JPH1137064A - Displacement type fluid machine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a displacement type fluid machine reduced in fluid loss in an intake process to the same degree as a scroll type fluid machine and higher,in intake efficiency than the scroll type fluid machine. SOLUTION: A displacer 5 and a cylinder 4 are arranged between end plates. When the center of the displacer 5 is aligned to the rotational center of a rotary shaft 6, one space is formed by the inner wall surface of the cylinder 4 and the external wall surface of the displacer 5, and when the position relation between the displacer 5 and the cylinder 4 is in a turning position, a plurality of spaces are formed. In a displacement type fluid machine l of such constitution, the curves of the internal wall surface of the cylinder 4 and the external wall surface of the displacer 5 are so formed that a rotation angle θc of the shaft in a stroke from the end of intake to the end of discharge in the plurality of spaces satisfies ((N-l)/N).360 deg.<θc<=360 deg., where N is the number of protruding parts protruded inward into the cylinder 4, and intake ports 7a communicating with the plurality of spaces are formed in positions of connecting two confinement end points formed by the cylinder 4 and turning motion of the displacer 5 arranged in the turning position.

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 that they are easy to manufacture and inexpensive because of their simple structure. There is a problem that the flow velocity in the process is high, and the performance is reduced due to an increase in pressure loss, and a problem that the rotation shaft system cannot be completely balanced due to the necessity of reciprocating the piston, resulting in large vibration and noise. .

[0004] In the rotary type fluid machine, although the stroke from the end of suction to the end of discharge is 360 ° of the shaft rotation angle, the problem of increased pressure loss in the discharge process is smaller than that of the reciprocating type fluid machine. Since the gas is discharged once per rotation, 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 is as long as 360 ° or more in terms of the shaft rotation angle.
Therefore, there is an advantage that the pressure loss in the discharge process is small, and generally a plurality of working chambers are formed, so that the fluctuation of the gas compression torque is small and the vibration and noise are 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. Further, since the stroke from the end of the suction to the end of the discharge is as long as 360 ° or more in terms of the shaft rotation angle, the compression process takes a long time and the internal leakage increases.

By the way, the displacer for moving the working fluid does not rotate relative to the cylinder into which the working fluid has been sucked, but revolves around a substantially constant radius, that is, makes a revolving motion to convey the working fluid. Japanese Patent Application Laid-Open No. 55-37b353 (Document 1)
U.S. Pat. No. 2,112,890 (Document 2);
JP-A-202869 (Document 3) and JP-A-6-2807
No. 58 (Reference 4). The displacement type fluid machine proposed here comprises a piston 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 piston, This piston moves the working fluid by revolving in the cylinder.

[0007]

Problems to be Solved by the Invention Documents 1 to 4 mentioned above
Does not have a reciprocating part unlike the reciprocating type, so that the rotating shaft system can be perfectly balanced. Therefore, the displacement type fluid machine has essentially advantageous features as a positive displacement fluid machine such that the vibration is small and the relative slip speed between the piston and the cylinder is small so that the 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 piston is as short as about 180 ° (210 °) at the shaft rotation angle θc (rotary). (Approximately half of the formula is about the same as the reciprocating formula), so there is a time lag (time lag) from the end of discharge of the working fluid to the start of the next (compression) stroke (end of suction). Since the working chamber from the end of suction to the end of discharge is formed to be biased around the rotation axis, the mechanical balance is poor, and the piston itself will be rotated as a reaction force from the compressed working fluid. Is excessively acting, and there is a disadvantage that reliability problems such as vane friction and wear are likely to occur. In addition, since the intake passages are dispersed and not optimally shaped, there is a disadvantage that performance is reduced due to fluid loss during the suction process.

SUMMARY OF THE INVENTION An object of the present invention is to provide a positive displacement type fluid machine which can reduce suction loss and relieve a rotational moment and has excellent performance and reliability.

[0010]

An object of the present invention is to dispose a displacer and a cylinder between end plates, and when the center of the displacer is aligned with the center of rotation of a rotating shaft, the inner wall surface of the cylinder and the outer wall surface of the displacer become one. In a positive displacement fluid machine in which two spaces are formed and a plurality of spaces are formed when the positional relationship between the displacer and the cylinder is set at a swiveling position, of the plurality of spaces,
Shaft rotation angle θc during the stroke from the end of suction to the end of discharge
And (((N−1) / N) · 360 °) <θc ≦ 360 ° (where N is the number of protrusions protruding inward of the cylinder) so that the inner wall surface of the cylinder and the A suction port communicating with the plurality of spaces is formed on the end plate, and the shape of the suction port communicates with the plurality of spaces and has a radial length of at least a turning radius or more. This is achieved by configuring the trajectory connecting the points with at least the contour curve of the cylinder.

[0011]

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 the present invention will be described with reference to FIGS. FIG. 1 (a) is a longitudinal sectional view of a hermetic compressor when a positive displacement fluid machine according to the present invention is used as a compressor ((b)
A), (b) is a BB sectional view of (a), FIG. 2 is an operation principle diagram of the displacement type compression element, and FIG. 3 is a hermetic seal when the displacement type fluid machine according to the present invention is used as a compressor. It is a component block diagram of a type compressor.

In FIG. 1A, a closed casing 3 accommodates a positive displacement element 1 according to the present invention and an electric element (not shown) for driving the same. The details of the positive displacement compression element 1 will be described. FIG. 1 (b) shows a triple 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 part, a plurality (in this case, 3
(There are three because of the strip wrap). The displacer 5 includes the cylinder 4
The inner peripheral wall 4a of the cylinder 4 (the vane 4b
And the vane 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 positive displacement compression element 1 will be described with reference to 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. Focusing on 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 is considered to be the start of the vortex), and the inner wall curve ( g-a) has a winding angle of about 3
60 ° (This is 360 ° in design, but does not exactly reach that value due to manufacturing errors. The same applies hereinafter.)
The winding angle is a vortex curve whose details will be described later), and the outer wall curve (g-b) is a vortex curve having a winding angle of about 360 °. The one inner peripheral contour shape is formed from an inner wall curve and an outer wall curve. Approximately equal pitch on these two curved circles (120
°), and the outer wall curve and the inner wall curve of the adjacent spiral body are connected by a smooth connection curve (bb ′) such as an arc to form the 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 the seven 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 to be described later and for manufacturing. In consideration of easiness, especially when these are not a problem, the pitch may be unequal.

