishing an optimization model by taking the linear section rigidity of a suspension as a design variable, the natural frequency of a power assembly as a constraint condition, the maximum energy decoupling rate of the power assembly and the minimum mass center longitudinal acceleration and impact amplitude of the power assembly under an impact working condition as optimization targets, and obtaining the design rigidity of the suspension according to the optimization model.
Further, the establishing of the reference coordinate system and the obtaining of the inertia parameters of the powertrain in the step (1) specifically includes: obtaining the mass m of the power assembly, establishing a fixed coordinate system O-XYZ by taking the centroid position O of the power assembly as a coordinate origin, wherein the Y axis is parallel to the output axis of the motor, the forward direction points to the free end of the output shaft of the motor, the forward direction of the Z axis is vertical upward, and the X axis is determined by a right-hand rule; obtaining the rotational inertia I of the power assembly to the X axis, the Y axis and the Z axis respectivelyxx、Iyy、IzzObtaining the inertia product I of the power assembly to the X axis and the Y axisxyProduct of inertia for Y-axis and Z-axis IyzAnd product of inertia I to X-axis and Z-axisxz。
Further, the mounting position in step (2) is defined as the intersection point of the elastic main axes of the suspension, and the mounting position of the ith suspension can be expressed as [ a ]xi ayi azi]。
Further, the undamped motion differential equation established in the step (3) is as follows:
Figure BDA0002878205530000021
wherein the mass matrix M is:
Figure BDA0002878205530000022
Mw＝diag([m m m])，
Figure BDA0002878205530000023
the stiffness matrix K is:
Figure BDA0002878205530000024
wherein:
Kgi＝RKsiRT
Figure BDA0002878205530000031
the matrix R is a cosine matrix between the local coordinate system of the suspension and the coordinate system of the center of mass of the power assembly, and the value of the matrix R is determined by the installation angle of the suspension;
in the differential equation of motion, q ═ wx wy wz θx θy θz]TIs a power assembly mass center displacement vector,
Figure BDA0002878205530000032
is an acceleration vector.
Further, in the step (4), when the parameterized dynamic model of the powertrain suspension system is established, rigid body models are adopted for both the electric drive assembly and the auxiliary frame.
Further, the stiffness of the suspension in the step (4) refers to the static stiffness of the elastic main shaft under the local coordinate system of the suspension: k is a radical ofui、kvi、kwiThe static stiffness of the suspension in the three elastic main shaft directions under the local coordinate system of the suspension, namely the static stiffness in the three directions of u, v and w, is respectively, the local coordinate system of the suspension is a coordinate system established by taking the elastic center of the suspension as the origin of coordinates and taking the three elastic main shafts as coordinate axes, and the static stiffness matrix form of the ith suspension can be expressed as follows: ksi＝diag([kui kvikwi])。
Further, the calculation method of the impact degree in the step (4) is as follows:
Figure BDA0002878205530000033
wherein, aβThe angular acceleration of the powertrain about the Y-axis.
Further, in the step (5), establishing an objective function with the maximum energy decoupling rate of the powertrain as follows:
in the center of mass coordinate of the power assembly according to the mass matrix M and the rigidity matrixK is solved to obtain six-order natural frequency omega of power assembly suspension systemj(j 1, 2.... 6) and the mode shape matrix Φ, then the energy distribution of the system in each order of primary vibration can be obtained, and when the system is in the j (j 1, 2.. 6) th order of primary vibration, the maximum kinetic energy is:
Figure BDA0002878205530000034
the expansion is as follows:
Figure BDA0002878205530000035
then, the energy ratio of the k (k ═ 1, 2.., 6) th generalized coordinate at the j-th order primary vibration is:
Figure BDA0002878205530000041
energy distribution of the power assembly in six main vibration directions can be obtained through the formula;
if ejMax E (k, j), k 1,2,3.. 6, then EjFor the decoupling rate of the power assembly in a certain vibration direction, the optimization target is as follows: by varying the stiffness of the suspension so that ejAs close to 1 as possible if the constraint is satisfied.
