JPS6324158B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to a refrigerant compressor for an automobile air conditioner, and particularly to a drive mechanism for a radial compressor.

There are various types of refrigerant compressors used in air conditioners for automobiles and the like, such as reciprocating type, swash plate type, and radial type, each of which has advantages and disadvantages, and is selected depending on the capacity of the engine.

FIG. 1 is a diagram schematically showing the main parts of a radial compressor as disclosed in, for example, Japanese Unexamined Patent Publication No. 51-124810. A crank pin 2 is eccentrically connected to a crankshaft 1 driven by an engine, and disc-shaped sliders 3a and 3b are eccentrically connected to the crank pin 2 in directions 180 degrees different from each other and are integrally connected to the crank pin 2. They are mated. Further, the sliders 3a and 3b are rotatably fitted into the center holes of double-headed pistons 4a and 4b, respectively.
Piston heads 5a, 5b formed at both ends of the double-headed piston 4a are loosely engaged with cylinders 6a, 6b,
Piston heads 5c, 5d formed at both ends of the double-headed piston 4b are loosely engaged with cylinders 6c, 6d.

The operation of this radial compressor will be explained next.
When the crankshaft 1 rotates clockwise as shown by the arrow, the crank pin 2 also rotates in the same direction. However, since the integral sliders 3a and 3b that are loosely engaged with the crank pin 2 are rotatably fitted into the center holes of the double-headed pistons 4a and 4b that are restricted by the cylinder 6, the dashed lines with arrows indicate It rotates counterclockwise as shown. Therefore, the double-headed pistons 4a and 4b are moved inside the cylinders 6a and 6b with their phases shifted by 90 degrees to suck and inhale the refrigerant gas.
Compression can be performed.

In order to perform such an operation smoothly, a torque must be applied to rotate the sliders 3a and 3b in the direction opposite to that of the crankshaft 1. The conditions will be discussed next.

FIG. 2 is an explanatory diagram showing the radial compressor of FIG. 1 in a balanced state. Let P ₁ and P ₃ be the pressures acting on the double-headed piston 4a, and P ₂ and P ₄ be the pressures acting on the double-headed piston 4b. Also,
The side force acting between the piston head 5 and the cylinder 6 is F _S , and the friction force between the piston head 5 and the cylinder 6 is F _f . When the compressor does not move in balance, the crank pin 2
This is a case where the force F _R acting on the slider 3 is equal to the combined force of F _S and F _f and the direction is opposite. That is, in order to rotate the slider 3 while overcoming the friction loss of each part of the compressor, a torque in the direction of the broken line arrow in FIG. 1 must further act on the slider 3. Such conditions are most necessary in a direction of about 45° to the cylinder 6.

The radial compressor described above has advantages such as a shorter axial length, higher energy efficiency, and fewer component parts than the swash plate type, etc. However, as mentioned above, the cylinder 6 On the other hand, if the crank pin 2 is oriented at 45 degrees, it has the disadvantage that rotation tends to be uneven.

An object of the present invention is to provide a radial compressor that can always operate smoothly, and is characterized in that at least one of the crank pin and the slider and the slider and the double-headed piston has a The reason lies in that it is constructed with a rotating bearing member interposed therebetween.

The present invention is based on the following study results conducted by the inventors in order to find the necessary conditions for solving the above-mentioned drawbacks.

FIG. 3 is a diagram illustrating conditions for rotating the slider from the static balance state of FIG. 2.
The rotation center of the crankshaft 1 is O ₁ , the axial center of the crank pin 2 is O ₂ , and the eccentric distance between C ₁ and O ₂ is r. This r is the rotation radius of the crank pin 2. Further, the diameter of the inner diameter portion of the slider 3 into which the crank pin 2 is fitted is d ₁ , and the outer diameter of the slider 3 is d ₂ . Double-headed piston 4 to slider 3
The force acting on is the force _FN of the double-headed piston 4 pushing the slider 3 perpendicularly to the cylindrical outer peripheral surface toward the center _O3 at point A, which is the point of contact between the two;
This is the resultant force F _t with the frictional force F _f2 generated by F _N.

