JPH09151848A - Reciprocation type compressor - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a reciprocation type compressor to reduce a rotation n-degree component of a torque fluctuation corresponding to the number of cylinders (n) and decrease the generation of noise and vibration. SOLUTION: The radial center Pa1 of a cylinder bore 12a (11a) to contain a single piston 28 is arranged on a radius line Ra passing an equal angular distance position specified on the arrangement periphery of the radius center of the cylinder bore. Cylinder bores 12b-12e (11b-11e) to contain remaining pistons 28 are arranged in such a manner that the radius centers Pb1-Pe1 thereof are positioned on radius lines rb1-re1 having a given deviation angle Δθwith radius lines Rb-Re passing an equal angular distance position. The half number in each lot of the radius lins rb1-re1 is moved from the radius line Rb-Re to the front side or the rear side in the direction of rotation of a drive shaft 32.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a reciprocating compressor used in, for example, a vehicle air conditioner.

[0002]

2. Description of the Related Art In a reciprocating compressor of this type, a drive shaft is supported inside a housing of the compressor.
A crank chamber is formed. A plurality of cylinder bores are arranged in parallel with each other around the drive shaft of a cylinder block forming a part of the housing. A piston is reciprocally housed in the cylinder bore to define a compression chamber. A swash plate is integrally rotatably attached to the drive shaft, and the piston is reciprocated in association with the rotation of the swash plate to compress the refrigerant gas in the compression chamber.

In such a reciprocating compressor, a compression reaction force is generated with the compression operation of each piston. This compression reaction force acts on the drive shaft via the swash plate, causing torque fluctuations. This torque fluctuation becomes an exciting force that causes torsional vibration of the drive shaft-clutch system. Here, when a fast Fourier transform (FFT) analysis is performed on the sum of the torque fluctuations, in other words, the sum of the compression reaction forces generated in the respective compression chambers, a wide range of frequency components from the 0th order to a considerably higher order is obtained. Of these frequency components, the main component is the n-th rotation component corresponding to the number of cylinders n. For example,
In a 10-cylinder type double-headed piston compressor, a rotating tenth-order component as shown in FIG. 9 is the main component. In the case of 10 cylinders, the phase of the curve showing the compression reaction force of each compression chamber is
It is offset by 36 °. On the other hand, as shown in FIG. 9 (a), the rotational tenth-order component repeats the same displacement ten times in the time of one rotation of the drive shaft,
One cycle is 360 ° / 10 = 36 °. Therefore, the phases of the torque fluctuations in the compression chambers completely match, and as shown in FIG. 9B, the amplitude of the torque fluctuations of the compressor as a whole is the amplitude a of the torque fluctuations per compression chamber. It will be 10 times. When the frequency of the n-th order rotation component and the like is close to the natural frequency of the compressor and the auxiliary machinery connected to the compressor, noise and vibration due to the resonance phenomenon occur, and the vehicle interior Was a cause of increasing the noise and vibration level of the.

In order to solve such a problem, for example, Japanese Patent Publication No. Sho 48-19121 discloses an axial plunger pump in which cylinder bores are arranged at unequal pitches in the circumferential direction. In this pump, the n cylinder bores accommodating the pistons respectively form an angle between the adjacent cylinder bores and the axial center of 360 ° / n.
It is arranged so as to be ± γ. This γ has a different value among the cylinder bores, and is set to be (360 ° / n ± γ) × n = 360 ° for one round of the pitch circle of each cylinder bore. Then, the discharge amount from each cylinder bore is made unequal, and the flow rate at the time of resonance,
It is intended to reduce fluctuations in torque, pressure and rotation speed.

[0005]

However, the above publication merely discloses changing the pitch of each cylinder bore in order to reduce the torsional vibration. That is, no specific measures against the torque fluctuation of the drive shaft are disclosed or suggested. For this reason, there has been a problem that the torque fluctuation cannot be sufficiently reduced, and the generation of noise and vibration may not be sufficiently suppressed.

It is an object of the present invention to provide a reciprocating compressor which is an exciting force of torsional vibration, in which the rotational nth order component of torque fluctuation corresponding to the number of cylinders n is reduced, and which causes less noise and vibration. To do.

[0007]

According to a first aspect of the present invention, a drive shaft is supported inside a housing and a crank chamber is formed.
A plurality of cylinder bores are arranged around the drive shaft in a cylinder block that forms a part of the housing, and a piston is reciprocally housed in the cylinder bore to partition and form a compression chamber. In a reciprocating compressor in which a cam plate is integrally rotatably attached and reciprocally moves in conjunction with the rotation of the cam plate to compress the refrigerant gas, the number of pistons is an odd number. , Arranging the equiangularly spaced positions corresponding to the number of the pistons on the circumference of the arrangement of the radial centers of the cylinder bores, arranging the radial centers of one cylinder bore at the equiangularly spaced positions, and the radius of the remaining half of the cylinder bores. From the equiangularly spaced positions to the center in front of the rotation direction of the drive shaft,
The radius centers of the other half of the remaining cylinder bores are displaced from the equiangularly spaced positions to the rear side in the rotational direction of the drive shaft with a predetermined displacement angle about the axis of the drive shaft.

According to a second aspect of the invention, in the invention of the first aspect, the shift angle satisfies the following equation. Δθ = arc cos (−1 / 2n) / n where Δθ offset angle and n indicate the number of cylinders in the compressor.

According to a third aspect of the present invention, a plurality of cylinder bores are provided around the drive shaft in a cylinder block that supports the drive shaft inside the housing and forms a crank chamber and forms a part of the housing. , A piston is reciprocally housed in the cylinder bore to define a compression chamber, a cam plate is integrally rotatably attached to the drive shaft, and the cam plate is interlocked with the rotation of the cam plate. In a reciprocating compressor in which a piston is reciprocated to compress a refrigerant gas, the number of pistons is an even number, and an equiangular interval corresponding to the number of the pistons is arranged on the circumference of the cylinder bore radial center. The positions are defined, and the radius centers of the other half of the cylinder bores are set from the radial center of the half of the cylinder bores to the front side in the rotation direction of the drive shaft from the equiangularly spaced positions. From equiangularly located heart in the rotation direction rear side of the drive shaft, respectively around the axis of the drive shaft is obtained by shifting with a predetermined shift angle.

