JP7697864B2 - Automotive Test Equipment - Google Patents

Description

The present invention relates to an automobile testing device, and more specifically, to a device capable of controlling the load of a dynamometer by simulating the turning motion caused by steering of an automobile (test vehicle).

For example, chassis dynamos have traditionally been used for evaluation tests of exhaust gases and fuel economy of automobiles such as motorcycles, passenger cars, and trucks. During evaluation tests, a dynamometer applies a load (running resistance) equivalent to that applied when the test vehicle is running on an actual road to a test vehicle placed on rollers in a test room, and, for example, modal exhaust gas amounts (g/km) and fuel economy values (km/L) are calculated (see, for example, Patent Documents 1 and 2). Such chassis dynamos are mechanical devices that are not designed to run test vehicles in turns. For this reason, they cannot meet the recent demand for evaluation tests using mode driving that simulates actual roads, including turns.

Meanwhile, a vehicle testing device has been proposed that removes the tires of the test vehicle, connects a dynamometer to each of the hubs that serve as the wheel connectors of the test vehicle, and applies power without using rollers. Among such vehicle testing devices, a system has been proposed that allows the actual steering angle of the wheels to be changed by operating the steering wheel of the test vehicle (see Patent Document 3). This makes it possible to perform evaluation tests of exhaust gas emissions and fuel efficiency based on the total driving force from the engine or motor of the test vehicle's engine to the hub during driving, including cornering, regenerative energy, and deceleration energy during braking.

However, when the test vehicle is steered while traveling on an actual road, it undergoes various independent movements (see Non-Patent Documents 1 and 2). The balance of the running resistance acting on each drive wheel and each driven wheel of the test vehicle changes accordingly, but the above Patent Document 3 has a problem in that it is not possible to take such a balance of running resistance into consideration. In other words, when performing evaluation tests without predetermined set values such as vehicle speed, such as mode driving simulating an actual road including turns, a front-loading test called front loading in an evaluation test in a test room including turns on a test course, an advanced driver assistance system (ADAS) when it is possible to simulate information represented by the visibility outside the vehicle, and driving during automatic driving (AD), there was no means of applying a set value (load) to the dynamometer in advance for turns.

Japanese Unexamined Patent Publication No. 1-173848 US Patent Publication No. 2011/0303000 Special Publication No. 2020-520457

Masato Abe, "Automobile Dynamics and Control", 2nd Edition, Tokyo Denki University Press, 2008 Yasuo Shimizu, "Advanced Automotive Engineering", First Edition, Tokyo Denki University Press, February 20, 2016

The present invention focuses on the fact that turning is caused by changes in steering angle due to steering wheel operation, and aims to provide an automobile testing device that assigns a set value to a dynamometer and can reproduce the power balance during driving and braking when the test vehicle is turning.

In order to solve the above problems, the present invention provides an automobile test vehicle comprising dynamometers each connected to a wheel coupling of a test vehicle, a control unit for controlling the load of each dynamometer, and a vehicle speed detection means for detecting the vehicle speed from the number of revolutions of the dynamometer, the control unit comprising a running resistance model that models the running resistance acting on the test vehicle in correspondence with preset test vehicle specifications and the vehicle speed of the test vehicle, and a balance control unit that determines the load to be applied to each dynamometer based on the calculation results of the running resistance model corresponding to the vehicle speed detected by the vehicle speed detection means, and each dynamometer is controlled according to the load determined by the balance control unit. The device further comprises a steering angle sensor that detects the steering angle associated with steering of the test vehicle, and a turning motion model that models the turning motion when the test vehicle is steered in correspondence with the specifications and vehicle speed of the test vehicle and the steering angle detected by the steering angle sensor, and is characterized in that the balance control unit is configured to input a balance command value as a calculation result of the turning motion model in addition to the calculation result of the running resistance model, and to determine the loads of each of the dynamometers by taking into account the balance control between the dynamometers based on the balance command value.

