JPH11264384A - Displacement fluid machine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce abrasion caused by a self-rotation moment applied on a displacer by forming profiles of a cylinder and the displacer so as to reverse a self-rotation moment of one operating chamber to that of the other operating chamber, by the self-rotation moment applied on the displacer per operating chamber. SOLUTION: In a positive displacement compressive element 1 driven by a. motor in a sealed container, an inner circumference of a cylinder 4 is formed so as to form hollow parts in the same shape per 120 deg., and a nearly circular arc shaped vane 4b is arranged on an end part of each hollow part. A displacer 5 engaged with a cylinder inner circumferential wall 4a and the vane 4b is housed in the cylinder 4. In this case, a negative self-rotation moment is acted on the displacer 5 by inner pressure of the operating fluid in an operating space while a rotary shaft 6 is rotated once, and an inner wall curve of the cylinder 4 and an outer wall curve of the displacer 5 are formed so as to set the total sum of the self-rotation moments by all operating spaces to be positive while the rotary shaft 6 is rotated once.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a pump, a compressor, an expander, and the like, and more particularly, to a positive displacement fluid machine.

[0002]

2. Description of the Related Art A reciprocating type fluid machine which moves a working fluid by repeating a reciprocating motion of a piston in a cylindrical cylinder has been used as a positive displacement type fluid machine, and a cylindrical piston has been eccentric in a cylindrical cylinder. A rotary (rolling piston type) fluid machine that moves a working fluid by rotating, meshing a pair of fixed scroll and orbiting scroll having an upright spiral wrap on an end plate, and orbiting the orbiting scroll. 2. Description of the Related Art A scroll type fluid machine that moves a working fluid by a hydraulic fluid is known.

[0003] Reciprocating fluid machines have the advantage of being easy and inexpensive to manufacture because of their simple structure, but the stroke from the end of suction to the end of discharge is as short as 180 ° in terms of the rotation angle of the rotating shaft. The problem is that the flow rate of the discharge process is high and the performance is reduced due to the increase in pressure loss, and the reciprocating motion of the piston is not enough to completely balance the rotating shaft system, resulting in large vibration and noise. There is.

[0004] In the rotary type fluid machine, although the stroke from the end of suction to the end of discharge is 360 ° in the rotation angle of the rotating shaft, the problem that the pressure loss in the discharge process increases is smaller than that of the reciprocating type fluid machine. Since the gas is discharged once per rotation of the shaft, the fluctuation of the gas compression torque is relatively large, and there is a problem of vibration and noise as in the reciprocating fluid machine.

Further, in the scroll type fluid machine, the stroke from the end of the suction to the end of the discharge requires a rotation angle of the rotary shaft of 360 °.
The pressure loss during the discharge process is small due to the long length as described above (usually about 900 ° for those used for air conditioning), and the fluctuation of gas compression torque is small due to the formation of a plurality of working chambers. Has the advantage of being small. However, it is necessary to manage the clearance between the spiral wrap in the lap meshing state and the clearance between the end plate and the tip of the lap, so that high-precision processing must be performed and the processing cost is high. There is. Also,
The stroke from the end of suction to the end of discharge is 3
There is a problem that the internal leakage increases as the compression stroke is longer, that is, as long as 60 ° or more.

By the way, the displacer for displacing the working fluid does not rotate relative to the cylinder into which the working fluid is sucked, but revolves around a substantially constant radius, that is, turns. Japanese Patent Application Laid-Open No. 55-23353 (Document 1), US Pat. No. 2,112,890 (Document 2), Japanese Patent Application Laid-Open No. 5-202869 (Document 3) and Japanese Patent Application Laid-Open No. 6-280758 disclose one type of positive displacement machine for transporting fluid. No. 4 (Document 4). The displacement type fluid machine proposed here comprises a displacer having a petal shape in which a plurality of members (vanes) extend radially from the center, and a cylinder having a hollow portion substantially similar to the displacer, The displacer moves the working fluid by revolving in the cylinder.

[0007] The positive displacement fluid machines disclosed in the above-mentioned documents 1 to 4 do not have a reciprocating portion unlike the reciprocating type, so that the imbalance of the rotating shaft system can be balanced. For this reason, vibration is small, and furthermore, since the relative sliding speed between the displacer and the cylinder is small, friction loss can be relatively reduced.

However, the stroke from the end of suction to the end of discharge of each working chamber formed by the plurality of vanes and cylinders constituting the displacer is about 180 ° (210 °) in terms of the rotation angle of the rotating shaft. Since it is short (about half of the rotary type and about the same as the reciprocating type), there is a problem in that the flow velocity of the fluid in the discharge process increases, the pressure loss increases, and the performance decreases. In the fluid machines disclosed in these documents, the rotation angle of the rotating shaft from the end of suction to the end of discharge of each working chamber is short, and the next (compression) stroke starts after the discharge of the working fluid is completed ( There is a time lag (time lag) from the end of the suction to the end of the suction, and the working chamber from the end of the suction to the end of the discharge is formed to be biased around the rotation axis, so that the mechanical balance is poor.

In order to solve this problem, Japanese Patent Laid-Open No. 9-26 is disclosed.
A positive displacement fluid machine disclosed in JP 8987 (Document 5) has been proposed. In the positive displacement fluid machine described in Document 5, the stroke from the end of suction to the end of discharge is as long as 360 degrees, and the suction stroke is simultaneously performed in the working chamber adjacent to the working chamber performing the compression stroke. The object of the present invention is to overcome the problems described in the above-mentioned documents 1 to 4 that the pressure loss is small and the mechanical balance is good.

[0010]

However, in the positive displacement type fluid machine described in the above-mentioned document 5, the displacer has a rotation moment for rotating the displacer itself as a reaction force from the compressed working fluid. It has a drawback that it acts excessively and easily causes reliability problems such as vane friction and wear. Further, in the prior art, no mention was made of such noise reduction of the positive displacement fluid machine.

SUMMARY OF THE INVENTION It is an object of the present invention to effectively reduce the rotation moment acting on the displacer under various conditions, and to reduce friction and friction.
It is an object of the present invention to solve the problem of wear and to provide a reliable and efficient displacement type fluid machine.

Another object of the present invention is to provide a positive displacement fluid machine capable of reducing noise caused by engagement between a cylinder and a displacer.

Another object of the present invention is to provide a rotation preventing mechanism that matches this displacement type fluid machine.

[0014]

SUMMARY OF THE INVENTION An object of the present invention is to provide a cylinder having an inner wall provided between end plates and having an inner wall constituted by a continuous curved surface, and an outer wall provided so as to face the inner wall of the cylinder. In the displacement type fluid machine including the inner wall, the outer wall, and a displacer forming a plurality of working spaces by the end plate when the rotary motion is caused by the rotation of the rotation shaft, the rotation direction of the rotation shaft is defined as a negative moment. Then, in the working space, a negative rotation moment acts on the displacer due to the internal pressure of the working fluid in the working space during one rotation of the rotation shaft, and the sum of the rotation moments in all the working spaces is equal to one rotation shaft. This is achieved by forming the cylinder inner wall curve and the displacer outer wall curve to be positive during rotation.

Further, the above object is achieved by disposing a displacer and a cylinder between end plates, and when the cylinder center and the displacer center are aligned, one space is formed by the cylinder inner wall surface and the displacer outer wall surface, In the displacement type fluid machine in which a plurality of working spaces are formed when the positional relationship between the displacer and the cylinder is set at the turning position, when the rotation direction of the rotating shaft for swiveling the displacer is defined as a negative moment, In each of the spaces, a negative rotation moment acts on the displacer due to the internal pressure of the working fluid in the working space during one rotation of the rotation shaft, and the sum of the rotation moments in all the working spaces is positive during one rotation of the rotation shaft. Forming the cylinder inner wall curve and the displacer outer wall curve so that It is achieved by.

