JPS5823267B2 - 2 wheel drive single track motor vehicle - Google Patents

Description

[Detailed description of the invention] The present invention relates to a two-wheel drive single-rail motor vehicle.

A single-rail transport vehicle is a transport vehicle that runs on a single track (mole rail) using a power vehicle, a bogie, and force.

Rails are easy to install and have the advantage of being able to pass through narrow spaces.

It is the best means of transportation for hilly areas such as orchards.

For example, single-rail transport vehicles with a capacity of 200 kg are widely used in mandarin orange orchards.

The structure of a power vehicle consists of a drifting mechanism, a power transmission mechanism, a drive pinion, upper and lower wheels that support the power vehicle on rails, and a frame that houses these.

A rack is provided continuously on the side or underside of the rail.

The drive pinion of the motor vehicle meshes with the rack.

When the drive pinion rotates under the rotational torque from the drifting mechanism, it pushes the rack rearward, allowing the motor vehicle to move forward.

In the case of a light load of about 200 kg, one drive pinion is sufficient.

However, single-gauge transport vehicles are beginning to be recognized for their usefulness in civil engineering applications.

This is because it faces slopes and runs unmanned.

For example, it can be used when building sand walls on slopes of mountains.

In this case, it is used to repeatedly transport ready-mixed concrete from the foothills to the site.

A conventional one-wheel drive single-rail transport vehicle weighs no more than 500 kg.
It can only transport ready-mixed concrete.

With a weight of 500 kg, it was poured before the next fresh concrete was transported; this was inconvenient as the mixed fresh concrete dried out.

Furthermore, even if the ready-mixed concrete truck weighs less than 2 tons, it takes too much time to empty it, making it uneconomical.

This cannot be said to be an indication of the labor-saving effects of using single-rail transport vehicles in civil engineering work.

Why is 500kg the limit? That being said, it lies in the strength of the rack.

A rack is welded to the rail of a single-gauge transport vehicle, and the drive pinion meshes with this rack.

A rack is a long board curved into a corrugated shape, and cannot withstand huge traction loads.

Due to the strength of the rack, approximately 500 kg has been considered the limit.

To increase the strength of the rack, use strong and thick steel.
All you have to do is make a full-fledged rack.

However, this would increase the cost of the rails and make no practical sense.

It is most desirable to be able to transport heavy objects while using rails with conventional racks.

For this purpose, it is sufficient to use two drive pinions instead of one.

Since there are two engagement points between the output pinion and the rack,
It should be able to withstand twice the traction force of conventional ones.

In other words, if a conventional rack can withstand 500kg in one place, increasing the number of drive pinions to two will support 1000kg.
It should be possible to create something that can withstand a load of 100 g.

But that wasn't the case.

Simply using two-wheel drive does not always apply equal force at the engagement point between the drive pinion and rack.

Plus, the pitch of the rack is not always accurate;
This is because there is always a negative deviation.

Of course, there is backlash at the meshing points of gears, pinions, and racks, so some errors can be absorbed.

However, one of the racks at the engagement point must be offset rearward from the other by a dimension equal to or greater than the backlash.

If so, the pinion and rack will separate at this point.

1. No driving force is transmitted.

Even if the dimensional error of the rack is less than the backlash, the forces applied to both will be unbalanced.

In any case, almost the entire load will be placed on only one pinion/rack combination.

This will cause rack damage.

I have no choice but to go with two-wheel drive.

A similar problem exists in a system in which a pinion is engaged with a hole in the rail instead of a rack.

It seems like it would be better to improve the dimensional accuracy of the rack.

However, since the rails are laid by connecting 4 or 6 m long rails, the pitch cannot be equal due to the phase of the joints.

It is impossible to strictly maintain equal pitch of the rack at points where the rack turns left and right or up and down.

Therefore, the inventor of the present invention realized that it would be better to use elastic gears (Japanese Patent Application No. 55-43748 (Japanese Unexamined Patent Application No. 56-142749)
)).

In an elastic gear, an output (boss part) and an input (gear part) are not rigidly connected, and an elastic body is interposed between them.

It is also called a cut ring and is extremely effective in alleviating momentary impact forces.

