JP3342622B2 - Loading status judgment device - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

[0001] 1. Field of the Invention [0002] The present invention relates to a loading state judging device for detecting a changing state of a loading amount in a vehicle.

[0002]

2. Description of the Related Art Conventionally, as a vehicle suspension device for controlling a damping force characteristic of a shock absorber, for example, Japanese Patent Laid-Open Publication No. Hei.
One described in Japanese Patent No. 500490 is known.

In this conventional vehicle suspension system, a dynamic vehicle running state is detected by a sensor, and a control signal for controlling a semi-active shock absorber provided on each wheel of the vehicle is formed. The vehicle control is performed according to the actual value of the force.

[0004]

However, in the conventional apparatus, since the change in the loading state in the vehicle is not considered at all, there are the following problems. That is, in this conventional device, a fixed loading state,
That is, a control gain for controlling the damping force characteristic of the shock absorber in each wheel is set in consideration of the vehicle weight, the wheel load acting on each wheel, the wheel load balance between the front wheel and the rear wheel, and the like. However, if the vehicle weight and the wheel load balance between the front and rear wheels change from the state at the time of design, the control gain for the running state of the vehicle is no longer appropriate, and optimum ride comfort and steering stability are obtained. May not be able to do so. Note that by using a vehicle height sensor separately, it is possible to detect a change in the loading state from a change in the vehicle height of the vehicle, but there is another problem that the cost increases. The present invention
It was made by focusing on the above-mentioned conventional problems, and without providing a separate vehicle height sensor, the vehicle vertical direction state quantity amplitudes of two places in the vehicle longitudinal direction detected by the vehicle vertical direction state quantity amplitude detecting means were used. It is an object of the present invention to provide a loading state determination device capable of detecting a change in a loading state in a vehicle.

[0005]

In order to achieve the above-mentioned object, a loading state judging device according to a first aspect of the present invention has a structure as shown in FIG.
As shown in the claim correspondence diagram, a pair of front and rear vehicle vertical state quantity amplitude detecting means provided at least in a predetermined distance in the front and rear direction of the vehicle and detecting the amplitude of the vertical state quantity of the vehicle respectively a ₁ , a _2, and the front-rear amplitude comparison value comparing the amplitudes of the two vehicle front-rear direction states detected by the pair of front-rear front-rear direction amplitude detectors a ₁ , a ₂ . A front / rear weight ratio change state detecting means b for detecting a front / rear weight ratio change state; and a load capacity for judging a vehicle load state change state from the front / rear weight ratio change state of the vehicle detected by the front / rear weight ratio change state detection means b. And a change judging means c. In addition, a loading state determination device according to a second aspect of the present invention is provided with a pair of vehicle front and rear which is provided at least in a front and rear direction of the vehicle at a predetermined distance and detects the amplitude of a state amount of the vehicle in the vertical direction. The position of the center of gravity of the vehicle is determined from the amplitudes of the direction state quantity amplitude detecting means a ₁ and a ₂ and the amplitudes of the _two vehicle front and rear direction states detected by the pair of front and rear vehicle vertical state amplitude detecting means a ₁ and a _2. Vehicle center-of-gravity position change state detecting means d for detecting a change state of the vehicle, and a load amount change-determining means c for determining a vehicle load amount change state from the vehicle center-of-gravity position change state detected by the vehicle center-of-gravity position change state detecting means d. Means. Further, in the loading state judging device according to the third aspect, a fixed point frequency for extracting only a component near a fixed point frequency of each vehicle vertical state amount signal to the pair of front and rear vehicle vertical state amount amplitude detecting means. Component extraction means e ₁ ,
including e _2. The fixed point frequency refers to a road surface input frequency at which the transmission rate of the road surface input to the sprung does not fluctuate due to the fluctuation of the damping coefficient of the shock absorber. Further, in the loading state judging device according to the fourth aspect, the vehicle vertical direction state quantity amplitude detected by the vehicle vertical direction state amplitude detecting means a ₁ and a ₂ is defined as the sprung vertical speed amplitude of the vehicle. Further, in the loading state judging device of the present invention, the vehicle vertical state amplitude detected by the vehicle vertical state amplitude detecting means a ₁ and a ₂ is defined as the sprung vertical acceleration amplitude of the vehicle. Further, in the loading state judging device according to the present invention, the vehicle vertical direction state amount amplitude detected by the vehicle vertical state amount amplitude detecting means a ₁ , a ₂ is defined as a vehicle sprung unsprung relative speed amplitude. did. Further, in the loading state judging device according to the seventh aspect, the vehicle vertical state amplitude detected by the vehicle vertical state amplitude detecting means a ₁ , a ₂ is calculated by:
The sprung unsprung relative acceleration amplitude of the vehicle was taken as the amplitude.

[0006]

According to the loading state determining device of the first aspect of the present invention,
As described above, the amplitude of the vertical state amount of the vehicle at both detection positions having a predetermined distance at least in the front-rear direction of the vehicle is detected, and the amplitudes of the vehicle vertical state amounts of the two front and rear portions are compared. From the front-rear amplitude comparison value, the front-rear weight ratio change state of the vehicle is obtained. When the vehicle is occupied by a single driver, and the vehicle has a minimum load capacity of zero in the luggage compartment, the front-rear weight ratio and the luggage load in the luggage compartment are large. Looking at the change in the weight ratio, there is a clearly distinguishable difference between the front and rear weight ratios depending on the load. This is a phenomenon that occurs when the luggage is loaded in the trunk room because the ratio of the load acting on the rear wheel side to the front wheel side increases, and the same phenomenon occurs when a person gets on the rear seat side. Therefore, the load amount change determining means c can determine the change in the load amount of the vehicle by checking the change in the front-rear weight ratio of the vehicle.

Further, in the loading state judging device according to the second aspect, the amplitude of the vertical state amount of the vehicle at both detection positions having a predetermined distance at least in the front-back direction of the vehicle is detected, as described above. The change state of the position of the center of gravity of the vehicle is detected from the amplitudes of the two vehicle front-rear direction states.

[0008] The center of gravity of the vehicle has a clearly distinguishable difference depending on the load of the vehicle. This is a phenomenon that occurs when the luggage is loaded in the trunk room because the ratio of the load acting on the rear wheel side to the front wheel side increases, and the same phenomenon occurs when a person gets on the rear seat side. Therefore, the load state change state of the vehicle can be determined by checking the change state of the position of the center of gravity of the vehicle by the load amount change determination unit c.