Another object of the present invention is to provide a cylinder having an inner wall formed by a continuous curved surface between end plates, and an outer wall provided to face the inner wall of the cylinder.
In a displacement type fluid machine including a displacer that forms a plurality of spaces with the inner wall and the outer wall when the orbital motion is performed, the cylinder inner wall and the cylinder inner wall are arranged such that a moment acting on the displacer due to an internal pressure of a working fluid or the like is in a fixed direction. While forming the displacer outer wall curve,
At least one of the inner wall of the cylinder and the outer wall of the displacer is provided with a surface treatment layer that is softer and more compliant than the base material.

Another object of the present invention is to dispose a displacer and a cylinder between end plates, and when the cylinder center and the displacer center are aligned, one space is formed by the cylinder inner wall surface and the displacer outer wall surface. A plurality of spaces are formed when the positional relationship between the displacer and the cylinder is set at a swivel position, and the displacement type fluid machine includes a rotation preventing mechanism that prevents the displacer itself from rotating relative to the cylinder. Forming an inner wall surface of the cylinder and an outer wall surface of the displacer such that a moment acting on the displacer due to a compression reaction force is in a fixed direction when the displacer is swirled to compress the working fluid; Is achieved by:

[0018]

DESCRIPTION OF THE PREFERRED EMBODIMENTS The features of the present invention described above will be further clarified by the following embodiments. Hereinafter, an embodiment of the present invention will be described with reference to the drawings. First, the structure of a positive displacement fluid machine according to an embodiment of the present invention will be described with reference to FIGS. FIG. 1A is a longitudinal sectional view of a hermetic compressor (a sectional view taken along line AA in FIG. 1B) when a positive displacement fluid machine according to an embodiment of the present invention is used as a compressor, and FIG. ) BB
FIG. 2 is a plan view showing a state in which a compression chamber is formed in the direction of an arrow, FIG. 2 is an operation principle diagram of a displacement type compression element, and FIG. 3 is a case where a displacement type fluid machine according to an embodiment of the present invention is used as a compressor. It is a longitudinal cross-sectional view of the hermetic compressor in FIG.

In FIG. 1, a positive displacement element 1 according to the present invention and an electric element 2 for driving the same are provided in a closed container 3.
(Not shown) are stored. The details of the positive displacement compression element 1 will be described. FIG. 1B shows a compression element (referred to as a three-line wrap) in which three sets of the same contour shape are combined. The inner peripheral shape of the cylinder 4 is formed such that the hollow portion has the same shape every 120 ° (center o ′). At the end of each hollow portion, there are a plurality of protruding inward (in this case, three wraps so that there are three wraps, and the number of the protrusions is referred to as the number of ridges and is represented by symbol N). ) Has a substantially arc-shaped vane 4b. Displacer 5
Are disposed inside the cylinder 4 and have an inner peripheral wall 4a (a portion having a larger curvature than the vane 4b) of the cylinder 4 and a base.
4b. When the center o 'of the cylinder 4 matches the center o of the displacer 5, a gap having a constant width is formed as a basic shape between the two contours.

Next, the operation principle of the displacement type compression element 1 is shown in FIG.
And FIG. The symbol o is the center of the displacer 5, and the symbol o 'is the center of the cylinder 4 (or the rotating shaft 6). Symbols a, b, c, d, e, and f denote contact points at which the inner peripheral wall 4a and the vane 4b of the cylinder 4 mesh with the displacer 5. Here, looking at the inner peripheral contour shape of the cylinder 4, three combinations of the same curves are connected smoothly in succession at three locations. Paying attention to one of these, the curve forming the inner peripheral wall 4a and the vane 4b can be regarded as one thick vortex curve (the tip of the vane 4b on the discharge port 8a side is regarded as the start of the vortex). The inner wall curve (ga) indicates that the winding angle is approximately 360 ° (the design concept is 360 °, but the value is not exactly the same value due to a manufacturing error).
The same applies hereinafter. Note that this winding angle is a vortex curve of which details will be described later. The outer wall curve (gb) has a winding angle of about 3
It is a 60 degree vortex curve. As described above, the one inner peripheral contour shape is formed from the inner wall curve and the outer wall curve. The two spiral curves are arranged at substantially equal pitches (120 ° because of three-line wrap), and the outer wall curve and the inner wall curve of the adjacent spiral body are connected smoothly with each other by a smooth connection curve (b−
The connection at b ′) forms the entire inner peripheral contour shape of the cylinder 4. The outer peripheral shape of the displacer 5 is also configured in the same principle as that of the cylinder 4.

The spiral body composed of three curves is arranged at a substantially equal pitch (120 °) on the circumference. This is for the purpose of uniformly distributing the load accompanying the compressing operation described later and for the purpose of manufacturing. In consideration of easiness, especially when these are not a problem, the pitch may be unequal.

The compression operation of the cylinder 4 and the displacer 5 configured as described above will be described with reference to FIG. 7a is a suction port, 8a is a discharge port, 3
It is provided on the corresponding end plate. By rotating the rotation shaft 6 in the clockwise direction, the displacer 5 revolves around the turning radius ε (= oo ′) without rotating around the center o ′ of the cylinder 4 on the fixed side, and the displacer 5 rotates. Of the plurality of working chambers 15 (in a plurality of hermetically sealed spaces surrounded by the inner peripheral contour (inner wall) of the cylinder and the outer peripheral contour (side wall) of the displacer), suction is completed and a compression (discharge) process is performed. In other words, the space from the end of inhalation to the end of discharge.
If the above-mentioned winding angle is limited to 360 °, this space is lost at the end of compression, but at that moment, the suction ends, so this space is counted as one. However, when used as a pump, a space that communicates with the outside through a discharge port is formed (in the present embodiment, there are always three working chambers). Hatched 1 surrounded by contact a and contact b
The following description focuses on two working chambers (they are divided into two at the end of suction, but the two working chambers are connected and become one as soon as the compression stroke is started). FIG. 2A shows a state in which the suction of the working gas from the suction port 7a into this working chamber is completed. FIG. 2 (2) shows a state in which the 90 ° rotation shaft 6 has rotated from this state, FIG. 2 (3) shows a state in which the rotation has advanced 180 ° from the beginning, and FIG. 2 (3) shows a state in which the rotation has further advanced and has rotated 270 ° from the beginning. It is FIG.2 (4). Fig. 2 (4)
When it is rotated 90 ° from, it returns to the initial state of FIG.
As a result, the working chamber 15 decreases its volume as the rotation proceeds, and the discharge port 8a is closed by the discharge valve 9 (shown in FIG. 1), so that the working fluid is compressed. When the pressure in the working chamber 15 becomes higher than the external discharge pressure, the discharge valve 9 opens automatically due to the pressure difference, and the compressed working gas is discharged through the discharge port 8a. The rotation angle of the rotating shaft from the end of suction (compression start) to the end of discharge is 360 °, and the next suction stroke is prepared while each compression and discharge stroke is being performed. Starts compression. For example, focusing on the space formed by the contacts a and d, the suction port 7a has already been formed in the stage of FIG.
Inhalation is started from, and the volume increases as the rotation proceeds, and when the state of FIG. 2D is reached, this space is divided. The fluid corresponding to the divided amount is the contacts b and e.
Supplemented by the space formed by

The manner of making up for this will be described in detail. FIG.
Focusing on the working chamber formed by the contacts a and b in the state (1), the suction formed by the space formed by the adjacent contacts a and d has begun. This space is once expanded as shown in FIG. 2 (3), and then divided as shown in FIG. 2 (4). Therefore, not all the fluid in the space formed by the contacts a and d is compressed in the space formed by the contacts a and b. 2 (4), the space formed by the contacts b and e in the suction process in FIG. 2 (4) has the same volume as the fluid volume that has been divided and not taken into the space formed by the contacts a and d. As shown in 1), it is divided and filled with the fluid flowing into the space formed by the contact point e and the contact point b near the discharge port. This is because, as described above, the wraps are arranged at a uniform pitch. That is, since the shapes of the displacer and the cylinder are formed by repeating the same contour shape,
Each working chamber can compress substantially the same amount of fluid even if fluid is obtained from different spaces. Although it is possible to perform processing so that the volumes formed in the respective spaces are equal even if the pitch is not uniform, the productivity is poor. In any of the prior arts described above, the space in the suction process is closed and compressed and discharged as it is, whereas the space in the suction process adjacent to the working chamber is divided to perform the compression operation. This is one of the features of the present embodiment.

