JP2006130985A - Vehicle body structure for automobile - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a vehicle body structure for an automobile capable of decreasing the weight of a vehicle while keeping the strength of reinforcement. <P>SOLUTION: The vehicle body structure for an automobile comprises a cabin 1 on which an occupant gets; side sills 61R, 62L respectively disposed on right and left sides of the cabin 1; shock absorbing members 69R, 69L respectively joined to right and left sides on a vehicle body front side of the side sills 61R, 61L, and absorbing collision energy from a vehicular front side; a front wheel 42 disposed between the shock absorbing members 69R, 69L; and rear wheels 2RR, 2RF disposed on right and left sides of a vehicle as driving wheels. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention belongs to a technical field of a vehicle body structure that effectively absorbs energy when a vehicle collides with an obstacle and suppresses an influence on an occupant.

The vehicle body structure that absorbs energy when the vehicle collides frontally with an obstacle and suppresses the influence on the occupant is provided with tires at the four corners of the vehicle to protect the occupant against impact from the front of the vehicle. For this reason, a configuration having side members on the inner left and right sides of the front wheel is known (for example, see Patent Document 1).
JP 2001-97048 A

  This body structure is a structure that absorbs collision energy when the side member is crushed when subjected to an impact from the front, and the side member is supported by a side sill via a member such as a structure or a reinforcing material. Take the shock.

  However, since the conventional vehicle body structure is configured to receive the impact received from the side member by the side sill on the side of the cabin, a member such as a structure or a reinforcing material is required, and the vehicle is maintained while maintaining the strength of the reinforcement. There was a problem that it was difficult to reduce the weight.

  The present invention has been made paying attention to the above problems, and an object of the present invention is to provide a vehicle body structure that can reduce the weight of a vehicle while maintaining the strength of reinforcement.

In order to achieve the above object, in the vehicle body structure of the present invention,
The cabin where the crew rides,
Structural members respectively disposed on the left and right sides of the cabin;
An impact absorbing member that is joined to the vehicle front side of the left and right structural members respectively to absorb collision energy from the vehicle front;
A front wheel disposed between the left and right impact absorbing members;
Driving wheels arranged on the left and right of the vehicle;
It is characterized by providing.

  Therefore, in the vehicle body structure of an automobile according to the present invention, the left and right impact absorbing members are directly joined to the structural member (side sill, etc.) on the side of the vehicle. Therefore, the front part of the vehicle body can be reduced in weight without requiring a reinforcing material. That is, it is possible to reduce the weight of the vehicle while maintaining the strength of reinforcement.

  BEST MODE FOR CARRYING OUT THE INVENTION The best mode for carrying out the vehicle body structure of the present invention will be described below based on Examples 1 to 5 shown in the drawings.

First, the configuration will be described.
FIG. 1 is a configuration diagram of a vehicle control system for running the vehicle of the first embodiment. The vehicle according to the first embodiment includes electric motors 3RL and 3RR as driving force generation sources, and the rotation shafts of the electric motors 3RL and 3RR are rear wheels 2RL of the electric vehicle via speed reducers 4RL and 4RR. , 2RR. Here, the output characteristics of the two electric motors 3RL and 3RR, the reduction ratios of the two speed reducers 4RL and 4RR, and the radii of the two left and right rear wheels 2RL and 2RR are all the same.

  Each of the electric motors 3RL and 3RR is a three-phase synchronous motor in which a permanent magnet is embedded in a rotor. Drive circuits 5RL and 5RR that control power exchange with lithium ion battery 6 receive power running and regenerative torque of electric motors 3RL and 3RR from integrated controller (drive force difference control means and vehicle attitude control means) 30. The torque command values tTRL (left rear wheel) and tTRR (right rear wheel) are adjusted to match. Then, the drive circuits 5RL and 5RR transmit the motor rotation speed detected by a rotation position sensor (not shown) attached to each motor rotation shaft to the integrated controller 30, respectively.

  The front wheel 42 is provided on the turning rotation shaft 41 of the front wheel 42, and details are shown in FIG. 2. The turning rotation shaft 41 of the front wheel 42 is inside the hollow support portion 45 of the front wheel 42 and rotates with respect to 45 through a bearing. The elements 44 and 43 and the front wheel 42 are all supported so as to rotate integrally around the turning shaft 41. Here, the intersection point P with the ground surface when the center of the rotation shaft 41 is extended and the point Q directly below the rotation center point 43 of the tire are configured such that the distance is ζ (> 0). A so-called caster structure in which the tire naturally steers so that the traveling direction of the turning rotation shaft 41 of the front wheel 42 coincides with the tire orientation A by traveling resistance during traveling is employed.

  The hollow support portion 45 of the front wheel 42 has a support shaft (not shown) in the vehicle longitudinal direction and the vehicle lateral direction so that the hollow support portion 45 is not easily deformed in the vehicle longitudinal and lateral directions. Further, the hollow support portion 45 is provided with a spring and a damper (not shown) with respect to the vertical direction, making it difficult to transmit the vertical force received by the front wheel 42 from the road surface to the vehicle body.

  Further, the front wheel 42 is provided with a friction brake by a hydraulic system (not shown) at a portion 44, and the hydraulic pressure of the brake system rises when the driver depresses the brake pedal, and the portion 44 according to the increase of the hydraulic pressure. The brake pad fixed to the front wheel 42 brakes the front wheel 42 by sandwiching a disk that rotates together with the front wheel 42.

  The rear wheels 2RL and 2RR are also provided with friction brakes (not shown), and the rear wheels 2RL and 2RR are braked in response to the driver's depression of the brake pedal, similarly to the front wheels 42. Furthermore, a link 51RL is connected to the rear wheel 2RL, and the rear wheel 2RL is steered by moving the link 51RL in the vehicle left-right direction by the steering motor 52RL.

  A motor drive circuit 53RL is connected to the steering motor 52RL, and the motor drive circuit 53RL receives the steering angle detection value from the actual steering angle sensor attached to the left rear wheel and the left rear wheel received from the integrated controller 30. Based on the steering angle target value tδrl, the torque of the steering motor 52RL is adjusted so that the actual left steering wheel actual turning angle matches the left rear wheel steering angle target value tδrl.