According to the invention described in claim 4, in the invention described in claim 3, the shift angle satisfies the following expression. Δθ = 180 ° / n where Δθ shift angle and n indicate the number of cylinders in the compressor.

According to the fifth aspect of the invention, the first to fourth aspects are provided.
In the invention described in any one of 1 to 3, the compression chamber is set to a dead volume of at least two kinds of values.

According to the sixth aspect of the present invention, the first to fifth aspects are provided.
In any one of the inventions, the dead volume of the compression chamber in the cylinder bore displaced from the equiangularly spaced position to the front side in the rotational direction of the drive shaft is enlarged as compared with a reference dead volume, and the equiangularly spaced position makes The dead volume of the compression chamber in the cylinder bore displaced to the rear side in the rotational direction of the drive shaft is reduced as compared with the reference dead volume.

According to a seventh aspect of the invention, in the sixth aspect of the invention, the reference dead volume is determined by a relationship diagram between the displacement angle and the dead volume obtained for each model of the compressor. Is.

According to the invention described in claim 8, in claims 1 to 7,
In any one of the inventions described above, the cylinder bores are formed so as to face each other in the front-rear direction, and the piston is configured as a double-headed type.

According to a ninth aspect of the invention, in the eighth aspect of the invention, the front dead volume and the rear dead volume are formed to have the same size with respect to one double-headed piston. is there.

In the reciprocating compressor having the above-described structure, the equiangular interval positions can be obtained by equally dividing the entire circumference by the number of pistons on the circumference of the array around the radius center of the cylinder bore. Stipulated.

First, when the number of pistons is odd,
The radial center of one cylinder bore is arranged at one of the equiangular intervals. The centers of the radii of the remaining half of the cylinder bores are displaced from the equiangularly spaced positions at the front side in the rotational direction of the drive shaft with a predetermined displacement angle about the axis thereof. On the contrary, the radial centers of the other half of the cylinder bores are displaced from the equiangularly spaced positions at the rear side in the rotational direction with a predetermined displacement angle.

Next, when the number of pistons is even,
The radial centers of the half of the cylinder bores are displaced from the equiangularly spaced positions at the front side in the rotation direction of the drive shaft with a predetermined displacement angle about the axis thereof. On the contrary, the radial centers of the other half of the cylinder bores are offset from the equiangular intervals with a predetermined offset angle on the rear side in the rotational direction.

Therefore, the phase of the torque fluctuation based on the compression reaction force generated in each compression chamber is displaced by the shift angle as compared with the case where the cylinder bores are arranged at equal angular intervals. Further, unlike the case where the cylinder bores are arranged at equiangular intervals, the phases of torque fluctuations occurring in the compression chambers do not completely match. Then, the rotational n-th order component corresponding to the number of cylinders n obtained by the fast Fourier transform analysis of the sum of the torque fluctuations is reduced as compared with the case where the cylinder bores are arranged at equal angular intervals.

The torque fluctuations generated in the compression chambers are sinusoidal waves having the same period. Therefore, by setting a predetermined phase difference in the plurality of sine waves, the phases of the plurality of sine waves can be inverted, and the amplitude of the torque fluctuation can be minimized as shown in FIG.

First, in order to superimpose an odd number (2m + 1, m is a natural number) of the sine waves to minimize the amplitude, the result of superimposing the 2m sine waves and the remaining one cancel each other. do it. That is, assuming that the sinusoidal wave of the torque fluctuation of the cylinder bores arranged at equal angular intervals is sin θ, and shifting 2m by half in the opposite direction with respect to the sin θ, m {sin (θ + Δψ) + sin (θ−Δψ)} It suffices to obtain the phase difference Δψ that satisfies + sin θ = 0 (1). This equation (1) is
It is transformed as follows.

2m · cos Δψ · sin θ + sin θ = 0 (2) When this equation (2) is solved for Δψ, Δψ = arc cos (−1 / 2m) (3). That is, the odd number of sine waves are canceled out by superimposing the odd number of sine waves by shifting the 2m sine waves in half in the opposite direction with respect to the sin θ with respect to the rotation angle of 360 ° of the drive shaft by Δψ satisfying the expression (3). be able to. Here, the rotation nth order component is a waveform in which the same phase is repeated n times within a time period corresponding to one rotation of the drive shaft,
One cycle of the phase is 360 ° / n. When the shift angle Δθ of each cylinder bore that minimizes the amplitude by canceling out the n-th order component of the rotation is obtained, Δθ = Δψ / n = arc cos (−1/2 m) / n (4).

Next, in order to superimpose an even number of the above-mentioned sine waves to minimize the amplitude, it suffices to shift half of the phases by 180 °. Here, one cycle of the rotational n-th order component in question is 360 ° / n as described above. Therefore, with respect to the rotation direction of the drive shaft, half of the cylinder bores are 180 ° / n in the front side and the other half are 18 in the rear side.
By shifting 0 ° / n, it is possible to secure the phase shift that minimizes the amplitude. That is, in order to minimize the amplitude, each cylinder bore may be displaced so that the displacement angle of each cylinder bore is Δθ and the following equation is satisfied.