Between the power transmission mechanism and the drive pinion of the single-rail transport vehicle,
By interposing an elastic gear mechanism, the influence of pitch error of the rack can be eliminated.

The elastic gear has a separate outer shell portion and a boss portion.

The outer shell part has a gear, and the boss part has a shaft hole.

The boss portion can rotate relative to the outer shell portion.

However, a plurality of springs are provided between the boss portion and the outer shell portion to elastically connect the two.

Even if there is a pitch deviation in the rack, the elastic gears are displaced to compensate for the deviation, so nearly equal thrust is applied to the two pinions of the two-wheel drive motor vehicle.

It was thought like this.

However, in reality, it is difficult to appropriately select the elastic constant of the spring, and the thrust forces of the two pinions are always not balanced.

The strain in a spring is proportional to the applied force.

The distortion is small for light loads, and large for heavy loads.

Increasing the strain increases the ability to absorb energy, but
This will force the spring to become weaker.

The relative rotation angle between the boss part and the outer shell part of the elastic gear is 3° to 5°.
It is difficult to design the angle to exceed 10 degrees.

There is pitch disturbance in the rack when the maximum thrust is applied.
Only in this case can the force imbalance be absorbed; in other cases the pitch irregularity cannot be corrected.

Let θm be the maximum relative angle change of the elastic gear.

This is often between 3° and 5°, and at most 10° or less.

Let R be the radius of the drive pinion.

If the pitch disturbance is δ, the difference Δθ between the front and rear angular changes of the pinion is Δθ=δ/R(1).

On the other hand, let K be the elastic modulus of the elastic gear.

In other words, T K--(2) dθ It is.

Here, T is the torque applied to the elastic gear, and θ is the angular change between the boss and the outer shell.

Difference in rack reaction force applied to front and rear 1 drive pinion (
If the difference in thrust force is ΔF, then this is proportional to the rack disturbance δ within a certain range.

That is, KK ΔF knee Δθ=−δ (3) RR2 Even if the rack disturbance δ is large, the thrust difference ΔF can be made small by reducing the elastic coefficient K or by increasing the radius R of the drive pinion.

Increasing the drive pinion radius R increases the size of the frame and increases the torque of the output shaft, so the vehicle does not become heavy.

If the elastic modulus K is reduced, this buffering effect will not work when a large thrust is applied.

In other words, by integrating equation (2), θ2〇〇＜T＜T.

T−To To≦T(4) where To is a constant and indicates a torque within a range in which no elastic action occurs.

Since the maximum value of θ is limited by θm, θ2θm,
Obtain θm(5) for maximum torque Tm=To+.

This must be able to bear the maximum thrust applied to the motor vehicle.

In other words, the maximum value of thrust is Fm, T m < F m, R(6) Must.

I will explain in more detail.

Let the maximum thrust be Fm and the rack disturbance be δ.

It is assumed that the maximum thrust Fm can be received when the angular displacement of the front and rear elastic gears is between 0<θ<θm.

This is an ideal condition. If one of the gears is θmθm, there is already no buffering effect, and the entire thrust is applied to one of the elastic gears.6 The same is true even if one of the gears is θm00.

The larger angular displacement θ1 is assumed to be 02, and the smaller one is 02.

RFm T□ δm θ 2□−−+−(7) ” 2K K 2R RF" 0.6" (8) ”=2KK−■ becomes.

δm is the maximum value of rack disturbance. Naturally, δm □ <θ (9) must be true, and when the thrust δ 2K F≦Fm-(θ--)= (10) R is in this range, the force is no longer distributed to both pinions.

The load on one pinion is O and the force is only on the other pinion rack.

K δm If −(θ・−□) is increased, such a field RR decreases.

However, if this is done, the buffering effect at the time of maximum thrust, where equations (7) and (8) hold, will be reduced.

In other words, the buffering effect can be evaluated as 0·−〇·1″R(1,) T,+T2 RFm, but since K is large, this also becomes large and the buffering effect is small.

