JPS6188043A - Prime number type planetary gear - Google Patents

Abstract

PURPOSE:To reduce noise of a planetary gear by determining the sum of the number of the teeth of a sun gear and the number of the teeth of a shell internal gear to be the multiplied number of the teeth of a planetary gear and also determining the number of the teeth of each of the gears to be unequal to the multiplied number of teeth on the planetary gear. CONSTITUTION:If the number of the teeth of a sun gear and that of the teeth of a shell internal gear are expressed by S and I, respectively, the sum of the teeth (S+I) is the multiplied number (n) of the teeth of a planetary gear. S and I are set to be unequal to the multiplied number (n). Each of S and I is determined not equal to the multiplied number of (n). Now, supposing that a reduction gear of a multiplied type with I=72, S=24 and P=24, and the reduction ratio of 1/4 and another reduction gear of a prime type with I=73, S=23 and P=25 and the reduction ratio of 1/4.17 are compared, for example, two planetary gears are made of the same material and has the same external periphery. But the gears has different diameters. Since highest noise level will be able to be measured at non-load time, so measurement is performed at that time. Revolving ratio of a motor is set to 1,800rpm. A measuring unit is sited 50cm apart from the gears in the opposite position to the motor and then noise level is measured. The obtained result was approximately 54 phons for the multiplied type while that was approximately 50 phons for the prime number type.

Description

[Detailed Description of the Invention] V) Technical Field This invention relates to a planetary gear device that generates little noise and smoothly transmits power.

Planetary gears include sun gears, planetary gears, outer shell internal gears,
It consists of a carrier on which a planetary gear is attached.

The input and output shafts are on the same axis, making it easy to use as a speed reducer or speed increaser. Also, it does not require much thickness,
It is an excellent power transmission device because the outer diameter can be made small.

However, planetary gears have many meshing points and a small degree of freedom, which can lead to tooth tip interference, abnormal eccentricity, and strong noise.

For this reason, there was a drawback in that each gear had to be of high precision.

However, the inventor believes that the use of highly accurate gears hinders the effective use of planetary gear systems due to cost considerations. Rather, we believe that we should create a planetary gear system with high power transmission efficiency and low noise using low-precision gears. 1, and I am proud that I have achieved great results in accordance with this belief.

(B) Prior Art The present inventor is of the opinion that there is no clear technical idea regarding the determination of the number of teeth of a planetary gear device.

The number of teeth of the planetary gear is P1, the number of teeth of the outer shell internal gear is M, and the number of teeth of the sun gear is S.

The number n of the planetary teeth 'Af is three or four. It is safe to say that there is nothing else like it. If there are two, the sun gear and outer shell internal gear cannot be determined.

If there are more than 5, the boat will not be able to obtain a large reduction ratio. The carrier is made up of two disks joined together by rivets, ridges, etc. Since space is also required for the joint, it is difficult to make a device with five planetary gears.

Therefore, it can be said that the number n of planetary gears is 3 or 4.

In this case, there was a vague opinion as follows regarding the determination of the number of teeth, which had become common knowledge. That is, the number of medium planetary gears is 3.
If so, the number of teeth of the sun gear, planetary gear, and outer shell internal gear must all be multiples of 3.

fl[+ If the number of planetary gears is 4, the number of teeth of the sun gear, planetary gear, and outer shell internal gear must all be multiples of 4.

This is the view.

The number of planetary gears n (3
Or 4), sun gear S, planetary gear P, tJ in the outer shell? J
S = ns between the number of teeth of gear I
(1) P = np
(2) I = ni
f8) holds true.

This is actually a heuristic rule of thumb when assembling planetary gears.

If the reduction ratio when used as a reduction gear is R, this is I
By and S, , , IJ (can be done.

For example, assume that the number n of planetary gears is three. Suppose you want to make a 1/7 reduction gear (R=7). For example, let I=96. Divide this by 6 to get 5=16, P=40. It seems like it would be possible to make a 1/7 reduction gear with this, but it wasn't.

The inventor actually made a planetary gear with this number of teeth.

Then I realized something terrible. When I tried to assemble this gear, I was unable to assemble it. After installing the carrier, planetary shaft, and planetary gear into the outer shell internal gear, when I tried to insert the sun gear, it just wouldn't fit.