As described above, when the test vehicle is turning, the steering angle can be detected and a set value (load) that takes into account the differential rotation speed and differential torque, which are an example of balance control for each drive wheel, can be applied to each dynamometer, making it possible to reproduce the power balance during driving and braking when the test vehicle is turning. In other words, because a set value for the steering angle is automatically applied to the dynamometer, evaluation tests can be carried out even when the set value is unknown in advance.

Here, when the test vehicle is steered to make a turn, the wheels located on the outside of the turning direction rotate faster than the wheels located on the inside. For this reason, as the steering angle increases, a deceleration torque called cornering drag is applied. Therefore, in the present invention, a configuration can be adopted in which the cornering drag is calculated by the turning motion model according to the steering angle detected by the steering angle sensor, and the calculation result is input to the running resistance model. This makes it possible to assign a set value (load) that takes into account the cornering drag, which is the turning drag resistance, to each dynamometer.

In the present invention, the turning motion model can be a model of the equation of motion of the test vehicle during turning, and the balance command value can be given as one of the following: the average number of revolutions following the vehicle speed, the angle difference based on the difference in the distance traveled, and the torque difference. Furthermore, the steering angle sensor may be provided on all steered wheels of the test vehicle, and a hub-connected type or roller type that is immovable in the steering direction can be used as a dynamometer connected to something other than the steered wheels. Furthermore, the handle angle of the test vehicle can be used instead of the steering angle of the steering angle sensor. The subject (test vehicle) that can be tested with the automotive test device of the present invention is any of two-wheeled vehicles, three-wheeled vehicles, four-wheeled vehicles, six-wheeled vehicles, eight-wheeled vehicles, or vehicles with trailers, including some or all of their wheels. In addition, when the test vehicle is a two-wheeled vehicle, the camber angle can be detected to reproduce the cornering drag during turning, and thus the front-rear differential rotation (balance) and cornering drag during turning of the two-wheeled vehicle can be added to the dynamometer.

FIG. 1 is a diagram for explaining the configuration of an automobile testing device according to an embodiment of the present invention. FIG. 11 is a diagram showing an example of a calculation method for generating a balance command in a turning motion model. An example of a method for calculating cornering drag in a turning motion model will be described. FIG. 2 is a diagram showing a model of the lateral slip angle of a steered front tire and the side force, cornering force, and cornering drag. FIG. 2 is a diagram showing an equivalent two-wheeled vehicle model of a four-wheeled vehicle. 5A and 5B are diagrams for explaining the turning radius of a test vehicle and the difference in the left and right wheel rotation speeds as an example of balance control. FIG. 2 is a block diagram of a vehicle sideslip angle model and a yaw moment model. FIG. 4A is a diagram for explaining a vehicle sideslip angle model, FIG. 4B is a diagram for explaining a yaw moment model, and FIG. 4C is a diagram for explaining a turning radius model. FIG. 13 is a diagram for explaining the configuration of an automobile testing device according to a modified embodiment. 13A and 13B are diagrams illustrating the configuration of an automobile testing device for a three-wheeled vehicle according to another modified embodiment. 13A and 13B are diagrams illustrating the configuration of an automobile testing device for a two-wheeled vehicle according to another modified embodiment, in which FIG. FIG. 13 is a diagram for explaining the configuration of a moving vehicle testing device for an automobile having six or more wheels according to an embodiment of another modified example. FIG. 13 is a diagram for explaining the configuration of a four-wheel steering automobile testing device according to another modified embodiment. FIG. 13 is a diagram for explaining the configuration of an automobile testing device according to a further modified embodiment. FIG. 13 is a diagram for explaining the configuration of an automobile testing device according to a further modified embodiment.

Below, with reference to the drawings, an embodiment of the automobile testing device TM of the present invention will be described, using a four-wheel drive, front-wheel steering automobile as the test vehicle Tc. In each drawing, unless otherwise specified, the same symbols and symbols are used for the same members, elements, and formulas, and terms indicating directions such as front, rear, up, and down are based on Figure 1.