Further, because the influence degree of vibration in each direction on the NVH performance of the whole vehicle is different, if the decoupling degree around the Y axis and in the vertical direction is the highest, a six-dimensional weight vector b ═ b is introduced for the purpose1 b2 b3 b4 b5 b6]To indicate the importance of the energy decoupling rate in each direction.
Further, in the step (5), the maximum energy decoupling rate of the power assembly and the minimum mass center longitudinal acceleration and impact amplitude of the power assembly under the impact working condition are set as optimization targets, and the established optimization model is as follows:
Figure BDA0002878205530000042
min f2＝max(Ax)
min f3＝max(J)
s.t.
kuil≤kui≤kuiu；
kvil≤kvi≤kviu；
kwil≤kwi≤kwiu；
ωj-ωj+1≤-1,(j＝1,2,...,6)；
5-ω1≤0；
ω6-40≤0；
90-e5≤0.
compared with the prior art, the invention at least has the following technical effects:
1) the design calculation and optimization method fully considers the inherent characteristics and transient response characteristics of the suspension system, takes the characteristics as an optimization target, establishes an optimization model for optimization, and gives consideration to both static and dynamic optimization effects;
2) the method simultaneously considers the static characteristic and the dynamic characteristic of the power assembly suspension system, can ensure that the natural frequency of the power assembly is reasonably distributed and has higher decoupling rate, and can also ensure that the power assembly has better vibration isolation effect under the transient impact working condition.
3) The method can predict the vibration characteristic and the optimization space of the suspension system in the initial stage of the design of the suspension system, can quickly and efficiently obtain an optimization scheme by modifying the relevant parameters of the suspension system, shortens the early-stage research and development period of the suspension system, reduces the research and development cost of enterprises, and has important engineering significance for ensuring the overall performance of the suspension system.
Drawings
FIG. 1 is a flow chart of a design calculation method for an electric vehicle powertrain suspension system.
FIG. 2 is a powertrain suspension system dynamics model.
Fig. 3 is a graph of the excitation torque applied in the simulation, which acts on the center of mass of the powertrain, the torque direction being positive about the Y-axis of the reference frame.
FIG. 4 is a comparison of transient responses for longitudinal acceleration of the powertrain centroid before and after optimization.
FIG. 5 is a comparison of transient responses for optimizing front and rear powertrain jerk.
Detailed Description
The invention is described in further detail below by way of an example with reference to the accompanying drawings.
The design calculation and optimization method for the suspension system of the power assembly of the electric automobile provided by the embodiment comprises the following steps:
(1) a center of mass coordinate system as shown in the dynamic model of the powertrain suspension system of fig. 2 is established. The center of mass position of the power assembly is a coordinate origin O, a fixed coordinate system O-XYZ is established by taking the O as the origin, the Y axis is parallel to the axis of the motor output shaft, the forward direction points to the free end of the motor output shaft, the forward direction of the Z axis is vertical upward, and the X axis is determined by the right-hand rule. Utilize three-wire pendulum inertial parameter measurement test bench to measure power assembly and respectively to X axle, Y axle, inertia I of Z axlexx、Iyy、IzzObtaining the inertia product I of the power assembly to the X axis and the Y axisxyProduct of inertia for Y-axis and Z-axis IyzAnd product of inertia I to X-axis and Z-axisxz. And simultaneously measuring the mass m of the power assembly. And obtaining inertia parameters of the subframe for subsequent dynamic simulation.
(2) And acquiring the installation position and the installation angle of the suspension. The mounting position of the suspension is defined as the intersection point of the elastic main shafts of the suspension, and the mounting position of the ith suspension is expressed as [ a ]xi ayi azi]。axi、ayi、aziRespectively, representing the coordinates of the mounting position of the suspension in the reference frame, i.e. the position of the elastic centre point of the suspension in the reference frame.