Δθ = 180 ° / n (5) By the way, the size of the compressor, the degree of proximity of each cylinder bore,
And considering its safety factor, the above equations (4) and (5)
In some cases, it is not possible to secure shift movement that satisfies the formula. In such a case, by changing the value of the dead volume of each compression chamber, it is possible to reduce the rotational n-th order component of the torque fluctuation that is the exciting force of the torsional vibration. The dead volume is the volume of the compression chamber when the piston reaches the top dead center. By changing the value of the dead volume in each compression chamber, the curve of the transition of the volume and the pressure during the compression operation in each compression chamber becomes different. Therefore, the phase of the torque fluctuation based on the compression reaction force generated in each compression chamber is deviated, and the n-th order rotation component corresponding to the number of cylinders n is not completely superimposed. Then, the same effect of reducing the torque fluctuation as that of the shift movement of the cylinder bores is exhibited.

Here, in the reciprocating compressor constructed as described above, the effect of reducing the torque fluctuation is determined by the shift movement of the cylinder bore and the change in the dead volume of each compression chamber. be able to. Therefore, when the displacement of the cylinder bores that satisfies the above formulas (4) and (5) cannot be ensured, the shortage of the displacement angle can be compensated by changing the value of the dead volume of each compression chamber.

Further, the dead volume of the compression chamber in the cylinder bore displaced from the equiangularly spaced positions to the front side in the rotational direction of the drive shaft is larger than the reference dead volume. Further, the dead volume of the compression chamber inside the cylinder bore, which is displaced rearward in the rotational direction of the drive shaft, is smaller than the reference dead volume. For this reason, the tendency of reducing the torque fluctuation due to the shift movement of the cylinder bore and the tendency of decreasing the torque fluctuation due to the change of the dead volume are consistent with each other.

Moreover, in the double-headed piston type compressor constructed as described above, in addition to the shift movement of the cylinder bore and the change of the dead volume as described above, the dead side on the front side of one double-headed piston is The volume and the dead volume on the rear side are formed to have the same size. The phase of the compression reaction force in this double-headed piston compressor has a difference of 180 ° between the sum total on the front side and the sum total on the rear side. Here, the rotational n-th order component corresponding to the number of cylinders n of the torque fluctuation which is the exciting force of the torsional vibration is an even-order component, and the phase thereof is the same displacement even within the time corresponding to one rotation of the drive shaft. It has to be repeated. For this reason, the front-side total and the rear-side total of the rotation n-order component are superimposed in phase. However, by shifting the cylinder bore and changing the dead volume as described above, the sum of the n-th order component of the rotation on the front side and the sum on the rear side are respectively reduced. Then, the rotational n-th order component of the entire compressor in which the front side total and the rear side total are superimposed is also reduced. Moreover, when the rotational n / 2-order component becomes an odd-order component, the odd-order component repeats the same displacement an odd number of times within the time corresponding to one rotation of the drive shaft.
Therefore, there is a phase difference of 180 ° between the front side and the rear side of the rotational n / 2 quadratic component, and the phases are opposite to each other. Then, the rotational n / 2nd order component cancels each other and disappears on the front side and the rear side of one piston.

[0028]

BEST MODE FOR CARRYING OUT THE INVENTION

(First Embodiment) A first embodiment of the present invention will be described below with reference to FIGS.

As shown in FIG. 1, the cylinder block 11 on the front side and the cylinder block 12 on the rear side are joined at the central portion. A front housing 15 is joined to the front end surface of the cylinder block 11 via a valve plate 13, and a rear housing 16 is joined to the rear end surface of the cylinder block 12 via a valve plate 14. The cylinder blocks 11 and 12 and the valve plates 1
Suction valve forming plates 17 and 18 which form suction valves 17a and 18a, respectively, are interposed between the valves 3 and 14. Discharge valve forming plates 19 and 20 that form discharge valves 19a and 20a are interposed between the valve plates 13 and 14 and the housings 15 and 16, respectively. Discharge valve forming plate 1
Retainer plates 21 and 22 for restricting the maximum openings of the discharge valves 19a and 20a are interposed between the housings 9 and 20 and the housings 15 and 16, respectively.

The cylinder blocks 11 and 12, the valve plate 1
3, 14, suction valve forming plates 17, 18 and discharge valve forming plate 1
9 and 20 are fastened and fixed to each other by a plurality of through bolts 23.

Suction chambers 24 and 25 are formed on the outer circumferences of the front housing 15 and the rear housing 16, and discharge chambers 26 and 27 are defined on the center side. FIG.
And as shown in FIG. 2, the cylinder blocks 11, 1
2, a plurality of cylinder bores 11a to 11e and 12a to
12e are penetratingly formed so as to be parallel to each other, and a double-headed piston 28 is inserted inside them. The compressor of this embodiment includes 10 double-headed pistons 28.
It is a cylinder type reciprocating compressor. A pair of front and rear compression chambers 29 and 30 are formed in the cylinder bores 11a to 11e and 12a to 12e. This compression chamber 29,
30 is an intake port 13 formed in the valve plates 13 and 14.
a, 14a into the suction chambers 24, 25, and also discharge ports 13b, 14b formed in the valve plates 13, 14
Is communicated with the discharge chambers 26, 27 via.

A crank chamber 31 is formed between the cylinder blocks 11 and 12. A drive shaft 32 is rotatably supported by radial bearings 33 in shaft holes 11f and 12f of both cylinder blocks 11 and 12, respectively. The drive shaft 32 is rotated by an external drive source such as a vehicle engine via a clutch (not shown).
A swash plate 34 as a cam plate is fitted and fixed to the intermediate outer peripheral portion of the drive shaft 32. The double-headed piston 28 is anchored to the swash plate 34 via shoes 35 and 36, and the double-headed piston 28 is reciprocated in the cylinder bores 11a to 11e and 12a to 12e by the rotation of the swash plate 34.

The boss portion 34a of the swash plate 34 is supported on both front and rear wall surfaces of the cylinder blocks 11 and 12 via thrust bearings 37 and 38. As shown in FIG. 1, the crank chamber 31 is in communication with the suction chambers 24, 25 by suction passages 39, 40 formed in the cylinder blocks 11, 12. The crank chamber 31 is the cylinder block 1
It is connected to the external refrigerant circuit via suction flanges (not shown) formed on the first and the second refrigerant passages. Further, the discharge chamber 2
Reference numerals 6 and 27 are connected to an external refrigerant circuit via discharge passages 41 and 42 formed in the cylinder blocks 11 and 12, both housings 15 and 16, and a discharge flange (not shown).