Further, in the loading state judging device according to the third aspect, as described above, the fixed point frequency component extracting means e ₁ , e
_{According to 2} , only the components near the fixed point frequency are extracted from the front and rear vehicle vertical state quantity signals, so that the change in the transmission rate of the road surface input to the sprung mass even when the damping coefficient of the shock absorber changes. Signal can be obtained.

[0010]

Embodiments of the present invention will be described with reference to the drawings.

(Embodiment 1) FIG. 2 is a structural explanatory view showing a vehicle suspension system to which a loading state judging device according to Embodiment 1 of the present invention is applied, wherein a vehicle body and four wheels are provided. Interposed, four shock absorbers SA _FL , S
A _FR , SA _RL , SA _RR (Note that in describing the shock absorber, when these four are collectively referred to, and when describing their common configuration, they are simply indicated as SA. Indicates the wheel position.
_FL indicates front wheel left, _FR indicates front wheel right, _RL indicates rear wheel left, and _RR indicates rear wheel right. ) Is provided. A vertical acceleration sensor (hereinafter referred to as a vertical acceleration sensor) that detects an acceleration G in the vertical direction is provided on the vehicle body in the vicinity (tower position) of the left and right shock absorbers SA _FL and SA _FR and the left and right shock absorbers SA _RL and SA _RR. , Upper and lower G sensor) 1
_FL , 1 _FR , 1 _RL , 1 _RR are provided, and the upper and lower G sensors 1 (1 _FL , 1 _FR , 1 _RL , 1 _RL ,
1 _RR ) and input the signal from each shock absorber S
A control unit 4 for outputting a drive control signal to the A pulse motor 3 is provided.

[0012] that shows the above structure a system block diagram of Figure 3, the control unit 4 includes an interface circuit 4a, CPU 4b, a drive circuit 4c, the interface circuit 4a, each vertical G sensors 1 _FL , 1 _FR , 1 _RL , 1 _RR , the sprung vertical acceleration G _FL ,
The _GFR , _GRL , and _GRR signals are input. 14, the sprung vertical speeds Δx _FL , Δx _FR , Δx _RL , at each tower position from the sprung vertical accelerations G _FL , G _FR , G _RL , and G _RR signals, as shown in FIG. Δx _RR and sprung-unsprung relative speed (Δx-Δx ₀ ) _FL ,
(Δx−Δx ₀ ) _FR , (Δx−Δx ₀ ) _RL , (Δx−Δ
x ₀ ) A signal processing circuit for _{obtaining the RR} and a signal processing circuit for _obtaining a determination signal for determining a change state of the load on the vehicle are provided as shown in FIG. The details of both signal processing circuits will be described later.

FIG. 4 is a sectional view showing the structure of the shock absorber SA.
A is a cylinder 30, a piston 31 that defines the cylinder 30 in an upper chamber A and a lower chamber B, an outer cylinder 33 in which a reservoir chamber 32 is formed on the outer periphery of the cylinder 30, a lower chamber B and a reservoir chamber 32. Base 34 and piston 31
A guide member 35 for guiding the sliding of the piston rod 7 connected to the outer cylinder 33, a suspension spring 36 interposed between the outer cylinder 33 and the vehicle body, and a bump rubber 37.

FIG. 5 is an enlarged sectional view showing a part of the piston 31. As shown in FIG. 5, the piston 31 has through holes 31a and 31b formed therein, and A compression-side damping valve 20 and an extension-side damping valve 12 for opening and closing 31a and 31b, respectively, are provided. A stud 38 that penetrates the piston 31 is screwed and fixed to a bound stopper 41 screwed to the tip of the piston rod 7, and the stud 38 bypasses the through holes 31a and 31b. Communication for forming flow paths (extension-side second flow paths E, expansion-side third flow paths F, bypass flow paths G, and compression-side second flow paths J to be described later) that communicate the upper chamber A and the lower chamber B. A hole 39 is formed.
An adjuster 40 for changing the flow path cross-sectional area of the flow path is rotatably provided in 9. Also, stud 38
The communication hole 3 is formed on the outer periphery of the communication hole 3 in accordance with the direction of fluid flow.
An expansion-side check valve 17 and a compression-side check valve 22 that allow and shut off the flow on the flow path side formed by 9 are provided. Note that the adjuster 40 is provided with the pulse motor 3.
Is rotated through the control rod 70 (see FIG. 4). Also, studs 38
A first port 21, a second port 13, a third port 18, a fourth port 14, and a fifth port 16 are formed in this order from the top.

On the other hand, the adjuster 40 has a hollow portion 19, a first horizontal hole 24 and a second horizontal hole 25 communicating between the inside and the outside, and a vertical groove 23 formed in the outer peripheral portion. I have.

Therefore, between the upper chamber A and the lower chamber B, a through-hole 31 is provided as a flow path through which fluid can flow in the extension stroke.
b, the inside of the extension side damping valve 12 is opened to open the lower chamber B
, The second port 13, the vertical groove 23,
Via the second port 13, the vertical groove 23, and the fifth port 16 via the fourth port 14, the outer peripheral side of the extension side damping valve 12 is opened to open the outer peripheral side of the extension side damping valve 12 to reach the lower chamber B, Then, the extension side check valve 17 is opened to open the extension side third flow path F to the lower chamber B, and the bypass to the lower chamber B via the third port 18, the second horizontal hole 25, and the hollow portion 19. There are four flow paths G. In addition, as a flow path through which a fluid can flow during the pressure stroke, the pressure side first valve that opens the pressure side damping valve 20 through the through hole 31a.
Channel H, hollow portion 19, first lateral hole 24, first port 21
, The pressure-side second flow path J that opens the pressure-side check valve 22 to reach the upper chamber A through the air passage, and the bypass flow that reaches the upper chamber A through the hollow portion 19, the second horizontal hole 25, and the third port 18. Road G
And three flow paths.

That is, the shock absorber SA is configured such that the damping force characteristic can be changed in multiple steps by rotating the adjuster 40 with the characteristics shown in FIG. 6 on both the extension side and the compression side. That is, as shown in FIG.
The state in which both pressure sides are soft (hereinafter, soft area SS
When the adjuster 40 is rotated counterclockwise from
When the damping force characteristic can be changed in multiple stages only on the extension side and the compression side is a region fixed to the low damping force characteristic (hereinafter referred to as the extension side hard region HS). Conversely, when the adjuster 40 is rotated clockwise, Only the compression side has a structure in which the damping force characteristic can be changed in multiple stages, and the extension side is a region fixed to the low damping force characteristic (hereinafter referred to as a compression side hard region SH).