As described above, the working chambers that perform a continuous compression operation are disposed at substantially equal pitches around the drive bearing 5a located at the center of the displacer 5, and each working chamber has a phase. And the compression is performed. That is,
Focusing on one space, the rotation angle of the rotating shaft is 360 ° from the suction to the discharge, but in the case of the present embodiment, three working chambers are formed and these discharge at a phase shifted by 120 °. As a compressor, the refrigerant is discharged three times during a rotation angle of the rotation shaft of 360 °.

Assuming that the space at the moment when the suction operation is completed (the space surrounded by the contact points a and b) is one space, when the winding angle is 360 ° as in this embodiment, any compression Even in the machine operating state, the space in the suction stroke and the space in the compression stroke are designed to be alternated, so that the next compression stroke is started immediately after the compression stroke is completed. Fluid can be compressed smoothly and continuously.

Next, a compressor incorporating the volumetric compression element 1 having such a shape will be described with reference to FIGS. In FIG. 3, in addition to the cylinder 4 and the displacer 5 described above in detail, the positive displacement element 1 includes a rotating shaft 6 for driving the displacer 5 by fitting a crank part 6a to a bearing at the center of the displacer 5, A main bearing 7 and a sub-bearing 8 which also serve as an end plate for closing the openings at both ends of the cylinder 4 and a bearing for supporting the rotary shaft 6; a suction port 7a formed in the end plate of the main bearing 7; And a discharge valve 9 that opens and closes the discharge port 8a with a differential pressure. However, the discharge valve 9 may be a mechanical or other forced valve. 5b
Is a through hole formed in the displacer 5. Also,
Reference numeral 10 denotes a suction cover attached to the main bearing 7, and reference numeral 11 denotes a discharge cover for forming a discharge chamber 8b integrally with the sub-bearing 8.

The electric element 2 includes a stator 2a and a rotor 2b. The rotor 2b is fixed to one end of the rotating shaft 6 by shrink fitting or the like. The electric element 2 is constituted by a brushless motor for improving electric motor efficiency, and is driven and controlled by a three-phase inverter. However, other motor types, for example,
A DC motor or an induction motor can be used.

Numeral 12 denotes lubricating oil stored at the bottom of the sealed container 3, in which the lower end of the rotary shaft 6 is immersed. 13
Is a suction pipe, 14 is a discharge pipe, and 15 is the above-described working chamber formed by meshing of the inner peripheral wall 4a and the vane 4b of the cylinder 4 with the displacer 5. The discharge chamber 8b is separated from the pressure in the sealed container 3 by a seal member 16 such as an O-ring.

The flow of the working gas (refrigerant gas) when the positive displacement fluid machine in this embodiment is used as an air conditioning compressor will be described with reference to FIG. As indicated by the arrow in the drawing, the working gas that has entered the closed casing 3 through the suction pipe 13 enters the suction cover 10 attached to the main bearing 7, and passes through the suction port 7a to perform positive displacement compression. Enter element 1,
Here, the rotation of the rotating shaft 6 causes the displacer 5 to perform a revolving motion, and the volume of the working chamber is reduced, so that the working chamber is compressed. The compressed working gas passes through a discharge port 8a formed on the end plate of the sub bearing 8, pushes up the discharge valve 9, and pressurizes the discharge chamber 8.
b, and flows out through the discharge pipe 14.
The reason why the gap is formed between the suction pipe 13 and the suction cover 10 is to cool the electric motor element by flowing the working gas into the electric motor element 2 as well.

The lubricating oil 12 stored inside is sent to each bearing sliding portion from the bottom through a hole provided in the rotary shaft 6 by a centrifugal pump or the like, and is lubricated. A part of the supplied lubricating oil is also supplied to the inside of the working chamber.

The effects of the positive displacement type fluid machine described with reference to FIGS. 1 to 3 will be described below. FIG. 4 shows the change in the working chamber volume (characterized by the ratio between the suction volume Vs and the working chamber volume V) in the present invention, using the rotation angle θ of the rotation shaft from the end of suction as the horizontal axis, compared with other types of compressors. Shown. Accordingly, the volume change characteristic of the positive displacement compression element 1 according to the present embodiment is based on a kind of condition of an air conditioner having a discharge start volume ratio of 0.37 (for example, when the working gas is Freon HCFC22, the suction pressure Ps = 0.6).
4MPa, discharge pressure Pd = 2.07MPa), the compression process is almost the same as the reciprocating type, and the compression process is completed in a short time, so that leakage of working gas is reduced, and the capacity and efficiency of the compressor are reduced. Can be improved. On the other hand, the discharge process is about 5 times longer than the rotary type (rolling piston type).
The pressure loss is reduced by 0% longer and the discharge flow rate is reduced, and the fluid loss (excess compression loss) in the discharge process can be greatly reduced to improve the performance.

FIG. 5 shows a change in the amount of work during one rotation of the shaft, that is, a change in the gas compression torque T in this embodiment, in comparison with other types of compressors (where Tm is an average torque). . Thus, the torque fluctuation of the positive displacement type compression element 1 of the present invention is very small, about 1/10 of the rotary type, and is equivalent to the scroll type, but is a swirling scroll like a scroll type Oldham ring. Since there is no reciprocating sliding mechanism for preventing rotation, the inertia balance of the rotating shaft system can be maintained and the vibration and noise of the compressor can be reduced. In addition, scroll
Since it does not have a long spiral shape like the conventional type, the processing time can be reduced and the cost can be reduced. In addition, since there is no end plate (end plate) for maintaining the spiral shape, the tool can be penetrated and processed, and the tool can be penetrated. It can be manufactured at the same processing cost as a rotary type as compared with a scroll type which could not be processed.

Further, since the thrust load due to the gas pressure does not act on the displacer, it is easy to manage the axial clearance which has an important effect on the performance of the compressor as seen in a scroll compressor, so that the performance can be improved. Further, as a result of the calculation, the thickness of the compression element can be reduced as compared with a scroll compressor having the same stroke volume and the same outer diameter, which can contribute to the reduction in size and weight of the compressor.

Next, the relationship between the above-mentioned winding angle and the rotation angle θc of the rotating shaft from the end of suction to the end of discharge will be described.
In the above-described embodiment, the winding angle is described as 360 °. However, it is possible to change the rotation angle θc of the rotating shaft in the compression stroke by changing the winding angle. For example, in FIG. 2, since the winding angle is 360 °, the rotation angle θc of the rotating shaft from the end of suction to the end of discharge returns to the original state when the rotation angle θc is 360 °. When the rotation angle θc of the rotating shaft from the end of suction to the end of discharge is reduced by making the winding angle smaller than 360 °, a state occurs in which the discharge port and the suction port communicate with each other,
There is a problem that the fluid once sucked flows backward due to the expansion action of the fluid in the discharge port. When the winding angle is larger than 360 °, the rotation angle θc of the rotating shaft in the compression stroke also becomes larger than 360 °, and two working chambers having different sizes from the end of the suction to the communication with the space having the discharge port are formed. It is formed. When this is used as a compressor, irreversible mixing loss occurs when the two working chambers merge because the pressure rises of these two working chambers are different from each other, resulting in an increase in compression power. Further, even if it is used as a liquid pump, it is difficult to apply the pump as a working chamber which is not communicated with the discharge port is formed. For this reason, it can be said that the winding angle is desirably 360 ° as much as possible within the range of allowable accuracy.