When elastic gears are used in this way, only near the maximum thrust, both elastic gears will undergo angular displacement and the thrust can be shared. However, when the thrust is small, if there is rack disturbance, one elastic gear will be forced to take its full load. Thrust is added.

Even at maximum thrust, the forces applied to both elastic gears are never the same.

It had these drawbacks.

What is the most ideal thing? This means that regardless of the magnitude of the thrust, the force is distributed to both pinions, and the force is divided in half.

These ideal requirements cannot be satisfied by elastic gears.

The reason for this is that the two drive pinions were completely independent with respect to phase changes.

It is sufficient if the phases of the two drive pinions can be correlated.

Since the rack pitch disturbance is a phase disturbance, the phase disturbance can be eliminated by correlating the phase of the drive pinion.

Phase, amplitude, and angular velocity are the three elements of rotational (oscillatory) motion.

How can I relate only the phase? The inventors have considered this point more deeply.

Then I realized that I could use a planetary gear system.

Rather than using planetary gears in the mundane way as a transmission, they are instead used as a phasing mechanism.

Planetary gears have one more degree of freedom than regular transmissions.

Usually, the outer shell gear is fixed and used as a transmission.

The inventor focused on the extra degree of freedom. For the first time, he realized that it was possible to adjust the phase by utilizing the degree of freedom of an external internal gear.

The planetary gear system has a sun gear in the center and 3 to 4 gears surrounding it.
There are two planetary gears, and surrounding them are external internal gears.

The planetary gears are supported by disks (called carriers).

When used as a speed reducer, attach the input shaft to the sun gear, the output shaft to the carrier, and fix the outer shell internal gear.

If the number of teeth of the sun gear is z3 and the number of teeth of the external internal gear is Zi, then the reduction ratio r=1+(Zi/Zs)
It is given by (12).

When using it as a speed increaser, replace the input and output shafts.

The point that the outer shell internal gear is fixed remains the same.

The present inventor interposed a planetary gear device between the two drive pinions and the power transmission system, supported the outer shell internal gear rotatably without fixing it, and made the outer shell internal gear mesh with each other. As a result, we were able to invent a two-wheel drive single-rail motor vehicle that can completely absorb rack disturbances.

EMBODIMENT OF THE INVENTION Hereinafter, the structure, operation, and effect of the present invention will be explained in detail with reference to drawings showing examples.

FIG. 1 is a right side view of a two-wheel drive single-rail power vehicle according to an embodiment of the present invention.

FIG. 2 is a left side view of the two-wheel drive single-rail motor vehicle excluding the engine and the like.

The power vehicle includes a frame 1, a drifting mechanism A, a power transmission mechanism B,
It consists of a drive pinion C1 wheel, etc.

The drifting mechanism A consists of an engine 2, a fuel tank 3, and the like.

It is placed at the upper front of the frame 1 and generates a motive force.

Power transmission mechanism B decelerates the driving force of the engine while
The signal is transmitted to the drive pinion C.

The power transmission mechanism B includes a centrifugal clutch 4 attached to the engine 2, a pulley 5 provided on the clutch output side, a belt 6,
It consists of a pulley 7 mounted on the frame 1 and a gear reduction mechanism connected to the pulley 7.

The belt cover 8 covers the pulley 5, belt 6, pulley 7, etc.

The braking mechanism includes a constant speed brake (descending brake) 9, an emergency stop brake 10, a stop brake drum 11, and the like.

These are provided on a suitable intermediate shaft (described later) in the power transmission mechanism B.

There are four wheels.

They are provided above and below, front and back, sandwiching the rail 12.

The rail 12 is, for example, a square hollow tube measuring 50 mm x 50 mm.

The front upper wheel 13 and the rear upper wheel 14 contact the upper surface of the rail 12 and receive the entire load of the motor vehicle.

The front lower wheels 15 and the rear lower wheels 16 abut against the lower surface of the rail 12, and cooperate with the upper wheels 13, 14 to sandwich the rail 12 from above and below.

The upper wheels 15,16 are arranged coaxially with the drive pinion C.

However, these are simply idlers and do not transmit torque.

When the gap between the rail 12 and the wheels D (13, 14, 15, 16) is widened, the bolts 19 of the eccentric plates 17, 18
, 20 and move it to one of the other holes 21, 22 in the plate and fix it to adjust the gap.