When the drive shaft 32 is rotated by an external drive source such as a vehicle engine, the swash plate 34 in the crank chamber 31 is rotated, and the plurality of double-headed pistons 28 are connected to the cylinder bores 11a through 11 through the shoes 35 and 36. 11e, 1
It is reciprocated within 2a to 12e. This double-headed piston 28
From the suction flange (not shown) by the movement of the crank chamber 3
The refrigerant gas guided to No. 1 is guided to the suction chambers 24, 25 from the crank chamber 31 via the suction passages 39, 40. In the suction stroke in which the double-headed piston 28 moves from the bottom dead center to the top dead center, the suction valves 17a and 18a are opened and the suction chamber 2
The refrigerant gas in 4 and 25 is sucked into the compression chambers 29 and 30 through the suction ports 13a and 14a. Next, the refrigerant gas in the compression chambers 29 and 30 is compressed in the compression / discharge stroke in which the double-headed piston 28 moves from the top dead center to the bottom dead center. When the refrigerant gas reaches a predetermined pressure, the high-pressure compressed refrigerant gas pushes out the discharge valves 19a, 20a and is discharged into the discharge chambers 26, 27 through the discharge ports 13b, 14b. Furthermore, the compressed refrigerant gas in the discharge chambers 26 and 27 is
A condenser forming an external refrigerant circuit via the discharge passages 41, 42,
It is supplied to the expansion valve and the evaporator and used for air conditioning in the vehicle interior.

Next, each of the cylinder bores 11a-11
The arrangement of e and 12a to 12e will be described. The front-side cylinder bores 11a to 11e and the rear-side cylinder bores 12a to 12e are paired with the double-headed piston 28, respectively. Therefore, the front-side cylinder bores 11a to 11e and the rear-side cylinder bores 12a to 12e have exactly the same arrangement. Here, only the rear cylinder bores 12a to 12e will be described.

As shown in FIG. 2, the cylinder bores 12a ...
The array circumference C of the radial centers Pa1 to Pe1 of 12e has the radial center Po on the axis of the drive shaft 32. Ra-R
e and rb1 to re1 are radial lines of the cylinder blocks 11 and 12 that pass through the radial center Po. Radial lines Ra to Re
Passes through equiangularly spaced positions where the array circumference C is equally divided by the number of pistons 28. In this embodiment, the number of pistons 28 is 5, and the radius lines Ra to Re are arranged at 72 ° intervals. The number of pistons 28 is an odd number, and the cylinder bore 12a accommodating one piston 28 has a radial center P
a1 is arranged on the radial line Ra, that is, at one of the equiangular intervals. The two cylinder bores 12b and 12d that respectively accommodate half of the remaining pistons 28 have their radial centers Pb1 and Pd1 on the radial lines rb1 and rd1. The radius lines rb1 and rd1 are displaced by a predetermined angle from the radius lines Rb and Rd to the front side in the rotation direction of the drive shaft 32. The two cylinder bores 12c and 12e that accommodate half of the other remaining pistons 28 have their radial centers Pc.
1 and Pe1 are on the radial lines rc1 and re1. The radius lines rc1 and re1 are shifted by a predetermined angle to the rear side in the rotation direction of the drive shaft 32 with reference to the radius lines Rc and Re.

The predetermined angle is calculated by the equation (4). Here, since the number of pistons is 5, 2m + 1 = 5 and m = 2.

Δθ = arc cos (−1/2 m) / n (4) = arc cos (−1/4) / 10 ≈104 ° / 10 = 10.4 ° That is, the cylinder bores are arranged at equiangular intervals. Cylinder bores 11b-11e, 12b-1 compared to
The torque fluctuations generated in the compression chambers 29 and 30 in 2e have a phase difference of about 104 °. here,
The 10th rotational component of the torque fluctuation, which is the main factor of noise and vibration, is 1 with respect to the rotation angle of 360 ° of the drive shaft 32.
Since the same displacement is repeated 0 times, one cycle becomes 36 °. Therefore, each cylinder bore 11b-11e, 12b-
In 12e, the shift angle for eliminating the rotational 10th order component is 10.4 °. Accordingly, the radius line rb
1 to re1 are respectively shifted by 10.4 ° with respect to the radial lines Rb to Re.

A compression reaction force is generated in each of the compression chambers 29 and 30 in accordance with the compression operation of the double-headed piston 28. The torque fluctuation based on this compression reaction force is a sine wave with the same period. Here, in the compressor of the present embodiment,
A pair of cylinder bores 11a, 12a accommodating one double-headed piston 28 are arranged at one of the equiangular intervals. On the other hand, each of the cylinder bores 11b to 11e and 12b to 12e corresponding to the other double-headed piston 28 is shifted by + 10.4 ° or -10.4 ° from the equiangular interval position by half in the rotational direction of the drive shaft 32. ing. Therefore, the cylinder bores 11b to 11e,
The rotation tenth-order component of the sum of the torque fluctuations generated in the compression chambers 29 and 30 in 12b to 12e is the cylinder bore 1
It has a phase difference of 104 ° with respect to the same components generated in the compression chambers 29 and 30 in 1a and 12a.
Therefore, the rotational 10th order component of the entire compressor is canceled,
The excitation force that causes torsional vibration is reduced.