In FIG. 7, the KK section, LL section, MM section, and NN section in FIG. 8, 9 and 10 and the damping force characteristics at each position are shown in FIGS.

Next, of the control operation of the control unit 4, the configuration of a signal processing circuit for obtaining the sprung vertical speed Δx and the sprung-unsprung relative speed (Δx−Δx ₀ ) is shown in FIG. Description will be made based on the drawings.

First, in B1, a phase delay compensation formula is used,
Each vertical G sensors _{1 (1 FL, 1 FR,} 1 RL, 1 RR) on each spring is detected by the vertical acceleration _{G (G FL, G FR,} G RL, G RR)
Is converted into a sprung vertical speed signal at each tower position. The general expression of the phase delay compensation can be expressed by the following transfer function expression (1).

G _(S) = (AS + 1) / (BS + 1) (1) (A <B) Then, a frequency band (0.5 Hz to 3) required for damping force characteristic control.
Hz), has the same phase and gain characteristics as the case of integration (1 / S), and as a phase lag compensation equation for lowering the gain on the low frequency side (up to 0.05 Hz), the following transfer function equation (2 ) Is used.

G _(S) = (0.001 S + 1) / (10S + 1) × γ (2) where γ is a signal and gain characteristic when speed conversion is performed by integration (1 / S). And γ = 10 in this embodiment. as a result,
The gain characteristic of the solid line in FIG.
As shown in the phase characteristic of the solid line in (b) of Fig. 5, the state where only the low-frequency gain is reduced without deteriorating the phase characteristic in the frequency band (0.5 Hz to 3 Hz) necessary for damping force characteristic control. Becomes The dotted lines (a) and (b) in FIG. 15 show the gain characteristic and phase characteristic of the sprung vertical velocity signal whose velocity has been converted by integration (1 / S).

In B2, band-pass filtering is performed to cut off components other than the target frequency band to be controlled. That is, the band-pass filter BPF is composed of a second-order high-pass filter HPF (0.3 Hz) and a second-order low-pass filter LPF (4 Hz), and has a sprung vertical velocity that targets a sprung resonance frequency band of the vehicle. Δx (Δx _FL ,
Δx _FR , Δx _RL , Δx _RR ) signals are obtained.

On the other hand, in B3, as shown in the following equation (3),
Using the transfer function Gu _(S) from each sprung vertical acceleration to the sprung-unsprung relative velocity, the vertical acceleration G (G _FL , G _FR , G _RL , G _{RR )} detected by each vertical G sensor 1. ) Signal from the sprung-unsprung relative velocity (Δx
−Δx ₀ ) [(Δx−Δx ₀ ) _FL , (Δx−Δx
₀ ) _FR , (Δx−Δx ₀ ) _RL , (Δx−Δx ₀ ) _RR ] signal is obtained. Gu _(S) = − ms / (cs + k) (3) As shown in the one-wheel model in FIG. 19, m is a sprung mass, c is a damping coefficient of a suspension, and k is a suspension. Is the Laplace operator.

Next, the content of the damping force characteristic control operation of the shock absorber SA in the control unit 4 will be described with reference to the flowchart of FIG. In addition,
This basic control is performed for each shock absorber SA _FL , SA _FR ,
This is performed for each of the SA _RL and the SA _RR .

In step 101, the sprung vertical speed Δx
Is determined to be a positive value. If YES, the flow proceeds to step 102 to control each shock absorber SA to the extension-side hard region HS. If NO, the flow proceeds to step 103.

In step 103, the sprung vertical speed Δx
Is determined to be a negative value. If YES, the routine proceeds to step 104, where each shock absorber SA is controlled to the pressure-side hard area SH, and if NO, the routine proceeds to step 105.

Step 105 is a processing step when NO is determined in steps 101 and 103, that is, when the value of the sprung vertical velocity Δx is 0,
At this time, each shock absorber SA is
To control.

Next, the operation of the damping force characteristic control will be described with reference to the time chart of FIG. The sprung vertical speed Δx is
When the value changes as shown in this figure, as shown in the figure, when the value of the sprung vertical velocity Δx is 0, the shock absorber SA is controlled to the soft area SS.

When the value of the sprung vertical velocity Δx becomes a positive value, the compression-side damping force characteristic is controlled to the expansion-side hard region HS to fix the compression-side damping force characteristic to the soft characteristic, while the expansion-side damping force characteristic (target The damping force characteristic position P _T ) is changed in proportion to the sprung vertical speed Δx based on the following equation (4).

P _T = α · Δx · K · δ (4) where α is a constant on the extension side, and K is sprung-unsprung. Gain variably set according to the relative speed (Δx−Δx ₀ ),
δ is a control gain variably set in accordance with a change in the load on the vehicle, and the details of the control for variably setting the control gain δ will be described later.

When the value of the sprung vertical velocity Δx becomes a negative value, the compression-side hard region SH is controlled to fix the expansion-side damping force characteristic to the soft characteristic while the compression-side damping force characteristic (target damping force) is set. The characteristic position P _C ) is calculated based on the following equation (5).
It changes in proportion to the sprung vertical speed Δx. P _C = β · Δx · K · δ (5) where β is a pressure-side constant.

Next, among the damping force characteristic control operations of the control unit 4, mainly the switching operation state of the control region of the shock absorber SA will be described with reference to the time chart of FIG.

In the time chart of FIG.
Is a state in which the sprung vertical speed Δx is reversed from a negative value (downward) to a positive value (upward). At this time, the relative speed (Δx−Δx ₀ ) is still a negative value (shock absorber SA).
Is a pressure stroke side), and at this time, the shock absorber SA is controlled to the extension side hard region HS based on the direction of the sprung vertical speed Δx,
Therefore, in this area, the shock absorber SA at that time is
The pressure stroke side, which is the stroke of, has soft characteristics.

In region b, the sprung vertical speed Δx remains a positive value (upward), and the sprung-unsprung relative speed (Δx−Δx ₀ ) changes from a negative value to a positive value (shock absorber). In this case, the shock absorber SA is controlled to the extension-side hard region HS based on the direction of the sprung vertical speed Δx, and the shock absorber is switched to the extension-side hard region HS. Is also an extension stroke. Therefore, in this region, the extension stroke which is the stroke of the shock absorber SA at that time has a hardware characteristic proportional to the value of the sprung vertical speed Δx.