Referring to FIG. 1, the automotive testing device TM of this embodiment includes dynamometers Dm1-Dm4 connected to the wheel couplings Wc1-Wc4 of the test vehicle Tc, and a control unit Cu that controls the load of each dynamometer Dm1-Dm4. Although the dynamometers Dm1-Dm4 differ depending on the structure of the vehicle suspension of the test vehicle Tc, known dynamometers that are directly attached to the wheel couplings Wc1-Wc4, which include the brake discs and wheel hubs, can be used, and detailed explanations are omitted here. In this case, the dynamometers Dm1 and Dm2 connected to the wheel couplings Wc1 and Wc2 of the steered front wheels are configured to move in the steering direction in response to steering (operation of the steering wheel of the test vehicle Tc).

The control unit Cu includes a running resistance model 11 that models the running resistance (rolling resistance, air resistance, inertial resistance) acting on the test vehicle Tc in correspondence with the specifications of the test vehicle Tc and the specifications of the tires that can be mounted on the test vehicle Tc (hereinafter referred to as "vehicle specifications Bs") that are set in advance (as calculation parameters) and the vehicle speed V of the test vehicle Tc, a driving and braking direction model 12 that models the load distribution to each dynamometer Dm1 to Dm4 and the driving and braking direction of the test vehicle Tc based on the calculation results of the running resistance model 11, a balance control unit 13 that determines the load to be applied to each dynamometer Dm1 to Dm4 based on the calculation results of the driving and braking direction model 12, and a vehicle speed detection means 14 that detects the vehicle speed of the test vehicle Tc. The driving and braking direction model 12 receives the rotation speed and torque measured by each dynamometer Dm1 to Dm4, and the calculation results are input to the vehicle speed detection means 14 to calculate the vehicle speed. The load determined by the balance control unit 13 is then input as a torque command to the power conversion devices Pc1 to Pc4 attached to each of the dynamometers Dm1 to Dm4, and the load of each of the dynamometers Dm1 to Dm4 is controlled accordingly. Note that since known components can be used for each of the above components, further detailed explanations will be omitted.

In the present embodiment, the vehicle testing device TM is configured to apply loads (set values) to the dynamometers Dm1 to Dm4, focusing on the fact that the turning of the test vehicle Tc is caused by changes in the steering angle due to steering operation, in order to reproduce the power balance during driving and braking when the test vehicle Tc is turning. This will be described in detail below. The dynamometers Dm1 and Dm2, which are respectively connected to the wheel connectors Wc1 and Wc2 of the steered front left and front right wheels, are fitted with known steering angle sensors 2a and 2b that detect the steering angles δ1 and δ2 associated with the steering of the test vehicle Tc. In this case, the mounting positions of the steering angle sensors 2a and 2b are not limited to this as long as they detect the steering angles δ1 and δ2, and the steering angle of the test vehicle Tc can be used instead of the steering angles δ1 and δ2 of the steering angle sensors 2a and 2b.

The control unit Cu also includes a turning motion model 15 that models the turning motion of the test vehicle Tc when it is steered in accordance with the vehicle specifications Bs, vehicle speed V, and steering angles δ1, δ2. The vehicle speed V calculated by the vehicle speed detection means 14 and the steering angles δ1, δ2 detected by the steering angle sensors 2a, 2b are input to the turning motion model 15, and the calculation results of the turning motion model 15 are input to the balance control unit 13 as a balance command. The balance control unit 13 then determines the loads for each of the dynamometers Dm1 to Dm4, taking into account the balance control (differential rotation speed and differential torque) between them based on the balance command.