According to the present embodiment configured as described above, the following excellent effects are exhibited. In the compressor of the present embodiment, the number of pistons 28 is 5, and the radial centers Pa of the cylinder bores 11a and 12a accommodating one piston 28 are arranged on the radial line Ra corresponding to equiangular intervals. The cylinder bores 12b and 12d that accommodate half of the remaining pistons 28 have their radial centers Pb1 and Pd.
1 is on radial lines rb1 and rd1 which are displaced from the equiangularly spaced positions by 10.4 ° to the front side in the rotational direction of the drive shaft 32. Cylinder bores 12c and 12e that accommodate half of the other remaining pistons 28 have radial centers rc1 and re1 whose radial centers Pc1 and Pe1 are displaced from the equiangularly spaced positions by 10.4 ° to the rear side in the rotational direction of the drive shaft 32. It is above.

Here, the phase of torque fluctuation based on the compression reaction force generated in each compression chamber 29, 30 will be considered. The phase of torque fluctuation caused by the compression reaction force in the cylinder bores 11a and 12a and the other cylinder bores 11b to 11e and 12
Between the sum of the phases of the torque fluctuations due to the compression reaction force at b to 12e, 104 for one rotation of the drive shaft 32.
This means that the phase difference of ° has been set. Then, the total sum of the torque fluctuations of the entire compressor becomes almost zero, and the rotational tenth-order component of the torque fluctuations is greatly reduced compared to the case where the pistons are arranged at equal angular intervals. Therefore, the excitation force for the torsional vibration between the drive shaft 32 and the clutch (not shown) is reduced, the noise caused by the resonance phenomenon excited by the excitation force is reduced, and the noise level in the vehicle interior can be reduced. .

(Second Embodiment) The second embodiment of the present invention will be described below.
The embodiment will be described with reference to FIGS. 4 to 6. In this second embodiment, for example, in a 10-cylinder type double-headed piston compressor, compared with the cylinder bores 12a to 12e of the first embodiment shown in FIG. 2, the cylinder bores 52a to 52e (51a to 51e) are Large diameter,
It is suitable when the predetermined displacement angle cannot be secured due to the interference of the adjacent cylinder bores.

Also in this second embodiment, the first
Similarly to the embodiment of FIG.
a to 51e and the rear cylinder bores 52a to 52e
Are formed in a similar manner to the double-headed piston 28, respectively. Therefore, the cylinder bores 52a to 52e on the rear side will be mainly described.

As shown in FIG. 4, the radial center Pa2 of the cylinder bore 52a accommodating one piston 28 is arranged on the radial line Ra corresponding to the equiangular intervals. The two cylinder bores 52b and 52d that respectively accommodate half of the remaining pistons 28 have their radial centers Pb.
2 and Pd2 are on the radial lines rb2 and rd2. The radius lines rb2 and rd2 are displaced from the radius lines Rb and Rd passing through the equiangularly spaced positions by an angle α1 toward the front side in the rotation direction of the drive shaft 32. Two cylinder bores 52c, 52, which each house half of the other piston 28
In e, the radial centers Pc2 and Pe2 have radial lines rc2 and r.
It is on e2. The radial lines rc2 and re2 are calculated from the radial lines Rc and Re passing through the equiangularly spaced positions from the drive shaft 3
It is offset by an angle α1 on the rear side in the rotational direction of 2. Note that adjacent cylinder bores, for example, 52b and 52
The angle α1 is set so that the adjacent cylinder bores do not interfere with each other and the desired strength is ensured in the partition wall between the adjacent cylinder bores even when c and the angle c1 are approached from each other at equal angular intervals. It

Further, as shown in FIGS. 4 and 5, the cylinder bores 52a to 52e are formed to have the same inner diameter. Then, each cylinder bore 52a-
The head of each double-headed piston 28 housed in 52e is scraped off by a predetermined length. Therefore, when each piston 28 reaches the top dead center position, the head end surface of the piston 28 and the cylinder bores 52a to 5a-5.
The distance between the inner end surface of 2e is different. Therefore, the dead volume in each compression chamber 30 is set to a different value. Here, the dead volume is the volume of the compression chamber 30 when the piston 28 reaches the top dead center position.

Further, the double-headed piston 28 is formed such that the front side and the rear side thereof have the same scraping amount. For this reason, the dead volume value of the compression chamber 29 on the front side and the dead volume value of the compression chamber 30 on the rear side of one double-headed piston 28 are the same. In other words, the compression chamber 29 and the cylinder bore 5 in the cylinder bore 51a that face each other in the axial direction of the drive shaft 32 via the double-headed piston 28.
The compression chamber 30 in 2a is set to the same dead volume. Similarly, cylinder bores 51b and 52b,
In 51c and 52c, 51d and 52d, 51e and 52e, the values of the dead volumes of the compression chamber 29 and the compression chamber 30 are the same. Therefore, the arrangement of the dead volumes of the front-side compression chambers 29 and the arrangement of the dead volumes of the rear-side compression chambers 30 are the same in the rotational direction of the drive shaft 32.

As shown in FIG. 5, the size of the dead volume of each compression chamber 30 is determined by the cylinder bore 5 whose radial center Pa2 is located on the radial line Ra corresponding to the equiangular intervals.
The value Vmid of the dead volume of the compression chamber 30 in 2a is used as a reference. The value of the dead volume in the cylinder bores 52b, 52d with the radial centers Pb2, Pd2 displaced to the front side in the rotational direction of the drive shaft 32 is obtained by adding a predetermined value V1 to the reference dead volume value Vmid. It has become a thing. Also, the radius center P
The values of the dead volume in the cylinder bores 52c, 52e in which c and Pe are displaced to the rear side in the rotational direction of the drive shaft 32 are the dead volume value Vmi serving as the reference.
It is a value obtained by subtracting a predetermined value V2 from d.

Next, the setting of the dead volume of each compression chamber 30 will be described. Now, as described above, by changing the dead volume of each compression chamber 30, the curve of the transition of the volume and the pressure in the compression operation in each compression chamber 30 becomes different. Then, the phase of the torque fluctuation generated in each compression chamber 30 is deviated. Therefore, even if the dead volume of each compression chamber 30 is changed, the rotational tenth-order component is reduced similarly to the case where the cylinder bore is displaced.