  Similarly, a link 51RR is connected to the rear wheel 2RR, and the rear wheel 2RR is steered by moving the link 51RR in the vehicle left-right direction by the steering motor 52RR. A motor drive circuit 53RR is connected to the steering motor 52RR, and the motor drive circuit 53RR receives the detected steering angle value from the actual steering angle sensor attached to the right rear wheel and the right rear wheel received from the integrated controller 30. Based on the steering angle target value tδrr, the torque of the steering motor 52RR is adjusted so that the actual right rear wheel steering angle matches the right rear wheel steering angle target value tδrr. The driving wheel turning means is constituted by the turning motors 52RL and 52RR and the links 51RL and 51RR.

  FIG. 3 is a diagram illustrating the vehicle body structure of the vehicle according to the first embodiment. As illustrated in FIG. 3, in the first embodiment, the impact on the vehicle front extension of the left and right side sills (structural members) 61L and 61R of the cabin 1 Absorbing members 69L and 69R are arranged. These shock absorbing members 69L and 69R have a hollow structure, and in particular, the tip (front of the vehicle) portion has a structure in which the strength is weakened, so that the impact is crushed from the tip against the impact from the front of the vehicle. . When the vehicle collides with an obstacle from the front, the collision energy is absorbed by the energy absorption by the crushing deformation, and the influence of the collision on the occupant is reduced.

  In the first embodiment, the electric motors 3RL, 3RR, a battery, and the like are arranged so that the front-rear center of gravity position of the electric vehicle is near the midpoint of the rear wheels 2RL, 2RR. For example, the front / rear center of gravity position is designed so that the ratio of the front wheel load to the sum of the rear left and right wheels is 2: 8.

Next, the integrated controller 30 will be described.
The integrated controller 30 includes an accelerator opening signal detected by the accelerator pedal sensor 23, a brake pedal force signal detected by the brake pedal sensor 22, and a displacement angle attached to a support point of the steering wheel 11 that can be displaced left and right. A displacement angle signal of the steering wheel 11 detected by the sensor 26, a rotation angle signal of the steering wheel 11 detected by the steering angle sensor 21 attached to the rotation shaft of the steering wheel 11, and an acceleration sensor 24 attached to the center of gravity of the vehicle. The vehicle body lateral acceleration (acceleration in the vehicle width direction) and the longitudinal acceleration signal detected by, the yaw rate signal detected by the yaw rate sensor 8, the state signal of the shift lever 25 operated by the driver, and the rotation center point 43 of the front wheel 42 Take Wheel speed signal detected by the front wheel rotation sensor 49 kicked is input.

  The shift position of the shift lever 25 can be selected only when the vehicle is stopped, is a position “P” used during parking, a position “D” used during normal forward travel, and a position “A” used during forward travel in Mode A. There is. Mode A is an operation mode in which the vehicle is rotated with the vicinity of the rear wheel of the turning inner wheel as the turning center.

  These shift positions are selected by the driver by operating the shift lever 25. Based on these signals, the integrated controller 30 calculates a torque command value tTRL for the rear left wheel motor 3RL and a torque command value tTRR for the rear right wheel motor 3RR, and supplies them to the drive circuits 5RL and 5RR for the motors 3RL and 3RR. Send. Here, the torque command value tTRL to the rear left wheel motor 3RL and the torque command value tTRR to the rear right wheel motor 3RR are both in units of Nm, and the direction in which the vehicle is accelerated forward is positive. The integrated controller 30 also calculates rear wheel steering angle target values tδrl and tδrr and transmits them to the motor drive circuits 53RL and 53RR, respectively. The rear wheel rudder angle target values tδrl and tδrr have a unit of rad, and the direction to turn left is positive.

Next, the operation will be described.
[Mode selection control processing]
FIG. 4 is a flowchart showing the flow of the mode selection control process executed by the integrated controller 30 according to the first embodiment. Each step will be described below. The integrated controller 30 includes peripheral components such as a RAM / ROM in addition to the microcomputer, and executes the flowchart of FIG. 4 at regular intervals, for example, every 5 ms.

  In step S401, each sensor signal and the received signals from the drive circuits 5RL and 5RR are stored in the RAM variable, and the process proceeds to step S402. Specifically, the accelerator opening signal is stored in a variable APS (the unit is% and fully opened is 100%), the brake pedal force signal is stored in a variable BRK (unit is Pa), and the steering wheel 26 The displacement angle signal is stored in variable Win (unit is rad, left is positive), and the rotation angle signal of steering wheel 11 is stored in variable δ (unit is rad, counterclockwise is positive). The vehicle body lateral acceleration signal is stored in a variable YG (takes the direction when turning left in FIG. 1 is positive), and the vehicle body yaw rate signal is stored in the variable γ (takes the direction when turning left in FIG. 1 is positive). The vehicle body lateral acceleration signal is stored in the variable YG (the direction when turning left in FIG. 1 is positive), and the shift lever signal is stored in the variable SFT. Further, the rotational speed signal from the front wheel rotation sensor 49 is stored in a variable NFL (unit is rad / s, and the direction in which the vehicle moves forward is positive). Further, for the signals received from the drive circuits 5RL and 5RR, the rotational speeds of the respective motors are stored in the variables NRL and NRR (both units are rad / s, and the direction in which the vehicle moves forward is positive).

In step S402, the speed V of the vehicle (unit is m / s and the direction in which the vehicle moves forward is positive) is calculated by the following equation, and the process proceeds to step S403.
V = (NFL * Rf + NRL / GG * Rr + NRR / GG * Rr) / 3
Here, Rf is the radius of the front wheel, Rr is the radius of the rear wheel, and GG is the reduction ratio of the reduction gear of the rear wheel.

  In step S403, it is determined whether or not the shift lever position is the position “P” used during parking. If “P”, the process proceeds to step S404, and this routine is executed with tTRL = tTRR = tδr1 = tδrr = 0. finish. Otherwise, the process proceeds to step S413.

  In step S413, it is determined whether or not the shift lever position is “A”. If the shift lever position is “A”, the process proceeds to step S414 to execute a calculation routine in mode A to be described later. End the routine. If not, the process proceeds to step S415, a calculation routine in mode D described later is executed, and this routine is terminated.

[About the design principle and operation form of the controller that realizes the reference model response]
Next, in the calculation routine in mode D (FIG. 6) and the calculation routine in mode A (step S414) executed in step S415 of FIG. 4, the torque command value tTRL to the rear left wheel motor 3RL is A method of calculating the torque command value tTRR and the rear wheel steering angle target values tδrl and tδrr to the right wheel motor 3RR will be described in order. Before describing the calculation processing in each calculation routine, first, the motor in mode D The calculation principle and realization method of the torque command value and the rear wheel steering angle target value will be described.