The upper wheels 13 and 14 are mounted on an eccentric shaft, and by rotating the eccentric plate little by little, the amount of eccentricity in the vertical direction changes.

In this way, the vertical positions of the upper wheels 13, 14 can be readjusted little by little.

Below the frame 1 is an oil reservoir 23, 23, in which the 1st drive pinion C2C is immersed.

Oil is gradually sent from the rail lubrication tank 24 to the oil reservoir 23.

The gear reduction mechanism of the power transmission mechanism B will be explained.

FIG. 3 is a schematic side view of the power transmission mechanism B.

FIG. 4 is a longitudinal sectional front view thereof.

The rotational force transmitted by the pulley 5, belt 6, and pulley 7 is transmitted to the input gear 29, 30 fixed to the input shaft 28.

One input gear 29 rotates an intermediate gear 32 mounted on an intermediate shaft 31 .

The switching shaft 33 is provided with a switching gear 34 that is slidable in the axial direction.

This is for switching between forward and forward movement.

When the input gear 30 directly meshes with the switching gear 34, the vehicle moves forward, and when the input gear 29 meshes with the switching gear 34 via the intermediate gear 32, the vehicle moves backward.

The switching shaft small gear 35 transmits the rotation of the switching shaft 330 to the brake small large gear 36 .

A large gear 36, a stop brake shoe 38, and a brake shaft small gear 39 are fixed to the stop brake shaft 31.

The other end is a constant speed brake shaft 40.

The stop brake shoe 38 is the emergency stop brake 10 described above.
It is activated by the operation of , and brakes the stop brake shaft 3γ.

A constant speed brake shoe 41 and a constant speed brake cooling fan 42 are attached to the constant speed brake shaft 40 .

The constant speed brake 9 is also called a descent brake, and automatically limits the upper limit of the speed so that the speed does not exceed a certain level when descending.

For example, the constant speed brake shoe 41 expands outward in diameter due to centrifugal force, contacts the inner wall of the brake drum 41', and receives a braking action due to frictional force.

The constant speed brake cooling fan 42 forcibly blows air to cool the brake shoes.

The brake shaft pinion 39 meshes with the distribution gear 43.

The above configuration is already known as a structure of a power vehicle of a single-rail transport vehicle.

In the case of conventional one-wheel drive, there is one output shaft gear that meshes with the distribution gear following the distribution gear, and the output shaft gear and drive pinion are attached to one output shaft, and the rotation of the output shaft gear This was directly reflected in the 1st drive pinion.

According to two-wheel drive, the distribution gear must provide equal rotational torque to the two output mechanisms.

In the aforementioned prior invention of the present inventor (Japanese Patent Application No. 55-43748), an elastic gear was provided between the distribution gear and the drive pinion.

In the present invention, a planetary gear train 44 is provided between the distribution gear and the drive pinion.

The planetary gear system 44 includes a sun gear 45 at the center, and planetary gears 46, 46 . ..., a carrier 47 that pivotally supports the planetary gear 46, and an outer shell internal gear 48 that surrounds the outside of the planetary gear 46 and meshes with it.

The configuration of the planetary gear system itself is well known.

In many cases, the outer shell internal gear 48 is fixed to the casing or frame.

In the present invention, the outer shell internal gear is not fixed and is rotatable.

External teeth 49 are further provided outside the outer shell internal gear 48.

The two outer shell internal gears 48, 48 each have external teeth 49,
They are engaged by 49.

This is the core of the invention.

Since the outer shell internal gears 48, 48 are not fixed but rotatable and are meshed with each other by the external teeth 49, 49, phase adjustment can be performed automatically.

Now, the distribution gear 43 meshes with the two relay gears 50.degree. 50 and transmits rotational torque.

The relay gear shaft 51 transmits the rotation of the relay gear 50 to the sun gear 45.

In the figure, the sun gear is small and the tip of the shaft 51 also serves as the sun gear.

However, more generally, the sun gear 45 can be separate and the relay gear shaft 51 can be attached to the shaft hole of the sun gear 45.