Region c is a state in which the sprung vertical speed Δx is reversed from a positive value (upward) to a negative value (downward). At this time, the sprung-unsprung relative speed (Δx
−Δx ₀ ) is a region having a positive value (the stroke of the shock absorber SA is on the extension stroke side). At this time, the shock absorber SA sets the compression-side hard region SH based on the direction of the sprung vertical speed Δx. , And therefore
In this region, the extension stroke side, which is the stroke of the shock absorber SA at that time, has soft characteristics.

In the area d, the sprung vertical speed Δx remains a negative value (downward), and the sprung-unsprung relative speed (Δx−Δx ₀ ) changes from a positive value to a negative value (shock absorber). At this time, the shock absorber SA is controlled to the compression-side hard region SH based on the direction of the sprung vertical velocity Δx, and the stroke of the shock absorber is also increased. Pressure stroke,
Therefore, in this area, the shock absorber SA at that time is
The pressure stroke side, which is the stroke of the above, has hardware characteristics proportional to the value of the sprung vertical speed Δx.

As described above, in this embodiment, the sprung vertical speed Δx and the sprung-unsprung relative speed (Δx−Δ
x ₀ ) has the same sign (region b, region d), the stroke side of the shock absorber SA at that time is controlled to a hard characteristic, and when the sign is different (region a, region c), the shock absorber at that time is controlled. The same control as the damping force characteristic control based on the skyhook control theory, in which the stroke side of the SA is controlled to the soft characteristic, is performed based only on the sprung vertical speed Δx signal. Further, in this embodiment, when the stroke of the shock absorber SA is switched, that is, when the transition is made from the area a to the area b and from the area c to the area d (from the soft characteristic to the hard characteristic), the switching step is performed. The damping force characteristic position on the side
c, the switching from the soft characteristic to the hard characteristic is performed without a time delay because the switching to the hardware characteristic has already been performed. As a result, a high control response can be obtained and the hardware characteristic can be switched from the soft characteristic to the hard characteristic. The switching to is performed without driving the pulse motor 3, thereby improving the durability of the pulse motor 3 and
Power consumption will be saved.

Next, among the control operations of the control unit 4, the contents of the variable setting control of the control gain δ based on the change in the load on the vehicle will be described. First, a configuration of a signal processing circuit for obtaining a judgment signal for judging a change state of the load capacity will be described with reference to a block diagram of FIG.

At C1, the sprung vertical speed Δx at the front wheel center position is calculated from the average values of the front left and right sprung vertical speeds Δx _FL and Δx _FR obtained by the signal processing circuit shown in the block diagram of FIG. On the other hand, in C5, _FS is obtained, and from the average value of the rear wheel side left and right sprung vertical speeds Δx _RL and Δx _RR signals,
The sprung vertical speed Δx _RS at the rear wheel center position is determined.

At C2 and C6, the components V _FL and V _RL below the sprung resonance frequency are calculated from the sprung vertical speed Δx _FS signal at the front wheel center position or the sprung vertical speed Δx _RS signal at the rear wheel center position. Low-pass filter processing for extraction is performed. That is, in this low-pass filter processing, a 0.6 Hz second-order low-pass filter LPF on the front wheel side and a rear-wheel side
A 0.8 Hz secondary low-pass filter LPF is used.

In the following C3 and C7, the peak values V _FP and V _RP of the absolute values | V _FL | and | V _RL | of the components V _FL and V _RL below the sprung resonance frequency are obtained. That is, this peak value V _FP ,
V _RP corresponds to the amplitude of the state quantity of the vehicle in the vertical direction. Therefore, C3, C7 and the signal processing circuit of FIG.

In the following C4 and C8, the absolute values of the components V _FL and V _RL below the sprung resonance frequency | V _FL | by passing through a second-order low-pass filter LPF (0.1 Hz).
, | V _RL | peak value V _FP , V _RP moving average value V _MF ,
_Find V _MR respectively.

Finally, in C9, the ratio of the rear wheel side moving average value V _{MR to} the front wheel side moving average value V _MF (V _MR /
V _MF ), the amplitude ratio (≒ weight ratio R _M ) of both sprung vertical speed signals at the front and rear positions of the vehicle is obtained.

That is, the C9 portion constitutes a front / rear weight ratio change state detecting means for detecting a front / rear weight ratio change state in the vehicle according to the first aspect of the present invention.

[0046] From the amplitude ratio R _M of both the vertical sprung mass velocity signals in the above manner the obtained vehicle longitudinal position, the front and rear weight ratio of the rear wheel center with respect to the front wheel center position of the vehicle can be determined in a simple manner The reason will be described below.

In the one-wheel model shown in FIG. 19, the amplitude X of the sprung behavior x can be expressed by the following equation (1).
Here, m is the sprung mass, k is the spring constant of the suspension, c is the damping coefficient of the suspension, and u is the road surface input (=
A sinωtω = 2πf).

[0048]

[Formula 1] In this equation 1, if k−mω ² > 0 (6), as the sprung mass m increases, the value of km−mω ² decreases, and the amplitude X of the sprung behavior x is increased, conversely, the sprung mass m is decreased, the value of k-milliohms ² increases, the amplitude X of the sprung mass behavior x is established a relationship that reduces.

The condition of the above equation (6) is given by the following equation (2).
3 can be replaced.

[0050]

[Formula 2]

[0051]

[Equation 3] Therefore, if the frequency components are limited to the sprung resonance frequency or lower, the “front wheel side sprung mass / rear wheel side sprung mass” is ...
························································································· The amplitude of the side sprung speed increases.

Contrary to the above, when “the front wheel side sprung mass / rear wheel side sprung mass” decreases.
"Amplitude of front wheel side sprung displacement / rear wheel side sprung displacement"
Decrease, and "amplitude of front wheel side sprung speed / amplitude of rear wheel side sprung speed" decreases.

The following relationship holds.

[0054] Thus, the front and rear weight ratio between the front wheel side weight and rear wheel weight of the vehicle can be simplified manner determined by the amplitude ratio R _M of the front wheel side sprung speed and the rear wheel side sprung velocity (weight Ratio ≒ amplitude ratio _RM ).