Now, the compression operation by the cylinder 4 and the displacer 5 configured as described above will be described with reference to FIG. 7a is a suction port and 8a is a discharge port, which are provided at three places respectively. By rotating the rotating shaft 6, the displacer 5 does not rotate around the center o 'of the cylinder 4 on the fixed side, and the turning radius ε
(= Oo ') and revolve around the center o of the displacer 5.
Around the plurality of working chambers 15 (a plurality of closed spaces surrounded by the inner peripheral contour (inner wall) of the cylinder and the outer peripheral contour (side wall) of the displacer), a space in which suction is completed and a compression (discharge) process is completed. In other words, this space is a period from the end of suction to the end of discharge. If the above-mentioned winding angle is limited to 360 °, this space disappears at the end of compression, but the suction ends at that moment. Therefore, this space is counted as one, but when used as a pump, a space that communicates with the outside via a discharge port is formed. In this embodiment, three working chambers are always formed.

One hatched working chamber surrounded by the contact points a and b (partitioned at the end of suction, the two working chambers are connected immediately after the compression stroke is started. ). FIG.
(1) is 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).
When rotated 90 ° from FIG. 2 (4), it returns to the initial state of FIG. 2 (1). As a result, the working chamber 15 reduces 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 automatically opens due to the pressure difference, and the compressed working gas is discharged through the discharge port 8a. The shaft rotation angle from the end of suction (compression start) to the end of discharge is 360 °, and the next suction stroke is prepared while each of the compression and discharge strokes is being carried out. It will be a start. For example, focusing on the space formed by the contacts a and d, the suction has already been started from the suction port 7a at the stage of FIG. 2A, and the volume increases as the rotation progresses. When it comes to a state,
This space is divided. Fluid corresponding to this divided amount is supplemented from the space formed by the contacts b and e.

Next, this operation will be described in detail. Paying attention to the working chamber formed by the contacts a and b in the state of FIG. 2A, the space formed by the adjacent contacts a and d has begun to be sucked, and the fluid in this space has a shaft rotation angle of 360. After that, it should be compressed by the space formed by the contact points a and b, but this space once spreads as shown in FIG. 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. The same amount of fluid as the volume of the fluid that has been divided and not taken into the space formed by the contacts a and d is shown in FIG.
The space formed by the contacts b and e in the suction process in (4) is divided as shown in FIG. 2A, and flows into the space formed by the contacts e and b near the discharge port. Is filled by the flowing fluid.

This is because, as described above, they are not arranged at an uneven pitch but are arranged at an equal pitch.
That is, since the shapes of the displacer and the cylinder are formed by repetition of the same contour shape, substantially the same amount of fluid can be compressed even if fluid is obtained from different spaces in any of the working chambers. 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 above-described prior arts, the space in the suction process is closed and compressed and discharged as it is, whereas the positive displacement compression element 1 of the present embodiment is initially in the suction process connected to one. One of the features of the present embodiment is a series of operations in which the working chamber is divided, and a compression operation is performed in spaces (working chambers) adjacent to each other with the suction port interposed therebetween and discharged from another discharge port.

Next, the suction port of the present invention is shown in FIG.
This will be described with reference to FIGS. FIG. 3 is a perspective view of a cylinder 4, a displacer 5, and a main bearing member 7 which are main components of the positive displacement compression element 1. FIG. The main bearing member 7 is opposed to the main bearing member 7 via the suction port 7a, the main bearing member 7b which supports the rotating shaft 6, the screw hole 7c for fixing the cylinder 4, and the main bearing member 7 via the cylinder 4 and the displacer 5. Discharge port 8 formed in sub-bearing member 8 attached by mounting
A counterbore-shaped pressure equalizing hole 7d having the same diameter as that of the discharge port is formed at a position opposed to a. 7e is a screw hole for fixing the cylinder 4, and 7f is the discharge cover 1.
1 is a screw hole for fixing the sub bearing member 8 and the cylinder 4 together.

Next, the suction port 7a of the present invention will be described with reference to FIGS. FIG. 4 shows a state in which the working gas has been sucked into the compression chamber 15a formed by the contact a (closing end point) and the contact b, and FIG. 5 shows a state formed by the contact c (closing end point) and the contact d. Compression chamber 15b
This is a state in which the suction of the working gas into has been completed. The contour shape of the suction port 7a is defined by the contact a, which is the closing end point of the compression chamber 15a, and the suction port 7a.
The section up to the section L1 up to the contact point c of the compression chamber 15b adjacent to the compression chamber 15b is constituted by the contour of the cylinder. Also, the compression chamber 15
A section L2 from the space a on the side a to the contact point a, which is the closing end point, connects the trajectory formed by the displacer 5 with the turning movement, that is, the contour line 5c of the displacer 5 from the contact point a, which is the closing end point in FIG. It consists of a locus. A section L3 from the space on the compression chamber 15b side to the closing end point c is a trajectory formed by the displacer 5 in accordance with the turning motion, that is, the contour of the displacer 5 from the contact point c which is the closing end point in FIG. It is composed of a locus connecting the lines 5d. The radial width of the suction port 7a is formed to be equal to the turning radius ε. Since the displacement type compression element 1 of the present invention has three wraps, the suction port 7a
Are arranged at equal pitches (120 °) on the circumference.

As a result, the compression chamber 15 adjacent to the compression chamber 15 from one suction port 7a, which is also a feature of the positive displacement compression element 1, is provided.
When the working gas is supplied to the valves a and 15b, a necessary and sufficient suction volume can be ensured, and the suction efficiency (volume efficiency) can be improved, and the (refrigeration) capacity of the compressor can be improved. Also,
Since the suction port 7a is disposed at the sliding portion between the displacer 5 and the cylinder 4 which receive the rotation moment (described in detail in FIG. 11), the sliding portion exists in a working gas atmosphere having a low temperature and a high oil viscosity. As a result, the sliding conditions are reduced, and a highly reliable positive displacement compressor that solves the problem of friction and wear can be provided. Further, by setting the width of the suction port 7a to the turning radius ε, fluid loss to the displacer 5 during the suction process can be minimized, and the behavior of the displacer 5 can be stabilized.