After much trial and error, I rearranged the sun gear to have 15 gears (S=15). This rearrangement made it possible to insert the sun gear.

Since we were unable to determine the cause of this, we investigated the number of teeth in planetary gears manufactured by other companies.

Then, when n=3, S, P, and I are always multiples of 30, and when n=4, s, p.

■It turns out that both numbers are multiples of 4.

This is what I mentioned earlier. It can be said that this is what is shown in formulas (1) to (3). It can be said that this has become common knowledge and an axiom in the planetary gear industry.

The inventor also thought that (1), ([+ is an axiom.

That's what I thought.

If it is not a gear with a shift, there has never been a planetary gear system that does not satisfy the +11 axiom. This can be said based on the research conducted by the present inventor. It is blind obedience to axioms.

The inventor also followed axiom (11, (Il)), but could not understand why it had to be so.

Since the inventor is an engineer, he wanted to know the basis of the axiom.

An inherent drawback of planetary gears is the generation of noise. Although noise is unavoidable due to the meshing of gears, in some cases It':b may be too large. Why is the noise so loud? This has also been a question for a long time.

Possible dislocation. Even if there is a dislocation, the sun gear S,
It has been thought that the outer shell internal gear I must be a multiple of n. It was thought that equations (1) and (3) must hold true anyway.

The reason was a mystery. I can only call it a rule of thumb.

There are still mysteries surrounding planetary gearing. It's a matter of engagement rate.

As the reduction ratio R was increased, the sun gear became smaller, but when the sun gear was overloaded, it often broke. In some cases, the head and foot gears were broken.

Although the engagement rate was always calculated as 1, it is clear that the engagement rate is not 1. However, there has been no formula for calculating the meshing ratio.

Contact ratio is the number of teeth that are in contact at the same time. This should be a function of the number of teeth. However, the formula for expressing the contact ratio as a function of the number of teeth has traditionally been
It didn't exist. For this reason, when calculating strength, the ambiguity must always be known, and the reality is that calculations cannot be made in the end, and destructive tests have been relied upon.

It was also unclear whether the engagement rate depended on the module or not.

As will be shown later, the inventor discovered that the contact ratio does not depend on the Moneur at all, but only on the number of teeth and the pressure angle.

(1) Proposal One phantom number in the field f tun is an integer.

The number of planetary gears is 3 or 4. Let us consider the case where n=3.

Every integer is a multiple of 3, 1 less than a multiple of 3, or 1 more than a multiple of 3.

The number of sun gears is as follows, where S is an integer: 1) S=8s (41++)
s = as + t (
5) iii) S = 38 − 1
(6) Either.

What on earth does a planetary gear do? Think of this skewer. On the diameter of the planetary gear, the sun gear and the outer shell internal teeth tJAtun mesh with the planetary gear.

FIG. 1 is a schematic diagram showing only the pitch circle of a planetary gear device with n=3.

The center of the whole is set to 0. The center of the planetary gear is Po, P,
, P2. The point where the sun gear meshes with the planetary gear is
Let Q, , , Q, , Q2. The points where the outer shell internal gear meshes with the planetary gear are Ro, R, and R2.The sun gear S and the outer shell internal gear face each other at Qo and Ro, with Po as the center.

Let's take Q as 0 point and count the number of teeth on the sun gear turning clockwise.

Let's count the number of teeth of the outer shell internal gear clockwise with Ro as 0 point.

The circumference is divided into three by po, p, and p2. Therefore, at point Q1, the third tooth of S7 is in mesh with the planetary tooth l[. Similarly, at point Q2, the 2S/3rd tooth of the sun gear is in mesh.

Since R0 point is counted as 0 point, R1 of outer shell internal gear engineering
The engagement at the point is I/3rd tooth. The engagement at point R2 is 2nd work/3rd piece.

The phase relationship must be considered.

It indirectly meshes with the sun gear and the shell internal power gear via the planetary gear.

The phase of a certain state of meshing is determined to be 0 for the sun gear and the outer shell internal gear. Suppose that the sun gear rotates clockwise by a certain minute phase. Then, the planetary gear rotates counterclockwise by the same phase amount, and the outer shell internal gear rotates counterclockwise by the same phase amount.