FIG. 21 shows the amplitude ratio (≒ weight ratio) R _M
In this figure, the time zone (I) is a time zone (I) when the vehicle is driven by a single driver, and the vehicle is running in the minimum loading capacity state where the luggage in the trunk room is 0. In the time zone (II), the amplitude ratio (≒ weight ratio) R when the vehicle runs with a large amount of luggage loaded in the luggage compartment is shown, while the amplitude ratio (≒ weight ratio) R _M is shown in a fluctuation state. _This shows the variation state of _M , and it can be seen that the level of the amplitude ratio (≒ weight ratio) RM has a clearly distinguishable difference depending on the loading amount. This is a phenomenon that occurs when the luggage is loaded in the trunk room because the ratio of the load acting on the rear wheel side to the front wheel side increases, and the same phenomenon occurs when a person gets on the rear seat side.

[0056] Therefore, by setting the payload decision threshold R _L to the middle position of both levels, it compares the amplitude ratio (≒ weight ratio) R _M and this payload decision threshold R _L, the vehicle It is possible to detect a change in the load amount. That is, the part that performs the above operation (step 20 in FIG. 20)
2) constitutes the loading amount change determination means in the first aspect of the present invention.

Hereinafter, the contents of the switching control operation of the damping force characteristic based on the change in the load amount will be described with reference to the flowchart of FIG. 20 and the time chart of FIG. First, FIG.
In the flowchart of 0, in step 201, it is determined whether or not the loading determination flag Flag has been reset to 0.0. If YES, the process proceeds to step 202.

[0058] In step 202, the amplitude ratio (≒ weight ratio) R _M, it is determined whether or not exceeds the payload decision threshold R _L, since the amount of loading is likely to have increased if YES Then, the process proceeds to step 203 to start the timer (Tt = Time−T _ON ), and then proceeds to step 204.

In this step 204, it is determined whether or not the timer count Tt has exceeded a predetermined determination time Δt. If YES, it is certain that the load has increased, so the flow proceeds to step 205 to determine the load. Flag
Is set to 1.0, and the routine proceeds to step 206. Then, in this step 206, after switching to the loading control parameters, the process proceeds to step 207.

In this step 207, it is determined whether or not the door of the vehicle is open. If the answer is YES, the load capacity may fluctuate. The loading determination flag Flag is reset to 0.0 in order to perform the loading amount determination of steps 202 to 205, and one control flow is completed.

In step 201, NO (Fla
g = 1.0), the loading amount determination in steps 202 to 205 is omitted to continue the loading control, and the process proceeds to step 206.

If the result of the determination in step 202 or step 204 is NO, there is no increase in the load amount, and the process proceeds to step 209 to switch to the normal control parameters. The control flow ends.

If the determination in step 207 is NO, there is no possibility that the load capacity fluctuates, so that one control flow is completed. Thereafter, the above control flow is repeated.

Next, the contents of the switching control operation of the damping force characteristic based on the change in the load amount will be described with reference to the time chart of FIG. (A) When the loading capacity is minimum When the vehicle is driven by a single driver and has the minimum loading capacity of zero luggage in the trunk room, as shown in the time chart of FIG. The amplitude ratio (≒ weight ratio) _RM is equal to or less than the loading amount determination threshold value _RL .
Is switched to the normal control parameter. That is, the control gain δ in the equations (4) and (5) for obtaining the target damping force characteristic positions P _T and P _C is set to the basic gain δ _M. In the above, the optimal damping force characteristic control based on the skyhook control theory is performed, and the riding comfort and the steering stability of the vehicle can be ensured.

(B) When the loading capacity is increased When traveling starts with a large amount of luggage loaded in the trunk room from the aforementioned minimum loading capacity, a loading capacity determination signal is generated as shown in the time chart of FIG. the amplitude ratio (≒ weight ratio) R _M exceeds the load amount determining threshold R _L, in which the state continues for a predetermined determination time Delta] t, at this time, to the loading time of the control parameter in step 206 Switching is performed. That is, the target damping force characteristic position P _T, the equation for obtaining the _{P C (4), (5} )
Control gain [delta] is, is variably set to the basic gain [delta] _M than the raised correction gain [delta] _H, thereby, the target damping force characteristic position P _T, a state of being set to be higher is P _C in. In this case, since becomes large increase in the weight of the rear wheel side from the front wheel side, the value of the correction gain [delta] _H is set higher is better for the rear wheels than the front wheel side. Further, once the switching to the loading control parameter is performed, the switching is continued at least until the door is opened thereafter.
Therefore, it is possible to prevent the deterioration of the riding comfort and the steering stability of the vehicle due to the increase in the load amount by using the high damping force characteristic which is automatically variably set.

As described above, the loading state determining apparatus according to the first embodiment of the present invention has the following effects. Without providing a separate vehicle height sensor, it is possible to detect a change in the loading state of the vehicle from two sprung vertical speed signals obtained from the sprung vertical acceleration signals detected by the respective vertical G sensors 1 in the vehicle longitudinal direction. Thus, the cost can be reduced.

As a means for converting the sprung vertical acceleration to the sprung vertical velocity, a phase lag compensation formula is used to prevent signal drift based on extra low-frequency signal input, such as during braking. Thus, it is possible to prevent the controllability of the damping force characteristic of the shock absorber SA from deteriorating and to secure the riding comfort of the vehicle.

The switching from the soft characteristic to the hard characteristic is performed without time delay, whereby a high control response is obtained, and the switching from the hardware characteristic to the soft characteristic is performed without driving the actuator. This makes it possible to improve the durability of the actuator and save power consumption.

Next, another embodiment of the present invention will be described. In the description of the other embodiments, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof will be omitted. Only different points will be described.

(Embodiment 2) Embodiment 2 of the present invention differs from Embodiment 1 of the present invention in the method of calculating a load amount determination signal for determining a change in the load amount. Hereinafter, a method of calculating the load amount determination signal will be described.

FIG. 22 is a diagram for explaining the state of movement of the position of the center of gravity G based on the change in the load on the vehicle.
As shown in this figure, the center of gravity G from the front tower position
The distance L to the position can be obtained by the following equation (7). Here, mf is a front wheel side sprung mass, mr is a rear wheel side sprung mass, and Lw is a wheel base. L = (mr / (mf + mr)) · Lw (7) Therefore, as in the case of the weight ratio, it can be approximated by the following equation (4).

[0072]

(Equation 4) That is, in the first embodiment of the invention, at C9 of the flowchart of FIG. 18, from the front wheel side moving average V _MF and the rear-wheel-side moving average value V _MR, the amplitude ratio (≒ weight ratio) to obtain the R _M However, in the second embodiment of the present invention, the following equation (8)
, An approximate value of a distance L (a state of fluctuation of the position of the center of gravity) from the position of the front wheel tower to the position of the center of gravity G is obtained.