In this embodiment, the section L1 of the suction port 7a is formed by the contour of the cylinder 4. However, there is no problem even if the section L1 extends over the contour of the cylinder 4. In addition, the shape of the suction port 7a only needs to be secured at least in the section L1 from the closing end point a of the compression chamber 15a to the closing end point c of the compression chamber 15b, and the shapes of the sections L2 and L3 are required. Any shape may be used as long as the displacer 5 is inside the trajectory formed by the turning motion.

The width of the suction port 7a may be larger than the turning radius ε within a range that does not affect the behavior of the displacer 5. Further, by forming the contour of the suction port 7a of the present invention, the oil supply groove 5b formed on the upper and lower end surfaces of the displacer 5 can be made as close as possible to the vicinity of the suction port 7a, and the oil supply hole ( (Not shown), it is possible to supply a necessary and sufficient amount of lubricating oil from the center of the lubrication groove 5b formed in the displacer 5 to the vicinity of the suction port 7a. Performance can be improved. Also,
Although the suction port 7a of the present invention is arranged on the main bearing member 7,
There is no problem even if the auxiliary bearing member 8 is attached to the main bearing member 7 so as to be opposed to the main bearing member 7 via the cylinder 4 and the displacer 5 depending on the specifications of the compressor, production equipment, and the like.

As described above, the working chambers for the continuous compression operation are distributed around the crank part 6a of the rotating shaft 6 located at the center of the displacer 5 at substantially equal pitches. The chambers are compressed out of phase. That is, when focusing on one space, the shaft rotation angle is 360 ° from the suction to the discharge, but in 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 within 360 ° of the shaft rotation angle. In this manner, the discharge pulsation of the refrigerant can be reduced, and the space adjacent to the working chamber in the suction process is divided so as to perform the compression operation, which is different from the reciprocating type, the rotary type and the scroll type.

If the space at the moment when the compression operation is completed (the space surrounded by the contact points a and b) is regarded as 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.

FIG. 6 shows a method of assembling the positive displacement compression element according to the present invention. In the figure, 7 is a main bearing member having both a main bearing 7b for supporting the rotary shaft 6 and a suction port 7a, 4 is a three-cylinder type cylinder, 6 is a rotary shaft having a double support structure,
Reference numeral 5 denotes a three-type displacer, 8 denotes an auxiliary bearing 8c that supports the rotating shaft 6, a discharge port, and a discharge valve 9 of a reed valve type.
Bearing member which also has a hole 11 and a discharge chamber 8b.
a, a discharge port 11b provided with a hole 11c for fixing the vane 4b of the cylinder 4 to the main bearing member 7; and 37b each of the cylinder 4, the sub-bearing member 8 and the discharge cover 11 in the main bearing member 7. Holes 4e, 8d, 11a
And 37a are bolts for fixing the vane 4b of the cylinder 4 to the main bearing member 7 via the sub bearing member 8 and the respective holes 4c, 8f, 11c of the discharge cover 11.

The components constituting the positive displacement compression element of this embodiment are almost end plate-shaped, and the main bearing member 7 having the suction port 7a and the cylinder 4 are separate components.
The main bearing member 7 is bolted through the hole 4d of the cylinder 4 so as to match the positional relationship between the suction port 7a and the cylinder 4.
In the direction of the arrow in the figure,
Displacer 5, auxiliary bearing member 8, and discharge cover 1
Stack the parts in the order of 1 and finally, bolt 3
By fixing them together at 7b, 37a, the displacement type compression element can be easily assembled. In this state, the positive displacement compression element can be turned upside down and fixed in the closed container.

As a result, it is possible to provide a positive displacement compression element having good productivity. Further, since each component constituting the positive displacement compression element has an end plate shape, the accuracy of the component alone can be improved, and a high performance positive displacement compression element can be provided. Furthermore, by shaping each component from a sintered material, the need for machining is eliminated, and a low-cost positive displacement compression element can be provided.

Next, a compressor incorporating the positive displacement compression element 1 having such a shape will be described with reference to FIGS. In FIG. 7, in addition to the cylinder 4 and the displacer 5 described in detail above, a positive displacement element 1 includes a rotary shaft 6 for driving the displacer 5 by driving a displacer 5 with a crank 6a fitted to a bearing at the center of the displacer 5. A main bearing member 7 and a sub-bearing member 8 serving also as an end plate for closing both end openings of the cylinder 4 and a bearing for supporting the rotary shaft 6; a suction port 7a formed in an end plate of the main bearing member 7; A discharge port 8a formed in the end plate of the sub-bearing member 8;
Reed valve 9 that opens and closes (opens and closes with differential pressure)
Having. 4c is a fixing hole for suppressing radial deformation of the vane 4b of the cylinder 4; 4d is a hole for temporarily fixing the cylinder 4 to the main bearing member 7;
Are fixed to the main bearing member 7. 5a is a through hole formed in the displacer 5, and 5b is an oil supply groove formed on both end surfaces of the displacer 5. Also, 10
Reference numeral denotes a suction cover attached to the main bearing member 7, and reference numeral 11 denotes a discharge cover for forming the discharge chamber 8b integrally with the sub bearing member 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 closed 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-mentioned working chamber formed by the engagement 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 closed container 3 by a seal member 16 such as an O-ring.

The flow of the working gas (refrigerant) 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 member 7, passes through the suction port 7a, and moves through the suction port 7a. Then, the displacer 5 performs a swiveling motion by the rotation of the rotating shaft 6 and the volume of the working chamber is reduced, so that the working chamber is compressed. The compressed working gas pushes up the discharge valve 9 through the discharge port 8a formed in the end plate of the sub bearing member 8, enters the discharge chamber 8b, and flows out through the discharge pipe 14 to the outside. The suction pipe 13
The reason why the gap is formed between the motor element and the suction cover 10 is to cool the motor element by flowing the working gas also into the motor element 2.

