CN116522597B - Wheel hub driving automobile wheel load estimation method - Google Patents
Wheel hub driving automobile wheel load estimation method Download PDFInfo
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Abstract
The invention discloses a method for estimating the wheel load of a hub-driven automobile, which comprises the following steps: analyzing vehicle load transfer mechanisms under different working conditions, and forming a wheel load calculation unified model; taking the mass center position in the vehicle load transfer stress analysis process as dynamic change, and forming a mass center position estimation model based on analysis of a mass center position dynamic change rule; fusing the wheel load calculation unified model and the centroid position estimation model to form a wheel load calculation optimization model fusing a dynamic centroid, and solving four-wheel load by using the wheel load calculation optimization model; the invention improves the estimation precision of the wheel load and saves the estimation cost.
Description
Technical Field
The invention relates to the technical field of control of steering stability and running safety of a hub-driven new energy automobile, in particular to a hub-driven automobile wheel load estimation method.
Background
The hub-driven automobile becomes a hot spot in industry research by virtue of outstanding driving energy conservation and flexible dynamics control. The introduction of the in-wheel motor increases the ratio of unsprung mass of the vehicle in the whole vehicle, and requires more advanced driving stability control to improve the driving safety of the vehicle. When the hub drives the automobile to run under the working conditions of acceleration, deceleration and steering, especially under the combined working condition of acceleration (deceleration) and steering, the whole wheel load can be transferred between the front axle, the rear axle, the left wheel and the right wheel due to the change of longitudinal acceleration and lateral acceleration of the automobile, so that the wheel load is redistributed on four wheels. The difference of the distribution of the wheel load can cause the difference of the ground adhesion force born by each wheel, and the upper limit value of the actual driving force of the wheel is influenced, so that the control efficiency of the driving moment of the wheel is further influenced. Meanwhile, the cornering stiffness of the wheels is related to the wheel load, and the cornering stiffness of the tires can be changed by load transfer, so that the running safety of the vehicle is affected. Therefore, the wheel load has an important influence on the driving stability control of the hub-driven automobile.
The real-time acquisition of the wheel load can be divided into two technical routes of direct estimation and indirect estimation. The direct estimation is to embed a sensor in the tire to monitor the dynamic information of the tire running, so as to obtain the tire state parameters, thereby providing a way of directly measuring the tire force [1][2] The core of the intelligent tire system is a tire internal acquisition device (comprising a sensor, a signal transmission device and the like). The method has high requirements on the precision and stability of the sensor, and meanwhile, the signals are generally transmitted to the vehicle-mounted terminal in a wireless transmission mode, and the signals are easily interfered by multiple physical field environments such as hub electricity, magnetism, a network and the like. Because of the limitations of the above-described technical factors, the tire load estimated by this method is currently less directly applied to vehicle dynamics control. The indirect estimation is derived through a vehicle dynamics model, and utilizes the characteristic parameters of the vehicle and collects the information such as the motion state parameters of the vehicle to realize the estimation of the tire wheel load [3][4] . The method is mainly based on a vehicle dynamics model, does not need to solve a complex calculus equation and does not need to introduce a new tire load test system, so the method has the advantages of strong estimation interpretation, high estimation efficiency and no additional system cost.
In conclusion, although the direct estimation method for acquiring parameters by implanting additional sensors in the wheel has a better estimation effect, the estimation cost is increased; the indirect estimation method is used for estimating by collecting information such as vehicle motion state parameters through characteristic parameters of the vehicle, and the estimation cost is not increased, but the estimation accuracy of the method cannot be guaranteed. In addition, the traditional indirect wheel load estimation usually assumes that the positions of the mass center of the whole vehicle, the mass center of the sprung mass and the like are unchanged, but in recent years, research shows that the position of the mass center of the vehicle can be obviously deviated under the working conditions of steering, braking and accelerating the vehicle, namely the position of the mass center of the vehicle under different working conditions is dynamically changed [5] 。
[1] Wang Yan, lin Bingqin, liang Guanqun, et al, applied research on the vertical force prediction of intelligent tires [ J ]. Rubber industry, 2019,39 (2): 117-121.