Planetary gear 46. .

A drive pinion C is fixed to the output shaft 53.

A drive pinion meshes with a rack 54 welded to the side force of the rail 12.

As mentioned above, the output shaft 53 includes the front and rear lower wheels 15,
16 is attached via a bearing 55.

The upper wheels 15 and 16 are idle wheels.

The output shaft 53 is rotatably supported by bearings 56 and 57 relative to the frame.

The relay gear shaft 51 is rotatably supported on the frame by bearings 58 and 59.

The outer shell internal gear 48 is rotatably supported by a bearing 60.

The operation of the above configuration will be explained.

The motive force of the driving mechanism A is transmitted to the distribution gear at the end of the power transmission mechanism B.

The distribution gear causes two relay gears 50.50 to rotate in unison.

The planetary gear device 44 provides a reduced output to the output shaft and drive pinion C.

The two drive pinions C push the rack at two places, so the motor vehicle moves forward or backward.

If there is a disturbance in the rack pitch, the outer shell internal gear 44
rotates and its meshing point X changes, so the contact pressure at the contact points between the two 1-drive pinions and the rack is always equal.

If the entire propulsion force is F, strictly speaking F/2. A force of F/2 is applied equally to the two drive pinions.

Conventional racks can withstand heavy loads because no local stress is applied to the racks.

This is a great advantage.

In a planetary gear system, the sun gear and the outer shell internal gear work together to rotate the carrier.

To turn the carrier to the right, the sun gear may be turned to the right or the outer shell internal gear may be turned to the right.

By the way, in a two-wheel drive power vehicle, the sun gears of the two planetary gear units rotate exactly at the same time.

The phases are the same.

However, since the carrier is connected to the drive pinion, if there is a positive or negative deviation in the pitch of the rack,
Its phase changes.

The phase shift of the carrier cannot be eliminated by the movement of the sun gear.

This is because the phase does not change. However, because the outer shell internal gears mesh with each other, they can rotate in opposite directions.

Therefore, the pitch disturbance of the rack can be absorbed by slight rotation of the outer shell internal gear.

The rotational torque applied to the outer shell internal gear is exactly the same.

This is self-evident due to Newton's law of action and reaction.

By the way, in a planetary gear device, the ratio of the torque applied to the carrier and the torque applied to the outer shell internal gear is determined by the ratio of the number of teeth.

In other words, it is constant.

In the two planetary gear systems, since the torque applied to the outer shell internal gear is the same, the torque applied to the carrier is also the same.

Therefore, in the present invention, the contact pressures at the meshing points between the two drive pinions and the racks are strictly the same.

The thrust is always divided into two and applied to the rack and pinion.

Prove it more strictly.

FIG. 5 is a schematic plan view of a general planetary gear device.

The angle shall be measured counterclockwise. The numbers of teeth of the sun gear, planetary gear, and outer shell internal gear are Zs and ZptZi.

The angular velocity of the sun gear, carrier, and outer shell internal gear is ΩS, Ω
Let it be C2Ωi.

If the relative angular velocities of the sun gear, planetary gear, and external internal gear seen from the carrier are Ωs', Ωp', and Ωi', then since the linear velocities at the meshing points are equal, ZsΩs'= -Z pΩp'= -Z iΩi '
(13).

The second term has a negative sign because the sun and planetary gears mesh with each other through external teeth.

The reason why the third term has the same sign as the second term is because the planet and the outer shell gear mesh with each other through their internal teeth, and the rotation direction is the same.

The angular velocity seen from the stationary system is the angular velocity beyond the carrier,
It is the sum of the angular velocity of the carrier.

ΩS2ΩS′+Ωc (14)Ωi−
Ωi'+Ωc (15) This is (13
) Substituting into the formula, we get the basic formula for a planetary gear device: % Formula % (16) In normal usage, the outer shell internal gear is fixed (Ω1=0) and used as a speed increaser or decelerator.

The reason why the reduction ratio r is as shown in equation (12) is Ω in equation (16).
You can confirm by setting i20.

Next, we turn to mechanical considerations.

Think about torque.