  “Motion and control of the body structure of an automobile” (Sankaido) shows a motion equation of vehicle behavior for steering front and rear wheels. For example, for p194, the front wheel rudder angle δf [rad] and the rear wheel rudder angle δr [rad] (the rear wheel left and right rudder angles are the same) are the manipulated variables, the vehicle yaw rate γ [rad / s] and the vehicle body center of gravity The equation of motion is shown with the vehicle body slip angle β [rad] at the position as the state quantity. This equation of motion shows that the vehicle speed V [m / s] is constant (dV = 0), V ≠ 0, and the slip angle (β [rad]) is very small (| β | << 1, sinβ ≒ β, cosβ ≒ 1). Derived on the premise of.

The concept of the equation of motion can be applied to the electric vehicle according to the first embodiment of the present invention. That is, an operation amount with the right wheel driving force u [N] and the left wheel driving force -u [N] is added, the steering angle of the rear wheel is set to the same value δr [rad], and the front wheel is As a function of the caster type, the lateral force generated at the front wheels is almost zero, and the equation of motion can be derived as follows.

  Here, Lr is the distance [m] between the rear wheel shaft and the center of gravity, Lt is the tread base distance / 2 [m] of the rear wheel, m is the vehicle weight [kg], and Iγ is the yaw moment of inertia [Nmss]. Kr is the rear wheel tire cornering stiffness [N / rad], which takes into account the decrease in cornering power with respect to the steering angle due to the influence of the rear wheel steering stiffness. V is the vehicle speed [m / s], γ is the yaw rate [rad / s], and β is the vehicle slip angle [rad] at the center of gravity of the vehicle.

The lateral force Y [N], yaw rate γ [rad / s], and slip angle β [rad] have the following relationship, as described in “Motion and Control of Car Body Structure” (Sankaido) p52.
Y = mV (dβ / dt + γ) (A2)

These equations of motion can be rewritten into the following form using the differential operator s.
β = {Q ₁₂ (s) / Qden (s)} · δr + {Q ₁₃ (s) / Qden (s)} · u
γ = {Q ₂₂ (s) / Qden (s)} · δr + {Q ₂₃ (s) / Qden (s)} · u
Y = {Q ₃₂ (s) / Qden (s)} · δr + {Q ₃₃ (s) / Qden (s)} · u (A3)

Q ₁₂ (s), Q ₁₃ (s), Q ₂₂ (s), Q ₂₃ (s), Q ₃₂ (s), Q ₃₃ (s), and Qden (s) are all functions of the vehicle speed V. It is expressed by the following formula.
Q ₁₂ (s) = 2VKr (I _γ s + mVLr)
Q ₁₃ (s) = -2Lt (mV ² -2LrKr)
Q ₂₂ (s) = -2mV ² KrLrs
Q ₂₃ (s) = 2VLt (mVs + 2Kr)
Q ₃₂ (s) = 2mV ² KrI _γ s ²
Q ₃₃ (s) = 4mVLtKr (Lrs + V)
Qden (s) = mV ² I _γ s ² + 2VKr (mLr ² + I _γ ) s + 2 mV ² LrKr (A4)

“Motion and control of the body structure of an automobile” (Sankaido) p203-p207 is designed so that the response of the yaw rate γ and slip angle β of the vehicle to the steering operation amount δ is a desirable transfer characteristic (reference model). A controller derivation method for generating the steering angle command value δf ^* and the rear wheel steering angle command value δr ^* is also shown. According to this method, in Example 1, the response of the yaw rate γ and the slip angle β to the steering operation amount δ becomes a desirable transfer characteristic (reference model), and the yaw rate γ and the slip angle β to the steering displacement angle Win are set. A controller (P1 (s) in FIG. 5 (a)) that calculates the command value u ^{* for} the driving force difference between the left and right wheels and the command value δr ^* for the rear wheel steering angle so that the response is a desirable transfer characteristic (standard model). , P2 (s), P3 (s), and P4 (s)) can be derived as follows.

Now, the desired transfer characteristic (reference model) of yaw rate γ with respect to the steering operation amount δ is G _γδ , the desired transfer characteristic (reference model) of slip angle β with respect to the steering operation amount δ is G _βδ , and the desired yaw rate γ with respect to the steering displacement angle Win. The transfer characteristic (normative model) is G _γw , and the desired transfer characteristic (normative model) of the slip angle β with respect to the steering displacement angle Win is G _βw .

G _γδ = m ₂ / (s ² + 2wns + wn ² )
G _βδ = 0
G _γw = 0
G _βw = m ₁ / (s ² + 2wns + wn ² )… (A5)

That is, desired response smooth second order response of the yaw rate γ for steering operation amount [delta] _{(e.g., wn = 4π, m 2 =} wn 2/4) and a smooth and desired response of the slip angle β with respect to steering wheel lateral displacement Win The secondary response characteristics are set (for example, wn = 4π, m ₁ = wn ² ). The slip angle β is set so that it is not affected by the steering operation amount δ, and the yaw rate γ is not affected by the steering displacement angle Win.

  Incidentally, in FIG. 5A, the relationship among the steering operation amount δ and the steering displacement angle Win, the yaw rate γ, and the slip angle β is as follows. Here, 1 / Td (s) is the servo delay of the rear wheel steering system.

_{β = ({Q 13 (s} ) / Qden (s)} p1 (s) + {Q 12 (s) / Qden (s)} · {p2 (s) / Td (s)}) δ +
_{({Q 13 (s) /} Qden (s)} p3 (s) + {Q 12 (s) / Qden (s)} · {p4 (s) / Td (s)}) Win
_{γ = ({Q 23 (s} ) / Qden (s)} p1 (s) + {Q 22 (s) / Qden (s)} · {p2 (s) / Td (s)}) δ +
_{({Q 23 (s) /} Qden (s)} p3 (s) + {Q 22 (s) / Qden (s)} · {p4 (s) / Td (s)}) Win ... (A6)