TABLE 1 Angle between the suspended elastic spindle and the reference coordinate axis
Figure BDA0002878205530000051
Figure BDA0002878205530000061
In the table, u, v and w are three elastic main shafts for suspension; alpha is alphaui、αvi、αwiRespectively representing the included angles beta between the u axis, the v axis and the w axis of the ith suspension and the X axis of the reference coordinate systemui、βvi、βwiRespectively representing the included angles between the u, v and w axes of the ith suspension and the Y axis of the reference coordinate system, gammaui、γvi、γwiRespectively representing the included angles between the u, v and w axes of the ith suspension and the Z axis of the reference coordinate system.
(3) According to the reference coordinate system established in the step (1) and the related parameters obtained in the step (2), the undamped motion differential equation of the power assembly suspension system can be derived as follows:
Figure BDA0002878205530000062
wherein the quality matrix is:
Figure BDA0002878205530000063
Mw＝diag([m m m])，
Figure BDA0002878205530000064
the stiffness matrix is:
Figure BDA0002878205530000065
in the above formula, the first and second carbon atoms are,
Kgi＝RKsiRT
Figure BDA0002878205530000066
the matrix R is a cosine matrix between the suspension local coordinate system and the powertrain centroid coordinate system, and can be obtained from the data in table 3.
F (t) represents the acting force of the power assembly and is a vector of 6 multiplied by 1; ksiRepresenting the stiffness of the ith suspension in its local coordinate system; kgiRepresenting a stiffness matrix of the ith suspension in a reference coordinate system, namely a stiffness matrix in a center-of-mass coordinate system of the powertrain; kiRepresents the equivalent stiffness of the ith suspension at the center of mass of the powertrain from KsiTo KiThe transformation process of (2) is a process of enabling the rigidity of the suspension to be equivalent to the center of mass of the power assembly from a local coordinate system of the suspension; b isiIs a transformation matrix determined by the mounting position of the suspension; n represents the number of suspensions.
In the differential equation of motion, q ═ wx wy wz θx θy θz]TIs a power assembly mass center displacement vector,
Figure BDA0002878205530000071
is an acceleration vector. w is ax、wy、wzRespectively representing the translational displacement of the center of mass of the power assembly in the x direction, the y direction and the z direction; thetax、θy、θzRespectively represent the rotational displacement of the center of mass of the power assembly in the x direction, the y direction and the z direction.
(4) Establishing a parameterized dynamic model (namely a simulation model) of the powertrain suspension system in multi-body dynamic software to facilitate subsequent optimization, increasing optimization efficiency, calculating the mass center longitudinal acceleration and the impact degree of the powertrain by taking a step function as excitation, and taking the suspension rigidity and the mounting position as model input parameters, wherein the mass center longitudinal acceleration (A) is a mass center longitudinal accelerationx) And the impact (J) as the output of the model, while taking these two parameters as one of the optimization objectives. The step function of this embodiment is a torque curve as shown in fig. 3. The multibody dynamics software used in this example was ADAMS, and it is understood that other ADAMS may be used in other examplesMultiple dynamics software of (1).
The rigidity of the suspension refers to the static rigidity of the elastic main shaft under a local coordinate system of the suspension: k is a radical ofui、kvi、kwiThe static stiffness of the suspension in the three elastic main shaft directions under the local coordinate system of the suspension, namely the static stiffness in the three directions of u, v and w, is respectively, the local coordinate system of the suspension is a coordinate system established by taking the elastic center of the suspension as the origin of coordinates and taking the three elastic main shafts as coordinate axes, and the static stiffness matrix form of the ith suspension can be expressed as follows: ksi＝diag([kui kvi kwi])。
In the step, when a parameterized dynamic model of the powertrain suspension system is established, rigid body models are adopted for both the electric drive assembly and the auxiliary frame.