Here, a method of forming the contours of the displacer 5 and the cylinder 4 which are the main parts constituting the positive displacement compression element 1 of the present invention will be described with reference to FIGS. For example). 8A and 8B are examples of the shape of a displacer having a planar shape formed by a combination of arcs as an example. FIG. 8A is a plan view, and FIG. 8B is a side view. (A) and (b) of FIG.
FIG. 4A is an example of a cylinder shape that meshes with a pair of displacers shown in FIG. 4, wherein FIG. 4A is a plan view and FIG. 4B is a side view. FIG. 10 is a diagram in which the center o of the displacer shown in FIG. 8 and the center o ′ of the cylinder shown in FIG. 9 are drawn (one set).

In FIG. 8A, in the planer shape of the displacer, the same contour shape is continuously connected at three places around the center o (the centroid of the equilateral triangle IJK). The contour shape is formed by a total of seven arcs from the radius R1 to the radius R7, and the points p, q, r, s, t, u, v, w
Are connection points of arcs of different radii. The curve pq is
An arc of radius R1 centered on one side IK of the equilateral triangle, where point p is at a distance R7 from vertex I. The curve qr is an arc of a radius R2 having a center on an extension of a straight line connecting the connection point q and the center of the radius R1, and the curve rs is a connection point r and a radius R2.
Similarly, the curve st is a circular arc having a radius R3 having a center on a straight line connecting the connection point s and the center of the radius R3. The curve tu is an arc of a radius R5 having a center on an extension of a straight line connecting the connection point t and the center of the radius R4, and the curve uv is a connection point u and a radius R5
An arc having a radius R6 centered on a centroid o on an extension of a straight line connecting the centers of the circles is a radius R7 centered on a vertex J on a straight line connecting the connection point v and the center (centroid o) of the radius R6. Is an arc.

The radii R1, R2, R3, R4, R5
The angle of each arc of R6 is determined by the condition that the connection is made smoothly at the connection point (the inclination of the tangent line at the connection point is the same). When the contour shape from the connection point p to the connection point w is rotated counterclockwise around the centroid o by 120 °, the connection point p overlaps the connection point w, and when further rotated by 120 °, the contour shape around the entire circumference is completed. I do. Thereby, the planar shape (thickness h) of the displacer is obtained.

When the plane shape of the displacer is determined, when the displacer makes a revolving motion with a revolving radius ε, the contour of the cylinder that meshes with the displacer is, as shown in FIG. An offset curve having a normal distance of ε is obtained.

The contour of the cylinder will be described with reference to FIG. Triangle IJK is the same equilateral triangle as in FIG. The contour shape is formed by a total of seven circular arcs like the displacer, and the points p ′, q ′, r ′, s ′, t ′,
u ′, v ′, w ′ are connection points of arcs having different radii. A curve p′q ′ is an arc having a radius (R1−ε) centered on one side IK of an equilateral triangle, where the point p ′ is at a distance of (R7 + ε) from the vertex I. The curve q'r 'is an arc of a radius (R2-?) Having a center on an extension of a straight line connecting the connection point q' and the center of the radius (R1-?), And the curve r's 'is a connection point r' and a radius An arc of a radius (R3-ε) having a center on a straight line connecting the center of (R2-ε) and a curve s′t ′ are similarly drawn on a straight line connecting the connection point s ′ and the center of the radius (R3-ε). This is an arc having a center and a radius (R4 + ε). Curve t'u '
Is an arc of radius (R5 + ε) centered on an extension of a straight line connecting the connection point t ′ and the center of radius (R4 + ε), curve u ′
v ′ is an arc of a radius (R6 + ε) centered on a centroid o ′ on an extension of a straight line connecting the connection point u ′ and the center of the radius (R5 + ε), and a curve v′w ′ is a arc of a connection point v ′ and a radius (R6 + ε). ) Is an arc of a radius (R7 + ε) centered on a vertex J on a straight line connecting the center (center o ′).

Note that the radii (R1-ε), (R2-ε),
(R3-ε), (R4 + ε), (R5 + ε), (R6 +
Like the displacer, the angle of each arc in ε) is determined by the condition that the connection is made smoothly at each connection point (the inclination of the tangent at the connection point is the same). When the contour shape from the connection point p 'to the connection point w' is rotated counterclockwise around the center o 'by 120 °, the connection point p' coincides with the connection point w '. Is completed. Thereby, the planar shape of the cylinder is obtained. The thickness H of the cylinder is slightly larger than the thickness h of the displacer.

FIG. 10 is a diagram in which the center o of the displacer and the center o 'of the cylinder are overlapped. As can be understood from the figure, the gap formed between the displacer and the cylinder has ε equal to the turning radius. It is desirable that this gap be ε over the entire circumference. However, within a range where the working chamber formed by the outer peripheral contour of the displacer and the inner peripheral contour of the cylinder operates normally, for some reason, this relationship is not satisfied. There is no problem even if there are collapsed parts.

Here, as the method of forming the contour shapes of the displacer and the cylinder, a method using a combination of multiple circular arcs has been described. However, the present invention is not limited to this, and the present invention is also applicable to a combination of arbitrary (higher-order) curves. A similar contour shape can be configured.

FIG. 11 is an explanatory diagram of loads and moments 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. Due to the deviation of the resultant force F from the center o of the displacer 5 (the length l of the arm), the rotation moment M for rotating the displacer 5
(= F · l) works. The reaction force R1 and the reaction force R2 at the contact points e and b of the displacer 5 and the cylinder 4 support the rotation moment M.

According to the present invention, at all times, 2 is close to the suction port 7a.
In addition, a moment is received at three or three contact points, and no reaction force acts on the other contact points. The positive displacement compression element 1 according to the present invention comprises:
Around the crank part 6a of the rotating shaft 6 fitted to the center of the displacer 5, working chambers in which the shaft rotation angle from the end of suction to the end of discharge becomes substantially 360 ° at a substantially equal pitch are dispersed and arranged. Therefore, the point of action of the resultant force F can be brought closer to the center o of the displacer 5, and the arm length l of the moment can be reduced to reduce the rotation moment M. Therefore, the reaction force R1 and the reaction force R2 are reduced.