Regarding the effect of reducing this 10th order rotation component,
A relationship diagram between the displacement angles Δθ of the cylinder bores 52b to 52e and the value of the dead volume of each compression chamber 30 as shown in FIG. 6 can be obtained for each compressor model. Here, Vmin represents the minimum dead volume value that is allowable in manufacturing, and Vmax represents the maximum dead volume value that is maximized to the extent that the compression performance of the compressor is not extremely reduced. Further, Δθmin and Δθmax are shift angles corresponding to the tenth-order component of rotation when the compression chambers 30 are set to the dead volume values of Vmin and Vmax, respectively. Δθm
id is defined by the following equation.

Δθmid = (Δθmin + Δθmax) / 2 (6) This Δθmid is made to correspond to the shift angle when the cylinder bores are arranged at equal angular intervals, that is, Δθ = 0 °. Then, the reference dead volume Vmid is set using the relationship diagram of FIG.

Then, in each of the cylinder bores 52b to 52e, the dead volume values V1 and V2 corresponding to the insufficient angle β1 with respect to the desired angle of the angle α1 are set based on the relationship diagram. Here, the cylinder bore 52b displaced to the front side in the rotation direction of the drive shaft 32,
The value of the dead volume in the compression chamber 30 in 52d is set to Vmid + V1 corresponding to the shift angle Δθmid + β1. In addition, the value of the dead volume in the compression chamber 30 in the cylinder bores 52c and 52e, which is displaced toward the rear side in the rotational direction of the drive shaft 32, is the displacement angle Δθmid−β.
It is set to Vmid-V2 corresponding to 1.

If Δθmid + β1> Δθmax or Δθmid−β1 <Δθmin, then
The dead volume value of the cylinder bore is V
It is set to max or Vmin.

According to the present embodiment configured as described above, the following excellent effects are obtained. (A) Due to the size of the compressor, the proximity of each cylinder bore, and its safety factor, the cylinder bores 51b to 5b are not secured to move to satisfy the desired displacement angle Δθ.
The dead volumes of the compression chambers 29 and 30 of 1e and 52b to 52e are changed. Therefore, the shortage of the phase difference of the torque fluctuation caused by the shortage of the shift angle Δθ is supplemented by the change of the dead volume. Then, in the 10-cylinder type double-headed piston type compressor, the rotational tenth-order component, which is the main component of the torque fluctuation that is the exciting force against the torsional vibration, is reliably reduced. Accordingly, due to the torsional vibration, the generation of noise due to the resonance phenomenon of the compressor and the auxiliary equipment connected thereto is reduced, and the noise level in the vehicle compartment is reduced.

(B) Cylinder bores 51b, 51d, 5 which are displaced forward with respect to the rotation direction of the drive shaft 32
The dead volume in each compression chamber 29, 30 in 2b, 52d is enlarged from the reference value Vmid by a predetermined value V1. On the other hand, the cylinder bores 51c, 51e, 52c, which are displaced rearward with respect to the rotation direction of the drive shaft 32,
The dead volume in each compression chamber 29, 30 in 52e is reduced by a predetermined value V2 from the reference value Vmid. Here, the larger the dead volume, the compression chamber 2
The compression reaction force generated at 9 and 30 is alleviated, and the torque fluctuation due to this compression reaction force is reduced. For this reason, there is a tendency to reduce the torque fluctuation due to the shift movement of the cylinder bore,
This is in agreement with the tendency of the torque fluctuation reduction due to the change of the dead volume. Therefore, the torque fluctuation is surely reduced by the shift movement of the cylinder bore and the change of the dead volume.

(C) The front and rear dead volumes of one double-headed piston 28 are formed to be the same. For this reason, the fifth order component of rotation disappears because the total sum on the front side and the total sum on the rear side cancel each other. Therefore, together with the effects described in the first embodiment and the terms (a) and (b), it is possible to suppress the generation of the rotational fifth-order component while reducing the rotational tenth-order component of the torque fluctuation. .

(Third Embodiment) The third embodiment of the present invention will be described below.
The embodiment will be described with reference to FIGS. 7 and 8.
The third embodiment is suitable for a compressor having an even number of pistons 28, for example, an 8-cylinder type double-headed piston type compressor, when the predetermined shift angle cannot be secured due to interference of adjacent cylinder bores.

Also in this third embodiment, the first
Similar to the embodiment of FIG.
a to 61d and rear cylinder bores 62a to 62d
Are formed in the same manner as the pair of double-headed pistons 28, respectively. Therefore, the rear-side cylinder bores 62a to 62d will be mainly described.

As shown in FIG. 7, the two cylinder bores 62a, 62c accommodating half of the pistons 28 have their radial centers Pa3, Pc3 passing through equiangularly spaced radial lines R.
a, Rc to the front side in the rotation direction of the drive shaft 32 by an angle α
It is on the radial lines ra3 and rc3 which are shifted by 2. Two cylinder bores 62 containing the other half of the pistons 28
b and 62d are radial lines rb3, whose radial centers Pb3 and Pd3 are displaced from the radial lines Rb and Rd passing through equiangularly spaced positions by an angle α2 to the rear side in the rotational direction of the drive shaft 32,
It is on rd3.

Here, in the 8-cylinder type compressor,
The shift angle which minimizes the problematic eighth-order rotation component is calculated by the above equation (5). Δθ = 180 ° / n (5) = 180 ° / 8 = 22.5 ° That is, in order to minimize the rotation 8th order component, the radius lines ra3 to rd3 are +2 with respect to the radius lines Ra to Rd.
It is necessary to shift it considerably by 2.5 ° or −22.5 °. However, when it is not possible to secure movement enough to satisfy the desired displacement angle Δθ due to the size of the compressor, the degree of proximity of each cylinder bore, and the safety factor thereof, the displacement angle is the same as in the second embodiment. α2 is set.
Further, in order to compensate for the shortage of the phase difference of the torque fluctuation due to the shortage of the shift angle, the dead volume of each compression chamber 29, 30 is changed as in the second embodiment.