Accordingly, the transfer characteristics, respectively desired transfer characteristic _{G βw, G γδ, G βδ} , from the condition that match the G _{Ganmadaburyu,} formula (A7) is guided,
G _βw =
m ₁ / (s ² + 2wns + wn ² ) = {Q ₁₃ (s) / Qden (s)} p3 (s) + {Q ₁₂ (s) / Qden (s)} · {p4 (s) / Td (s) }
_{G βδ = 0 = {Q 13} (s) / Qden (s)} p1 (s) + {Q 12 (s) / Qden (s)} · {p2 (s) / Td (s)}
G _γδ =
^{_{m 2 / (s 2 + 2wns}} + wn 2) = {Q 23 (s) / Qden (s)} p1 (s) + {Q 22 (s) / Qden (s)} · {p2 (s) / Td (s) }
_{G γw = 0 = {Q 23} (s) / Qden (s)} p3 (s) + {Q 22 (s) / Qden (s)} · {p4 (s) / Td (s)} ... (A7)
By solving these simultaneous equations and using the relational expression (A4), the controllers P1 (s), P2 (s), P3 (s) and P4 (s) can be derived as shown in the expression (A8). Here, the servo delay of the rear wheel steering is a first-order delay of a time constant τ (for example, τ = 0.1 [s]), that is, Td = τs + 1.

p1 (s) = {m ₂ / (2Lt)} {(I _γ s + mVLr) / (s ² + 2wns + wn ² )}
p2 (s) = {m ₂ (mV ² -2LrKr) / (2VKr)} {(τs + 1) / (s ² + 2wns + wn ² )}
p3 (s) = {m ₁ mVLr / (2 * Lt)} {s / (s ² + 2wns + wn ² )}
p4 (s) = {m ₁ (mVs + 2Kr) / (2Kr)} {(τs + 1) / (s ² + 2wns + wn ² )}… (A8)

As described above, in the first embodiment, the command for the driving force difference between the left and right wheels so that the responses of the yaw rate γ and the slip angle β to the steering operation amount δ and the steering displacement angle Win are respectively desirable transmission characteristics (reference model). A controller (P1 (s), P2 (s), P3 (s), and P4 (s) in FIG. 5A) for calculating the value u _* and the rear wheel steering angle command value δr ^* can be derived. .

Next, a method for realizing the controllers P1 (s), P2 (s), P3 (s), and P4 (s) in the equation (A8) will be described. Since P1 (s), P2 (s), P3 (s), and P4 (s) can be rewritten by the following equation (A9), a method for realizing the equation (A9) will be described. b0, b1, and b2 are functions such as the vehicle speed V.
yx = (b2s ² + b1s + b0) / (s ² + 2wns + wn ² )… (A9)

  Equation (A9) can be rewritten as shown in FIG. Therefore, at every predetermined time (for example, every 5 ms), first, X2 and X1 are updated by performing the integration operation in FIG. 5B by, for example, Euler approximation, and then b0, b1, and b2 are changed. This is realized by sequentially updating the output yx from X2, X1, b0, b1, b2, and ux from moment to moment after updating sequentially according to the vehicle speed V.

[Calculation routine in mode D]
In the calculation routine in mode D in step S415 in FIG. 4, the flowchart in FIG. 6 is executed.

  In step S501, the target driving force tTD of the vehicle is calculated. The calculation is performed by referring to a map MAP_tTD (V, APS) stored in advance in the ROM. The map MAP_tTD (V, APS) is characteristic data with the vehicle speed V and the accelerator opening APS as axes, and is set as shown in FIG. 7, for example.

At step S502, it calculates a steady-state value m ₁ of the vehicle body slip angle to steering wheel lateral displacement Win. For example, a fixed value of m ₁ = wn ² (wn is 4π, for example). Alternatively, in order to realize characteristics according to the vehicle speed V and the steering operation amount δ, a method of multiplying the steering displacement angle Win by a value obtained by table map MAP_m1 (V, δ) stored in advance in ROM is used. May be.

  In steps S503 to S512, the target torque tU [Nm] and the rear wheel steering command values tδrl, tδrr of the driving force difference generated in the rear wheel left and right motors according to the steering wheel rotation angle δ, the steering displacement angle Win, and the vehicle speed V are calculated. Calculate. Based on the above design principle, the calculation is such that the response of the yaw rate γ and the slip angle β to the steering operation amount δ is a desirable transfer characteristic (reference model), and the response of the yaw rate γ and the slip angle β to the steering displacement angle Win. Is calculated to have a desirable transfer characteristic (normative model).

Desirable transfer characteristic of yaw rate γ with respect to steering operation amount δ (reference model) G _γδ , Desirable transfer characteristic of slip angle β with respect to steering operation amount δ (reference model) G _βδ , Desirable transfer characteristic of yaw rate γ with respect to steering displacement angle Win (reference) Model) G _γw , a desirable transmission characteristic of the slip angle β with respect to the steering displacement angle Win (normative model) G _βw will be described as the characteristic shown in the equation (A5).

In step S503, the vehicle speed V calculated in step S420 is used to calculate the values corresponding to b0, b1, and b2 in equation (A9) for the first equation in equation (A8) as follows.
b2 = 0
b1 ＝ (m ₂ Ir) / (2Lt)
b0 = (V * m ₂ * m * Lr) / (2Lt) ... (B1)
Here, m ₂ is such that the steady value of the yaw rate with respect to the steering wheel rotation angle δ is, for example, δ / 4.
m ^₂ = wn _2/4
Keep it as Vehicle design values are used for m, Ir, Lr, and Lt.

  In step S504, X2 and X1 are updated by Euler approximation of the integration calculation of FIG. 5B using X2 and X1 obtained when the previous step S504 is executed. In FIG. 5B, ux is the steering wheel rotation angle δ, and the output yx is substituted into the variable yx1. When calculating, as X2 and X1 in FIG. 5B, variables X2a and X1a are used as variables used in step S504. After updating X2 and X1 in FIG. 41 (b), the output yx is calculated from the relational expression shown in FIG. 5 (b) in accordance with those values and b0, b1 and b2 obtained in step S503, and assigned to yx1. To do.

In step S505, using the vehicle speed V calculated in step S420, the values corresponding to b0, b1, and b2 in equation (A9) are calculated as follows for the equation in the second stage of equation (A8). In order to prevent 0%, the vehicle speed V is calculated by limiting the minimum value to Vmin (for example, 1 m / s).
b2 = 0
b1 ＝ m ₂ (mV ² -2LrKr) τ / (2VKr)
b0 ＝ m ₂ (mV ² -2LrKr) / (2VKr)… (B2)
Here, m, Lr, and Kr use vehicle design values. Also, τ is set to, for example, about 0.1 in accordance with the servo delay of the rear wheel steering system.