FIG. 12 shows a comparison of the rotational moment M acting on the displacer during one rotation of the shaft due to the internal pressure of the working fluid between the compression element shown in FIG. 19 and the compression element shown in FIG. Calculation conditions are working fluid HFC1
34a refrigeration conditions (suction pressure Ps = 0.095 MPa,
The discharge pressure Pd is 1.043 MPa). Accordingly, in the compression element according to the present embodiment in which the maximum value of the number n of the working chambers is equal to or greater than the number of the working chambers, the working chambers from the end of the suction to the end of the discharge are dispersed around the rotation axis at a substantially equal pitch. It is possible to provide a structure in which the mechanical balance is improved and the load vector due to compression is 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 suction port 7a and the discharge port 8a
The relationship between the period in which is communicated and the rotation angle of the compression stroke rotation shaft will be described. The time lag Δθ represented by the rotation angle of the rotating shaft between the time when the suction port and the discharge port communicate with each other, that is, from the end of the discharge of the working fluid to the start of the next compression stroke (end of suction) is the rotation axis of the compression stroke. Is represented as Δθ = 360 ° −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 clearance 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 is reduced due to the re-expansion of the gas in the clearance volume of the discharge port, and the (refrigeration) capacity of the compressor is reduced. Will decrease. 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 rotating 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 cylinder is set to 360 °, the rotation angle θc of the rotating shaft in the compression stroke is 36 °.
The angle can be set to 0 ° and θc <360 ° by moving the seal point of the suction port or the discharge port. 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. 18 described later can be changed to θc = 360 ° by changing the position and size of the discharge port. This can be realized by increasing the discharge port so that the working chamber 15a and the working chamber 15b communicate with each other immediately after the suction end state in FIG. 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.

FIG. 13 shows the suction port of the present invention for a four-line wrap in which four sets of the same contour shape are combined at a 90 ° pitch. The basic concept of the method of forming the contour shapes of the displacer 18 and the cylinder 19 and the compression operation and the like are the same as those of the aforementioned three-line wrap. The drawing shows a state where the suction of the working gas into the compression chamber 15a formed by the contact a (closing end point) and the contact c and the contact b '(closing end point) and the compression chamber 15b formed by the contact c is completed. It is. The contour shape of the suction port 21a is constituted by the contour line of the cylinder from the contact point a, which is the closing end point of the compression chamber 15a, to the section L1 from the contact point b 'of the adjacent compression chamber 15c. Also, the compression chamber 15a
L from the space on the side to the contact point a, which is the end point of confinement
Reference numeral 2 denotes a trajectory formed by the displacer 18 along with the turning motion. Similarly, a section L3 from the space on the side of the compression chamber 15c to the closing end point b ′ is also
The displacer 18 is configured by a trajectory formed by the turning motion. The radial width of the suction port 7a is formed to be equal to the turning radius ε.

Since the displacement type compression element 1 of the present invention has four wraps, the suction ports 21a are arranged at equal pitches (90 °) on the circumference.
At four locations. As a result, when supplying the working gas from the one suction port 21a to the adjacent compression chambers 15a and 15c as in the case of the above-described three-line wrap, it is possible to secure a necessary and sufficient suction volume, and the suction efficiency (volume efficiency) As a result, the (refrigeration) capacity of the compressor can be improved.
As described with reference to FIG. 6, the components forming the positive displacement compression element of this embodiment are almost end plate-shaped, and the main bearing member 21 having the suction port 21a and the cylinder 1 are formed.
9 is a separate part, the suction port 21a and the cylinder 1
9 is fixed to the main bearing member 21 with a bolt or the like through a hole of the cylinder 21 in accordance with the positional relationship of 9 and finally the outer peripheral portion is integrally fixed with a bolt or the like to assemble the displacement type compression element. Can be easily performed.

Further, the shape of the suction port of the present invention can be expanded to a positive displacement fluid machine having two or more vanes 19a (N is practically 8 to 10 or less). Further, the suction port of the present invention can be arranged according to the number of vanes, and the suction efficiency can be ensured according to the number of vanes. As described above, 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 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 between the sliding portion of 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.

FIG. 14 shows the change in volume of the working chamber (indicated by the ratio between the suction volume Vs and the working chamber volume V) in the present invention, with the shaft rotation angle θ from the end of the suction being taken on the horizontal axis. Shown in comparison. Thus, the volume change characteristics of the positive displacement compression element 1 according to the present embodiment are based on the condition of the air conditioner with the discharge start volume ratio of 0.37 (for example, when the working gas is Freon HCFC22, the suction pressure Ps = 0.64 MPa, the discharge pressure Pd). =
Comparing with 2.07 MPa), the compression process is almost the same as that of the reciprocating type, and since the compression process is completed in a short time, leakage of working gas is reduced, and the capacity and efficiency of the compressor can be improved. On the other hand, the discharge process is about 50% longer than the rotary type (rolling piston type), the pressure loss is reduced due to the slower discharge flow rate, and the fluid loss (excess compression loss) in the discharge process is greatly reduced to improve performance. Can be achieved.

FIG. 15 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 displacement type compression element 1 of the present invention is very small, about 1/10 of that of the rotary type, and is equivalent to that of the scroll type. Since no rotating mechanism is provided, the balance of the rotating shaft system can be completely maintained, and the vibration and noise of the compressor can be reduced. In addition, since it is not a long spiral shape like the scroll type, processing time and cost can be reduced, and since there is no end plate (end plate) for maintaining the spiral shape, it is possible to work by penetrating a jig. Compared to the scroll type, which can not be manufactured, it can be manufactured by the same processing as the rotary type, and the thrust load does not act, and it is easier to manage the axial clearance which has an important effect on the performance of the compressor, so the performance can be improved . Furthermore, it can contribute to the reduction in size and weight of the compressor.