L ≒ (V _MF / (V _MF + V _MR )) · Lw (8) That is, the C9 portion is the center of gravity of the vehicle according to the second aspect of the present invention. It constitutes a vehicle center-of-gravity position change state detecting means for detecting a change state of the position.

Therefore, the content of the determination at step 202 in the flowchart of FIG.
By comparing the distance L to the position with a reference value, it is possible to determine the state of change in the load amount. That is, the part that performs the operation of step 202 constitutes the load amount change determining means in the second aspect of the present invention. Therefore, also in the second embodiment of the present invention, the same effect as in the first embodiment of the present invention can be obtained.

(Embodiment 3) Embodiment 3 of the present invention differs from Embodiments 1 and 2 of the invention in the method of calculating a load amount determination signal for determining a change in load amount. intended to, hereinafter, the configuration of a signal processing circuit for determining the loading state determination signal R _M used in the payload of the changing state determination will be described based on the block diagram of FIG. 23.

First, at D1, an average value of the front-wheel-side left and right sprung vertical accelerations G _FL and G _FR detected by the front wheel-side left and right up-and-down G sensors _1FL and _1FR is calculated from the average value of the front-wheel-side center position at the front wheel side. On the other hand, the acceleration GFS is obtained. At D5, the center value of the rear wheel side left and right upper and lower sprung vertical accelerations _GRL and _GRR detected by the rear wheel side left and right upper and lower G sensors 1 _RL and 1 _RR is calculated. The sprung vertical acceleration GRS of the position is obtained.

In subsequent D2 and D6, a two-stage band pass for extracting the sprung vertical acceleration GFS at the front wheel center position or the rear wheel center position, and the fixed point frequency components GFS- _H and GRS- _H of the GRS signal. Filter BPF1 and BPF2 processes are respectively performed. In other words, the band-pass filter processing includes a fixed point frequency of 1.4 Hz when the vehicle is normally loaded.
-Order band-pass filter B whose cut-off frequency is
PF1 and the fixed point frequency when the vehicle is heavily loaded 1.
A second-order bandpass filter BPF2 having a cutoff frequency of 2 Hz is used.

Then, the two filters BPF1, BPF
As shown in FIG. 24, as shown in FIG. 24, a small damping value (Q = 5) is used in order to emphasize the intended gain in the 1.2 to 1.4 Hz portion, as shown in the gain characteristic diagram with respect to frequency.

In subsequent D3 and D7, low-frequency processing signals GFS- _L and GRS indicating the fluctuation state of the amplitude between the positive peak value and the negative peak value in the fixed point frequency components GFS- _H and GRS- _H . _-L is requested. That is, as shown in FIG. 25A, the positive and negative peak values of the fixed point frequency components GFS- _H and GRS- _H were respectively detected and stored, and the next peak value was detected. At that time, the peak value is sequentially updated. Then, each time a positive or negative peak value is detected, the negative peak value is subtracted from the positive peak value, and the resulting value is stored in a memory to store a low frequency signal as shown in FIG. Processing signal GF
S- _L and GRS- _L are obtained.

In D4 and D8, the low-frequency processed signals GFS- _L and GRS- _L are processed by a low-pass filter LPF having a cut-off frequency of 0.05 Hz to thereby obtain a moving-averaged low-frequency processed signal Af, Ar is obtained.

Finally, at D9, the two low-frequency processed signals A
f, from Ar, on the basis of the following equation (9), determining the loading state determination signal R _M corresponding to the amplitude ratio of both sprung mass vertical acceleration signals in the vehicle front and rear positions. R _M = Af / Ar (9) That is, the signal processing circuit shown in FIG. 23 constitutes a loading amount determination signal detecting means in the claims.

From the amplitude ratio of the two sprung vertical acceleration signals at the vehicle front-rear position (loading state determination signal _RM ) obtained as described above, the front-rear weight ratio of the rear wheel center position to the front wheel center position of the vehicle can be simply calculated. The reason will be described below.

When the behavior of the vehicle is considered in the model of FIG. 19, the transmission rate x / u from the road surface input u to the sprung mass is as shown in FIG.

FIG. 26 (a) shows the sprung transmissibility characteristics with respect to the road surface input frequency on the front wheel side, and FIG.
Indicates the transmissibility characteristics of the sprung mass with respect to the road input frequency on the rear wheel side.In both figures, the solid line shows the transmissivity characteristics at the time of standard loading, and the dotted line shows the transmissivity characteristics at the time of full loading. It is.

As is clear from these characteristic diagrams, the sprung transmissibility, that is, the sprung behavior level is compared between the standard loading and the full loading, as shown in FIG. 26A for the front wheels. 26, while the rear wheel side shows a large level difference as shown in FIG. 26 (b). In particular, when looking at the road surface input frequency, the high-frequency side of 1 Hz or higher has a high standard load (solid line) and a full load (dotted line).
The level difference occurs in the direction in which is lower.

[0086] Then, for example, when a 1.4 Hz component of the sprung behavior x is extracted, if the sprung mass m becomes large, ...
When the transmissibility (x / u) decreases and the sprung mass m decreases, the relationship is such that the transmissivity (x / u) increases. Assuming that the front wheel side road surface input uf and the rear wheel side road surface input ur are the same, the load ratio (mr / mf) ≒ (xf / uf) / (xr / ur)
) = Xf / xr.

The low-frequency processed signal Ar on the rear wheel side corresponds to the large level fluctuation at 1.2 to 1.4 Hz in (b) of FIG. 26, and the level change on the front wheel side in (a) of FIG. The low-frequency processing signal Af on the front wheel side corresponds to a small level fluctuation, and therefore, the loading state determination signal _RM
The value of (= Af / Ar) changes in proportion to the change in the load of the vehicle.

Therefore, the front / rear weight ratio between the front wheel side weight and the rear wheel side weight of the vehicle is determined by the amplitude ratio between the front wheel side sprung vertical acceleration and the rear wheel side sprung vertical acceleration (loading state determination signal R
_M ) can be obtained simply (weight ratio / amplitude ratio).

Next, the meaning of the fixed point frequency used in the third embodiment of the present invention will be described. This fixed point frequency means that the fixed point frequency is applied to the spring of the road surface input even if the damping coefficient of the shock absorber SA fluctuates. Is the road input frequency at which the transmissibility of the road does not change.