[2] Huang Xiaojing, zhang Feng, zhang Shiwen, etc. vertical load measurement of intelligent tires [ J ]. Automobile engineering, 2020,42 (9): 1272-1283.
[3] Zhang Lipeng, qi Bingnan simulation calculation of vertical load of tyre under automobile turning braking condition [ J ]. Agricultural machinery research, 2006 (1): 104-106.
[4]KIM S W,YANG W J,JIN S K,et al.Vehicular vertical tire forces estimation using unscented kalman filter[C]//2019 12th Asian Control Conference(ASCC),2019:325-330.
[5] Sun Yongfeng, yang Qing, wu Shuo, etc. methods for calculating dynamic centroids of vehicles, methods and systems for calculating yaw moments, china, CN201110080679.6[ P ],2012-10-17.
Disclosure of Invention
In order to solve the problems in the prior art, the invention aims to provide a wheel load estimation method of a wheel hub driven automobile, which improves the wheel load estimation precision and saves the estimation cost.
In order to achieve the above purpose, the invention adopts the following technical scheme: a method for estimating the wheel load of a hub-driven automobile comprises the following steps:
step 1, analyzing vehicle load transfer mechanisms under different working conditions, and forming a wheel load calculation unified model;
step 2, regarding the mass center position in the vehicle load transfer stress analysis process as dynamic change, and forming a mass center position estimation model based on analysis of a mass center position dynamic change rule;
and 3, fusing the wheel load calculation unified model and the centroid position estimation model to form a wheel load calculation optimization model fusing the dynamic centroids, and solving the four-wheel load by using the wheel load calculation optimization model.
As a further improvement of the present invention, in step 1, a vehicle load transfer mechanism is analyzed from three aspects of a stress condition under quasi-static state of the vehicle, a stress condition under acceleration and deceleration running, and a stress condition under steering running, and a wheel load estimation unified model is obtained therefrom; in the analysis process, influence of air resistance and tire deformation on vehicle movement is ignored; assuming that the vertical load acts on the wheel center plane, the wheel tread and the wheel base are fixed values, and the vehicle mass distribution is bilateral symmetry; regardless of the roll and pitch motions of the unsprung mass.
As a further improvement of the present invention, the step 1 specifically includes the steps of:
step 1.1, analyzing a load transfer mechanism under a quasi-static driving condition of a vehicle to obtain a four-wheel vertical load under the condition:
wherein F is zfl,nom 、F zfr,nom 、F zrl,nom 、F zrr,nom The four-wheel vertical load under the working condition of static or uniform straight running of the vehicle is respectively represented by fl, fr, rl, rr, wherein the four-wheel vertical load is respectively represented by front left, front right, rear left and rear right wheels, B is the wheel base, L is the wheel base, a and B are the distances from the mass center of the whole vehicle to the front and rear axles, c and d are the distances from the mass center of the whole vehicle to the left and right wheels, m is the mass of the whole vehicle, and g is the gravity acceleration;
step 1.2, analyzing a load transfer mechanism under the acceleration and deceleration running state of the vehicle to obtain a load transfer amount under the state:
ΔF zi,acc ＝F si,acc +F di,acc +F Tiθ
wherein DeltaF zi,acc Is the sum of the vertical load variation of the wheels caused by the longitudinal acceleration, F si,acc 、F di,acc Additional inertial forces, F, generated by longitudinal acceleration of the sprung and unsprung masses, respectively Tiθ An additional inertial force generated for the sprung mass pitching motion; m is m s 、m d The sprung and unsprung masses, h, respectively s 、h d The sprung mass and unsprung mass centroid heights, respectively, θ being the pitch angle, k θ For pitch stiffness, a xs 、a xd Longitudinal accelerations of the sprung mass and the unsprung mass, respectively, phi being the roll angle;
step 1.3, analyzing a load transfer mechanism in a steering running state of the vehicle to obtain a load transfer amount in the state:
ΔF zi,roll ＝F si,roll +F di,roll +F Tiφ
wherein DeltaF zi,roll Is the sum of the changes of the vertical load of the wheels caused by the lateral acceleration, F si,roll 、F di,roll Inertial forces generated by acceleration movements of the sprung and unsprung masses, respectively, F Tiφ An additional inertial force generated for the sprung mass roll motion; phi is the side-tipping angle, k φ For the rigidity of the side inclination angle of the suspension, including the rigidity k of the side inclination angle of the front suspension φf And rear suspension roll stiffness k φr ，a ys 、a yd Lateral acceleration of the sprung and unsprung masses, respectively;
step 1.4, based on analysis of the vehicle load transfer mechanism under different states, dynamic vertical load F of the wheels zi And (3) carrying out comprehensive expression by using the following formula to obtain a wheel load calculation unified model:
wherein a is x And a y The vehicle longitudinal acceleration and the lateral acceleration, respectively.