The rotation angle θc of the rotating shaft during the compression stroke in the fluid machine described in the above-mentioned Japanese Patent Application Laid-Open No. 55-23353 (Document 1) is θc = 180 °.
The rotation angle θc of the rotating shaft during the compression stroke in the fluid machine described in Japanese Patent Application Laid-Open No. 869 (Reference 3) and Japanese Patent Application Laid-Open No. 6-280758 (Reference 4) is θc = 210 °. The period from the end of discharge of the working fluid to the start of the next compression stroke (end of suction) is 180 ° in the rotation angle of the rotating shaft in Reference 1, and 150 ° in References 3 and 4.

When the rotation angle θc of the rotating shaft in the compression stroke is 210 °, each working chamber (reference numerals I, II, I
FIG. 6 (a) shows a compression stroke diagram of the compression strokes indicated by II and IV). However, the number of rows N = 4. Four working chambers are formed when the rotation angle θc of the rotating shaft is 360 °, and the number n of working chambers simultaneously formed at a certain angle is n = 2 or 3. The maximum value of the number of working chambers formed simultaneously is 3, which is smaller than the number of working chambers.

Similarly, FIG. 7A shows the case where the number of threads N = 3 and the rotation angle θc of the rotating shaft in the compression stroke is 210 °. Also in this case, the number n of working chambers formed at the same time is n = 1 or 2, and the maximum value of the number of working chambers formed at the same time is 2, which is smaller than the number of rows.

In such a state, the working chamber is formed so as to be deviated around the rotation axis, so that a mechanical imbalance occurs, the rotation moment acting on the displacer becomes excessive, and the contact load between the displacer and the cylinder is reduced. There is a problem in that the performance is reduced due to an increase in mechanical friction loss and the reliability is reduced due to vane wear.

In order to solve this problem, in the present embodiment, the rotation angle θc of the rotating shaft in the compression stroke is given by: ◆ (((N−1) / N) · 360 °) <θc ≦ 360 ° (Equation 1) The outer contour of the displacer and the inner contour of the cylinder are formed so as to satisfy ◆. In other words, the above-mentioned winding angle is in the range of Expression 1. Referring to FIG. 6B, when the rotation angle θc of the rotation shaft in the compression stroke is 270
゜, the number of working chambers formed simultaneously n
Is n = 3 or 4, and the maximum value of the number of working chambers is 4. This value corresponds to the number N of rows (= 4). In FIG. 7B, the rotation angle θc of the rotation shaft during the compression stroke is 24 degrees.
0 °, the number n of working chambers formed simultaneously is n = 2 or 3, and the maximum value of the number of working chambers is 3
It is. This value is equal to the number N of rows (= 3).

As described above, the rotation angle θc of the rotating shaft in the compression stroke
Since the maximum value of the number of working chambers becomes equal to or more than the number N of rows by making the lower limit value of the left side of Equation 1 larger than the value on the left side of Equation 1, the working chambers are dispersedly arranged around the rotation axis. In addition, the rotational balance acting on the displacer is reduced, the contact load between the displacer and the cylinder is reduced, and the reliability of the contact portion can be improved as well as the performance is improved by reducing the mechanical friction loss.

On the other hand, the upper limit of the rotation angle θc of the rotation shaft in the compression stroke is 360 ° according to the equation (1). The upper limit of the rotation angle θc of the rotation shaft in the compression stroke is ideally 360 °. As described above, the time lag from the end of discharge of the working fluid to the start of the next compression stroke (end of suction) is set to 0.
It is possible to prevent a decrease in the suction efficiency due to the re-expansion of the gas in the gap volume that occurs when θc <360 °, and to reduce the suction efficiency when θc> 360 °.
Since the pressure rises of the two working chambers are different from each other, irreversible mixing loss that occurs when the two are combined can be prevented. The latter will be described with reference to FIG.

FIG. 8 shows a displacement type fluid machine in which the rotation angle θc of the rotating shaft in the compression stroke is 375 °. FIG. 8A shows a state in which the suction of the two working chambers 15a and 15b in the figure has been completed. At this time, the pressures of the two working chambers 15a and 15b are equal at the suction pressure Ps. The discharge port 8a is located between the working chambers 15a and 15b, and does not communicate with both working chambers. FIG. 8B shows a state in which the rotation of the rotary shaft has been advanced by 15 ° from this state. This is a state immediately before the discharge port 8a and both working chambers 15a and 15b communicate with each other. At this time, the volume of the working chamber 15a is smaller than that at the end of the suction in FIG. 8A, the compression is progressing, and the pressure is higher than the suction pressure Ps. On the other hand, the volume of the working chamber 15b is larger than that at the end of suction, and the pressure is also lower than the suction pressure Ps due to the expansion action. When the next instantaneous working chambers 15a and 15b join (communicate),
The irreversible mixing as shown by the arrow in FIG. 8C occurs, and the performance decreases due to the increase in the compression power. Therefore, the upper limit of the rotation angle θc of the rotation shaft in the compression stroke is 3
60 ° is a desirable state.

FIGS. 9A and 9B show a compression element of a positive displacement fluid machine described in Document 3 or 4, wherein FIG. 9A is a plan view and FIG. 9B is a side view. The number of threads N is 3, and the rotation angle θc (winding angle) of the rotating shaft in the compression stroke is 210 °. In this figure, the number n of working chambers is n as shown in FIG.
= 1 or 2. This figure shows that the rotation angle θ of the rotating shaft is 0.
The state of ゜ is shown, and the number n of working chambers is two. As is clear from this drawing, the space on the right side of the space formed by the outer peripheral contour of the displacer and the inner peripheral contour of the cylinder does not serve as a working chamber, but instead has a suction port 7 a and a discharge port 8.
a communicates. Therefore, there is a problem that the gas once flowing into the cylinder 4 from the suction port 8a flows backward due to the re-expansion of the gas in the clearance volume of the discharge port 8a, and the suction efficiency is reduced.

Now, consider a case where the rotation angle θc of the rotation shaft in the compression stroke of the displacement type fluid machine shown in FIG. In order to increase the rotation angle θc of the rotation shaft during the compression stroke, the winding angle of the contour curve of the cylinder 4 must be increased as shown by a two-dot chain line. It is difficult to make the rotation angle θc of the rotation shaft of the compression stroke larger than 240 ° so that the thickness becomes thin and the maximum value of the number n of the working chambers becomes more than the number N of rows (N = 3).

FIG. 10 shows an embodiment of a compression element of a positive displacement fluid machine having the same stroke volume (suction volume), the same outer diameter, and the same turning radius as the positive displacement fluid machine shown in FIG. The rotation angle θc of the rotation shaft in the compression stroke of the compression element shown in FIG. 10 realizes 360 ° larger than 240 °. This is because, in the compression element shown in FIG. 9, the interval between the seal points forming the working chamber is constituted by a uniform curve. Even if the rotation angle θc is to be increased, the maximum is 2
Although the limit is 40 °, in the compression element according to the present embodiment shown in FIG. 10, the space between the seal points (ac) is not a uniform curve, and the shape near the contact point b projects from the displacer. There is a constriction on the way from the center to the tip of each displacer strip. These are also applicable to the embodiment shown in FIG. With these shapes, the winding angle from the contact point a to the contact point b in FIG. 10 can be set to 360 ° larger than 240 °, and the winding angle from the contact point b to the contact point c is 240 °.
It can be greater than {360}. As a result,
The rotation angle θc of the rotation shaft in the compression stroke should be
60 °, and the maximum value of the number n of working chambers is
The above can be considered. For this reason, the working chambers are dispersed and the rotation moment can be reduced.

Further, by increasing the number of working chambers that can function effectively, assuming that the cylinder height (thickness) of the compression element shown in FIG. 9 is H, the compression chamber shown in FIG. Since the cylinder height is 0.7H, which is 30% lower, the size of the compression element can be reduced.