Next, the relationship between the above-mentioned winding angle and the shaft rotation angle θc from the end of suction to the end of discharge will be described in detail.
The shaft rotation angle θc can be changed by changing the winding angle. For example, when the shaft rotation angle 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 communicates with the suction port, and the fluid expands in the discharge port once. A problem occurs such that the sucked fluid flows backward. Further, when the shaft rotation angle from the end of suction to the end of discharge is made larger than the winding angle of 360 ° to increase the shaft rotation angle, the size differs from the end of suction to the time of communication with the space having the discharge port. When the two working chambers are formed and used as a compressor, irreversible mixing loss occurs when the two working chambers are combined because the pressure rises of the two working chambers are different from each other, increasing the compression power and increasing the rigidity of the displacer. Decrease. Further, even if an attempt is made to use the liquid pump, a working chamber that does not communicate 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 shaft rotation angle θc of the compression stroke in the fluid machine described in the above-mentioned Japanese Patent Application Laid-Open No. 55-37b353 (Reference 1) is θc = 180 °, and
No. 69 (Reference 3) and Japanese Patent Application Laid-Open No. 6-280758 (Reference 4) have a shaft rotation angle θc of 210 ° during the compression stroke in the fluid machine. The period from the end of the discharge of the working fluid to the start of the next compression stroke (end of suction) is 180 ° in the shaft rotation angle θc in Reference 1, and 150 ° in References 3 and 4.

When the shaft rotation angle θc of the compression stroke is 210 °, each working chamber (reference numerals I, II, III, IV) during one rotation of the shaft
16) is shown in FIG. 16 (a). However,
The number of threads N = 4. Four working chambers are formed when the shaft rotation angle θc 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. 17A shows the case where the number of threads N = 3 and the shaft rotation angle θc in the compression stroke is 210 °. Also in this case, the number n of working chambers formed simultaneously is n = 1 or 2
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 to be deviated around the rotation axis, so that a mechanical imbalance occurs, the rotation moment acting on the displacer becomes excessive, the contact load between the displacer and the cylinder increases, and There is a problem that performance is reduced due to an increase in friction loss and reliability is reduced due to wear of the vane.

In order to solve this problem, in the present embodiment, the shaft rotation angle θc in the compression stroke is (((N−1) / N)
(360 °) <θc ≦ 360 ° (Equation 1) The outer contour of the displacer and the inner contour of the cylinder are formed. In other words, the above-mentioned winding angle is in the range of Expression 1. Referring to FIG. 16B, the shaft rotation angle θc in the compression stroke is larger than 270 °, the number n of working chambers formed simultaneously 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. 17B, the shaft rotation angle θc in the compression stroke is larger than 240 °, 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. This value is equal to the number N of rows (= 3).

By setting the lower limit value of the shaft rotation angle θc of the compression stroke larger than the value on the left side of the equation 1, the maximum value of the number of working chambers becomes equal to or more than the number N, and the working chamber is rotated around the rotating shaft. The mechanical balance is improved, the rotational moment acting on the displacer is reduced, the contact load between the displacer and the cylinder is reduced, the performance is improved by reducing mechanical friction loss, and the contact area is improved. Reliability can be improved.

On the other hand, the upper limit of the shaft rotation angle θc in the compression stroke is 360 ° according to the equation (1). The upper limit of the shaft rotation angle θc in this compression stroke is ideally 360 °. As described above, the time lag from the end of the discharge of the working fluid to the start of the next compression stroke (end of suction) can be made zero, and the gas in the gap volume occurring when θc <360 ° can be re-established. It is possible to prevent a reduction in suction efficiency due to expansion, and to prevent irreversible mixing loss that occurs when the two working chambers are combined due to different pressure increases in the two working chambers when θc> 360 °. Figure 18 for the latter
This will be described with reference to FIG.

The shaft rotation angle θc in the compression stroke of the displacement type fluid machine shown in FIG. 18 is 375 °. FIG.
(A) shows a state in which the suction of the two working chambers 15a and 15b shaded 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. 18 shows a state in which 15 ° rotation has progressed at this axis rotation angle θc from this state.
(B). 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. 18A, 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. Next moment working chamber 15a
When 15b and 15b are united (communicated), irreversible mixing occurs as indicated by an arrow in FIG. 18C, and performance is reduced due to an increase in compression power. Therefore, it is concluded that the upper limit of the shaft rotation angle θc in the compression stroke is ideally 360 °.

FIGS. 19A and 19B show a compression element of a positive displacement fluid machine described in Reference 3 or 4, wherein FIG. 19A is a plan view and FIG. 19B is a side view. The number of threads N is 3, and the shaft rotation angle θc (winding angle) in the compression stroke is 210 °. In this figure, the number n of working chambers is n = 1 as shown in FIG.
Or it becomes 2. This figure shows a state where the shaft rotation angle θc is 0 °, and the number n of working chambers is two. As is apparent from this figure, 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 7a and a discharge port 8a.
Are in communication. Therefore, there is a problem that the gas once flowing into the cylinder 4 from the suction port 7a 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. By the way, consider a case where the shaft rotation angle θc of the compression stroke of the positive displacement type fluid machine shown in this figure is expanded by using the concept of the present embodiment. In order to enlarge the shaft rotation angle θc during the compression stroke, the cylinder 4
Must be increased, the thickness of the vane 4b becomes extremely thin as shown in FIG.
It is difficult to make the shaft rotation angle θc of the compression stroke larger than 240 ° so that the maximum value of the rotation number is equal to or greater than the number N of rows (N = 3).

FIG. 20 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 shaft rotation angle θc in the compression stroke of the compression element shown in FIG.
Realizes 360 ° larger than 240 °. This is because, in the compression element shown in FIG. 19, the space between the seal points forming the working chamber is formed by a smooth curve, and therefore, for example, the shaft rotation angle θc of the compression stroke based on the concept of the present embodiment. In the compression element according to the present embodiment shown in FIG. 20, the gap between the seal points (ac) is not smooth (not a uniform curve) even if it is attempted to enlarge. The shape in the vicinity of the contact point b is formed so as to protrude when viewed from the displacer, and a constriction exists in the middle of each strip of the displacer from the center to the tip. These are also true for the embodiment shown in FIG. With these shapes, the winding angle from the contact a to the contact b can be set to 360 ° larger than 240 °, and the winding angle from the contact b to the contact c can be set to 360 ° larger than 240 °. As a result, the shaft rotation angle θc in the compression stroke can be set to 360 ° which is larger than 240 °, and the maximum value of the number n of working chambers can be set to the number N or more. For this reason, the working chambers are dispersed and the rotation moment can be reduced.