As a further improvement of the present invention, the step 2 is specifically as follows:
when the combination of the longitudinal acceleration and the lateral acceleration exists, the longitudinal variation and the transverse variation of the vehicle centroid coordinates relative to the vehicle when the vehicle runs at a constant speed are respectively delta x and delta y, and the moment balance is obtained:
ma x h+mgΔx＝0
ma y h+mgΔy＝0
simplifying a centroid position estimation model to quantify the dynamic change of the centroid in the longitudinal direction and the transverse direction:
as a further improvement of the present invention, the step 3 is specifically as follows:
based on centroid position estimation model, at F zi,nom In consideration of changes in longitudinal and lateral acceleration induced a, b, c, d, F zi,nom Will no longer be a fixed value but will change with centroid position changes; at the same time, deltaF zi,acc And DeltaF zi,roll The change of the vertical load of the wheel caused by longitudinal and lateral acceleration is not considered any more, but only the change of the vertical load of the wheel caused by the additional inertia moment generated by the pitching and rolling movement of the sprung mass is considered; obtaining a wheel load estimation optimization model after fusing the dynamic centroids:
the beneficial effects of the invention are as follows:
the invention provides a hub-driven automobile wheel load estimation method considering the dynamic change of a mass center on the basis of the existing indirect estimation method mechanism; in terms of precision, the new estimation method controls the calculation error to be within 10% under all working conditions, has higher precision and fewer required parameters, and meets the dynamic control requirement of the hub-driven automobile; in terms of cost, the new estimation method is based on the concept of indirect estimation of vehicle dynamics parameters, does not increase a hardware system required by estimation independently, has the advantages of high stability, low time ductility and strong practicability, and is lower in calculation cost than a direct estimation method.
Drawings
FIG. 1 is a schematic diagram of a vehicle stress analysis in an embodiment of the present invention, where (a) is a stress condition under quasi-static conditions, (b) is a stress condition under acceleration and deceleration running, and (c) is a stress condition under steering running;
FIG. 2 is a schematic diagram of a dynamic centroid in an embodiment of the invention;
FIG. 3 is a schematic diagram illustrating a simulation test in an embodiment of the present invention;
fig. 4 is a schematic view of the vertical wheel load under 45 ° steering braking conditions in an embodiment of the present invention.
Detailed Description
Embodiments of the present invention will be described in detail below with reference to the accompanying drawings.