Here, the phase does not mean a rotation angle, but a value from 0 to 1 assigned to each tooth.

After all, if we take the clockwise phase to be positive, the phase of the sun wheel φ
The sum of S and the minute displacement of the phase φi of the external internal gear is always 0
It turns out that.

dφ5−)−dφi=o (71 Here, the phase is counted as a fractional part between 0 and 1.

Here, let the parentheses [...] represent the decimal part of the number inside.

Since point Q is the 0 point of the sun gear, the phase is 0. The phase at 91 points is (S/3). The phase at point Q2 is (23/8).

Since the Ro point is the 0 point of the external internal gear, the phase is 0 here. The phase at point R1 is CI/8). R
The phase at the two points is C21/8).

The phase refers to the Qo SRo point at the meshing point.
This shows a slight deviation within one tooth from the above state. Therefore, the fractional part of the number of teeth counted from the Qo-Ro point gives the phase.

Using formula (7), [S/3) + (:I/3) = 0
(81 and C2S/8) + [2I/3) = 0
The formula (9) holds true.

There are three combinations to satisfy these requirements.

(Both als and I are multiples of 3.

(b) S is 1 greater than a multiple of 3, and I is 1 greater than a multiple of 3.
small.

ICI S is 1 less than the 30 times number, and ■ is 1 more than the 3 times number.

There are three cases, but comprehensively speaking, it is sufficient that the sum of S and 1 is a multiple of 3.

inclusive expression % formula % (10) It can be summarized and expressed as

Conventionally, among (a), direction, and (C), only (a) was known. It was thought that the gears of the planetary gear system could not be fitted without relying on (2).

But that's not the case. The gear fitting conditions include Mata), (C
) is possible.

(In the case of al, it can be written as 5=33.I=3i (11), and in the case of direction, S=3s+1.I=3i-1(1
2) In the case of (C), S = as-1, I = 3i+1
(13) and 1! ? ＜I can do it. Consider the case of a planetary gear system without dislocation. At this time, it must be S + 2P=I (14).

In any case of (11) to (13), (i-
It is only required that s) be a multiple of 2.

The reduction ratio of planetary gears is often 3 times or more.The number of sun gear teeth is often 12 or more.

An example is shown. (a) to (c), that is, (11) to (
In equation 18), consider an example where 1=22. s=4.6.
8.10 may be the case.

Table 1 shows these basic values.

Based on the combinations in Table 1, 3), ℃), and (C)
The number of teeth of I, S, and P in the case of
), (13). The reduction ratio R can also be calculated using equation (4).

Table 2 shows the case of ■), which is calculated using equation (11).

Table 2: When tal I is a multiple of 8 Table 3 shows the case (b), which is calculated using equation (12).

Table 4 shows the case of (C1), which is calculated using equation (13).

Table 4 (c) Such a combination is possible when I is a multiple of 3 and greater by 1.

(b), (It was completely unknown in the past that combinations of the number of teeth such as C1 were possible.The inventor of the present invention came up with this idea for the first time after many years of experience and consideration.

Inclusive expression (l is the easiest to understand and important expression.

As can be seen from these considerations, no direct conditions are imposed on the number of teeth P of the planetary gear.

If there is no dislocation, conditions can be determined for P as well as S and I. Corresponding to fal, (b), and (C) above, P is a multiple of 3, 1 less than a multiple of 3, and 3, respectively.
This means that it is 1 larger than the @ number.

Since P is an integer, this imposes another new condition on the teeth of S and I.

However, when using a shifted gear system, there is no longer any condition imposed on P. Therefore, there is no additional condition regarding the number of teeth S and I.

However, even if shifted gears are used, comprehensive expression (10) and basic expressions (11) to (13) must necessarily hold true. This is an important point.

The above is a case where the number n of planetary gears is 3, but the same can be said even when n is 4.

Since there are four planetary gears, the sun gear meshes with the planetary gear at 0, S/4.2S/4.3S/4th gear, counting from point 0, and the outer shell internal gear meshes from point 0. ,0,I/4
．． 2I/4.3I/4th gear meshes with the planetary gear.