  In step S506, X2 and X1 obtained by executing the previous step S506 are used, and X2 and X1 are updated by Euler approximation of the integral calculation of FIG. In FIG. 5B, ux is the steering wheel rotation angle δ, and the output yx is substituted into the variable yx2. When calculating, as X2 and X1 in FIG. 5B, variables X2b and X1b are used as variables used in step S506. After updating X2 and X1 in FIG. 5 (b), the output yx is calculated from the relational expression shown in FIG. 5 (b) according to those values and b0, b1 and b2 obtained in step S505, and assigned to yx2. To do.

In step S507, the vehicle speed V calculated in step S420 and m ₁ calculated in step S502 are used, and the values corresponding to b0, b1, and b2 in equation (A9) are expressed as follows for equation (A8) in the third stage: Calculate as follows.
b2 = 0
b1 ＝ m ₁ mVLr / (2Lt)
b0 = 0 ... (B1 ')
Here, m, Lr, and Lt use vehicle design values.

  In step S508, X2 and X1 obtained by executing the previous step S508 are used, and X2 and X1 are updated by Euler approximation of the integral calculation of FIG. In FIG. 5B, ux is the steering displacement angle Win, and the output yx is substituted into the variable yx3. When calculating, as X2 and X1 in FIG. 5B, variables X2c and X1c are used as variables used in step S508. After updating X2 and X1 in FIG. 5 (b), the output yx is calculated from the relational expression shown in FIG. 5 (b) according to those values and b0, b1 and b2 obtained in step S507, and substituted into yx3. To do. In this way, by setting b0 = 0, it is possible to realize an operation in which the left / right driving force difference operation is not constantly performed with respect to the steering displacement angle Win (the same as the steady gain of p3 (s) being 0). .

At step S509, the use of a m ₁ computed by the vehicle speed V and the step S502 calculated in step S420, for formula of formula (A8) the fourth stage, the formula (A9) of b0, b1, the value of the next corresponding to b2 Calculate as follows.
b2 ＝ Vm ₁ mτ / (2Kr)
b1 ＝ (m ₁ mV + 2 * m ₁ τKr) / 2
b0 ＝ m ₁ … (B2 ')
Here, m and Kr use vehicle design values. Τ is set to the same value as described above.

  In step S510, X2 and X1 obtained by executing the previous step S510 are used, and X2 and X1 are updated by Euler approximation of the integral calculation of FIG. In FIG. 5B, ux is the steering displacement angle Win, and the output yx is substituted into the variable yx4. When calculating, as X2 and X1 in FIG. 5B, variables X2d and X1d are used as variables used in step S510. After updating X2 and X1 in FIG. 5 (b), the output yx is calculated from the relational expression shown in FIG. 5 (b) according to those values and b0, b1 and b2 obtained in step S509 and substituted into yx4. To do.

In step S511, the target left / right driving force difference tU is calculated as the sum of the yx1 value calculated in step S504 and the yx3 value calculated in step S508.
tU = yx1 + yx3

In step S512, rear wheel steering command values tδrl and tδrr are calculated as the sum of the yx2 value calculated in step S506 and the yx4 value calculated in step S510.
tδrl = tδrr = yx2 + yx4

In step S518, torque command values tTRL and tTRR for the rear wheels are calculated from the target driving force tTD and the target left / right driving force difference tU by the following equation, and this routine is terminated.
TTRL = tTD * Rr / GG / 2-tU * Rr / GG
TTRR = tTD * Rr / GG / 2 + tU * Rr / GG… (B3)

[Calculation routine in mode A]
In the calculation routine in mode A in step S414 in FIG. 4, the flowchart of FIG. 8 is executed.

  In step S801, a target speed tV at the center position of the rear wheel shaft is calculated. For example, when the accelerator calculation APS is 0, tV = 0, and when the APS is 100%, tV = 3 [m / s] is assigned proportionally.

  In step S802, a target value tρ that is an inverse value of the turning radius of the center position of the rear wheel shaft (point P in FIG. 9) is calculated. The target value tρ is a positive value when turning left, that is, when the vehicle's turning center is on the left, and negative when turning right, that is, when the vehicle's turning center is on the right. For example, the characteristics shown in FIG. 10 are set according to δ. Here, for example, when the length of the half of the tire width is Wt and the length of the half of the rear wheel tread base distance is Lt, the target value tρ is −1 / (Lt + Wt) and 1 / (Lt + Wt) Set to be in the range between.

  In step S803, the target slip angle βc of the center position of the rear wheel shaft (point P in FIG. 9) (the angle of the direction in which the point P advances with respect to the direction of the vehicle body, the counterclockwise direction being taken positively, the unit is [rad ]). The calculation is performed with reference to the table previously stored in the ROM as shown in FIG. 11 according to tρ and the steering displacement angle Win.

In step S804, the left rear wheel motor and the right rear wheel are obtained from the target speed tV of the center position of the rear wheel shaft (point P in FIG. 9), the target value tρ of the reciprocal value of the turning radius of the center position of the rear wheel shaft and the target slip angle βc. Calculate the target rotational speed tNRL and tNRR of the wheel motor using the following formula. At that time, pay attention to the following relationship. That is, when the vehicle turns so as to realize tρ obtained in step S802 and the target slip angle βc obtained in step S803, the rotation center of the vehicle is point Q in FIG. 9, and therefore the moving speed of point P is tV The moving speed of the left rear wheel is tV * R1 / R0 (R1 is the distance between the left rear wheel and point Q, R0 is the distance between point P and point Q), and the moving speed of the right rear wheel is tV * R2 / R0 (R2 is the distance between the right rear wheel and point Q). Using this relationship, the target rotational speeds tNRL and tNRR of the left rear wheel motor and the right rear wheel motor are calculated by the following equations.
tNRL = tV / Rr * GG * sqrt (1-2tρLt cos (βc) + tρ ² Lt ² )
tNRR = tV / Rr * GG * sqrt (1 + 2tρLt cos (βc) + tρ ² Lt ² )… (C1)
Here, it is utilized that the values of R1 / R0 and R2 / R0 can be derived as follows from the geometric relationship shown in FIG.
R1 / R0 = sqrt (1-2tρLt cos (βc) + tρ ² Lt ² )
R2 / R0 = sqrt (1 + 2tρLt cos (βc) + tρ ² Lt ² )
Here, sqrt means a square root function.