FIG. 11 is an explanatory diagram of the load and moment acting on the displacer 5 in the present embodiment.
The symbol θ is the rotation angle of the rotating shaft 6, and ε is the turning radius. With the compression of the working gas, a tangential force Ft perpendicular to the eccentric direction and a radial force Fr corresponding to the eccentric direction act on the displacer 5 due to the internal pressure of each working chamber 15 as shown in the figure. The resultant force of Ft and Fr is F. A rotational moment M that attempts to rotate the displacer due to the deviation of the resultant force F from the center o of the displacer 5 (arm length 1).
(= F · l) works. Supporting the rotation moment M are reaction forces R1 and R2 at the contact points e and b of the displacer 5 and the cylinder 4.

In the displacement type fluid machine of this embodiment, the moment is always received at two or three contact points near the suction port 7a, and the other contact points have a contour shape in which no reaction force acts (details). Will be described later). The displacement type compression element 1 according to the present embodiment has a substantially equal pitch around the crank part 6a of the rotating shaft 6 fitted to the center of the displacer 5, and the rotation angle of the rotating shaft from the end of suction to the end of discharge is substantially equal. 360
°, the working point of the resultant force F can be closer to the center o of the displacer 5, and the length l of the arm of the moment can be reduced to rotate the moment. M can be reduced. Therefore, the reaction force R1 and the reaction force R2 are reduced. Further, as can be seen from the positions of the contact points e and b, the sliding portion of the displacer 5 and the cylinder 4 receiving the rotation moment M is set so as to be near the working gas inlet port 7a of a low temperature and high oil viscosity. Therefore, an oil film on the sliding portion is secured, and a highly reliable positive displacement compressor that solves the problem of friction and wear can be provided.

FIG. 12 shows a comparison between the compression element shown in FIG. 9 and the compression element shown in FIG. 13 in terms of the rotational moment M acting on the displacer during one rotation of the shaft due to the internal pressure of the working fluid. The calculation conditions are the refrigeration conditions of the working fluid HFC134a (intake pressure Ps = 0.095 Mpa, discharge pressure Ps
d = 1.043 Mpa). Thus, the number of working chambers n
In the compression element according to the present embodiment in which the maximum value is equal to or greater than the number of rows, the working chambers in the compression stroke from the end of suction to the end of discharge are dispersed at substantially equal pitches around the rotation axis, so that The balance can be improved, and the load vector due to compression can be configured to be substantially directed to the center. Therefore, the rotation moment M acting on the displacer can be reduced. As a result, the contact load between the displacer and the cylinder is reduced, so that the mechanical efficiency can be improved and the reliability as a compressor can be improved.

Here, the relationship between the period in which the suction port 7a and the discharge port 8a communicate with each other and the rotation angle of the rotating shaft in the compression stroke will be described. A time lag Δ represented by the rotation angle of the rotating shaft during a period in which the suction port and the discharge port communicate with each other, that is, from the end of discharge of the working fluid to the start of the next compression stroke (end of suction)
θ is the rotation angle θc of the rotation shaft during the compression stroke, and Δθ = 3
It is expressed by 60 ° -θc.

When Δθ ≦ 0 °, there is no period in which the suction port and the discharge port communicate with each other, so that the suction efficiency does not decrease due to the re-expansion of the gas in the gap volume of the discharge port.

When Δθ> 0 °, since there is a period in which the suction port and the discharge port communicate with each other, the suction efficiency decreases due to the re-expansion of the gas in the clearance volume of the discharge port, and the compressor (refrigeration) ) Capability will be reduced. Further, a decrease in the suction efficiency (volume efficiency) leads to a decrease in the adiabatic efficiency, which is the energy efficiency of the compressor, or the coefficient of performance.

The rotation angle θc of the rotation shaft in the compression stroke is determined by the winding angle of the contour curve of the displacer or cylinder and the positions of the suction port and the discharge port. When the winding angle of the contour curve of the displacer or the cylinder is set to 360 °, the rotation angle θc of the rotating shaft in the compression stroke can be set to 360 ° and θc <360 ° by moving the seal point of the suction port or the discharge port. Can also be. However, θc> 360 ° cannot be achieved. For example, the rotation angle θc = 375 ° of the rotary shaft in the compression stroke of the compression element shown in FIG. 8 can be changed to θc = 360 ° by changing the position and size of the discharge port. This is because immediately after the suction end state in FIG.
It can be realized by enlarging the discharge port so that the and the working chamber 15b communicate with each other. By performing such a change, it is possible to reduce irreversible mixing loss that occurs due to the difference in pressure rise between the two working chambers that occurred when θc = 375 °. Therefore, it can be said that the winding angle of the contour curve is a necessary condition for determining the rotation angle θc of the rotating shaft in the compression stroke, but is not a sufficient condition.

In the above-described embodiment, the closed type compressor in which the pressure in the closed casing 3 is low (suction pressure) has been described. However, the use of the low pressure type has the following advantages.

(1) Since the heating of the electric element 2 by the compressed high-temperature working gas is small, the stator 2a and the rotor 2b
, The motor efficiency is improved and the performance can be improved.

(2) With a working fluid compatible with the lubricating oil 12 such as chlorofluorocarbon, the pressure is low, so that the proportion of the working gas dissolved in the lubricating oil 12 decreases, and the oil bubbling phenomenon occurs in bearings and the like. And reliability can be improved.

(3) The pressure resistance of the sealed container 3 can be reduced, and the thickness and weight can be reduced.

Next, a case where the pressure in the closed container 3 is high (discharge pressure) will be described. FIG. 13 is an enlarged sectional view of a main part of a high-pressure hermetic compressor using a positive displacement fluid machine as a compressor according to another embodiment of the present invention. In FIG. 13, components denoted by the same reference numerals as those in FIGS. 1 to 3 are the same components and perform the same operations. In the figure, reference numeral 7b denotes a suction chamber formed integrally with the main bearing 7 by a suction cover 10, which is separated from the pressure (discharge pressure) in the sealed container 3 by a seal member 16 or the like. Reference numeral 17 denotes a discharge passage communicating between the inside of the discharge chamber 8b and the closed container 3. The operation principle and the like of the positive displacement compression element 1 are the same as those of the low pressure (suction pressure) type described above.

The flow of the working gas, as indicated by the arrow in the drawing, flows through the suction pipe 13 and enters the suction chamber 7b. The working gas passes through a suction port 7a formed in the main bearing 7 and is displaced by a positive displacement element. 1, where the rotation of the rotating shaft 6 causes the displacer 5 to make a revolving motion, and the volume of the working chamber 15 is reduced, thereby being compressed. The compressed working gas pushes up the discharge valve 9 through the discharge port 8 a formed in the end plate of the sub bearing 8, enters the discharge chamber 8 b, passes through the discharge passage 17, and enters the closed container 3. It flows out from a discharge pipe (not shown) connected to the container 3.

The advantage of such a high pressure type is that lubricating oil 1
2 is at a high pressure, the lubricating oil 12 supplied to each bearing sliding portion by the centrifugal pump action or the differential pressure due to the rotation of the rotating shaft 6 passes through the gap at the end face of the displacer 5 and enters the cylinder 4. The working chamber 1
5 in that the sealability and the lubricity of the sliding portion can be improved.

As described above, in the compressor using the positive displacement fluid machine of the present invention, it is possible to select either the low-pressure type or the high-pressure type according to the specifications and use of the equipment, the production equipment, etc., and the degree of freedom in design is greatly increased. To expand.