A formula very similar to formulas (8) and (9), [S/4) + CI/4) = O(15)(2S/
4) + (:2I/4) = 0 (16)
(3S/4) + (:8I/4) = 0
We obtain (17). There are four cases that satisfy this requirement:

(d) Both S and I are multiples of 4.

S = 4s, I = 4i (1
8) (el S is 1 greater than a multiple of 4, and Engineering is 1 less than a multiple of 4.

S = 45 + 1, I = 4i -1 (19) (
fl S is 2 greater than a multiple of 40, and ■ is 2 smaller than a multiple of 4.

S = 414+ + 2, I = 4i -2(2
0) Ri S is 1 less than a multiple of 40, and I is a multiple of 4.
1 greater than

5=4s-1, I=4i+1 (21)
Similarly, the comprehensive expression % expression % (22) is obtained. That is, it is sufficient that the total number of teeth of the sun gear and the number of teeth of the outer shell internal gear is a multiple of four.

In the comprehensive expressions (10) and (22), the number of planetary gears is n
It can be further expressed as the general comprehensive expression % expression %) (23).

Conventionally, among those expressed by equation (23), S = n
It was thought that only those with s 1 = n i (24) could be manufactured. However, the present inventor is expressed by the (Yu) formula, and (24
It was for the first time that I realized that it was still possible to produce something other than ). In other words, (28) is good, and S〆ns or Tsu If! n i (25
).

The inventor believes that such a planetary gear system has superior advantages.

In order to easily distinguish between the conventional gear and the gear proposed by the present inventor, the planetary gear device with the conventional number of teeth as shown in equation (24) will be tentatively named "multiple type". .

On the other hand, the equation (25) holds true and the planetary gear device newly invented by the present inventor will be referred to as a "prime number type". This name is derived from the fact that relatively many prime numbers appear in ISS, P, etc. This does not mean that all numbers are prime.

On the other hand, in the conventional type, a prime number never appears in the number of teeth (unless there is a dislocation).

These things can be easily seen from Tables 2 to 4.

The gist of the present invention is to provide a novel prime number type planetary gear device.

It has features and effects superior to those of the multiple type, but in order to invent this, we must first discuss the meshing process.

2) Engagement rate Engagement rate is not an accurate term, but it is commonly used in this industry, so this term will be used here as well.

This term refers to how many teeth are in contact at the same time when two teeth are meshing. It should also be expressed as the number of teeth in simultaneous engagement, but here it is referred to as the engagement ratio. Although it is called a ratio, it is not a ratio value.

Normally, the strength is calculated as one sheet. The reason is,
This is because there was no formula for easily calculating the engagement ratio E as a function of the number of teeth.

The inventor first derives a formula that gives the engagement ratio E.

FIG. 2 is a schematic diagram showing a state in which two gears are meshed.

01.0□ is the center of the gear, and will be called gear 1 and gear 2.

The base circle of the gear is B, the pitch circle is P1, the tip circle is J, and the root circle is H, and the subscript of each gear is attached.

The root circle is different from the root circle, and the pitch circle is 1 module (instead of 1,25 modules).
It is a circle with a small radius.

Because they mesh, the two pitch circles touch at point A.

H, P, and J differ by one module in terms of radius. Hl
and J2 touch at point N, and H2 and Jl touch at point. J,,J
Let the intersection of 2 be M.

There is a relationship between the diameter B of the base circle and the diameter of the pitch circle as follows, where α is the pressure angle: B=Dcosα (26).

If we draw a common tangent to the two base circles and let the points of contact be F and G, this will pass through the point of contact A of the pitch circle. And /AoIF== α (
27)/AO□G=α (2
8).

The basic circle means that if you pull out the thread wound around this circle and then unwind it, the involute tooth surface will become better.
It is a yen.

The common tangent line FG is also called a line of action. The point of contact between the teeth of the two gears is always above this, and the direction of the force is parallel to FG.

The two gears touch the tooth surface of the other gear between the root circle H and the tip circle J. They touch only in the shaded arcuate portion LKMNO.

Moreover, since the trace of the contact point is on FG, in the end, the intersection point between FG and the circle Jl, J2O, '/! :
The contact point of the tooth surface is on the line segment UV. As the gear rotates, the contact point moves from υ→A→V.

After all, the length of the line segment U'/ is the engagement length.