Examples
The accurate estimation of the wheel load has a non-negligible impact on vehicle ride, steering stability and safety. In order to improve the wheel load estimation precision and save the estimation cost, a wheel hub driving automobile wheel load estimation method considering the dynamic change of the mass center is provided. Firstly, vehicle load transfer mechanisms under different working conditions are analyzed, and a unified wheel load calculation model is formed. And then, regarding the mass center position in the vehicle load transfer stress analysis process as dynamic change, and forming a mass center position estimation model based on analysis of a mass center position dynamic change rule. Finally, the wheel load calculation unified model and the mass center position estimation model are fused to form a wheel load calculation optimization model of a fusion dynamic mass center, so that a novel method for indirectly estimating the vertical load of the wheel of the hub-driven automobile by considering the dynamic change of the mass center of the vehicle is provided; the method specifically comprises the following steps:
1. vehicle wheel load estimation unified model:
load transfer during vehicle movement is the primary cause of wheel load variation. When the vehicle runs under the working conditions of acceleration, deceleration and steering, especially under the combined working conditions of acceleration and deceleration, the whole wheel load can be transferred between the front axle, the rear axle, the left wheel and the right wheel due to the change of longitudinal acceleration and lateral acceleration of the vehicle, and the wheel load is redistributed on four wheels. In general, the load transfer mechanism of a vehicle is discussed from three aspects of stress analysis under quasi-static state of the vehicle and stress analysis under steering running, and a unified model of wheel load estimation is obtained. A schematic diagram of the stress analysis in each running state is shown in fig. 1. In the stress analysis, the influence of air resistance and tire deformation on the movement of the vehicle is ignored; assuming that the vertical load acts on the wheel center plane, the wheel tread and the wheel base are fixed values, and the vehicle mass distribution is bilateral symmetry; regardless of the roll and pitch motions of the unsprung mass.
In FIG. 1, C is the mass center of the whole vehicle, m is the mass of the whole vehicle, a and b are the distances from the mass center of the whole vehicle to the front and rear axles, h is the height of the mass center of the whole vehicle, L is the wheelbase, F zf And F zr The vertical force applied by the front wheel and the rear wheel respectively, g is the gravitational acceleration, a x For vehicle longitudinal acceleration, m s 、m d Sprung and unsprung masses, respectively, C s 、C d The mass center of mass of the sprung mass and the unsprung mass, h s 、h d The sprung mass and the unsprung mass are respectively the mass center height, theta is the pitch angle, a xs 、a xd Longitudinal acceleration of sprung and unsprung masses, respectively, phi being the roll angle, k φ For the camber stiffness of the suspension (the camber stiffness k of the sub-front suspension) φf And rear suspension roll stiffness k φr )，a ys 、a yd The sprung and unsprung masses are laterally accelerated, respectively.
Quasi-stationary driving of a vehicle means that the vehicle is in a stationary or constant speed straight driving state. With reference to fig. 1 (a), the load transfer mechanism under the quasi-static driving condition of the vehicle is analyzed to obtain the vertical load of the four wheels under the condition:
wherein F is zfl,nom 、F zfr,nom 、F zrl,nom 、F zrr,nom The vertical loads of four wheels under the working condition of static or uniform straight running of the vehicle are respectively represented by fl, fr, rl, rr, wherein the vertical loads of the four wheels are respectively represented by front left, front right, rear left and rear right wheels, B is the wheel track, and c and d are the distances from the mass center of the whole vehicle to the left and right wheels.
Load transfer during acceleration and deceleration of the vehicle is mainly caused by additional inertial forces generated by longitudinal acceleration movements of the sprung and unsprung masses and additional inertial forces (moments) generated by pitching movements of the sprung mass during acceleration and braking of the vehicle. Referring to fig. 1 (b), the load transfer mechanism in the acceleration/deceleration running state of the vehicle is analyzed to obtain the load transfer amount in this state:
ΔF zi,acc ＝F si,acc +F di,acc +F Tiθ
wherein DeltaF zi,acc Is the sum of the vertical load variation of the wheels caused by the longitudinal acceleration, F si,acc 、F di,acc Additional inertial forces, F, generated by longitudinal acceleration of the sprung and unsprung masses, respectively Tiθ Additional inertial force is generated for the sprung mass pitching motion.
Load transfer when the vehicle is turning is mainly caused by the additional inertial force generated by the lateral acceleration motion of the sprung and unsprung masses and the additional inertial force (moment) generated by the roll motion of the sprung mass. Referring to fig. 1 (c), the load transfer mechanism in the steering travel state of the vehicle is analyzed to obtain the load transfer amount in this state:
ΔF zi,roll ＝F si,roll +F di,roll +F Tiφ
wherein DeltaF zi,roll Wheel drop for lateral accelerationTo the sum of the load variation amounts, F si,roll 、F di,roll Inertial forces generated by acceleration movements of the sprung and unsprung masses, respectively, F Tiφ Additional inertial force is generated for the sprung mass roll motion.