Incidentally, the fact that the rotational moment acting on the displacer is reduced by adopting a contour shape in which the working chambers are arranged in a dispersed manner as compared with the positive displacement fluid machines described in Documents 1 to 4 is described above. did. However, it has been found that simply dispersing the working chambers still has a problem in that the rotation moment is still large and the contact receiving the rotation moment has large wear. An embodiment for solving this problem will be described with reference to FIG. In FIG. 14, components denoted by the same reference numerals as those in FIG. 1 are the same components and perform the same operations. Even under various pressure conditions under which the compressor is operated, particularly under severe conditions such as a high pressure ratio represented by the ratio of the discharge pressure Pd to the suction pressure Ps, the rotation moment acting on the displacer 5 as the compression reaction force of the working gas is obtained. Fig. 3 shows a compression element shape that can be effectively reduced. In the figure, three working chambers 15a, 15b, and 15c are formed, and the three-row wrap where the number of threads N = 3 satisfies the above-described basic relationship of Equation 1.

As the compression element shape, the contour shapes of the inner wall of the cylinder 4 and the outer wall of the displacer 5 near the discharge port 8a are changed as compared with the wrap shape (contour shape) shown in FIG. That is, the tip of the contour of the inner wall of the cylinder 4 near the discharge port 8a is directed more outward (in the direction away from the axis center) than the cylinder 4 shown in FIG. The thickness is changed inward (in a direction closer to the axis center) than the displacer 5 shown in FIG. Thus, a reverse (negative) moment against the rotation moment acting on the displacer 5 is generated, and the total rotation moment is reduced.
Hereinafter, this operation will be described.

The load F1 due to the internal pressure of the working chamber 15a is obtained by multiplying the area obtained by multiplying the length of the line connecting the two contacts a and b constituting the working chamber 15a by the height of the displacer 5 and the pressure in the working chamber 15a and the suction pressure Ps. The action position is represented by a vertical bisector of a line connecting the contact points a and b. Similarly, a load F2 due to the internal pressure of the working chamber 15b and a load F3 due to the internal pressure of the working chamber 15c are also obtained. The resultant force of these three loads is F, and the position where the resultant force F is applied is indicated by the contour shown in FIGS. , The rotation moment M can be reduced. FIG. 15 shows the refrigerating condition of the working fluid HFC134a (the suction pressure Ps = 0.0
(95 MPa, discharge pressure Pd = 1.043 MPa), the rotation moment M during one rotation of the shaft, and the loads F1, F2, F due to the internal pressures of the working chambers constituting the rotation moment M.
3 is an exploded view of the moments M1, M2, and M3.

When the moment M1 due to the internal pressure of one working chamber, ie, the working chamber 15a shown by a thin solid line, is changed, the moment becomes negative in the latter half of the shaft rotation angle θ, and the moment in the opposite direction to the rotation moment M is obtained. It can be seen that has occurred. The other two working chambers also increase the moment M1 by 120 °
Since the phases are shifted (shown by broken lines and dashed-dotted lines, respectively), the sum of the rotation moments at a certain shaft rotation angle θ becomes smaller due to the action of the moment in the opposite direction. The reaction force (contact load) at the contact points (contact points a, b, and g in FIG. 14) with the cylinder 4 that supports the rotation moment M is reduced, so that the reliability of the contact portion can be improved and the mechanical friction loss is reduced. High efficiency of the positive displacement fluid machine can be achieved. Further, since the contact load between the cylinder 4 and the displacer 5 is reduced, the magnitude and variation of the bearing load of the bearing 5a located at the center of the displacer 5 are reduced, and the variation of the gas compression torque of the rotating shaft 6 is further reduced. It can contribute to the reduction of vibration and noise.

FIG. 16 shows a compression element of a positive displacement fluid machine according to another embodiment of the present invention. The embodiment of FIG.
The inner wall of the cylinder 4 and the outer wall of the displacer 5 near the discharge port 8a are further enlarged than the compression element shown in FIG. The moment due to the load F3 due to the internal pressure of the working chamber 15c in the latter half of the compression stroke is in the opposite direction to the rotation moment M, and the acting position of the resultant force F is substantially at the center o of the displacer 5.
, It can be seen that the rotational moment is reduced.

Here, the contour shape shown in FIG.
Will be described in detail. The displacer position in both figures is several degrees ahead of the displacer position shown in FIG. Since the working chamber surrounded by the contact points a and b has just finished the suction stroke, the internal pressure load F1 is small, and the direction of the vector is counterclockwise opposite to the turning direction. Since the working chamber surrounded by the contact points c and d is in the discharge stroke, the internal pressure load F
The size of 2 is relatively large and the direction is the rotation direction.

On the other hand, since the working chamber surrounded by the contact points e and f is in the discharge stroke, the internal pressure load F3 has the largest value at the shaft rotation angle shown in FIG. The direction of the vector of the internal pressure load F3 is almost in the center direction at the shaft rotation angle position in FIG. 14, but when turning proceeds from this state, the direction of the normal line of the straight line connecting the contact point e and the contact point f is obtained. Is closer to the right than the axis center in the figure, and a negative (clockwise) rotation moment is generated by this internal pressure vector. On the other hand, in the compressor having the contour shape shown in FIG.
In spite of the shaft rotation angle shown in FIG. 4, the direction of the internal pressure load F3 of the working chamber surrounded by the contacts e and f has already a negative rotation moment. This is because, as described above, the shape of the tip of the displacer and the tip of the cylinder near the discharge port 8a are formed such that the tip of the vane faces outward and the tip of the displacer faces inward. The magnitude of the rotation moment M due to the resultant force F can be designed to be changed by variously changing the shape of this portion.

FIG. 17 shows an embodiment of the present invention under the refrigeration conditions of HFC134a described above, and FIG. 10, FIG.
6 is a comparison of the rotation moment during one rotation of the shaft acting on the displacer of the compression element shown in FIG. The magnitude of the rotation moment is determined by the compression element in FIG. 10, the compression element in FIG.
The compression element shown in FIG. 16 is in order, and in the compression element shown in FIG. 16, the rotation moment is an alternating moment in which the rotation moment swings positively and negatively during one rotation of the shaft.

FIG. 18 further shows a change in pressure ratio of these three compression elements (working fluid HFC134a, suction pressure Ps =
The change in the rotation moment of 0.095 MPa is calculated. Here, the rotation moment is compared with the average value during one rotation of the shaft. Thus, the compression element shown in FIG. 14 has a smaller rotation moment than the compression element shown in FIG. 10 under the condition of a high pressure ratio. It can be said that this is a compression element of a positive displacement compressor suitable for a refrigeration system. On the other hand, FIG.
In the compression element shown in FIG. 6, the rotation moment becomes negative when the pressure ratio becomes high, but it turns most when the pressure ratio is 10 or less (3-5 for both the room air conditioner and the package air conditioner) which is the pressure condition of the ordinary air conditioning system. The moment is small, and can be said to be a compression element of a positive displacement compressor suitable for an air conditioning system. Note that only the condition where the suction pressure Ps is low is shown here, but it is confirmed that the same tendency as in FIG. 18 is obtained even under the air conditioning condition where the suction pressure is high.

According to one embodiment of the present invention described above,
Since it is possible to arbitrarily select the compression element shape so as to conform to the required specifications of the system such as operating pressure conditions, etc., the degree of freedom of design is wide, and it is possible to provide the most suitable positive displacement compressor for each system. System performance can be further improved.