In step S805, feedback control is performed so that the rotation speed NRL of the left rear wheel motor approaches the target rotation speed tNRL of the left rear wheel motor. For example, proportional control is performed as in the following equation to calculate a torque command value tTRL for the left rear wheel motor. Kp1 is a proportional gain.
tTRL = Kp1 * (tNRL-NRL)

In step S806, feedback control is performed so that the rotational speed NRR of the right rear wheel motor approaches the target rotational speed tNRR of the right rear wheel motor. For example, proportional control is performed as in the following equation to calculate the torque command value tTRR for the right rear wheel motor. Kp1 is the same proportional gain as in step S805.
tTRR = Kp1 * (tNRR-NRR)

In step S807, rear wheel steering command values tδrl and tδrr are calculated. The rear wheel steering command values tδrl and tδrr are calculated based on the vehicle rotation center (point Q in FIG. 9) corresponding to the tρ obtained in step S802 and the target slip angle βc obtained in step S803. The left rear wheel turning angle tδrl is calculated by the following equation so that the line connecting the rotation center point Q and the left rear wheel center is perpendicular to the rolling direction of the tire.
tδrl = tan ⁻¹ {sin (βc) / (cos (βc) -tρLt)}
The right rear wheel turning angle tδrr is calculated by the following equation so that the line connecting the rotation center point Q and the right rear wheel center is perpendicular to the rolling direction of the tire.
tδrr = tan ^-1 {sin (βc) / (cos (βc) + tρLt)}

  By performing the above calculation, as shown in FIG. 12, it is possible to perform an operation of turning the vehicle in a small turn with the vicinity of the rear wheel of the turning inner wheel as the turning center.

[Problems of conventional car body structure]
A typical example of an automobile that effectively absorbs energy when a vehicle collides with an obstacle and suppresses the influence on an occupant is shown in FIG. This vehicle has a configuration in which tires are arranged at four corners of the vehicle, and side members 71L and 71R are provided on the left and right sides of the front wheels in order to protect the occupant against impact from the front of the vehicle. When receiving an impact from the front, the side members 71L and 71R are crushed to absorb energy. The side members 71L and 71R are supported by the side sills 76R and 76L via members such as the structures 72R and 72L, the reinforcing materials 73R and 73L, and the reinforcing materials 74R and 74L, and receive the impact.

  However, in the conventional vehicle body structure, in order to receive the impact received from the side members 71L, 71R by the side sills 76R, 76L on the side of the cabin 1, the structures 72R, 72L, the reinforcing materials 73R, 73L, the reinforcing materials 74R, 74L, etc. A member is required, and it has been difficult to reduce the weight of the vehicle while maintaining the strength of reinforcement.

[Shock absorbing action of Example 1]
On the other hand, in the vehicle body structure of the first embodiment, as shown in FIG. 3, the impact absorbing members 69L and 69R are arranged on the vehicle front side surface, so that not only the impact received from the front but also the impact received from the oblique front is effective. Can be absorbed. Particularly, since the shock absorbing members 69L and 69R are directly joined to the side sill 61L and 61R which are high strength members on the side of the vehicle, a reinforcing material is required between the shock absorbing members 69L and 69R and the side sills 61L and 61R. Therefore, the front part of the vehicle body can be reduced in weight, and the effect of improving the running efficiency of the vehicle is also produced.

  As for the front wheels 42, one wheel does not have a steering mechanism, so that the manufacturing cost can be reduced. Moreover, the front part of the vehicle can be made lighter by being one wheel, and the running efficiency of the vehicle can be further improved.

  Furthermore, since the front wheels 42 are arranged between the left and right shock absorbing members 69L and 69R, the overall length of the vehicle is greatly suppressed as compared with a vehicle in which the front wheels are arranged in front of the left and right shock absorbing members 69L and 69R. As a result, the driver can easily drive the vehicle.

  Regarding the vehicle movement, the front wheel 42 is a single wheel and does not have a steering mechanism because the turning operation is performed by making the driving force of the outer wheels of the rear wheels 2RL and 2RR larger than the driving force of the inner wheel. It is possible to turn. A small turning operation as shown in FIG. 12 is also possible. Further, the rear wheels 2RL and 2RR have a turning function, and the turning angle can be adjusted according to the operation of the driver, so that the turning posture can be adjusted.

  Furthermore, because the front part of the vehicle body is light for the reasons described above, the center of gravity position of the vehicle body can be designed near the rear wheel, which also contributes to the improvement of the vehicle performance (the center of gravity position). The closer to the rear wheel is, the larger the turning force can be generated (in Formula A4, the value of Q33 (0) / Qden (0) increases as Lr decreases, so that the exercise performance improves).

Next, the effect will be described.
In the vehicle body structure of the automobile according to the first embodiment, the following effects can be obtained.

  (1) Since the shock absorbing members 69L and 69R are directly joined to the side sills 61L and 61R on the side of the vehicle, no reinforcing material is required between the shock absorbing members 69L and 69R and the side sills 61L and 61R. The part can be reduced in weight, and the effect of improving the running efficiency of the vehicle is also born. Furthermore, since the front wheels 42 are arranged between the left and right impact absorbing members 69L and 69R, the overall length of the vehicle can be significantly reduced as compared with a vehicle in which wheels are arranged in front of the left and right impact absorbing members 69L and 69R. As a result, the driver can easily drive the vehicle.

  (2) The rear wheels 2RL and 2RR can be independently adjusted in driving force by the electric motors 3RL and 3RR, and have an integrated controller 30 that turns the vehicle by adjusting the difference in driving force between the rear wheels 2RL and 2RR. A desired turning operation according to the operation of the driver can be realized while the front wheel 42 is a single wheel and the steering mechanism is not provided.

  (3) The integrated controller 30 that includes the steering motors 52RL and 52RR for turning the rear wheels 2RL and 2RR and the links 51RL and 51RR, adjusts the steering angle of the rear wheels 2RL and 2RR, and controls the posture of the vehicle. Therefore, the vehicle posture according to the driver's operation can be obtained.

  (4) Since the position of the center of gravity of the vehicle is set near the midpoint of the rear wheels 2RL and 2RR, the turning force can be increased and the motion performance of the vehicle can be improved.

  (5) Since the left and right impact absorbing members 69L and 69R are structural members that absorb the collision energy from the oblique direction of the vehicle, the influence on the cabin 1 with respect to the forward collision can be effectively suppressed.