FIG. 27 shows the sprung transmissivity characteristics with respect to the road surface input frequency. The dotted line shows the transmissivity characteristics when the damping coefficient is large, and the solid line shows the transmissivity when the damping coefficient is small. It is a characteristic. Therefore, when the model shown in FIG. 19 is considered, the fixed point frequency can be obtained by the following equation 5 when the sprung resonance frequency is ωn.

[0091]

(Equation 5) Then, the sprung transmission rate at that time is 0 [dB].

Therefore, when the loading state is determined based on the sprung amplitude level, the loading is more accurately determined by making a determination using a component near the fixed point frequency which is not affected by the damping coefficient of the shock absorber SA. The effect is obtained that the quantity can be determined.

(Embodiment 4) In Embodiment 4 of the present invention, Embodiments 1 to 3 of the present invention perform switching in only two stages of control parameters during normal operation and control parameters during loading. On the other hand, the control parameters are continuously switched in accordance with the fluctuation of the amplitude ratio (≒ weight ratio) R _M (or the distance L from the position of the front wheel tower to the position of the center of gravity G) which is the load amount determination signal. FIG. 2
8 is a flowchart showing the contents of the control parameter switching setting control, and the amplitude ratio (≒ weight ratio) R in FIG.
_A description will be given based on a time chart showing the content of the control setting for switching the control gain with respect to _M.

First, in the flowchart of FIG.
In step 301, the control gain δf for the front wheels is
The value obtained by multiplying the amplitude ratio (≒ weight ratio) _RM by the front wheel side constant a (=
a · R _M ), and in step 302, a value obtained by multiplying the amplitude ratio (≒ weight ratio) R _M by a rear wheel-side constant b (= b · R _M ) as the rear wheel-side control gain δr. Set. Both constants a and b are positive numbers, and the rear wheel constant b is set to a value larger than the front wheel constant a (0 <a <
b).

In the following step 303, the above equations (4) and (5)
By applying the control gains δf and δr to the target damping force characteristic positions P _T and P _C of the front and rear shock absorbers SA in both the expansion strokes, respectively. Control gain delta] f As described above, by performing switching setting of [delta] r, as shown in the time chart of FIG. 29, the amplitude ratio steplessly relative variation of (≒ weight ratio) R _M and continuously controlled gain delta] f, δr is switched. Therefore, according to the fourth embodiment of the present invention, it is possible to perform fine control of correction of control parameters according to a change in the load amount.

Although the embodiments of the present invention have been described above, the specific configuration is not limited to the embodiments of the present invention, and there are design changes and the like within a range not departing from the gist of the present invention. This is also included in the present invention.

For example, in the embodiment of the invention, the case where the sprung vertical velocity amplitude is used as the vehicle vertical state variable amplitude detected by the vehicle vertical direction variable amplitude detecting means has been described. Vertical acceleration amplitude, sprung unsprung relative velocity amplitude, sprung unsprung relative acceleration amplitude, and the like can also be used.

Using the relationship in the case of A (Formula 6 below) in Formula 1, the component may be extracted by extracting a component on the higher frequency side (for example, 3 Hz) than the fixed point frequency.

[0099]

(Equation 6) Further, in the embodiment of the invention, the change in the load amount is always determined until the control parameter is switched to the control parameter at the time of loading. However, the control parameter is switched to the normal control parameter until the door and / or the trunk of the vehicle is opened. The state may be maintained, and when the door and / or the trunk of the vehicle is opened, the vehicle may travel thereafter to start the change in the load capacity.

Further, the rate of change of the vehicle speed is obtained from the vehicle speed signal detected by the vehicle speed sensor. When the rate of change of the vehicle speed exceeds a predetermined threshold value, the determination of the change in the load is stopped. Is also good.

Further, in the embodiment of the present invention, the variable setting of the control gain is performed for all the shock absorbers on the front wheel side and the rear wheel side. However, the control gain can be changed only on either the front wheel side or the rear wheel side. Good.

Further, in the embodiment of the present invention, the soft region SS is controlled only when the sprung vertical speed signal is 0. However, a predetermined dead zone centered on 0 is provided and the spring is controlled within the range of the dead zone. By maintaining the damping force characteristic in the soft region SS while the up-down speed is changing, control hunting can be prevented.

[0103]

As described above, the first aspect of the present invention is as follows.
In the loading state determination device described above, as described above, at least a predetermined distance is provided in the front-rear direction of the vehicle, and each of the loading state determination devices detects the amplitude of the vertical state amount of the vehicle.
A pair of vehicle vertical state magnitude amplitude detecting means and a pair of vehicle vertical direction state quantity amplitude detecting means and a pair of vehicle vertical direction state quantity amplitude detecting means are used to compare the amplitudes of the two vehicle front and rear vertical state quantities with the longitudinal amplitude comparison value. A front / rear weight ratio change state detecting means for detecting a front / rear weight ratio change state, and a load capacity change determination for judging a vehicle load amount change state from the front / rear weight ratio change state of the vehicle detected by the front / rear weight ratio change state detection means. Means, and without providing a separate vehicle height sensor, the vehicle vertical direction state quantity amplitude detected at two vehicle longitudinal directions detected by the vehicle vertical direction state amplitude detection means in the vehicle. The effect that the change of the loading state can be detected can be obtained.

Further, in the loading state judging device according to the second aspect of the present invention, as described above, at least a predetermined distance is provided in the front-rear direction of the vehicle, and the amplitude of the vertical state amount of each of the vehicles is determined. A change in the center of gravity of the vehicle based on a pair of front and rear vehicle vertical state amplitude detection means, and two vehicle front and rear vehicle vertical state amplitudes detected by the pair of front and rear vehicle vertical state amplitude detectors. A vehicle center-of-gravity position change state detecting means d for detecting a state; and a load amount change determining means c for determining a vehicle load-amount change state from the vehicle center-of-gravity position change state detected by the vehicle center-of-gravity position change state detecting means d.
The same effects as those of the loading state determination device according to the first aspect can be obtained.

In the loading state judging device according to the third aspect of the present invention, the pair of front and rear vehicle vertical state state amplitude detecting means detects only a component near a fixed point frequency in each vehicle vertical state state signal. By adopting the means including the fixed point frequency component extracting means to be extracted, a signal without fluctuation of the transmissibility of the road surface input to the sprung even if the damping coefficient in the shock absorber changes, is obtained. The effect is obtained that accurate loading state determination can be performed.

[Brief description of the drawings]

FIG. 1 is a diagram corresponding to a claim showing a loading state determination device of the present invention.