In other words, the shift angle α2 is such that even if the adjacent cylinder bores, for example, 62a and 62b are brought closer to each other by the angle α2 from the equiangular spacing positions, the adjacent cylinder bores do not interfere with each other and the distance between the adjacent cylinder bores is small. The partition walls are set so that a desired strength is secured.

As shown in FIG. 8, the drive shaft 3
Cylinder bores 62a, 6 which are displaced to the front side in the rotation direction of 2
The dead volume of the compression chamber 20 in 2c is a predetermined value V3 added to the reference dead volume value Vmid. On the other hand, the dead volume value of the compression chamber 30 in the cylinder bores 62b and 62d, which is displaced rearward in the rotational direction of the drive shaft 32, is obtained by subtracting a predetermined value V4 from the reference dead volume value Vmid. Has become.

With such a configuration, in the compressor accommodating the even number of pistons 28, the rotation n of torque fluctuation is
Secondary components can be reduced. Therefore, the excitation force for the torsional vibration between the drive shaft 32 and the clutch (not shown) is reduced, and the noise and vibration based on the resonance phenomenon excited by it are reduced, and the noise and vibration level in the vehicle interior is lowered. Can be made.

The present invention can be modified and embodied as follows. (1) In the first embodiment, the cylinder bore 1
2b, 12d (11b, 11d) are displaced to the rear side in the rotation direction of the drive shaft 32, and the cylinder bores 12c, 12e
To shift (11c, 11e) to the front side in the rotation direction of the drive shaft 32.

(2) In the second embodiment, the cylinder bores 52b, 52d (51b, 51d) are displaced to the rear side in the rotational direction of the drive shaft 32, and the compression chamber 29 in the cylinder bores 52b, 52d (51b, 51d) is moved. Three
The dead volume value of 0 is set to Vmid-V2, and the cylinder bores 52c and 52e (51c and 51e) are set.
e) is shifted to the front side in the rotational direction of the drive shaft 32, and the value of the dead volume of the compression chambers 29, 30 in the cylinder bores 52c, 52e (51c, 51e) is Vmid + V.
Set to 1.

(3) The present invention is embodied in a double-headed piston compressor having the number of cylinders other than those described in the above embodiment, for example, 6 and 12 cylinders. (4) Embodying the present invention in a single-head piston compressor.

(5) The present invention is embodied in a wave cam plate type piston type compressor. Even with the above configuration, it is possible to reduce the n-th order rotation component corresponding to the number of cylinders n.

(6) The dead volume of each compression chamber 29, 30 is changed by providing a recess in the head of the double-headed piston 28. (7) The dead volume of each compression chamber 29, 30 is changed by providing a groove on the head of the double-headed piston 28.

(8) The dead volumes of the compression chambers 29 and 30 are changed by changing the cylinder bores 11a to 11e and 12a.
It should be done by providing a notch on the inner peripheral surface of ~ 12e. (9) The dead volumes of the compression chambers 29 and 30 are changed by changing the lengths of the cylinder bores 11a to 11e and 12a to 12e.

(10) The dead volume of each compression chamber 29, 30 is changed by changing the thickness of the valve plates 13, 14. (11) The dead volume of each compression chamber 29, 30 is changed by changing the thickness of the suction valves 17a, 18a.

Even with the above structure, the dead volume of each compression chamber 29, 30 can be changed with a simple structure.

[0071]

As described above, according to the present invention, the following excellent effects can be obtained. In the case of an odd number of pistons, the radial centers of one cylinder bore are arranged at equiangular intervals defined on the circumference of the array of the radial centers of one cylinder bore. The radius center of the remaining half of the cylinder bores is from the equiangular distance position to the front side in the rotation direction,
The radial centers of the other half of the cylinder bores are respectively displaced rearward in the rotational direction by a predetermined displacement angle. In the case of an even number of cylinders, the center of radius of half of the cylinder bores is forward from the equiangular position in the rotational direction of the drive shaft, and the center of radius of the other half of the cylinder bores is conversely toward the rear in the rotational direction. It is displaced with a predetermined displacement angle. Therefore, the rotational n-th order component corresponding to the number of cylinders n of the torque fluctuation based on the compression reaction force of each compression chamber is reduced as compared with the case where the cylinder bores are arranged at equal angular intervals. Therefore, the exciting force of the torsional vibration of the drive shaft-clutch system is suppressed, the resonance phenomenon excited by the torsional vibration is reduced in the compressor and the auxiliary machine connected thereto, and the noise / vibration level in the vehicle interior is reduced. Can be reduced.

Further, in the case where the displacement of each cylinder bore that satisfies the desired displacement angle cannot be secured due to the size of the compressor, the degree of proximity of each cylinder bore, and the safety factor thereof, the compression in each cylinder bore is compressed. The dead volume of the room is changed. Therefore, the shortage of the torque fluctuation reducing effect due to the shift movement of each cylinder bore is supplemented by the torque fluctuation reducing effect due to the change of the dead volume of each compression chamber.

Moreover, the dead volume of the compression chamber in the cylinder bore displaced from the equiangularly spaced position to the front side in the rotational direction of the drive shaft is larger than the reference dead volume. Further, the dead volume of the compression chamber inside the cylinder bore, which is displaced rearward in the rotational direction of the drive shaft, is smaller than the reference dead volume. Therefore, the tendency of reducing the torque fluctuation due to the shift movement of the cylinder bore and the tendency of reducing the torque fluctuation due to the change of the dead volume are in agreement. Therefore, even when the displacement of each cylinder bore that satisfies the desired displacement angle cannot be ensured, the torque fluctuation based on the compression reaction force of each compression chamber can be reliably reduced.