  FIG. 14 is a diagram illustrating the vehicle body structure of the second embodiment, and shock absorbing members 62L and 62R are arranged on the vehicle front extension of the side sills 61L and 61R on the left and right sides of the cabin 1. FIG. The shock absorbing members 62L and 62R have a pyramid shape and have a hollow structure. In particular, the tip (front of the vehicle) portion has a structure in which the strength is weakened so that the tip is crushed from the tip.

  When the vehicle collides with an obstacle from the front, the collision energy is absorbed by the energy absorbing action by this crushing deformation, and the influence of the collision on the occupant is reduced. If the impact absorbing members 62L and 62R are in the shape of a pyramid as in this structure, it is possible to more effectively absorb the impact from obliquely forward of the vehicle.

FIG. 15 is a diagram illustrating a vehicle body structure of the third embodiment.
In the third embodiment, four wheels are arranged in a diamond shape so that the driving wheels 9L and 9R at the center in the vehicle front-rear direction can be independently motor-driven and steered, and the remaining two wheels 42 and 2 are rotary casters. It is. In addition to the shock absorbing members 62FL and 62FR provided at the front end of the vehicle, the same shock absorbing member 62RL and the like at the rear end of the vehicle are added to the three-wheel configuration shown in FIG. 62RR is provided.

  Since the caster 2 at the rear end of the vehicle hardly affects the turning operation of the vehicle, the turning operation and the turning posture of the vehicle are adjusted in the same manner as in the first embodiment by the left and right driving forces of the middle two wheels 9L and 9R and the turning amount. Can be realized. In the third embodiment, since the impact absorbing members 62RL and 62RR similar to the front end of the vehicle are provided at the rear end of the vehicle, not only the impact from the front and oblique front of the vehicle but also the impact from the rear and oblique rear of the vehicle is also effective. It absorbs automatically and protects passengers.

  In addition, since the drive wheels 9L and 9R are arranged closer to the center of the vehicle, the position of the front and rear center of gravity of the vehicle can be made closer to the vicinity of the drive wheel, the degree of freedom as a vehicle is high, and the motion performance of the vehicle is improved. There is also a feature that can be.

Next, the effect will be described.
In the vehicle body structure of the automobile according to the third embodiment, the following effects can be obtained in addition to the effects (2) to (5) of the first embodiment.

  (6) Since the shock absorbing members 62FL and 62FR on the front side of the vehicle and the shock absorbing members 62RL and 62RR on the rear side of the vehicle are directly joined to the side sills 61L and 61R on the side of the vehicle, the shock absorbing members 62FL, 62FR and 62RL , 62RR and the side sills 61L, 61R do not require a reinforcing material, the weight of the front part of the vehicle body and the rear part of the vehicle body can be reduced, and the effect of improving the running efficiency of the vehicle is also produced. Further, the front wheel 42 is disposed between the shock absorbing members 62FL and 62FR on the front side of the vehicle, and the rear wheel 2 is disposed between the shock absorbing members 62RL and 62RR on the rear side of the vehicle. Compared with a vehicle in which wheels are arranged at the rear, the overall length of the vehicle can be greatly reduced, and an effect that makes it easier for the driver to drive the vehicle can be obtained.

  FIG. 16 is a diagram illustrating a vehicle body structure of the fourth embodiment. In the fourth embodiment, the front wheels and the rear wheels in FIG. 14 are interchanged, and the turning operation and turning posture adjustment of the vehicle can be realized in the same manner as in the first embodiment by the left and right driving forces of the front wheels 42L and 42R and the turning amount. . In the fourth embodiment, the rear end of the vehicle is provided with an impact absorbing structure (impact absorbing members 62L and 62R) similar to the front end of the first embodiment. Absorb and protect passengers.

  As shown in FIG. 17, the vehicle behavior of the fourth embodiment has a feature that the rear end of the vehicle protrudes outside during turning. This is suitable as an implementation form of a vehicle that does not like the rear end of the vehicle passing inside the turn (a so-called inner ring difference) during turning.

Next, the effect will be described.
In the vehicle body structure of the automobile according to the fourth embodiment, the following effects can be obtained in addition to the effects (2) to (5) of the first embodiment.

  (7) Since the shock absorbing members 62L and 62R are directly joined to the side sills 61L and 61R on the side of the vehicle, the rear part of the vehicle body does not require a reinforcing material between the shock absorbing members 62L and 62R and the side sills 61L and 61R. The weight of the vehicle can be reduced, and the effect of improving the running efficiency of the vehicle is also produced. Furthermore, since the rear wheel 2 is disposed between the left and right shock absorbing members 62L and 62R, the vehicle can be significantly reduced in overall length as compared with a vehicle having wheels disposed behind the left and right shock absorbing members 62L and 62R. As a result, the driver can easily drive the vehicle.

  In the first to fourth embodiments, the wheels other than the drive wheels are rotating casters, but in the fifth embodiment, a case where a so-called steer-by-wire system in which the turning angle can be adjusted with an electric motor will be described.

  In this case, the turning angle may be adjusted in the direction in which the wheel position should proceed. For example, when the front wheel 42 in the first embodiment is a steer-by-wire system, it can be realized by changing the calculation routine of mode D and mode A as follows.

[Calculation routine in mode D]
FIG. 18 is a flowchart showing the flow of the calculation routine in mode D of the fifth embodiment, which differs from the first embodiment shown in FIG. 6 in that the next step S513 is executed following step S512.

In step S513, the direction βf that the front wheel 42 should travel is calculated. The forward direction βf of the front wheel 42 is calculated by the following equation as βf of the vehicle body slip angle at the front wheel position when the responses of the yaw rate γ and the slip angle β each have desirable transfer characteristics (reference model).
βf = (β + (L-Lr) γ / V)
= {m ₁ / (s ² + 2wns + wn ² )} Win + {(L-Lr) m ₂ / (s ² + 2wns + wn ² ) / V} δ… (E1)
Here, L is the distance (wheelbase length) between the front wheel 42 and the rear wheels 2L, 2R. In addition, this calculation uses the terms {m ₁ / (s ² + 2wns + wn ² )} Win and {(L-Lr) m ₂ / (s ² + 2wns + wn ² ) / V} δ as described above for yx1, yx2 , yx3, yx4, etc., respectively, and calculating by adding the sum of the terms. In order to prevent 0%, the vehicle speed V is calculated by limiting the minimum value to Vmin (for example, 1 m / s).

  Next, the electric motor for adjusting the turning angle is feedback-controlled so that the turning angle of the front wheel 42 coincides with βf.