Based on analysis of the vehicle load transfer mechanism under different conditions, the dynamic vertical load F of the wheels zi The integrated expression can be performed by using the formula (10) to obtain a unified model of wheel load estimation:
2. centroid position estimation model:
the traditional indirect estimation of the vertical load of the wheel usually assumes that the positions of the mass center of the whole vehicle, the mass center of the sprung mass and the like are unchanged, but in recent years, research shows that the position of the mass center of the vehicle can be obviously deviated under the working conditions of steering, braking and accelerating of the vehicle, namely, the position of the mass center of the vehicle is dynamically changed under different working conditions. The sprung mass centroid position change is completely consistent with the whole vehicle centroid position change rule through analysis, and the sprung mass centroid position change directly leads to the whole vehicle centroid position change; the vertical relative distance between the center of mass position of the whole vehicle and the center of mass position of the sprung mass is almost unchanged, and the fact that the height of the center of mass of the whole vehicle is unchanged in a vehicle coordinate system under the full working condition can be understood; in the longitudinal and transverse planes, the relative distance of the mass center gradually increases along with the occurrence of steering and compound working conditions; the longitudinal and transverse variation of the mass center of the whole vehicle cannot be always regarded as a constant value in the actual running process of the vehicle, and the influence of the dynamic variation on the vertical load of the wheels is considered. Therefore, when the vertical load of the wheel is estimated indirectly, the position of the center of mass of the whole vehicle in the vertical direction is almost unchanged and can be ignored, and the longitudinal and transverse changes are not negligible.
And constructing a dynamic change schematic diagram of the mass center of the whole vehicle as shown in fig. 2, and reasonably quantifying the change of the mass center of the whole vehicle in the longitudinal direction and the transverse direction. When the vehicle runs at a constant speed, the longitudinal acceleration and the lateral acceleration are both zero, and the center of mass of the whole vehicle is positioned at a coordinate origin S0 in a plane; when longitudinal acceleration exists, the distance from the mass center to the front and rear axes dynamically changes, the mass center coordinate moves to the position S1, and the longitudinal change amount is deltax; when the transverse acceleration exists, the distance from the mass center to the left and right sides is dynamically changed, the mass center coordinate is moved to the S2, and the transverse change amount is delta y; when there is a combination of longitudinal and lateral acceleration, the centroid coordinates will move to S3.
From the moment balance, it is possible to:
ma x h+mgΔx＝0
ma y h+mgΔy＝0
simplifying a centroid position estimation model to quantify the dynamic change of the centroid in the longitudinal direction and the transverse direction:
3. wheel load estimation optimization model:
based on centroid position estimation model, at F zi,nom In consideration of changes in longitudinal and lateral acceleration induced a, b, c, d, F zi,nom Will no longer be a fixed value but will change with centroid position changes. At the same time, deltaF zi,acc And DeltaF zi,roll The wheel vertical load change due to longitudinal and lateral acceleration is not considered, but only the wheel vertical load change due to the additional inertial force (moment) generated by the sprung mass pitching and rolling motion. Obtaining a wheel load estimation optimization model after fusing the dynamic centroids:
the method comprises the steps of obtaining unsprung mass distribution data on a vehicle, vehicle wheelbase and wheelbase, vehicle roll angle rigidity and pitch angle rigidity, and ground clearance of a vehicle mass center, distance between front and rear axles, distance between left and right wheels and the like under static state based on basic parameters of the whole vehicle; acquiring vehicle longitudinal acceleration a in real time through acceleration sensor x And lateral acceleration a y Data; acquiring a vehicle body pitch angle and a roll angle of the whole vehicle posture by using a gyroscope and the like; solving the four-wheel load by using a wheel load estimation model by combining the basic parameters of the whole vehicle and the real-time parameters acquired based on the sensors; vehicle state identification and vehicle performance control related research can be carried out based on the wheel load estimated value.