By dividing the meshing length UV by the spacing between meshing teeth, the number of meshing teeth, that is, the meshing ratio E, can be determined.

The tooth surface is an involute curve formed by the base circle B. If all tooth surfaces are extended along involute curves and intersected by the line of action FC, these curves are always orthogonal to the line of action FG.

The circumference of the base circle divided by the number of teeth Z is called the normal pitch.

The tooth surface of a gear is a group of involute curves drawn from the base circle at each normal pitch.

Then, when the adjacent tooth surfaces are extended along the involute curve, they intersect perpendicularly to the line of action FG, and the interval therebetween is equal to the normal pitch q. A schematic diagram is shown in FIG. It is assumed that the starting point of the involute curve is 01, CG, etc. on the base circle. The interval 2s 3 along this circumference is the normal pitch. Let the points where these involute curves intersect with the line of action FG be mountains T, , T2, and T3. From the definition of an involute curve, arc GC1 and line segment GT1
are equal. Therefore, the intervals T1T2, T2T3, . . . are equal to the normal pitch q.

The tooth flanks always contact on the line of action, and the interval between the intersection of the extension of the adjacent tooth flank and FG is the normal pitch q, so the interval between the contacting tooth flanks is equal to the normal pitch q.

The engagement ratio E is the value obtained by dividing the line segment υV by the normal pitch q.

CI (29) Normal pitch q
has a relationship with module m as follows: q = myrcOsα (30).

All you have to do is find UV, but this is the sum of AU and AV. Both can be determined independently as a function of the number of teeth.

I decided to look for AV. This has nothing to do with gear 2 and only depends on gear 1.

FIG. 4 shows only the necessary parts of FIG. 2. The base circle B1 and the dedendum circle H1 are omitted.

Only the pitch circle P1 and tip circle J are shown.

The line of action is FAY. For the sake of simplicity, we will consider the scale with the module as 1.

The diameter of the pitch circle P is Z, and the diameter of the tip circle J is (
Z+2).

When the horizontal axis is x and the vertical axis is y, the formula for the tip circle is as follows, and the formula for the line of action FAY is y = -x plow α - (32). Since we are looking for the intersection of two equations, we can solve them by combining them.

What we want to find is the length of AV = t, AV = t = xseca
(18), so t + Ztlanα-(ZI1) = O(34). Solve this and get .

This gives the meshing length AV as a function of the number of teeth Z. It should be replaced with z Fiz.

AU is obtained by replacing Z with 22. A
The meshing ratio is the sum of tJ and AV divided by q.

However, here we consider the value obtained by dividing AM by q.

It is called the number of partial engagements. The normal pitch q is πCOSα from equation (30) on a scale where the module is 1.

The number of partial meshes is a function of Z, 2πωSQ It is determined by

The pressure angle α is included as a parameter, and with this, the engagement ratio E (Z1Z2) can be determined.

However, equation (36) is a formula for external gears.

In the case of an internal gear, the number of partial meshes is M (Z).

The mesh number L (Z) of the external gear is a monotonically increasing function with respect to Z, and Z; ω gives the maximum value. This is.

The number of meshes M(Z) of the internal gear is a monotonically decreasing function with respect to Z, and 2=■ gives the minimum value.

and is equal to the desired limit of L(Z). This is natural. At Z→■, the distinction between external and internal teeth disappears, and the tooth becomes just a rack.

The meshing number M(Z) of the internal gear is always larger than (39), and the meshing number L(Z) of the external gear is always smaller than (89).

When two gears Z1 and Zl are in mesh, and both are external gears, the meshing ratio E'(Z,, Zl) is (36
) using the definition of E (Z,, Zl) = L(ZI) + L
(Zl) (40) If Zl is an external tooth and Zl is an internal tooth, then E(Z,, Zl) = L(ZI)
+ M(Zl) (41).

This is an accurate formula for meshing ratio given for the first time by the inventor.

Equations (86) and (37) cannot be used as they are to obtain a unified equation. However, if the curvature W is defined by the reciprocal of Z, and if W is determined to be positive for external teeth and negative for internal teeth, a unified formula can be obtained for the number of partial meshes L(w).

External teeth → W = - (42) The unified formula L(w) is It is.