In this step, the method for calculating the impact degree is as follows:
Figure BDA0002878205530000072
wherein, aβThe angular acceleration of the powertrain about the Y-axis.
The longitudinal acceleration of the center of mass of the power assembly can be directly acquired through ADAMS.
(5) The method comprises the steps of establishing an optimization model by taking the linear section rigidity of a suspension as a design variable, the natural frequency of a power assembly as a constraint condition, the maximum energy decoupling rate of the power assembly and the minimum mass center longitudinal acceleration and impact amplitude of the power assembly under an impact working condition as optimization targets, and obtaining the design rigidity of the suspension according to the optimization model. The method comprises the following specific steps:
in the center of mass coordinate of the power assembly, the six-order natural frequency omega can be solved according to the mass matrix M and the rigidity matrix Kj(j ═ 1, 2.. 6) and the mode shape matrix Φ, it is further possible to obtain the energy distribution of the system when making primary vibrations of each order. When the system is subjected to j (j ═ 1, 2.., 6) th order primary vibration, the maximum kinetic energy is as follows:
Figure BDA0002878205530000081
φjrepresenting a j-th order mode vector of the power assembly,
Figure BDA0002878205530000082
representing its transposed vector for the purpose of calculating the kinetic energy of the vibration; the expansion is as follows:
Figure BDA0002878205530000083
k is the number of rows in the mass matrix,lis the number of columns, mlkIs the element of the ith row and k columns in the quality matrix.
Then, the energy ratio of the k (k ═ 1, 2.., 6) generalized coordinates at the i-th order primary vibration is:
Figure BDA0002878205530000084
Tlmeaning represented is the l column vector, m, in the vibration energy matrixklThe element representing the kth row and the l column in the mass matrix M has the unit kg.
Energy distribution of the power assembly in six main vibration directions can be obtained through the formula;
if ejmaxE (k, j), k 1,2,3.. 6, then ejFor the decoupling rate of the power assembly in a certain vibration direction, the optimization target is as follows: by varying the stiffness of the suspension so that ejAs close to 1 as possible if the constraint is satisfied.
Because the vibration in each direction has different influence on the NVH performance of the whole vehicle, generally speaking, the decoupling degree in the Y-axis and the vertical direction is expected to be the highest. For this purpose, a six-dimensional weight vector b ═ b is introduced1 b2 b3 b4 b5 b6]To indicate the importance of the energy decoupling rate in each direction.
An optimization model can thus be obtained as follows:
Figure BDA0002878205530000091
min f2＝max(Ax)
min f3＝max(J)
s.t.
kuil≤kui≤kuiu；
kvil≤kvi≤kviu；
kwil≤kwi≤kwiu；
ωj-ωj+1≤-1,(j＝1,2,...,6)；
5-ω1≤0；
ω6-40≤0；
90-e5≤0.
wherein, minf1: the sum of six directions (1-energy decoupling rate) is minimized, namely the total energy decoupling rate is maximized; minf2: minimizing the peak value of the longitudinal acceleration curve of the center of mass of the power assembly; minf3: minimizing the peak value of the impact curve; k is a radical ofuil: represents the lower limit of the change of the u-direction stiffness of the ith suspension; k is a radical ofuiu: denotes the upper limit, k, of the change in the u-direction stiffness of the i-th suspensionvil、kviuRespectively representing the lower limit and the upper limit of the variation of the v-direction stiffness of the ith suspension; k is a radical ofwil、kwiuRespectively representing the lower limit and the upper limit of the variation of the w-direction rigidity of the ith suspension; e.g. of the type5The decoupling rate in the direction of the powertrain Pitch, i.e., the decoupling rate in the Pitch direction, is indicated.