In this way, the resultant force F _t of the force of the double-headed piston 4 pushing the slider 3 is as follows:
Therefore, the side force in the direction perpendicular to the wall of cylinder 6 shown by the broken line is
It can be decomposed into F _S and parallel friction force F _f1 . In addition, the resultant force F _t is the force F _d that causes the crank pin 2 to push the inner diameter part of the slider 3 at point B due to the drive torque newly applied to the crankshaft 1.
Since it is generated as a reaction force, F _t and F _d are equal in magnitude and opposite in direction. In addition, the above
F _d is the resultant force of the force of the crank pin 2 pushing the slider 3 at point B perpendicular to the inner circumferential surface, _FM , and the frictional force F _f3 generated by this _FM .

Here, the coefficients of friction on the sliding surfaces between the piston head 5 and the cylinder 6, between the double-headed piston 4 and the slider 3, and between the slider 3 and the crank pin 2 are expressed as μ ₁ , μ ² , μ, respectively. If it is ₃ ,
It is clear that the following relationship exists.

F _f1 ≦μ ₁ F _S F _f2 ≦μ ₂ F _N F _f3 ≦μ ₃ F _M ...(1) That is, just before this mechanism starts to move due to the driving torque applied to the crankshaft 1,
The equality sign in each equation (1) holds true.

Now, when μ ₁ , μ ₂ , and μ ₃ are sufficiently small, even if each frictional force reaches its maximum just before the piston drive mechanism starts moving from a stationary state, the forces F _t and F _d acting on the slider 3 and each point of application A , B are as shown in FIG. 3, and F _t and F _d form a couple that rotates the slider 3 in the direction of the arrow. At this time, this piston drive mechanism operates smoothly in the forward direction.

On the other hand, when the static friction coefficients μ ₁ , μ ₂ , μ ₃ increase to a certain extent, the respective frictional forces F _f1 , F _f2 , F _f3 increase to μ ₁ F _S , μ ₂ F _N , μ ₃ F _M , respectively. Before growing up, F _t and
Since the force acts on the same line as F _d and does not form a couple that attempts to rotate the slider 3 in the forward direction in a completely balanced manner, the piston drive mechanism does not operate in this case.

FIG. 4 is an explanatory diagram of a state in which the piston drive mechanism is not in operation, and the same parts as in FIG. 3 are given the same reference numerals. In this case, μ ₁ , μ ₂ , μ ₃ are large, so F _t and
The points of action A and B of F _d have the positional relationship shown in the figure.
That is, F _t and F _d are equal in magnitude and opposite in direction, and act on the same line.
At this time, F _t and F _d are perfectly balanced, the slider 3 does not move, and this piston drive mechanism does not operate.

In this state, the static friction coefficient μ, the eccentricity γ of the crank pin 2, the inner diameter d ₁ and the outer diameter of the slider 3
It is determined by the value of d ₂ and is unrelated to the magnitude of the torque that attempts to rotate the crankshaft 1, so it will not operate no matter how large the driving torque is applied. However, the friction coefficients μ ₁ , μ ₂ , μ ₃ of each part, the eccentricity γ of the crank pin 2, and the inner diameter of the slider 3
By selecting P ₁ and the outer diameter d ₂ , it is possible to generate a force couple that rotates the slider 3 in the forward direction as shown in FIG. Let's consider this condition further.

As shown in Figure 3, connect the center O ₁ of crankshaft 1 and the center O ₂ of crank pin 2.
The O ₁ O ₂ line makes an angle of 45° to the horizontal center axis of the double-headed piston 4b, and the center O ₂ of the crank pin 2
In order to operate smoothly even when the O ₂ O ₃ line connecting O 3 and the center O ₃ of the slider 3 forms an angle of 45° with respect to the vertical axis of the double-headed piston 4a, μ ₁ ,
The following relationship must hold between μ ₂ , μ ₃ , γ, d ₁ , and d ₂ .

Rsin (θ ₄ −θ ₁ )＞d ₁ /2sinθ ₃ ...(2) However, θ ₁ = tan ^-1 μ ₁ , θ ₂ = tan ^-1 μ ₂ , θ ₃ = tan ^-1 μ ₃
, θ ₄ = tan ^-1 [{d ₂ /2sin (θ ₁ −θ ₂ ) + γsin45゜}/
{d ₂ /2cos(θ ₁ −θ ₂ )+γcos45゜}] It is.