Next, referring to Figure 2, a calculation method for generating a balance command in the turning motion model 15 will be described using an example modeled based on the equation of motion of the test vehicle Tc during turning. Here, the turning motion model 15 has a first turning model 15a, a second turning model 15b, and a coefficient calculation unit 15c, and the first turning model 15a receives the vehicle speed V, which is a state quantity calculated by the vehicle speed detection means 14, and the average steering angle δ calculated from the steering angle δ1 detected by the steering angle sensor 2a of the front left wheel and the steering angle δ2 detected by the steering angle sensor 2b of the front right wheel. Further, as the vehicle specifications Bs, the wheelbase l, the length from the vehicle center of gravity to the front wheels lf, the length from the vehicle center of gravity to the rear wheels lr, the tread d, the vehicle weight m, the tire radius rt, the cornering power kf of the front wheels, and the cornering power kr of the rear wheels are set in advance, and from these, the coefficient calculation unit 15c calculates and sets a coefficient A, which is a stability factor, in advance. Then, the turning radius R is calculated from the coefficient A, vehicle speed V, wheelbase 1, and average steering angle δ, and the average steering angle δ, turning radius R, vehicle speed V, tread d, and tire radius rt obtained in the turning model 15a are input to the second turning model 15b, and the front left and right wheel rotation speed difference ΔNfrl, the rear left and right wheel rotation speed difference ΔNrrl, and the front and rear wheel rotation speed difference ΔNfr caused by the turning motion of the test vehicle Tc are calculated, and these calculation results (ΔNfrl, ΔNrrl, ΔNfr) are output to the balance control unit 13 as balance commands.

Next, referring to FIG. 3, a method of calculating cornering drag in the turning motion model 15 will be described using an example modeled based on the equation of motion of the test vehicle Tc during turning. The turning motion model 15 also includes a third turning model 15d and a fourth turning model 15e. In the coefficient calculation unit 15c, the vehicle speed V, the average steering angle δ calculated from the steering angle δ1 detected by the steering angle sensor 2a of the front left wheel and the steering angle δ2 detected by the steering angle sensor 2b of the front right wheel, the wheelbase l, the length lf from the center of gravity of the vehicle to the front wheels, the length lr from the center of gravity of the vehicle to the rear wheels, and the coefficients A and B are input to the third turning model 15d, and the vehicle side slip angle β and the vehicle yaw moment γ are calculated, and further the front wheel side slip angle βf and the rear wheel side slip angle βr are calculated. Then, the front wheel side slip angle βf and the rear wheel side slip angle βr are input to the tire side slip characteristic 151 of the fourth turning model 15e, which serves as a cornering drag calculation unit, and the cornering drag Rc as the calculation result is input to the running resistance model 11.

Here, from FIG. 4 showing a model of the steered front wheel side slip angle βf, side force Fs, cornering force Fc, and cornering drag Rc, the side force Fs is vector-decomposed into cornering force Fc and cornering drag Rc. Therefore, the tire side slip characteristic 151 in FIG. 3 is an XY characteristic that outputs the side force Fs and cornering force Fc according to the front wheel side slip angle βf (or rear wheel side slip angle βr). This characteristic function presets the relationship between the front wheel side slip angle βf, side force Fs, and cornering force Fc experimentally obtained in advance as a tire characteristic function, and outputs the side force Ffs and cornering force Ffc by correcting the measurement points according to the input front wheel side slip angle βf. In general commercial vehicles, the same tires are sometimes used to make the wear of the tires even by rotating them. For this reason, it is sufficient to prepare one type of tire characteristic function, but if different tires are used on the front and rear wheels, or if modeling is required to take into account changes in characteristics such as tire wear, it is sufficient to prepare two types of this function.

For the rear wheels, the side force Frs and cornering force Frc are output by correcting the distance between the measurement points according to the rear tire side slip angle βr. The cornering drag Rfc of the front wheels is calculated from the side force Ffs and cornering force Ffc of the front wheels, and the cornering drag Rrc of the rear wheels is calculated in the same manner. The cornering drag Rc applied to the entire vehicle can be calculated for all four wheels from the above two values. Note that the entire vehicle is calculated using the running resistance model 11, so it is input here. In general, when the steering angles δ1 and δ2 are small, the front wheel side slip angle βf also becomes small, and since the side force Fs and the cornering force Fc are almost the same, the cornering drag Rc becomes almost zero. And when the steering angles δ1 and δ2 become large, the side force Fs increases accordingly, while the cornering force Fc decreases and the cornering drag Rc increases, which affects the drive and braking system test. Incidentally, "cornering drag" is often described as an increase in rolling resistance due to tire characteristics relative to the above steering angle. Meanwhile, windage loss also affects the increase in rolling resistance, as the projected cross-sectional area and CD value of the vehicle change when turning. In this case too, it is sufficient to input this to the dynamometer as an increase in rolling resistance due to wind from an oblique direction based on the vehicle sideslip angle β. This increase in wind loss can also be considered cornering drag in a broad sense. However, to implement this model, it is necessary to obtain characteristic values using a limited number of wind tunnel test devices, but as there is not much data available, practical use depends on the availability of such data.