TABLE 2 initial stiffness of each suspension and its variation range
Figure BDA0002878205530000092
The system transient response pairs before and after optimization are shown in fig. 4 and 5, for example. FIG. 4 is a transient response of optimized fore and aft powertrain longitudinal acceleration, and FIG. 5 is a comparison of optimized fore and aft powertrain jerk. Through comparison, the energy decoupling rate of the power assembly suspension system is improved after the design calculation and optimization method is used for optimization; meanwhile, the longitudinal acceleration and the impact degree of the power assembly under impact excitation are both reduced, and the transient response is improved, so that the effectiveness of the design calculation and optimization method is demonstrated.
It should be understood that the above-mentioned embodiments of the present invention are only examples for clearly illustrating the present invention, and are not intended to limit the embodiments of the present invention. Other variations and modifications will be apparent to persons skilled in the art in light of the above description. And are neither required nor exhaustive of all embodiments. Any modification, equivalent replacement, and improvement made within the spirit and principle of the present invention should be included in the protection scope of the claims of the present invention.

Claims (10)

1.一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于，包含以下步骤：1. a design calculation and optimization method of an electric vehicle powertrain mounting system, is characterized in that, comprises the following steps: (1)建立参考坐标系并获取动力总成和副车架的惯性参数；(1) Establish a reference coordinate system and obtain the inertial parameters of the powertrain and subframe; (2)获取悬置的安装位置与安装角度；(2) Obtain the installation position and installation angle of the suspension; (3)建立动力总成悬置系统的无阻尼运动微分方程；(3) Establish the undamped motion differential equation of the powertrain mount system; (4)建立动力总成悬置系统的参数化动力学模型并以阶跃函数为激励，计算动力总成的质心纵向加速度和冲击度，将悬置刚度、安装位置作为模型输入参数，质心纵向加速度(Ax)和冲击度(J)作为模型的输出；(4) Establish a parametric dynamic model of the powertrain mounting system and use the step function as an excitation to calculate the longitudinal acceleration and impact of the powertrain's center of mass, and use the mounting stiffness and installation position as model input parameters. Acceleration (A x ) and shock (J) as output of the model; (5)以悬置的线性段刚度为设计变量，动力总成的固有频率为约束条件，动力总成的能量解耦率最大、冲击工况下动力总成的质心纵向加速度和冲击度幅值最小为优化目标，建立优化模型，根据所述优化模型得到悬置的设计刚度。(5) Taking the stiffness of the linear segment of the mount as the design variable, the natural frequency of the powertrain as the constraint condition, the energy decoupling rate of the powertrain is the largest, the longitudinal acceleration of the center of mass of the powertrain and the magnitude of the impact degree under the impact condition The minimum is the optimization objective, an optimization model is established, and the design stiffness of the suspension is obtained according to the optimization model. 2.根据权利要求1所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于：步骤(1)中所述建立参考坐标系并获取动力总成的惯性参数，具体包括：获得动力总成的质量m，以动力总成的质心位置O为坐标原点建立固定坐标系O-XYZ，Y轴平行于电机输出轴线、正向指向电机输出轴自由端，Z轴正向竖直向上，X轴以右手定则确定；获得动力总成分别对X轴、Y轴、Z轴的转动惯量Ixx、Iyy、Izz，获得动力总成对X轴和Y轴的惯性积Ixy、对Y轴和Z轴的惯性积Iyz以及对X轴和Z轴的惯性积Ixz。