To verify the effectiveness of the proposed method, a simulation test was performed by designing the operating conditions as shown in fig. 3: 0s-5s, and the vehicle runs at a constant speed of 5 km/h; 5s-18s, accelerating the vehicle from 5km/h to 50km/h;18s-20s, steering wheel turns left 45 degrees, and vehicle speed is kept 50km/h;20s-25s, maintaining a steering wheel angle of 45 degrees and a vehicle speed of 50km/h; the steering wheel angle is kept at 45 degrees for 25s-30s, and the vehicle speed is reduced from 50km/h to 5km/h. The conventional estimation method without fusion of dynamic centroid variation is compared with the new wheel load estimation method with fusion of dynamic centroid, and the vertical load estimation comparison result shown in fig. 4 and table 1 is obtained.
Table 1 wheel load calculation error contrast under joint simulation conditions
From the comparison results, it can be seen that: under the above simulation working conditions, in the estimation results of the traditional estimation method without fusion of dynamic mass centers, the maximum estimation error of part of the vertical loads of the wheels can be ensured to be within 10%, the engineering usable range is met, but as the complexity of the working conditions is improved, the wheel load estimation error is increased, and the maximum wheel load estimation error of part of the wheels reaches about 23%, so that the wheel load estimation is invalid. The new wheel load estimation method integrating the dynamic mass center has good four-wheel vertical load estimation effect under all working conditions, the maximum estimation error is not more than 9%, and the dynamic control requirement of the hub-driven automobile is met.
The foregoing examples merely illustrate specific embodiments of the invention, which are described in greater detail and are not to be construed as limiting the scope of the invention. It should be noted that it will be apparent to those skilled in the art that several variations and modifications can be made without departing from the spirit of the invention, which are all within the scope of the invention.
Claims (1)
1. The method for estimating the wheel load of the hub-driven automobile is characterized by comprising the following steps of:
step 1, analyzing vehicle load transfer mechanisms under different working conditions, and forming a wheel load calculation unified model;
step 2, regarding the mass center position in the vehicle load transfer stress analysis process as dynamic change, and forming a mass center position estimation model based on analysis of a mass center position dynamic change rule;
step 3, fusing the wheel load calculation unified model and the centroid position estimation model to form a wheel load calculation optimization model fusing a dynamic centroid, and solving four-wheel load by using the wheel load calculation optimization model;
in the step 1, analyzing a vehicle load transfer mechanism from three aspects of a stress condition under a quasi-static state of a vehicle, a stress condition under acceleration and deceleration running and a stress condition under steering running, and obtaining a wheel load estimation unified model; in the analysis process, influence of air resistance and tire deformation on vehicle movement is ignored; assuming that the vertical load acts on the wheel center plane, the wheel tread and the wheel base are fixed values, and the vehicle mass distribution is bilateral symmetry; regardless of the roll and pitch motions of the unsprung mass;
the step 1 specifically comprises the following steps:
step 1.1, analyzing a load transfer mechanism under a quasi-static driving condition of a vehicle to obtain a four-wheel vertical load under the condition:
wherein F is zfl,nom 、F zfr,nom 、F zrl,nom 、F zrr,nom The four-wheel vertical load under the working condition of static or uniform straight running of the vehicle is respectively represented by fl, fr, rl, rr, wherein the four-wheel vertical load is respectively represented by front left, front right, rear left and rear right wheels, B is the wheel base, L is the wheel base, a and B are the distances from the mass center of the whole vehicle to the front and rear axles, c and d are the distances from the mass center of the whole vehicle to the left and right wheels, m is the mass of the whole vehicle, and g is the gravity acceleration;
step 1.2, analyzing a load transfer mechanism under the acceleration and deceleration running state of the vehicle to obtain a load transfer amount under the state:
ΔF zi,acc ＝F si,acc +F di,acc +F Tiθ