[Calculation routine in mode A]
FIG. 19 is a flowchart showing the flow of the calculation routine in mode A of the fifth embodiment, which differs from the first embodiment shown in FIG. 8 in that the next step S808 is executed following step S807.

In step S808, the direction βf that the front wheel 42 should travel is calculated. βf is calculated as the angle C shown in FIG.
βf = tan ^-1 {tρL + sin (βc) / cos (βc)}

  Next, feedback control is performed on the electric motor for adjusting the turning angle so that the turning angle of the front wheel 42 coincides with βf.

  By performing the above processing, an equivalent movement can be realized even when the rotary caster (front wheel 42) is used as a steer-by-wire system. In the case of the steer-by-wire system, the turning angle of the front wheels 42 can be placed under control, so that the turning angle can be kept within a certain limit (for example, limited to 90 degrees left and right). Therefore, the wheel space of the front wheel 42 can be reduced, and the cabin 1 can be widened accordingly. Also, even when the vehicle attitude tends to deviate from the desired response due to changes in the road surface friction coefficient, etc., a force that approaches the desired response of the vehicle attitude acts on the front wheels, so that more stable vehicle motion performance can be achieved. Will be able to.

(Other examples)
As mentioned above, although the best form for implement | achieving this invention was demonstrated based on Examples 1-5, the concrete structure of this invention is not limited to each Example, The summary of invention is shown. Design changes and the like within a range that does not deviate are also included in the present invention.

  For example, in the embodiment, a configuration in which the force received by the front and rear shock absorbers is received by a side sill is not limited thereto, but may be a structure that is received by a structure of an A pillar or a door.

  In the embodiment, the front wheels and the rear wheels corresponding to the caster wheels are formed one by one. However, a plurality of wheels (for example, two wheels) may be arranged side by side between the left and right shock absorbers. Further, the driving mode of the vehicle is not limited to the vehicle body structure of the electric vehicle shown in the embodiment. The present invention can also be applied to a fuel cell vehicle equipped with a fuel cell and a hybrid electric vehicle.

It is a vehicle control system block diagram for making the vehicle of Example 1 drive. It is the side view and top view which show the front wheel of Example 1. FIG. It is a figure which shows the vehicle body structure of the vehicle of Example 1. FIG. 3 is a flowchart illustrating a flow of mode selection control processing executed by the integrated controller 30 of the first embodiment. It is a figure explaining the calculation method at the time of mode D in the integrated controller of Example 1. FIG. 6 is a flowchart illustrating a calculation routine in mode D according to the first embodiment. 6 is a ROM data characteristic diagram of a target driving force tTD used in a calculation routine in mode D in Embodiment 1. FIG. 6 is a flowchart illustrating a calculation routine in mode A according to the first embodiment. It is a figure explaining the calculation method at the time of mode A in Example 1. FIG. It is a ROM data characteristic view of the reciprocal value target value tρ of the turning radius used in the calculation routine in mode A in the first embodiment. FIG. 6 is a ROM data characteristic diagram of each target slip βc used in the calculation routine in mode A in the first embodiment. It is a figure which shows the turning behavior example in Example 1. FIG. It is a figure explaining a prior art. It is a figure which shows the vehicle body structure of the vehicle of Example 2. FIG. It is a figure which shows the vehicle body structure of the vehicle of Example 3. FIG. It is a figure which shows the vehicle body structure of the vehicle of Example 4. FIG. It is a figure which shows the turning behavior example of Example 4. 14 is a flowchart illustrating a calculation routine in mode D according to the fifth embodiment. 10 is a flowchart illustrating a calculation routine in mode A according to the fifth embodiment.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Cabin 2RL, 2RR Rear wheel 3RL, 3RR Electric motor 4RL, 4RR Reducer 5RL, 5RR Drive circuit 6 Lithium ion battery 8 Yaw rate sensor 11 Steering wheel 21 Steering angle sensor 22 Brake pedal sensor 23 Accelerator pedal sensor 24 Acceleration sensor 25 Shift bar 26 Displacement angle sensor 30 Integrated controller 41 Steering rotation shaft 42 Front wheel 43 Rotation center point 44 Site 45 Hollow support part 49 Front wheel rotation sensor 61L, 61R Side sill 69L, 69R Shock absorbing member

Claims (7)

The cabin where the crew rides,
Structural members respectively arranged in the longitudinal direction of the vehicle body on the left and right sides of the cabin;
A shock absorbing member that is joined to the left and right sides of the left and right structural members of the vehicle body and absorbs collision energy from the front of the vehicle,
A front wheel disposed between the left and right impact absorbing members;
Driving wheels arranged on the left and right of the rear side of the vehicle body,
A vehicle body structure characterized by comprising: The cabin where the crew rides,
Structural members respectively arranged in the longitudinal direction of the vehicle body on the left and right sides of the cabin;
A shock absorbing member that is joined to the left and right sides of the left and right structural members of the left and right structural members and absorbs collision energy from the rear of the vehicle,
A rear wheel disposed between the left and right impact absorbing members;
Driving wheels arranged on the left and right of the front side of the vehicle body;
A vehicle body structure characterized by comprising: The cabin where the crew rides,
Structural members respectively arranged in the longitudinal direction of the vehicle body on the left and right sides of the cabin;
A shock absorbing member that is joined to the left and right sides of the left and right structural members of the left and right structural members, and absorbs collision energy from the front of the vehicle body;
Front wheels disposed between left and right impact absorbing members on the vehicle front side;
A rear wheel disposed between the left and right impact absorbing members on the vehicle rear side;
Driving wheels disposed between the front wheels and the rear wheels and on the left and right sides of the vehicle body;
A vehicle body structure characterized by comprising: The vehicle body structure according to any one of claims 1 to 3,
The left and right driving wheels can independently adjust the driving force,
A vehicle body structure for an automobile, comprising driving force difference control means for adjusting the driving force difference between the left and right driving wheels to turn the vehicle. The vehicle body structure according to any one of claims 1 to 4,
Drive wheel steering means for steering the left and right drive wheels,
A vehicle body structure for an automobile, comprising vehicle attitude control means for adjusting a turning angle of the drive wheel to control the attitude of the vehicle. The vehicle body structure according to any one of claims 1 to 5,
An automobile body structure characterized in that the center of gravity of the vehicle is set near the midpoint of the left and right drive wheels. The vehicle body structure according to any one of claims 1 to 6,
The left and right impact absorbing member is a structural member that absorbs collision energy from an oblique direction of the vehicle.

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