2. the design calculation and optimization method of a kind of electric vehicle powertrain mount system according to claim 1, is characterized in that: described in step (1), establish reference coordinate system and obtain the inertial parameter of powertrain, Specifically, it includes: obtaining the mass m of the powertrain, establishing a fixed coordinate system O-XYZ with the center of mass position O of the powertrain as the coordinate origin, the Y axis is parallel to the motor output axis, and the positive direction points to the free end of the motor output shaft, and the Z axis is positive. Vertically upward, the X-axis is determined by the right-hand rule; the moment of inertia I xx , I yy , and I zz of the powertrain for the X-axis, Y-axis, and Z-axis are obtained, and the powertrain for the X-axis and the Y-axis is obtained. The product of inertia I xy , the product of inertia I yz for the Y and Z axes, and the product of inertia I xz for the X and Z axes. 3.根据权利要求1所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于：步骤(2)中所述安装位置定义为悬置各弹性主轴的交点，第i个悬置的安装位置可以表示为[axi ayi azi]。3. The design calculation and optimization method of an electric vehicle powertrain suspension system according to claim 1, wherein the installation position described in step (2) is defined as the intersection of each elastic main shaft of the suspension, and the first The installation positions of the i suspensions can be expressed as [a xi a yi a zi ]. 4.根据权利要求1所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于，步骤(3)所建立的无阻尼运动微分方程为：4. the design calculation and optimization method of a kind of electric vehicle powertrain mount system according to claim 1, is characterized in that, the undamped motion differential equation established by step (3) is:
Figure FDA0002878205520000011
Figure FDA0002878205520000011
 其中质量矩阵M为：where the mass matrix M is:
Figure FDA0002878205520000012
Mw＝diag([m m m])，
Figure FDA0002878205520000013
Figure FDA0002878205520000012
M w =diag([mmmm]),
Figure FDA0002878205520000013
 刚度矩阵K为：The stiffness matrix K is:
Figure FDA0002878205520000021
Figure FDA0002878205520000021
 其中：in: Kgi＝RKsiRT K gi = RK si R T
Figure FDA0002878205520000022
Figure FDA0002878205520000022
 矩阵R为悬置局部坐标系与动力总成质心坐标系之间的余弦矩阵，其值由悬置的安装角度决定；The matrix R is the cosine matrix between the local coordinate system of the mount and the coordinate system of the center of mass of the powertrain, and its value is determined by the mounting angle of the mount; 运动微分方程中，q＝[wx wy wz θx θy θz]T为动力总成质心位移向量，
Figure FDA0002878205520000024
为加速度向量。In the differential equation of motion, q=[w x w y w z θ x θ y θ z ] T is the displacement vector of the center of mass of the powertrain,
Figure FDA0002878205520000024
is the acceleration vector. 5.根据权利要求1所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于：步骤(4)中在建立所述动力总成悬置系统的参数化动力学模型时，电驱动总成和副车架均采用刚体模型。5. The design calculation and optimization method of an electric vehicle powertrain mounting system according to claim 1, wherein in step (4), the parametric dynamics of the powertrain mounting system is established in the step (4). When modeling, the electric drive assembly and subframe are rigid body models. 6.根据权利要求1所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于：步骤(4)中所述悬置的刚度是指在悬置局部坐标系下的弹性主轴静刚度：kui、kvi、kwi，分别为悬置在其局部坐标系下三个弹性主轴方向上的静刚度，即分别为u、v、w三个方向上的静刚度，悬置局部坐标系是以悬置的弹性中心为坐标原点，三个弹性主轴为坐标轴建立的坐标系，第i个悬置的静刚度矩阵形式可以表示为：Ksi＝diag([kui kvi kwi])。6. The design calculation and optimization method of an electric vehicle powertrain mount system according to claim 1, wherein: the stiffness of the mount described in step (4) refers to the mount in the local coordinate system of the mount The elastic spindle static stiffness of : k ui , k vi , k wi , which are the static stiffnesses in the three elastic spindle directions suspended in its local coordinate system, namely the static stiffnesses in the three directions of u, v, and w respectively , the suspension local coordinate system is a coordinate system established with the elastic center of the suspension as the coordinate origin, and the three elastic main axes as the coordinate axes. The static stiffness matrix form of the i-th suspension can be expressed as: K si =diag([k ui k vi k wi ]). 7.根据权利要求1所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于：步骤(4)中所述冲击度的计算方法为：7. The design calculation and optimization method of an electric vehicle powertrain suspension system according to claim 1, wherein: the calculation method of the impact degree described in step (4) is:
Figure FDA0002878205520000023
Figure FDA0002878205520000023
 其中，aβ为动力总成绕Y轴转动的角加速度。Among them, a β is the angular acceleration of the powertrain rotating around the Y axis. 8.根据权利要求1所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于，步骤(5)中，建立所述动力总成的能量解耦率最大的目标函数为：8 . The design calculation and optimization method of an electric vehicle powertrain mount system according to claim 1 , wherein in step (5), the goal of establishing the maximum energy decoupling rate of the powertrain is established. 9 . The function is: 在动力总成质心坐标中，根据质量矩阵M和刚度矩阵K求解出动力总成悬置系统的六阶固有频率ωj(j＝1,2,...,6)和振型矩阵Φ，然后可获得系统在作各阶主振动时的能量分布，当系统作第j(j＝1,2,...,6)阶主振动时，其最大动能为：In the coordinates of the center of mass of the powertrain, the sixth-order natural frequency ωj ( j =1,2,...,6) and mode matrix Φ of the powertrain mount system are solved according to the mass matrix M and the stiffness matrix K, Then the energy distribution of the system under each order of main vibration can be obtained. When the system is subjected to the jth (j=1,2,...,6) order main vibration, its maximum kinetic energy is:
Figure FDA0002878205520000031
Figure FDA0002878205520000031
 展开式为：The expansion is:
Figure FDA0002878205520000032
Figure FDA0002878205520000032
 那么，第k(k＝1,2,...,6)个广义坐标在第j阶主振动时能量占比为：Then, the energy ratio of the kth (k=1,2,...,6) generalized coordinate in the jth order main vibration is:
Figure FDA0002878205520000033
Figure FDA0002878205520000033
 通过该公式可以获得动力总成六个主振动方向上的能量分布；The energy distribution in the six main vibration directions of the powertrain can be obtained by this formula; 若ej＝maxE(k,j),k＝1,2,3...6，则ej为动力总成某个振动方向上的解耦率，则优化的目标便是：通过改变悬置的刚度，使得ej在满足约束条件的情况下尽可能接近1。If e j =maxE(k,j),k=1,2,3...6, then e j is the decoupling rate in a certain vibration direction of the powertrain, and the optimization goal is: by changing the suspension Set the stiffness so that e j is as close to 1 as possible while satisfying the constraints. 9.根据权利要求8所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，其特征在于，由于各个方向的振动对整车NVH性能影响程度不同，通常使绕Y轴和垂直方向上的解耦程度最高，为此引入一个六维的权重向量b＝[b1 b2 b3 b4 b5 b6]来表示各个方向上能量解耦率的重要程度。9. the design calculation and optimization method of a kind of electric vehicle powertrain suspension system according to claim 8, it is characterized in that, because the vibration of each direction has different influence degree on the NVH performance of the whole vehicle, usually make around the Y-axis and The degree of decoupling in the vertical direction is the highest, so a six-dimensional weight vector b=[b 1 b 2 b 3 b 4 b 5 b 6 ] is introduced to represent the importance of the energy decoupling rate in each direction. 10.根据权利要求1-9所述的一种电动汽车动力总成悬置系统的设计计算与优化方法，步骤(5)所述动力总成的能量解耦率最大、冲击工况下动力总成的质心纵向加速度和冲击度幅值最小为优化目标中，建立的优化模型如下：10. A design calculation and optimization method for an electric vehicle powertrain mounting system according to claims 1-9, wherein the energy decoupling rate of the powertrain in step (5) is the largest, and the powertrain under impact conditions is the largest. In the optimization objective, the minimum longitudinal acceleration of the center of mass and the magnitude of the impact degree are established, and the established optimization model is as follows:
Figure FDA0002878205520000041
Figure FDA0002878205520000041
 min f2＝max(Ax)min f 2 =max(A x ) min f3＝max(J)min f 3 =max(J) s.t.s.t. kuil≤kui≤kuiu；k uil ≤k ui ≤k uiu ; kvil≤kvi≤kviu；k vil ≤k vi ≤k viu ; kwil≤kwi≤kwiu；k wil ≤k wi ≤k wiu ; ωj-ωj+1≤-1,(j＝1,2,...,6)；ω j -ω j+1 ≤-1,(j=1,2,...,6); 5-ω1≤0；5-ω 1 ≤ 0; ω6-40≤0；ω 6 -40≤0; 90-e5≤0。90-e 5 ≤0.