wherein DeltaF zi,acc Is the sum of the vertical load variation of the wheels caused by the longitudinal acceleration, F si,acc 、F di,acc Additional inertial forces, F, generated by longitudinal acceleration of the sprung and unsprung masses, respectively Tiθ An additional inertial force generated for the sprung mass pitching motion; m is m s 、m d The sprung and unsprung masses, h, respectively s 、h d The sprung mass and unsprung mass centroid heights, respectively, θ being the pitch angle, k θ For pitch stiffness, a xs 、a xd Longitudinal accelerations of the sprung mass and the unsprung mass, respectively, phi being the roll angle;
step 1.3, analyzing a load transfer mechanism in a steering running state of the vehicle to obtain a load transfer amount in the state:
ΔF zi,roll ＝F si,roll +F di,roll +F Tiφ
wherein DeltaF zi,roll Is the sum of the changes of the vertical load of the wheels caused by the lateral acceleration, F si,roll 、F di,roll Inertial forces generated by acceleration movements of the sprung and unsprung masses, respectively, F Tiφ An additional inertial force generated for the sprung mass roll motion; phi is the side-tipping angle, k φ For the rigidity of the side inclination angle of the suspension, including the rigidity k of the side inclination angle of the front suspension φf And rear suspension roll stiffness k φr ，a ys 、a yd Lateral acceleration of the sprung and unsprung masses, respectively;
step 1.4, based on analysis of the vehicle load transfer mechanism under different states, dynamic vertical load F of the wheels zi And (3) carrying out comprehensive expression by using the following formula to obtain a wheel load calculation unified model:
wherein a is x And a y Longitudinal acceleration and lateral acceleration of the vehicle, respectively;
the step 2 specifically comprises the following steps:
when the combination of the longitudinal acceleration and the lateral acceleration exists, the longitudinal variation and the transverse variation of the vehicle centroid coordinates relative to the vehicle when the vehicle runs at a constant speed are respectively delta x and delta y, and the moment balance is obtained:
ma x h+mgΔx＝0
ma y h+mgΔy＝0
simplifying a centroid position estimation model to quantify the dynamic change of the centroid in the longitudinal direction and the transverse direction:
the step 3 specifically comprises the following steps:
based on centroid position estimation model, at F zi,nom In consideration of changes in longitudinal and lateral acceleration induced a, b, c, d, F zi,nom Will no longer be a fixed value but will change with centroid position changes; at the same time, deltaF zi,acc And DeltaF zi,roll The change of the vertical load of the wheel caused by longitudinal and lateral acceleration is not considered any more, but only the change of the vertical load of the wheel caused by the additional inertia moment generated by the pitching and rolling movement of the sprung mass is considered; obtaining a wheel load estimation optimization model after fusing the dynamic centroids:
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| CN108556850A (en) * | 2018-04-18 | 2018-09-21 | 吉林大学 | A kind of multi-axle heavy type wheel of vehicle vertical load general calculation method |
| CN113449378A (en) * | 2021-06-23 | 2021-09-28 | 中国人民解放***箭军工程大学 | Transverse load transfer rate calculation method based on vehicle suspension deformation |
| KR20220072952A (en) * | 2020-11-25 | 2022-06-03 | 한국타이어앤테크놀로지 주식회사 | Method for estimating a vehicle's center of gravity and wheel load, and system for vehicle's center of gravity and wheel load estimation by performing the same |
| CN114889445A (en) * | 2022-05-09 | 2022-08-12 | 武汉理工大学 | Vehicle driving force control method and system based on working condition identification and storage medium |
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| Publication number | Priority date | Publication date | Assignee | Title |
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| CN108556850A (en) * | 2018-04-18 | 2018-09-21 | 吉林大学 | A kind of multi-axle heavy type wheel of vehicle vertical load general calculation method |
| KR20220072952A (en) * | 2020-11-25 | 2022-06-03 | 한국타이어앤테크놀로지 주식회사 | Method for estimating a vehicle's center of gravity and wheel load, and system for vehicle's center of gravity and wheel load estimation by performing the same |
| CN113449378A (en) * | 2021-06-23 | 2021-09-28 | 中国人民解放***箭军工程大学 | Transverse load transfer rate calculation method based on vehicle suspension deformation |
| CN114889445A (en) * | 2022-05-09 | 2022-08-12 | 武汉理工大学 | Vehicle driving force control method and system based on working condition identification and storage medium |
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