Next, in order to explain the vehicle motion model of the test vehicle Tc, FIG. 5 shows an equivalent two-wheeled vehicle model of a four-wheeled vehicle, and various quantities related to turning can be derived. In FIG. 5, as tire characteristics, βf is the front wheel side slip angle, βr is the rear wheel side slip angle, Ff is the front wheel cornering force, and Fr is the rear wheel cornering force. In addition, as vehicle body motion, δ is the steering angle, β is the vehicle side slip angle, γ is the vehicle yaw moment, and V is the vehicle speed. As vehicle specifications, l is the wheelbase, lf is the distance between the vehicle center of gravity G and the front wheel axis, lr is the distance between the vehicle center of gravity G and the rear wheel axis, m is the vehicle weight, and J is the vehicle yaw moment of inertia. When solving for turning, centrifugal force and camber thrust are also related to the roll direction during turning, but in this embodiment, since the vehicle motion model itself during turning is known, the camber of the four wheels is omitted, and only an example when solving from the motion in the vehicle yaw direction is described.

The equations of motion are given below from the above specifications and state quantities. That is, the yaw moment γ at the vehicle's center of gravity G is calculated from Equation 1, the equation for the vehicle's lateral axis direction at the vehicle's center of gravity G is calculated from Equation 2, and the relationship between the steering angle δ, vehicle side slip angle β, front wheel side slip angle βf, and rear wheel side slip angle βr is calculated from Equation 3.

Formula 1

Formula 2

Formula 3

If we first summarise the coefficients A and B found from the parameters corresponding to the above specifications, we get Equation 4. Coefficient A is a parameter generally known as the stability factor Sf.

Formula 4

From the above, when calculating the vehicle side slip angle β and vehicle yaw moment γ for the steady state of dV/dt=0, dβ/dt=0, dγ/dt=0, we obtain the following Equation 5.

Formula 5

From the above formula 5, the turning radius R is given by the following formula 6.

Formula 6

Next, referring to FIG. 6, in order to explain the fourth turning model 15e in FIG. 3, the turning radius of the test vehicle Tc and the difference in the left and right wheel rotation speeds, which is an example of balance control, will be explained. In FIG. 6, R is the turning radius, R' is the inner wheel turning radius, d is the vehicle tread, θ is the turning angle, L1 is the inner wheel travel distance, L2 is the outer wheel travel distance, △L is the inner and outer wheel travel distance difference, V is the speed, △V is the inner and outer wheel speed difference, V-△V/2 is the inner wheel speed, V+△V/2 is the outer wheel speed, and ω is the angular velocity. Also, rt is the tire radius. Then, △t is the time until the travel difference of △L occurs, and the inner and outer wheel speed difference △V can be calculated from the following equation 7.

Formula 7

The following formula 8 is used to convert the inner/outer wheel speed difference ΔV into the left/right rotation speed difference ΔNrl. Note that since the turning radius difference is the same for the front/left/right wheels and the rear/left/right wheels, the left/right rotation speed difference ΔNrl is also the same.

Formula 8

Similarly, the difference in rotation speed between the front and rear wheels, ΔNfr, can be geometrically approximated as follows using the difference in turning radius:

Formula 9

Here, in the above, the vehicle side slip angle β and vehicle yaw moment γ were calculated for the steady state of dV/dt=0, dβ/dt=0, dγ/dt=0, but it is possible to calculate in real time using a calculator to determine the transients as well. Then, by substituting the front wheel side slip angle βf and the rear wheel side slip angle βr for the front wheel cornering force Ff and the rear wheel cornering force Fr described above and substituting these into the above formulas 1 and 2 and integrating all the terms, formulas 10 and 11 are obtained.