Equation (2) above is easier to hold as μ ₁ , μ ₂ , μ ₃ , d ₁ , d ₂ are smaller and γ is larger, but since it is necessary to maintain mechanical strength, d ₁ , d ₂ , γ cannot be made too small. Also, since μ ₁ is the relationship between the piston head 5 and cylinder 6, it is of course effective to improve the lubrication of this, but it is appropriate to reduce μ ₂ and μ ₃ for this piston drive mechanism. We have reached the conclusion that this is an improvement measure.

FIG. 5 is a vertical sectional view of a radial compressor according to an embodiment of the present invention, in which the same parts as in FIG. 3 are given the same reference numerals. Crankshaft 1a, 1b
is driven by an engine, and the crank pin 2 is formed parallel to the middle part thereof with an offset of γ. This crank pin 2 is fitted into a slider 3 via a pair of needle bearings 11a, 11b, and the slider 3 is fitted with needle bearings 10a, 1
It is fitted into the center hole of the double-headed piston 4a via 0b. A double-headed piston 4b of the same shape is installed in a direction perpendicular to this double-headed piston 4a (direction perpendicular to the plane of the paper). Piston heads 5a and 5b are fixed to both ends of the double-headed piston 4a and slide within the respective cylinders 6a and 6b, and refrigerant flows through a refrigerant passage 14 provided in the piston head 5 and is compressed.

This type of radial compression mechanism is mounted on the mounting plate 1.
The crankshaft 1a and the crankshaft 1b, which is an extension of the crankshaft 1a, are supported by the cylinder block 8 via ball bearings 7a and 7b. ing. Further, a presser spring 9 and a mechanical seat 15 are installed in the narrow diameter portion at the left end of the crankshaft 1 to prevent leakage of refrigerant. Note that balance weights 16a and 16b are fixed to the crankshafts 1a and 1b, respectively, to compensate for unbalance due to eccentricity of the crank pin 2.

The above needle bearings 10 and 11 are 2.5 to 3
mm diameter and rolling friction, the friction coefficients μ ₂ and μ 3 between the crank pin 2 and the slider 3 and between the slider 3 and the double-headed piston ₄ are significantly reduced. Therefore, the operation of the double-headed piston 4 is made smooth and the reliability of this piston drive mechanism is increased.
It becomes possible to improve the compression efficiency of the refrigerant.

The radial compressor of this embodiment has needle bearings interposed between the crank pin and the slider and between the slider and the center hole of the double-headed piston to ensure smooth operation of the double-headed piston and improve compression efficiency. This has the effect of improving.

In the above embodiment, needle bearings 10 and 11 are interposed on the inside and outside of the slider 3, but similar effects can be obtained even if either the needle bearing 10 or the needle bearing 11 is omitted.

The radial compressor of the present invention has the advantage that the double-headed piston always operates smoothly.

[Brief explanation of drawings]

Figure 1 is a diagram schematically showing the main parts of a radial compressor, Figure 2 is an explanatory diagram showing the radial compressor in Figure 1 in a balanced state, and Figure 3 is a diagram showing the static state of Figure 2. A diagram explaining the conditions for rotating the slider from a balanced state, FIG. 4 is an explanatory diagram of a state in which the piston drive mechanism of FIG. 3 does not operate, and FIG. FIG. 1...Crankshaft, 2...Crank pin, 3...Slider, 4...Double-headed piston, 5
...Piston head, 6...Cylinder, 7...Ball bearing, 8...Cylinder block, 9...
... Pressure spring, 10, 11 ... Needle bearing, 12 ... Side cover, 13 ... Mounting plate,
14... Refrigerant passage, 15... Mechanical seat,
16...Balance weight.

Claims (1)

[Claims] 1. A crank pin eccentrically connected in parallel to a crankshaft driven by an engine;
A pair of disc-shaped sliders that fit eccentrically in directions 180 degrees different from each other on the crank pin, a pair of double-headed pistons having holes into which the pair of sliders fit, and a pair of double-headed pistons that are slidably connected to each other. The radial type has four moving cylinders formed at right angles to each other, and as the crankshaft rotates, the double-headed piston slides inside the cylinder with a phase difference of 90 degrees, compressing the refrigerant. A radial compressor, characterized in that a rotary bearing member is interposed between at least one of the crank pin and the slider and between the slider and the double-headed piston. 2. The radial compressor according to claim 1, wherein the rotary bearing member is a needle bearing.