In addition, in the double-headed piston type compressor, the dead volume on the front side and the dead volume on the rear side have the same size for the same piston. Therefore, even if the rotational n / 2nd order component is an odd order component when the number of cylinders is n, the pistons having the same rotational n / 2nd order component cancel each other out on the front side and the rear side. It disappears together. Therefore, in combination with the effect of the invention described above, it is possible to suppress the generation of the rotation n / 2-order component while reducing the rotation n-order component corresponding to the number of cylinders n. Then, the occurrence of a new vibration generation factor due to the countermeasure for the rotation n-order component is prevented.

[Brief description of the drawings]

FIG. 1 is a sectional view showing a compressor according to a first embodiment.

FIG. 2 is a sectional view taken along line 2-2 in FIG. 1;

FIG. 3 is an explanatory diagram showing a relationship between a phase difference and an amplitude of a rotation nth order component.

FIG. 4 is a sectional view showing a compressor according to a second embodiment.

FIG. 5 is an explanatory diagram related to changing the dead volume of each compression chamber.

FIG. 6 is an explanatory diagram showing a relationship between a shift angle and a dead volume.

FIG. 7 is a sectional view showing a compressor according to a third embodiment.

FIG. 8 is an explanatory diagram related to changing the dead volume of each compression chamber.

FIG. 9 is an explanatory diagram regarding a superimposing phenomenon of a rotation tenth-order component.

[Explanation of symbols]

11, 12 ... Cylinder block, 11a to 11e, 12
a-12e, 51a-51e, 52a-52e, 61a
-61d, 62a-62d ... Cylinder bore, 15 ... Front housing which comprises a part of housing, 16 ... Rear housing which comprises a part of housing, 28 ... Double-headed piston, 29, 30 ... Compression chamber, 32 ... Drive shaft,
34 ... Swash plate as a cam plate, Pa1 to Pe1, Pa2
Pe2, Pa3 to Pd3 ... Radius center of cylinder bore, C
... Arrangement circumference of radius center of cylinder bore, Δθ, α1, α
2 ... Sliding angle.

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Tomoji Tarutani, 2-chome Toyota-cho, Kariya city, Aichi Prefecture Toyota Industries Corporation (72) Inventor Satoshi Konomura 2-chome, Toyota-cho, Kariya city, Aichi prefecture Stock company Toyota Industries Corporation (72) Inventor Junichi Toyama 1-1, Showa-cho, Kariya city, Aichi prefecture Nihon Denso Co., Ltd. (72) Inventor Yasuo Takahara 1-1-chome, Showa town, Kariya city, Aichi prefecture NIDEC Co., Ltd.

Claims (9)

[Claims] 1. A drive shaft is supported inside a housing, a crank chamber is formed, and a plurality of cylinder bores are arranged around the drive shaft in a cylinder block forming a part of the housing. A piston is housed reciprocally in the compression chamber to form a partition, and a cam plate is integrally rotatably attached to the drive shaft, and the piston is reciprocated in conjunction with the rotation of the cam plate, In a reciprocating compressor that compresses a refrigerant gas, the number of pistons is an odd number, and equiangularly spaced positions corresponding to the number of the pistons are defined on an array circumference of a radial center of a cylinder bore, and The radial centers of the cylinder bores are arranged at the equiangular intervals, and the remaining half of the cylinder bores' radial centers are moved from the equiangular intervals to the drive shafts. Reciprocating the center of the radius of the other half of the remaining cylinder bores to the front side in the rolling direction and to the rear side in the rotational direction of the drive shaft from the equiangularly spaced positions with respect to the axial line of the drive shaft with a predetermined displacement angle. Dynamic compressor. 2. The reciprocating compressor according to claim 1, wherein the shift angle satisfies the following equation. Δθ = arc cos (−1 / 2n) / n where Δθ: shift angle n: number of compressor cylinders 3. A drive shaft is supported inside the housing, a crank chamber is formed, and a plurality of cylinder bores are arranged around the drive shaft in a cylinder block forming a part of the housing. A piston is housed reciprocally in the compression chamber to form a partition, and a cam plate is integrally rotatably attached to the drive shaft, and the piston is reciprocated in conjunction with the rotation of the cam plate, In a reciprocating compressor configured to compress a refrigerant gas, the number of pistons is an even number, the equiangular interval position corresponding to the number of the pistons is defined on the array circumference of the radial center of the cylinder bore, and half of the pistons are defined. The radial centers of the cylinder bores are located at the equiangularly spaced positions from the equiangularly spaced positions to the front side in the rotational direction of the drive shaft, and the radial centers of the other half of the cylinder bores are equiangularly spaced positions. The rotation direction rear side of al the drive shaft, a reciprocating compressor, respectively shifted by a predetermined shift angle about the axis of the drive shaft. 4. The reciprocating compressor according to claim 3, wherein the shift angle satisfies the following equation. Δθ = 180 ° / n where Δθ: shift angle n: number of cylinders in the compressor 5. The reciprocating compressor according to claim 1, wherein the compression chamber is set to a dead volume of at least two kinds of values. 6. The dead volume of the compression chamber in the cylinder bore displaced from the equiangularly spaced position to the front in the rotational direction of the drive shaft is enlarged as compared with a reference dead volume, and the drive shaft is rotated from the equiangularly spaced position. The reciprocating compressor according to any one of claims 1 to 5, wherein the dead volume of the compression chamber in the cylinder bore displaced rearward in the direction is reduced as compared with the reference dead volume. 7. The reciprocating compressor according to claim 6, wherein the reference dead volume is determined by a relationship diagram between a displacement angle and a dead volume obtained for each model of the compressor. 8. The reciprocating compressor according to claim 1, wherein the cylinder bores are formed so as to face each other in the front-rear direction, and the piston is configured as a double-headed type. 9. The reciprocating compressor according to claim 8, wherein a dead volume on the front side and a dead volume on the rear side are formed to have the same size with respect to one of the double-headed pistons.