The limits of (38) and (39) indicate the maximum number of meshes of the external gear. Even when external teeth and internal teeth mesh, the force of the external teeth is small depending on the number of teeth, so the upper limit of the meshing ratio E (Z,, Z2) shown in (4o) and (41) is (38). , (39).

It is.

When the pressure angle α is 20°, the upper limit of E is 1.98.

If the pressure angle is 146, the upper limit of E is 2.7.

(36), (37), and (44) are exact formulas, but since they are not intuitively clear, we can use approximate calculations to
Think about the meaning of these expressions.

- Consider the case where the curvature W is sufficiently small. When α is 20°, W is 0.02 or less, and formula (44) is approximated as follows.

In the case of external teeth (W is positive), it can be seen that as the number of teeth increases, the curvature W decreases, and the number of engagement increases, approaching the limit value (38).

In the case of internal teeth (W is negative), it can be seen that as the number of teeth decreases, (-W) increases and the number of meshes increases.

An exact calculated value was obtained using equation (44). α=20
'.

In the case of external gears, the curvature W and the number of partial meshes are shown in the fifth example.
Shown in the table.

Table 5: Number of partial meshes of external gears Table 6 shows examples of the curvature W and the number of partial meshes M(w) for internal gears.

Table 6: Number of partial meshes for internal gears Tables 5 and 6 list only some values for W, but the number of meshes can be determined strictly for any tooth reason. The contact ratio can be determined for any combination of gears.

For example, consider the meshing of 20 and 25 external gears. Since W = 0.05 and 0.04, from Table 5, EC20,25) = 0.778 + 0.80
6-; 1.6 (47).

For example, when 20 external teeth mesh with 63 internal gears, E C20, -63') = 0.778 + 1.1
58'': 1.9 (48).

The interlocking ratio of external teeth and internal teeth is higher than that of external teeth and external teeth.

(It has already been stated that when 1 = 20', the upper limit of the meshing ratio is 1.98, and equation (48) is close to that.

The engagement rate is a value of 1 or more, but if it is written as (1 + J) (0 × δ × 1), it is expressed as 1 at the rate of (l - δ) in time.
Two sheets mesh at a ratio of δ.

The meshing ratio of 1.9 means that the time it takes for the two sheets to mesh is -, and almost two sheets are engaged. Planetary gear (
The meshing ratio between the external gear (external teeth) and the external shell internal gear is thus high.

However, since both the sun gear and the planet gear are external gears, their meshing ratio is low.

In particular, when the reduction ratio R is set to a large value, the number of teeth of the sun wheel decreases, so the meshing ratio becomes particularly low. The number of single tooth meshing increases, making the teeth more likely to break.

(a) Disadvantages of multiple type planetary gears The number of sun teeth S and the number of teeth in the outer shell I are
) The disadvantages of the multiple type (conventional example) planetary gear system, which is a multiple of ), will be explained.

In Fig. 1 (n=3), meshing points Qo, Qs, Q2
The phases of the sun and planetary gears are all equal.

The phases of the planets and outer shell gears at Ro, Rt, and R2 are also all equal.

The intensity of noise I(t) generated in gears varies with time, and this is as follows: I(t) = Iksin(Qlt + d)
(49). However, ω is the reciprocal of the period T in which one tooth moves by 9 normal pitches multiplied by 2π. It is natural that the strength of the noise will differ if the phase between the meshing teeth differs.

However, ω is a kana and has a large value, and the number of teeth of the gear is 21.
．． 2. , when the rotational angular velocity is 01 and Ω2, ω= Ztnt = Z2Ω,
(50).

In Figure 1, the meshing points are Q, ,Q, between the sun and the planets.
There are three points: , Q2.

What is the noise generated at three points? . (t)=Iks11(ωt+φ.)
(S,)I 1 (t) = I km(o)t
+φ1) (52)I,
(t) = Iksin(cut+φ2)
(Guo) and 11).

In the case of multiple type, the phases at 8 points are equal, so φ0= φ1 = φ2=o (5
4). The sum of noise is 1(t)=81ksInωt(5EI), which is eight times the noise at one point.

ψ) Advantages of prime type planetary gears The present invention proposes a prime type planetary gear device. In the case of a prime number type, the phases at three points differ by 1/3.