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Cited By (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN113536477A (en) * 2021-06-30 2021-10-22 英轩重工有限公司 A Parameter Design and Vibration Isolation Performance Analysis and Evaluation Method of a Loader Vibration Isolation Device
CN114357612A (en) * 2021-12-20 2022-04-15 中国第一汽车股份有限公司 Processing method and device for characteristic parameters of suspension and vehicle
CN115982890A (en) * 2023-01-18 2023-04-18 重庆赛力斯新能源汽车设计院有限公司 Design method, device, equipment and medium for automobile range extender suspension system
CN116383972A (en) * 2023-06-05 2023-07-04 江铃汽车股份有限公司 Suspension system rigidity curve structure design method and device
CN116561901A (en) * 2023-07-06 2023-08-08 徐州徐工汽车制造有限公司 Frequency avoidance optimization design method and design system for power assembly suspension system

Citations (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN110701235A (en) * 2019-09-06 2020-01-17 佛山科学技术学院 Design method of semi-active damping tie rod based on vehicle power suspension system

Patent Citations (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN110701235A (en) * 2019-09-06 2020-01-17 佛山科学技术学院 Design method of semi-active damping tie rod based on vehicle power suspension system

Non-Patent Citations (8)

* Cited by examiner, † Cited by third party
Title
席智伦: "某型电动客车悬置系统多目标优化", 《噪声与振动控制》 *
徐中明等: "纯电动汽车动力总成悬置系统的优化", 《汽车工程》 *
王毅等: "发动机总成悬置系统的仿真优化设计", 《轻型汽车技术》 *
王治新: "基于集成仿真的客车动力总成悬置系统优化", 《农业装备与车辆工程》 *
王道勇: "动机启停时动力总成悬置系统的设计方法研究", 《振动与冲击》 *
盛精等: "XMQ6838Y客车动力总成悬置系统参数的优化设计", 《华侨大学学报(自然科学版)》 *
赵士超等: "基于ADAMS的动力总成悬置系统灵敏度分析与优化", 《桂林电子科技大学学报》 *
黄鼎友等: "动力总成悬置系统建模及振动仿真", 《江苏大学学报(自然科学版)》 *

Cited By (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN113536477A (en) * 2021-06-30 2021-10-22 英轩重工有限公司 A Parameter Design and Vibration Isolation Performance Analysis and Evaluation Method of a Loader Vibration Isolation Device
CN114357612A (en) * 2021-12-20 2022-04-15 中国第一汽车股份有限公司 Processing method and device for characteristic parameters of suspension and vehicle
CN115982890A (en) * 2023-01-18 2023-04-18 重庆赛力斯新能源汽车设计院有限公司 Design method, device, equipment and medium for automobile range extender suspension system
WO2024152806A1 (en) * 2023-01-18 2024-07-25 重庆赛力斯新能源汽车设计院有限公司 Automobile range extender suspension system design method and apparatus, device and medium
CN116383972A (en) * 2023-06-05 2023-07-04 江铃汽车股份有限公司 Suspension system rigidity curve structure design method and device
CN116383972B (en) * 2023-06-05 2023-08-11 江铃汽车股份有限公司 Suspension system rigidity curve structure design method and device
CN116561901A (en) * 2023-07-06 2023-08-08 徐州徐工汽车制造有限公司 Frequency avoidance optimization design method and design system for power assembly suspension system
CN116561901B (en) * 2023-07-06 2023-10-13 徐州徐工汽车制造有限公司 Frequency avoidance optimization design method and design system for power assembly suspension system

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