FIG. 19 shows that the compression elements shown in FIGS. 10 and 16 are each integrated with an electric element and assembled into a hermetic compressor, and the refrigeration conditions of the working fluid HFC134a (suction pressure Ps = 0.095 MPa, Discharge pressure Pd =
The experimental result comparing the noise of the compressor alone at 1.043 MPa) is shown. As a result of comparing the sound pressure levels of both compressors with the horizontal axis representing the rotational speed of the compressor, the compression element (solid line) shown in FIG. 10 is compared with the compression element (dashed line) shown in FIG.
It has been found that the noise can be reduced by about 7 to 13 dB (decibel). It was also confirmed that this difference in noise was caused by the difference in rotation moment acting on the displacer. That is, referring to FIG. 17, the compression element shown in FIG. 10 always has a positive rotation moment and a moment in a fixed direction acts on the displacer, whereas the compression element shown in FIG. It is an alternating rotation moment that switches between positive and negative three times during rotation. Therefore, FIG.
In the compression element shown in FIG. 16, the contact between the cylinder and the displacer is kept only in the rotation prevention section which is the contact portion between the cylinder and the displacer. In the compression element shown in FIG. In the middle of the rotation prevention section, the contact between the cylinder and the displacer is cut off, and the cylinder contacts the displacer on the opposite side to the direction of the previous rotation moment,
It was also confirmed from the observation of the sliding portion after the experiment that the tooth surface separation vibration caused by the impact of the contact was caused.
Therefore, the noise of the displacement type fluid machine can be reduced by forming the inner wall surface of the cylinder and the outer wall surface of the displacer so that the moment acting on the displacer by the internal pressure of the working fluid is in a certain direction. That is, as a rotation preventing mechanism of this type of compressor, the rotation moment always acts in one direction, and the rotation moment caused by the internal pressure load of at least one of the plurality of working chambers is reduced to the internal pressure of another working chamber. By forming the contour of the wrap so that the rotation moment is opposite to the rotation moment due to the load, the rotation prevention mechanism can be constructed without using the rotation prevention mechanism of another member such as the Oldham coupling. Here, the case where the rotation preventing mechanism receives the rotation moment acting on the displacer by contact with the cylinder has been described, but the present invention is not limited to this, and the case where a rotation preventing mechanism such as an Oldham coupling is separately provided is provided. In addition, the vibration of the anti-rotation mechanism is reduced, and the same noise reduction effect is achieved.

FIG. 19 also shows the results of noise experiments when the inner wall surface of the cylinder of the compression element shown in FIG. 10 is subjected to a surface treatment by using broken lines. Thus, by forming the surface treatment layer on the sliding portion of the inner wall of the cylinder that comes into contact with the displacer, an additional 10 dB can be obtained especially on the high-speed rotation side.
It was found that noise reduction of (dB) or more could be achieved.
FIG. 20 is a schematic diagram of the surface treatment layer formed on the inner wall surface of the cylinder at this time. In the figure, reference numeral 18 denotes a surface treatment layer formed on the inner wall surface of the cylinder 4, which is softer and more conformable than the base material (in the experiment, the material of the cylinder and the displacer is cast iron or iron-based metal), A film having a high oil retention, such as a manganese phosphate film.
By forming such a surface treatment layer between the contact part of the cylinder and the displacer, the sliding surface is smoothened by the effect of the conformity of the surface treatment film, and the lubricating property of the sliding part is also improved. It was confirmed that the contact was reduced, and the sliding noise caused by the engagement between the cylinder and the displacer was greatly reduced due to the damping action of the oil film and the like. Other surface treatment methods that exhibit the same effect include a sulfuration and nitriding treatment for a base material of cast iron and steel, and a fluororesin impregnation treatment for anodization of an aluminum alloy. The surface treatment layer may be formed not on the inner wall of the cylinder but on the outer wall of the displacer, and may be formed not only on the sliding portion but also on the entire surface.

Although the displacement type fluid machine having three vanes 4b on the inner periphery of the cylinder 4 has been described above, the present invention is not limited to this, and the number of vanes 4b is two. This is expanded to at least N positive displacement fluid machines (the value of N is practically 8 to 10 or less). At this time, the rotation moment can be reduced if the number of working chambers that generate the negative rotation moment is at least one. However, if the number is insufficient, the number of working chambers that generate the negative rotation moment may be increased. There are the following advantages as the number N of vanes increases as much as practical.

(1) Torque fluctuation is reduced, and vibration and noise are reduced. (2) When the cylinders are compared with the same outer diameter, the cylinder height for securing the same suction volume Vs is reduced, and the size of the compression element can be reduced. (3) Since the rotation moment acting on the displacer is reduced, the mechanical friction loss of the sliding portion between the displacer and the cylinder can be reduced and the reliability can be improved. (4) The pressure pulsation in the suction / discharge pipe is reduced, and the vibration and noise can be further reduced. This allows
A non-pulsating flow fluid machine (compressor, pump, etc.) required for medical or industrial use can be realized.

The positive displacement fluid machine described above can be applied to an air conditioning system of a heat pump cycle capable of cooling and heating. It comprises the positive displacement compressor 30, the outdoor heat exchanger and its outdoor fan, the expansion valve, the indoor heat exchanger and its indoor fan, and the four-way valve described in FIG. Displacement compressor 3
0 operates in accordance with the operation principle diagram shown in FIG. 2, and operates the working fluid (for example, CFC HCFC22, R407C, or the like) between the cylinder 4 and the displacer 5 by starting the compressor.
R410A).

In the cooling operation, the compressed high-temperature and high-pressure working gas flows into the outdoor heat exchanger from the discharge pipe 14 through the four-way valve, and radiates and liquefies by the blowing action of the outdoor fan.
It is throttled by an expansion valve, adiabatically expanded to a low temperature and a low pressure, absorbed by the indoor heat exchanger and gasified, and then sucked into the positive displacement compressor 30 through the suction pipe 13.

On the other hand, in the heating operation, the refrigerant flows in the opposite direction to the cooling operation by switching the four-way valve, and the compressed high-temperature and high-pressure working gas passes through the four-way valve from the discharge pipe 14 to exchange indoor heat. Into the chamber, radiate heat into the room by the action of the indoor fan, liquefy, squeeze by the expansion valve, adiabatically expand to low temperature and low pressure, and absorb heat from the outside air to gasify by the outdoor heat exchanger After that, it is sucked into the positive displacement compressor 30 through the suction pipe 13.

The displacement compressor can be applied to a refrigerating (cooling) dedicated cycle such as a refrigerator.

By activating the positive displacement compressor 30, the working fluid is compressed between the cylinder 4 and the displacer 5, and the compressed high-temperature and high-pressure working gas flows from the discharge pipe 14 into the condenser. Dissipation of heat by the blowing action of the fan,
It is liquefied, squeezed by an expansion valve, adiabatically expanded to a low temperature and low pressure, is endothermic gasified by an evaporator, and is sucked into a positive displacement compressor 30 through a suction pipe 13. Here, since both of the positive displacement compressors of the present invention are mounted, a highly reliable refrigeration / air-conditioning system with excellent energy efficiency, low vibration and low noise can be obtained. Although the low-pressure type is described as an example of the positive displacement compressor 30 here, the high-pressure type also functions in the same manner and has the same effect.

In the embodiments described above, a compressor is described as an example of a positive displacement fluid machine. However, the present invention can be applied to a pump, an expander, and a power machine. Further, in the present invention, one (cylinder side) is fixed and the other (displacer) performs a revolving motion without rotating at a substantially constant turning radius, but is relatively equivalent to the above motion. The present invention can also be applied to a double-rotation type positive displacement fluid machine that has a simple movement form.

[0082]

As described above in detail, according to the present invention, the rotational moment acting on the displacer is reduced to reduce the friction loss between the displacer and the cylinder, thereby improving the performance and reducing the noise. And a highly reliable positive displacement fluid machine can be obtained.

[Brief description of the drawings]

FIG. 1 is a longitudinal sectional view and a plan view of a compression element of a hermetic compressor in which a positive displacement fluid machine according to the present invention is applied to a compressor.

FIG. 2 is an explanatory view of the operation principle of the positive displacement fluid machine according to the present invention.

FIG. 3 is a longitudinal sectional view of a positive displacement fluid machine according to the present invention.

FIG. 4 is a diagram showing a volume change characteristic of a working chamber according to the present invention.

FIG. 5 is a graph showing a change in gas compression torque according to the present invention.

FIG. 6 is a diagram showing a relationship between a rotation angle of a rotation shaft and a working chamber in a four-row wrap.