Formula 10

Formula 11

In the above equations 10 and 11, s is the Laplace operator, and 1/s has the same meaning as the integral symbol. If the above is expressed as a block diagram, the whole is shown in FIG. 7, with the above equation 10 being the vehicle side slip angle model, and equation 11 being the yaw moment model. If the vehicle side slip angle model is described in detail according to equation 10, it becomes as shown in FIG. 8(a), and if the yaw moment model is described in detail according to equation 11, it becomes as shown in FIG. 8(b). Furthermore, the turning radius model becomes as shown in FIG. 8(c). As a result, if the first turning model 15a is replaced with the dynamic vehicle turning model of FIG. 7 and FIG. 8(a) to (c), the dynamic front wheel side slip angle βf, rear wheel side slip angle βr, turning radius R, vehicle side slip angle β, and vehicle yaw moment γ can each be obtained.

According to the above embodiment, when the test vehicle Tc is making a turn, the steering angle δ is detected and a set value (load) that takes into account the differential rotation speed and differential torque, which is an example of balance control for each wheel, can be applied to each dynamometer Dm1 to Dm4. Therefore, the power balance during driving and braking when the test vehicle Tc is making a turn can be reproduced. In other words, since the set value for the steering angle δ can be automatically applied to the dynamometers Dm1 to Dm4, an evaluation test can be performed even if the set value is unknown in advance.

Although the embodiment of the present invention has been described above, various modifications are possible without departing from the scope of the technical concept of the present invention. In the above embodiment, the setting value of the balance control is assumed to be the passive differential gear of the test vehicle Tc, and the representative state quantity is the difference rotation speed setting and the balance control of the dynamometer control system, but this is not limited to this. For example, the difference may be related to the difference in the differential mechanism or the presence or absence of active vehicle control, but even in this case, the travel distance difference and difference rotation speed of each wheel calculated from the steering angle δ during steering do not change, so the balance setting on the transmitting side can be output as is, and the control system on the receiving side balance control can be adapted so that the vehicle difference is the same as the actual travel.

In the above embodiment, the test vehicle Tc is a four-wheel drive and front-wheel steering vehicle, but the present invention is not limited to this. For example, the present invention can be applied to two-wheel drive vehicles, and the present invention can be applied to cases where recent vehicles control brakes such as ABS (antilock braking system), TRC (traction control), and ESC (electronic stability control), which are standard preventive safety technologies for vehicle stability devices. The driven wheels are the driving wheels, and the non-driven wheels are the trailing wheels. When the test vehicle (actual vehicle) runs on the actual road, the trailing wheels also rotate because they are in contact with the road surface. In this case, when the evaluation test is performed using the vehicle test device TM of the present invention, there is no road surface, so the trailing wheels must be rotated so that the vehicle stability device judges the vehicle to be spinning and abnormal, and the vehicle cannot be driven. In such a case, as shown in FIG. 9, the dynamometers Dm3 and Dm4 connected to the trailing wheel coupling are also controlled in rotation speed so that the vehicle speed is the same as the driving wheels, so that the vehicle stability device can drive without judging the vehicle to be abnormal. The control function of these dynamometers Dm3 and Dm4 is called the "vehicle speed tracking control unit." However, if the vehicle stabilization device is disabled and testing is performed to obtain the desired evaluation, then dynamometers can be provided only for the drive wheels (two front-wheel drive wheels, or two rear-wheel drive wheels). The control system can also be configured to switch the load distribution for two-wheel drive. These control functions can be provided in the load distribution unit of the load distribution and tire drive control direction model described above.

Furthermore, in the above embodiment, the front left/right wheel rotation speed difference ΔNfrl, the rear left/right wheel rotation speed difference ΔNrrl, and the front/rear wheel rotation speed difference ΔNfr are described as examples of balance commands. This assumes that the vehicle's differential mechanism is a differential gear, and is generally performed when the vehicle side does not perform control. In the case of a differential mechanism other than a differential gear or when the vehicle side performs balance control for each wheel such as active yaw control, the balance command is not limited to the rotation speed, but rather gives a model of the angle difference or torque difference based on the difference in travel distance. Alternatively, the balance control may be disabled, and the evaluation test may be performed using only the slip model in the model of the tire's driving/braking direction.