The torque applied to the sun gear, carrier, and external internal gear is
Let Ts and Tcs Ti (including the symbols).

Since the sum of all torques in steady state is O T s +T c +T i =O(17).

According to the law of conservation of energy, TsΩs+TcΩc+TiΩi=0 (18).

Further, the mutual constraint condition for the number of teeth Z i =Z s+2Z p (19) holds true.

From equations (16) to (19), regarding torque, Ts −Tc Ti (2o)Zs
This can be done by finding a simple proportional expression: Zs+Zi Zi.

If the thrust applied to the front and rear drive pinions is Fl, F2, and the total thrust is F, then F-F1,000 F2 (21) Based on the meshing point of the external internal gear when there is no pitch disturbance of the rack, Let the counterclockwise rotation angle of the internal shell gear be 01 and θ2.

This is not an independent variable.

Since the two external internal gears mesh, they can only move in opposite directions.

In other words, θ +0220 (22).

Since the outer shell internal gears mesh with each other, the torque Ti1
t Tt2 is always equal.

In other words, T i1= T i2 (
23) holds true.

Assume that there is a pitch disturbance δ in the rack.

This appears as a phase difference ΔθC between the carriers of the front and rear planetary gear devices.

As a result, the outer shell internal gear rotates by Δ and θ.

From equation (16), Zi△θ・= (Zs +Z i )△θc (2
4) R△θC−δ (25) However, R
is the radius of the drive pinion.

In other words, the rotation angle △θ of the external shell internal gear is (Zs+Zi)δ △θ= (26)i
It is given by R.

The outer shell internal gear rotates only by a small angle even if there is pitch disturbance δ of the rack.

Therefore, there is no need to cut the entire outer periphery of the outer shell internal gear.

It is sufficient to provide teeth only in the range determined by equation (26).

Furthermore, Δθ is θ −θ −2θ −−20(27)1
2-1-2 holds true.

In other words, if there is a deviation δ in the pitch of the rack, one of the outer shell gears will rotate by half the angle of equation (26), and the other outer shell gear will rotate by the same angle in the opposite direction, which is due to the difference in pitch. The effect is immediately erased.

From equations (20) and (23), the torque Tc1. applied to the carrier. If Tc2, Tc1=Tc2 (28).

The torques of the carriers of the front and rear planetary gears are equal.

The thrust force F1 is divided by the radius R of the drive pinion.
, F2, so Fl = F2 (29).

From formula (21), F1=F/2 (
30) F2: F/2 (31) This means that the torques applied to the front and rear drive pinions are exactly equal, and the thrust applied to the rack is also half of the total thrust.

Even if there is a discrepancy in the pitch of the rack, the front and rear thrust is always equally distributed in half.

This is derived from equation (23).

This is because the outer shell internal gear is rotatable and meshed together.

According to the law of action and reaction, the torques are equal and the thrusts applied to the rack are also equal.

The meshing of the outer shell internal gears always equalizes the distribution of force.

In the two-wheel drive single-rail power vehicle according to the present invention, even if there is pitch disturbance in the rack, equal thrust and torque can always be distributed to the front and rear drive pinions.

No local force is applied to the rack.

Therefore, a transport vehicle for civil engineering can be constructed using a conventional rack.

Furthermore, during assembly, there is no need to worry about the phase relationship between the shaft, planetary gear set, and pinion.

This is because if the outer shell internal gear is rotated, it will be balanced somewhere.

Even if the pitch disturbance of the rack becomes permanent, there is no problem at all.

This is because the outer shell internal gear rotates freely.

From equation (26), the rotation angle of the outer shell internal gear is small.

Therefore, it is sufficient to cut the external teeth 49 only in a portion.

As shown in FIG. 6, an extension lever 61 may be provided from the outer shell internal gear 48, and an external tooth 49 may be provided at the end thereof at a narrow center angle corresponding to the range of equation (26).

In this way, a planetary gear system with small dimensions can be used.

As shown in Fig. 7, only a part of the outer shell internal gear has external teeth 4.
9 may be provided, and the remaining peripheral surface may be a circular surface and supported rotatably on the inner peripheral surface of the frame 1.