In other words, φ. = 0 (56)φ, =
2π/8 (57')φ2=4π
/8 (58) can be written. The sum of noise It(t) is obtained by adding (S, ) to (53). Suddenly, the noise disappears.

Actually, there is a DC component in equation (49). There is a DC part and a part that vibrates at ω.

The DC component does not disappear even in the case of a prime number type, but the part that oscillates at ω completely disappears. Therefore, the amount of noise is drastically reduced.

If the sun gear is small and has a low meshing ratio, one gear will be meshing for a long time, so the ratio of the vibrating part will be particularly large. In such a case, the noise canceling effect due to the phase shift is particularly noticeable.

Although the effect of noise has been explained, the planetary gear system also has the drawback of generating vibration.

The magnitude of vibration is also controlled by the prime number type planetary gear system.
It can be significantly reduced. The reason is the same as for noise.

In this way, it is possible to obtain a planetary gear device that rotates quietly and smoothly.

Various proposals have been made to eliminate noise and vibration in planetary gear systems, but the present invention differs from these proposals in that it aims to eliminate the most fundamental cause of noise and vibration. This is an excellent invention in principle.

B) Example: Multiple type and prime type planetary gear devices made of iron were manufactured and the noise levels were measured.

The sun gear, planet gear, and outer shell internal gear are all made of steel. The planetary shaft is 5 MM diameter steel. The carrier was made of chromium molybdenum steel.

The number of planetary gears is 3, and the modulus p is 0.8.

Kl Multiple type (conventional type) I=72 S=24 P=24 It is a 1/4 speed reducer.

(bl Prime number type (present invention) ! =73 S=23 P=25 4.It is a 1/17th reduction gear.

These two planetary gear devices are made of the same material and have the same outer circumferential dimensions. The diameter of the gears differs slightly.

Since noise generation is greatest when there is no load, it is assumed that there is no load. In other words, the output shaft was not connected to the carrier, but the motor was connected to the input shaft. Motor rotation is 1800
Chill in times/minute.

A measuring device was placed at a position opposite to the motor at a distance of 50 α, and the noise level μ was measured.

■) The noise level was approximately 8 phon. The noise level in (b) was about the same. Of course, there is temporal fluctuation in both cases, and the average value is the upper value.

It can be seen that the planetary gear system of the present invention is fT effective in reducing noise.

[Brief explanation of drawings]

FIG. 1 is a schematic diagram showing only the pitch circle of the planetary gear system. Po
-P is the planetary axis, q0 to q2 are the meshing points of the sun and planetary gears, and Ro-R is the meshing point of the planetary and outer shell internal gears. S,
P and I represent the sun, planet, and outer shell internal gear, and may also represent the number of teeth. Figure 2 is a schematic diagram of the base circle, tip circle, pitch circle, root circle, line of action, etc. when two gears are in mesh. 0 is the center of the gear, H is the root circle, P is the pitch circle, J is the tip circle, and B is the base circle. It corresponds to gear 1.2 and has a suffix. The line of action FC and the center line O1%02 intersect at point A. The intersections of FG and tip circles J2 and Jl are U and V. FIG. 3 is an explanatory diagram in which an involute curve is drawn for each normal pitch from the base circle. C1T1, etc. are involute curves, and C1, C2, and C3 are starting points taken for each normal pitch on the base circle. FC is a line of action, orthogonal to the involute curve. FIG. 4 is an explanatory diagram for determining the number of partial meshes L(Z) of external gears. P is the pitch circle, J is the tip circle, and FAY is the line of action. a is the pressure angle, and the line of action FAY makes an angle α with the X axis. Inventor Hidetaka Matoba Patent applicant Matex Co., Ltd. Figure 1 〇9

Claims (1)

[Claims] It consists of a sun gear, n planetary gears that surround and mesh with the sun gear, a carrier that pivotally supports the planetary gear by a planetary shaft, and an outer shell internal gear that meshes with the planetary gear. The number of teeth is S, and the number of teeth of the outer shell internal gear is I, and the sum of the numbers of both teeth (S + I) is a multiple of the number n of planetary gears, but it is characterized in that neither S nor I is a multiple of n. Prime number type planetary gear device.