Further, as the number of working chambers that can function effectively increases, assuming that the cylinder height (thickness) of the compression element shown in FIG. 19 is H, the compression element shown in FIG. Since the cylinder height is 0.7H, which is 30% lower, the size of the compression element can be reduced. Further, by arranging the suction port 7a of the present embodiment having a circular shape in the optimum shape of the present invention, the suction efficiency can be improved.

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. 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 temperatures of the stator 2a and the rotor 2b are reduced, and the motor efficiency is improved and the performance can be improved. (2) In a working fluid compatible with the lubricating oil 12 such as Freon, the pressure is low, so that the proportion of the working gas dissolved in the lubricating oil 12 is reduced, so that the oil bubbling phenomenon in bearings and the like is less likely to occur, and the reliability is improved. Performance 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. 21 is an enlarged sectional view of a main part of a high-pressure hermetic compressor using a positive displacement fluid machine according to another embodiment of the present invention as a compressor. In FIG. 21, components denoted by the same reference numerals as those in FIGS. 1 to 7 are the same components and perform the same operations. FIG.
In FIG. 1, reference numeral 7b denotes a suction chamber formed integrally with the main bearing member 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 flows through the suction pipe 13 and enters the suction chamber 7b as shown by the arrow in the drawing. The working gas flows through the suction port 7a formed in the main bearing member 7 to be positively compressed. The element 1 is entered, in which the rotation of the rotating shaft 6 causes the displacer 5 to make a pivoting movement and to be compressed by reducing the volume of the working chamber 15. The compressed working gas is supplied to a discharge port 8 a formed in an end plate of the sub bearing member 8.
, The discharge valve 9 is pushed up into the discharge chamber 8 b, and through the discharge passage 17 into the closed container 3.
Flows out from a discharge pipe (not shown) connected to the outside. The advantage of such a high-pressure type is that the lubricating oil 12 is at a high pressure, so that the lubricating oil 12 supplied to each bearing sliding portion by the centrifugal pump action or the like due to the rotation of the rotary shaft 6 has a gap between the end faces of the displacer 5. And the like, it is easy to be supplied into the cylinder 4 through the like, and thus the sealing property of the working chamber 15 and the lubrication property 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 or the production equipment, etc. To expand. Further, it is possible to apply the suction port shape of the present invention to a high pressure type,
Similar effects can be obtained.

FIG. 22 shows an air conditioning system to which the positive displacement compressor of the present invention is applied. This cycle is a heat pump cycle capable of cooling and heating, and includes the positive displacement compressor 22, the outdoor heat exchanger 23, and the fan 2 of the present invention described with reference to FIG.
4, expansion valve 25, indoor heat exchanger 26 and its fan 27,
It comprises a four-way valve 28. The chain line indicates the outdoor units 29 and 30 are indoor units. Positive displacement compressor 22
Operates in accordance with the operation principle diagram shown in FIG. 2, and operates the working fluid (for example, CFC HCFC22 or R4) between the cylinder 4 and the displacer 5 by activating the positive displacement compressor 22.
07C, R410A, etc.).

In the cooling operation, the compressed high-temperature and high-pressure working gas flows from the discharge pipe 14 through the discharge pipe 14 as indicated by the solid arrow.
The air flows into the outdoor heat exchanger 23 through the direction valve 28, is radiated and liquefied by the blowing action of the fan 24, and is throttled by the expansion valve 25.
After being adiabatically expanded to a low temperature and low pressure, the indoor heat is absorbed by the indoor heat exchanger 26 and gasified, and then is sucked into the positive displacement compressor 22 through the suction pipe 13. On the other hand, in the case of the heating operation, as indicated by the dashed arrow, the air flows in the opposite direction to the cooling operation, and the compressed high-temperature and high-pressure working gas flows from the discharge pipe 14 through the discharge pipe 14.
Flows into the indoor heat exchanger 26 through the
7, the heat is radiated into the room, liquefied, squeezed by the expansion valve 25, adiabatically expanded to a low temperature and low pressure, and the outdoor heat exchanger 23 absorbs heat from the outside air to be gasified and then sucked. It is sucked into the positive displacement compressor 22 through the pipe 13.

FIG. 23 shows a refrigeration system equipped with the positive displacement compressor of the present invention. This cycle is a cycle dedicated to freezing (cooling). In the figure, 31 is a condenser, 3
2 is a condenser fan, 33 is an expansion valve, 34 is an evaporator, 35
Is an evaporator fan. By starting the positive displacement compressor 22, 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 to the condenser 31 as shown by a solid line arrow. After flowing in, the heat is radiated and liquefied by the blowing action of the fan 32, is throttled by the expansion valve 33, adiabatically expanded to a low temperature and low pressure, is endothermic gasified by the evaporator 34, and is volumetrically compressed through the suction pipe 13. Machine 22. Here, FIG. 22, FIG.
Since both are equipped with the positive displacement compressor 22 of the present invention,
A highly reliable refrigeration / air-conditioning system with excellent energy efficiency, low vibration and low noise can be obtained. Although the low-pressure type compressor 22 has been described as an example here, the high-pressure type also functions in the same manner and can achieve the same effect. Further, by mounting the positive displacement compressor 22 of the present invention, a silencer or the like is not required, and the cost of the system can be reduced.

[0069]

As described in detail above, according to the present invention, the displacer and the cylinder are arranged between the end plates,
When the center of the displacer is aligned with the center of rotation of the rotating shaft, one space is formed by the inner wall surface of the cylinder and the outer wall surface of the displacer, and the shaft rotation angle from the end of suction to the end of discharge of each working chamber is approximately 360 °. A suction port communicating with the plurality of spaces is disposed on the end plate, and has a shape having a radius of at least a turning radius or more in a radial direction, and connects the closing end points of different spaces. By forming the trajectory at least with the contour curve of the cylinder, the suction efficiency to the adjacent compression chambers can be improved, the overcompression loss in the discharge process can be greatly reduced, and the rotation moment acting on the displacer has been reduced to reduce the rotation moment between the displacer and the cylinder. By reducing the friction loss, a positive displacement type fluid machine with improved performance and high reliability can be obtained. In addition, by mounting such a positive displacement fluid machine in a refrigeration / air-conditioning cycle, a highly reliable refrigeration / air-conditioning system with excellent energy efficiency 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 perspective view of a positive displacement compression element according to the present invention.