Although there are not many of them at present, even if the test vehicle Tc is a three-wheeled vehicle, evaluation tests can be performed by preparing the necessary number of dynamometers and using steering angle sensors on the front wheels, just like with four-wheeled vehicles. For such an automobile test device, FIG. 10(a) shows an example of the configuration for a "test vehicle Tc1" with one front wheel and two rear wheels, and FIG. 10(b) shows an example of the configuration for a "test vehicle Tc2" with two front wheels and one rear wheel. Also, if a test is performed in which a two-wheeled vehicle (test vehicle Tc3) such as a motorcycle is turned by the steering angle, as shown in FIG. 11(a), the necessary number of dynamometers Dm1 and Dm3 can be prepared as above. However, as shown in FIG. 11(b), if the vehicle body is tilted and cornering is performed by a cornering force called a camber force corresponding to an angle called a camber angle using the automobile test device of the present invention, the dynamometer Dm2 used for this also needs to be angled to correspond to the camber angle. In such cases, a camber angle sensor can be installed and its output signal can be input to the turning model to provide this functionality. Since the vehicle does not have a differential mechanism on the front and rear wheels, balance control is not required for evaluation. Furthermore, if ABS abnormality detection is to be prevented in the same way as in four-wheeled vehicles, vehicle speed tracking control and front and rear wheel balance control can be added, and a cornering drag load can be applied to perform evaluation tests of drive and braking during turns.

The present invention can be applied not only to four-wheel commercial vehicles such as buses and trucks, but also to vehicles with six or more trailing wheels. However, as shown in FIG. 12, for trailer vehicles in which the tractor part with the driving wheels and the trailer part are connected via a coupling device, the vehicle side slip angles β1 and β2 of the tractor part and the trailer part are not the same, so the "turning model (see FIGS. 2 and 3)" of the above embodiment can be rewritten for trailers to be valid. In the case of a four-wheel vehicle with four-wheel steering function, as shown in FIG. 13, if steering angle sensors 2c and 2d are provided on the rear wheel side and input to the tire lateral force model and turning motion model 15, evaluation tests can be performed in the same way. In the vehicle motion model, if the rear wheel steering angle in the equation for the rear wheel side slip angle βr is set to δr in the same way as the front wheel steering angle δ in the equation for the front wheel side slip angle βf, the turning model can be solved in the same way using the following equation 12.

Formula 12

In the above embodiment, the dynamometers Dm1 to Dm4 connected to the wheel couplings (hubs) Wc1 to Wc4 and movable in response to steering are described in the same manner for all wheels. However, for the dynamometers of the wheel axles that do not need to be movable in response to steering, a hub-coupled dynamometer that does not move in response to steering or a roller-type dynamometer that is independent on the left and right can be combined to provide balance control and cornering drag Rc for each wheel. Even if a roller-type dynamometer that is not independent on the left and right is combined, it is possible to provide cornering drag Rc, although left and right balance control is not possible. In addition, if it is used not for evaluation of driving and braking forces, but for example, when steering with the wheels rotating without applying a load, for example, to check the operation of the vehicle stability system excluding the ESC, or to check the partial operation of other electronic control devices, it functions as a test device for that purpose.