FIG. 2 is a configuration explanatory view showing a first embodiment of a loading state judging device of the present invention.

FIG. 3 is a system block diagram showing a first embodiment of a loading state judging device according to the present invention;

FIG. 4 is a sectional view showing a shock absorber applied to the first embodiment of the invention.

FIG. 5 is an enlarged sectional view showing a main part of the shock absorber.

FIG. 6 is a damping force characteristic diagram corresponding to a piston speed of the shock absorber.

FIG. 7 is a damping force characteristic diagram corresponding to a step position of a pulse motor of the shock absorber.

FIG. 8 is a perspective view of the shock absorber shown in FIG.
It is -K sectional drawing.

FIG. 9 is a perspective view of the shock absorber shown in FIG.
It is an L sectional view and MM sectional view.

FIG. 10 is a sectional view taken along line NN of FIG. 5 showing a main part of the shock absorber.

FIG. 11 is a damping force characteristic diagram when the shock absorber is on the extension side hard.

FIG. 12 is a damping force characteristic diagram of the shock absorber in a soft state on an extension side and a compression side.

FIG. 13 is a damping force characteristic diagram of the shock absorber in a pressure-side hard state.

FIG. 14 is a block diagram showing a signal processing circuit for obtaining a sprung vertical speed and a sprung-unsprung relative speed according to the first embodiment of the present invention;

FIG. 15 is a diagram showing a gain characteristic (a) and a phase characteristic (b) of a sprung vertical velocity signal obtained by the signal processing circuit according to the first embodiment of the present invention.

FIG. 16 is a flowchart showing the details of the damping force characteristic control operation of the control unit according to the first embodiment of the present invention.

FIG. 17 is a time chart showing the details of the damping force characteristic control operation of the control unit according to Embodiment 1 of the present invention.

FIG. 18 is a block diagram showing a signal processing circuit for obtaining an amplitude ratio (≒ weight ratio) as a load amount determination signal according to the first embodiment of the present invention.

FIG. 19 is a one-wheel model diagram for deriving a transfer function from each sprung vertical acceleration to the sprung-unsprung relative velocity in the first embodiment of the invention.

FIG. 20 is a flowchart showing the contents of a switching control operation of a damping force characteristic (control parameter) based on a change in a load amount according to the first embodiment of the invention;

FIG. 21 is a time chart showing the contents of a switching control operation of a damping force characteristic (control parameter) based on a change in a load amount in the first embodiment of the present invention.

FIG. 22 is an explanatory diagram for deriving an equation for calculating a distance to the position of the center of gravity in the second embodiment of the present invention.

FIG. 23 is a block diagram showing a signal processing circuit for obtaining an amplitude ratio (≒ weight ratio) as a load amount determination signal according to Embodiment 3 of the present invention.

FIG. 24 is a graph showing gain characteristics with respect to frequency of bandpass filters BPF1 and BPF2 used in the third embodiment of the present invention.

FIG. 25 is a time chart for explaining how to obtain a low-frequency processing signal in the third embodiment of the present invention.

FIG. 26 is a graph showing sprung transmissivity characteristics with respect to a road surface input frequency on the front wheel side (a) and the rear wheel side (b) in the third embodiment of the invention.

FIG. 27 is a graph showing a sprung transmissivity characteristic with respect to a road surface input frequency for explaining a fixed point frequency in the third embodiment of the present invention.

FIG. 28 is a flowchart showing the contents of a switching control operation of a damping force characteristic (control parameter) based on a change in a load amount according to Embodiment 4 of the present invention.

FIG. 29 is a time chart showing the contents of a switching control operation of a damping force characteristic (control parameter) based on a change in a loaded amount in Embodiment 4 of the present invention.

[Explanation of symbols]

a ₁ vehicle vertical direction state amplitude detecting means a ₂ vehicle vertical direction state amplitude detecting means b front / rear weight ratio change state detecting means c loading amount change determining means d vehicle center of gravity position change state detecting means e ₁ fixed point frequency component extracting means e ₂ fixed point frequency component extracting means

──────────────────────────────────────────────────続 き Continued on front page (58) Field surveyed (Int. Cl. ⁷ , DB name) B60G 17/00 B60G 17/015

Claims (7)

(57) [Claims] 1. A pair of front and rear vehicle vertical state amount amplitude detecting means provided at a predetermined distance in at least the front and rear direction of the vehicle and detecting the amplitude of the vertical state amount of the vehicle, respectively. A front / rear weight ratio change state detection for detecting a front / rear weight ratio change state of a vehicle from a front / rear amplitude comparison value obtained by comparing amplitudes of two front / rear vehicle vertical state quantities detected by a pair of vehicle vertical state amplitude detection means. Means, and a load change determining means for determining a load change state of the vehicle based on the change in the front / rear weight ratio of the vehicle detected by the front / rear weight ratio change state detecting means. Judgment device. 2. A pair of front and rear vehicle vertical state amplitude detecting means provided at a predetermined distance at least in the front and rear direction of the vehicle and detecting the amplitude of the vertical state amount of the vehicle, respectively. Vehicle center-of-gravity position change state detection means for detecting a change state of the vehicle center-of-gravity position from the amplitudes of the two vehicle front-rear direction state quantities detected by the pair of vehicle vertical state-amount amplitude detection means; A load state change judging device, comprising: a load state change judging means for judging a load state change state of the vehicle from a vehicle center of gravity position change state detected by the state detection means. 3. A fixed point frequency component extracting means for extracting only a component near a fixed point frequency of each vehicle vertical state amount signal in said pair of front and rear vehicle vertical state amount amplitude detecting means. The loading state determination device according to claim 1 or 2, wherein: 4. The loading state judging device according to claim 1, wherein the vehicle vertical state amplitude detected by the vehicle vertical state amplitude detecting means is a sprung vertical speed amplitude of the vehicle. . 5. The loading state judging device according to claim 1, wherein the vehicle vertical state amplitude detected by the vehicle vertical state amplitude detecting means is a sprung vertical acceleration amplitude of the vehicle. . 6. The load according to claim 1, wherein the vehicle vertical direction state amplitude detected by the vehicle vertical direction state amplitude detecting means is a relative sprung unsprung speed amplitude of the vehicle. State determination device. 7. The load according to claim 1, wherein the vehicle vertical state amplitude detected by the vehicle vertical state amplitude detecting means is a relative sprung unsprung acceleration amplitude of the vehicle. State determination device.