It becomes a sliding bearing.

As shown in FIG. 8, from the outer internal gear, the lever 61'
It is also possible to lengthen it, make a long hole at the end, and insert the pin p.

This has the disadvantage that the torque is not always the same, but has the advantage of being simple in structure.

Furthermore, the outer shell internal gear is not necessarily supported on the intermediate gear shaft.

It may also be brought into sliding contact with the inner peripheral surface of the casing.

This makes the outer shell internal gear more stable. Although the carrier 47 is shown as being on one side, the carrier 47 is not limited to this, and may be on both sides to support the planetary gear.

This makes the output shaft more stable.

A disk or a ring having the same pitch circle may be attached to both sides of the planetary gear and the outer shell internal gear, respectively.

In this way, centrifugal force can be suppressed by the contact between the disc and ring, resulting in a planetary gear system with excellent transmission efficiency and low noise.

FIG. 9 is a sectional view showing such another embodiment.

The external shell internal gear is provided with external teeth 49 only near the meshing point, and the rest is simply a circumferential surface, which is rotatably supported by a circumferential bearing 62 of the frame.

A disk 63 and a ring 64 are provided on both sides of the planet and the outer shell internal gear.

As explained in detail above, according to the present invention, even if there is a pitch disturbance in the rail rack in a two-wheel, one-drive, single-rail power vehicle, the thrust can always be carried in half, and the rack No local force acts on it.

Therefore, it can be said that it is completely ideal as a single-rail power vehicle that carries heavy loads.

In this way, it is a useful invention.

[Brief explanation of drawings]

Fig. 1 is a right side view of a two-wheel drive single-rail power vehicle according to an embodiment of the present invention, Fig. 2 is a left side view of the two-wheel drive single-rail motor vehicle excluding the engine, and Fig. 3 is a power transmission mechanism. a schematic side view of
FIG. 4 is a front view in four longitudinal sections of the power transmission mechanism, FIG. 5 is a schematic plan view of the planetary gear device, and FIGS. 6, 7, and 8 are views of external shell internal gears according to other embodiments. The side view and FIG. 9 are longitudinal sectional views of a planetary gear device according to another embodiment. 1...Frame, 2...Engine, 3
... Fuel tank, 4 ... Centrifugal clutch, 5 ... Pulley, 6 ... Belt, 7.
...Pulley, 8...Belt cover, 9.
... Constant speed brake, 10 ... Emergency stop brake, 11 ... Emergency stop brake drum, 1
2...Rail, 13...Front upper wheel,
14... Rear upper wheel, 15... Front lower wheel, 16... Rear lower wheel, 17°18...
...Eccentric plate, 28...Input shaft, 29°30...Input gear, 31...Intermediate shaft,
32... Intermediate gear, 33... Switching shaft,
34... Switching gear, 35... Switching shaft small gear, 36... Brake fine large gear, 37...
... Stop brake shaft, 38 ... Stop brake shoe, 39 ... Brake shaft small gear, 40 ...
...Constant speed brake shaft, 41...Constant speed brake shoe, 42...Constant speed brake cooling fan,
43...Distribution gear, 44...Planetary gear device, 45...Sun gear, 46...Planetary gear, 47...Carrier, 48...
Outer shell internal gear, 49...External teeth, 50...
・Relay gear, 51...Relay gear shaft, 53...
...Output shaft, 54...Rack, 55-60.
...Bearing, 61...Lever, 62...
... Circumferential bearing, 63 ... Disc, 64 ...
··ring.

Claims (1)

[Claims] 1 Consists of a drifting mechanism A, a power transmission mechanism B, two drive pinions C that mesh with the rack, a wheel rail that clamps the rail, and a frame 1 that houses these mechanisms A, B, C, and D. In a two-wheel drive single-rail power vehicle, two planetary gear systems are interposed between the power transmission mechanism B and the second drive pinion C, and the sun gear of the planetary gear system is connected to the power transmission mechanism B, and the carrier is connected to the drive pinion C. A two-wheel drive single-rail power vehicle characterized in that the two outer shell internal gears are connected to pinions C and are configured to be rotatable and mesh with each other.