Furthermore, in the above embodiment, the steering angle is output from the dynamometers Dm1 to Dm4 that move in the steering direction (i.e., the steering angle sensors 2a and 2b are provided on the dynamometers Dm1 to Dm4), but the steering angle of the wheel part, called the actual steering angle, may be detected. Since the steering is performed by the steering wheel, the power balance during turning and the cornering drag Rc can also be reproduced by using the steering wheel steering angle as the steering angle. Note that steering is often performed using a steering system called a power assist, which assists the driver's muscle power. For example, in the case of the electric type shown in FIG. 14, the steering force is transmitted to the vehicle coupling parts Wc1 and Wc2 by the steering wheel Hs, motor, reduction gear, rack and pinion, and steering shaft mechanism, so that the ratio of the steering wheel operation angle θ to the actual steering angle δ is a roughly constant reduction ratio G, excluding the play angle in the steering system. The test vehicle Tc is also provided with an angle sensor (not shown) for detecting the angle of the steering wheel Hs, which is connected to a vehicle control device called an ECU, which is connected to a communication system between ECUs called a CAN in the test vehicle Tc, and angle data of the steering wheel Hs is assigned to the vehicle, a part of which is connected to an on-board fault diagnosis device called an OBD external to the vehicle, and the steering wheel angle θ may be obtained by connecting dynamometers Dm1, Dm2 (their control units) and the ECU via the OBD port. Then, the steering wheel angle θ is divided by the reduction ratio G with respect to the actual steering angle δ in the first turning model 15a to obtain the actual steering angle δ, which is input to the subsequent turning model 15b, thereby controlling the dynamometers Dm1, Dm2 during turning. In addition to using data held by the control device of the test vehicle Tc as the steering angle θ, as an example shown in FIG. 15, a measuring device such as a steering angle meter may be attached to detect the steering angle θ without connecting to the vehicle's ECU, output the signal, and connect it to the turning motion model 15.

TM... automobile test device, Tc... test vehicle, Wc1 to Wc4... wheel coupling parts, Dm1 to Dm4... dynamometers, Cu... control unit, 11... running resistance model, Bs... vehicle specifications, V... vehicle speed, 13... balance control part, δ... steering angle, 2a, 2b... steering angle sensor, 15... turning motion model, θ... steering wheel operation angle (steering wheel angle).

Claims (7)

The test vehicle is provided with dynamometers each connected to a wheel coupling portion of the test vehicle, a control unit for controlling the load of each dynamometer, and a vehicle speed detection means for detecting the vehicle speed from the number of revolutions of the dynamometer,
A vehicle testing device, comprising: a control unit, which is provided with a running resistance model that models the running resistance acting on a test vehicle in correspondence with preset vehicle specifications and a vehicle speed of the test vehicle; and a balance control unit that determines a load to be applied to each dynamometer based on a calculation result of the running resistance model corresponding to a vehicle speed detected by a vehicle speed detection means, and in which each dynamometer is controlled in accordance with the load determined by the balance control unit,
In a vehicle in which a dynamometer connected to a steered wheel coupling is movable in a steering direction in response to steering,
The vehicle further includes a steering angle sensor that detects a steering angle associated with steering of the test vehicle, and a turning motion model that models turning motion when the test vehicle is steered in accordance with vehicle specifications, vehicle speed, and the steering angle detected by the steering angle sensor of the test vehicle,
a balance command value as a result of calculation in a cornering motion model is input to a balance control unit, in addition to the result of calculation in a running resistance model, and the load is determined by taking into account balance control between dynamometers based on this balance command value ; a cornering drag is calculated by the cornering motion model in accordance with the steering angle detected by the steering angle sensor, and the result of this calculation is input to the running resistance model . 2. The vehicle testing device according to claim 1 , wherein the cornering motion model is a model of an equation of motion of the test vehicle when cornering. 3. The vehicle testing device according to claim 1, wherein the steering angle sensor is provided for each of the wheels of the test vehicle which are steered. 3. An automobile testing apparatus according to claim 1 , wherein the dynamometer connected to something other than the steered wheels is of a hub-connected type or roller type which is immovable in the steering direction. 5. The automobile testing device according to claim 1 , wherein a steering wheel angle of a test vehicle is used instead of the steering angle detected by the steering angle sensor. 5. The automobile testing device according to claim 1 , wherein the test vehicle is any one of a two-wheeled vehicle, a three-wheeled vehicle, a four-wheeled vehicle, a six-wheeled vehicle, an eight-wheeled vehicle, and a vehicle having a trailer. 2. An automobile testing apparatus according to claim 1 , wherein, when the test vehicle is a two-wheeled vehicle, the automobile testing apparatus is capable of detecting a camber angle and reproducing cornering drag during cornering.

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