FIG. 7 is a diagram showing a relationship between a rotation angle of a rotation shaft and a working chamber in the three-row wrap.

FIG. 8 is an operation explanatory diagram when the winding angle of the compression element is larger than 360 °.

FIG. 9 is a view for explaining the expansion of the winding angle of the compression element.

FIG. 10 is a modified example of the positive displacement fluid machine shown in FIG. 1;

FIG. 11 is an explanatory view of a load and a moment acting on the displacer of the present invention.

FIG. 12 is a diagram showing a relationship between a rotation angle of a rotation shaft of a compression element and a rotation moment ratio.

FIG. 13 is a longitudinal sectional view of a main part of a hermetic compressor according to another embodiment of the present invention.

FIG. 14 is a compression element diagram of a positive displacement fluid machine according to another embodiment of the present invention.

FIG. 15 is a rotation moment diagram during one rotation of a shaft acting on a displacer of the positive displacement fluid machine shown in FIG. 14;

FIG. 16 is a compression element diagram of a positive displacement fluid machine according to another embodiment of the present invention.

FIG. 17 is a comparison diagram of the rotation moment of the compression element shown in FIGS. 10, 14, and 16 during one rotation of the shaft.

FIG. 18 is a comparison diagram of the rotation moment when the pressure ratio of the compression element shown in FIGS. 10, 14, and 16 changes.

FIG. 19 is a comparison diagram of the sound pressure level with respect to the rotation speed of the compression element shown in FIGS. 10 and 16;

FIG. 20 is a schematic view of a surface treatment layer formed on an inner wall surface of a cylinder.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Displacement type compression element, 2 ... Electric element, 3 ... Closed container, 4
... Cylinder, 4a ... Inner peripheral wall, 4b ... Vane, 5 ... Displacer, 5a ... Bearing, 5b ... Through hole, 6 ... Rotating shaft, 6
a ... Crank part, 7 ... Main bearing, 7a ... Suction port, 8 ... Sub bearing, 8a ... Discharge port, 8b ... Discharge chamber, 9 ... Discharge valve, 10 ...
Suction cover, 11 ... Discharge cover, 12 ... Lubricating oil, 13 ...
Suction pipe, 14 ... discharge pipe, 15 ... working chamber, 16 ...
Seal member, 17: discharge passage, 18: surface treatment layer, 30
... displacement compressor, o ... center of displacer, o '... center of cylinder.

 ──────────────────────────────────────────────────続 き Continuing on the front page (72) Kunihiko Takao 502, Kandachi-cho, Tsuchiura-shi, Ibaraki Pref. Machinery Research Laboratory, Hitachi, Ltd. Within Hitachi, Ltd.Cooling Division (72) Inventor, Shigetaro Tagawa 800, Oda-machi, Ohira-cho, Shimotsuga-gun, Tochigi Prefecture Inside of Hitachi, Ltd.Cooling Division (72) Cooling and Heating Division (72) Inventor Kenji Tojo 390 Muramatsu, Shimizu-shi, Shizuoka Prefecture Inside Air Conditioning Systems Division, Hitachi, Ltd.

Claims (9)

[Claims] 1. A cylinder having an inner wall provided between end plates and having an inner wall composed of a continuous curved plane, and an outer wall provided so as to face the inner wall of the cylinder. In the displacement type fluid machine including the inner wall, the outer wall, and the displacer that forms a plurality of working spaces by the end plate when moved, when the rotation direction of the rotation shaft is a negative moment, the working space is A negative rotation moment acts on the displacer due to the internal pressure of the working fluid in the working space during one rotation of the rotation shaft, and the sum of the rotation moments of all the working spaces becomes positive during one rotation of the rotation shaft. A positive displacement fluid machine in which the cylinder inner wall curve and the displacer outer wall curve are formed. 2. A displacer and a cylinder are arranged between end plates, and when the cylinder center and the displacer center are aligned, one space is formed by the cylinder inner wall surface and the displacer outer wall surface, and the displacer and the cylinder are formed. In the displacement type fluid machine in which a plurality of working spaces are formed when the positional relationship with the turning position is set, when the rotating direction of the rotating shaft for swiveling the displacer is set to a negative moment, the working spaces rotate. A negative rotation moment acts on the displacer due to the internal pressure of the working fluid in the working space during one rotation of the shaft, and the sum of the rotation moments of all the working spaces becomes positive during one rotation of the rotating shaft. A positive displacement fluid machine in which a cylinder inner wall curve and the displacer outer wall curve are formed. 3. A displacer and a cylinder are disposed between end plates, and when the cylinder center and the displacer center are aligned, one space is formed by the cylinder inner wall surface and the displacer outer wall surface, and the displacer and the cylinder are formed. In the displacement type fluid machine in which a plurality of working spaces are formed when the positional relationship with the rotation position is provided, a plurality of rotation preventing sections are provided on the inner wall surface of the cylinder, and the cylinder and the displacer are provided only in the rotation preventing section. Is a positive displacement fluid machine that slides in contact. 4. A displacer and a cylinder are disposed between end plates, and when the cylinder center and the displacer center are aligned, one space is formed by the cylinder inner wall surface and the displacer outer wall surface, and the displacer and the cylinder are formed. In the displacement type fluid machine in which a plurality of working spaces are formed when the positional relationship with the swirling position is set, at least one of the plurality of working spaces is operated at a rotation angle of a rotating shaft for rotating the displacer. A positive displacement fluid machine having the cylinder inner wall shape and the displacer outer wall shape such that a rotational moment to the displacer due to an internal pressure load in a space has a rotational moment opposite to a rotational moment in another working space. 5. A displacer and a cylinder are arranged between end plates, and when the cylinder center and the displacer center are aligned, one space is formed by the cylinder inner wall surface and the displacer outer wall surface, and the displacer and the cylinder are formed. In a positive displacement fluid machine having a rotation preventing mechanism for preventing relative rotation of the displacer itself with respect to the cylinder, a plurality of spaces are formed when the positional relationship with When the working fluid is compressed in such a manner, the inner wall surface of the cylinder and the outer wall surface of the displacer are formed so that the moment acting on the displacer due to the compression reaction force is in a fixed direction, and the positive displacement fluid machine is used as the rotation preventing mechanism. . 6. A cylinder having an inner wall formed between a pair of end plates and having a continuous planar shape, and an outer wall provided so as to oppose the inner wall of the cylinder. In a displacement type fluid machine including an outer wall and a displacer forming a plurality of working spaces by the end plate, the cylinder inner wall and the displacer outer wall curve are such that a moment acting on the displacer due to an internal pressure of a working fluid or the like becomes a fixed direction. A positive displacement fluid machine formed as a compressor for a refrigerator. 7. A cylinder having an inner wall between a pair of end plates, the inner surface of which is formed by a continuous curved line, and an outer wall provided so as to face the inner wall of the cylinder. In a displacement type fluid machine including an outer wall and a displacer forming a plurality of working spaces by the end plate, the cylinder inner wall and the displacer outer wall curve are changed so that a direction of a moment acting on the displacer changes due to an internal pressure of a working fluid or the like. A positive displacement fluid machine formed as a compressor for a refrigerator. 8. A cylinder having an inner wall formed between a pair of end plates and having a continuous curved plane, and an outer wall provided so as to face the inner wall of the cylinder. In a displacement type fluid machine including an outer wall and a displacer forming a plurality of spaces by the end plate, a displacement type fluid machine provided with a surface treatment layer on at least one of the inner wall of the cylinder and the outer wall of the displacer. . 9. A cylinder having an inner wall between end plates and having a plane shape formed by a continuous curve, and an outer wall provided so as to oppose the inner wall of the cylinder. In a displacement type fluid machine including an outer wall and a displacer forming a plurality of spaces by the end plate, a film softer than a base material is formed on at least one of the inner wall of the cylinder and the outer wall of the displacer. Positive displacement fluid machinery.