FIG. 4 is a plan view of a compression element of a hermetic compressor in which the positive displacement fluid machine according to the present invention is applied to a compressor.

FIG. 5 is a plan view of a compression element of a hermetic compressor in which the positive displacement fluid machine according to the present invention is applied to a compressor.

FIG. 6 is a perspective view showing a method of assembling the positive displacement compression element according to the present invention.

FIG. 7 is a longitudinal sectional view of a positive displacement fluid machine according to the present invention.

FIG. 8 is a view showing a contour configuration method of a displacer of a positive displacement fluid machine according to the present invention.

FIG. 9 is a diagram showing a contour configuration method of a cylinder of the positive displacement fluid machine according to the present invention.

FIG. 10 is a diagram in which the displacer and the cylinder shown in FIGS. 8 and 9 are overlapped.

FIG. 11 is a diagram illustrating loads and moments 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 plan view of a compression element of a hermetic compressor in which the positive displacement fluid machine according to the present invention is applied to a compressor.

FIG. 14 is a diagram showing a volume change characteristic of a working chamber in the present invention.

FIG. 15 is a graph showing a change in gas compression torque in the present invention.

FIG. 16 is a diagram showing a relationship between a shaft rotation angle and a working chamber in a four-row wrap.

FIG. 17 is a view showing a relationship between a shaft rotation angle and a working chamber in a three-row wrap.

FIG. 18 is an operation explanatory diagram when the winding angle of the compression element is larger than 360 °.

FIG. 19 is a view for explaining an enlargement of a winding angle of a compression element.

FIG. 20 is a modification of the positive displacement fluid machine shown in FIG. 1;

FIG. 21 is a longitudinal sectional view of a main part of a hermetic compressor according to another embodiment of the present invention.

FIG. 22 is a diagram showing an air conditioning system to which the positive displacement compressor of the present invention is applied.

FIG. 23 is a diagram showing a refrigeration system to which the positive displacement compressor of the present invention is applied.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Displacement type compression element 2 Electric element 2a Stator 2b Rotor 3 Airtight container 4 Cylinder 4a Inner peripheral wall 4b Vane 4c, 4d, 4e Hole 5 Displacer 5a Through hole 5b Oil supply groove 5c, 5d Outline 6 Rotating shaft 6a Crank Reference Signs List 7 Main bearing member 7a Suction port 7b Main bearing member 7c Screw hole 7d Equalizing hole 7e Screw hole 7f Screw hole 8 Sub bearing member 8a Discharge port 8b Discharge chamber 9 Discharge valve 10 Suction cover 11 Discharge cover 12 Lubricating oil Reference Signs List 13 suction pipe 14 discharge pipe 15 working chamber 16 seal member 17 displacement type compression element 18 displacer 19 cylinder 20 rotation shaft 21a suction port 22 displacement type compressor 23 outdoor heat exchanger 24 fan 25 expansion valve 26 indoor heat exchanger 27 fan 28 4-way valve 29 Outdoor unit 30 Indoor unit 31 Vessel 32 condenser fan 33 expansion valve 34 evaporator 35 evaporator fan 36 discharge passage 37 volts 37a volts 37b bolt 38 bolt

 ──────────────────────────────────────────────────の Continuing on the front page (72) Inventor Hiroaki Hata 800, Tomita, Ohira-machi, Ohira-cho, Shimotsuga-gun, Tochigi Prefecture Inside the Refrigeration Division, Hitachi, Ltd. (72) Inventor Kenji Tojo 390 Muramatsu, Shimizu-shi, Shizuoka Prefecture In-house Air Conditioning Systems Division, Hitachi, Ltd. (72) Inventor Shigeru Machida 502, Kandachicho, Tsuchiura-shi, Ibaraki Machinery Research, Hitachi, Ltd. Inside

Claims (5)

[Claims] 1. A displacer and a cylinder are arranged between end plates, and when the center of the displacer is aligned with the center of rotation of a rotating shaft, one space is formed by the inner wall surface of the cylinder and the outer wall surface of the displacer. A displacement type fluid machine in which a plurality of spaces are formed when a positional relationship with the cylinder is set at a turning position, wherein a suction port communicating with the plurality of spaces is formed in the end plate. 2. A displacer and a cylinder are arranged between end plates, and when the displacer center is aligned with a rotation center of a rotating shaft, one space is formed by the inner wall surface of the cylinder and the outer wall surface of the displacer. When the positional relationship between the cylinder and the cylinder is at the swirling position, in the displacement type fluid machine in which a plurality of spaces are formed, the closing of two places formed by the swirling motion of the cylinder and the displacer arranged at the swirling position is completed. A positive displacement fluid machine with a suction port formed at the point connecting the points. 3. A displacer and a cylinder are disposed between end plates, and when the displacer center is aligned with the center of rotation of a rotating shaft, one space is formed by the inner wall surface of the cylinder and the outer wall surface of the displacer. In a displacement type fluid machine in which a plurality of spaces are formed when a positional relationship with the cylinder is set at a turning position, an outer peripheral shape of a suction port is formed by a turning motion of the cylinder and the displacer disposed at the turning position. A positive displacement fluid machine constituted by a substantially contour curve of the cylinder connecting two closing end points, and having an inner peripheral shape formed by a substantially curved line having a smaller turning radius than the substantially contour curve of the cylinder. 4. The displacement type fluid machine according to claim 3, wherein the outer peripheral shape of the suction port is such that each of the contour curves of the displacer located at two of the closing end points and two of the closing end points. And the inner peripheral shape is smaller than the approximate curve of the cylinder connecting the respective approximate curves smaller than the contour curve of the displacer by a turning radius and the approximate contour curve of the cylinder connecting the two closing points. A positive displacement fluid machine composed of approximately curved lines. 5. A refrigeration and air-conditioning system using the positive displacement fluid machine according to claim 1 as a positive displacement compressor and comprising a condenser, an expansion valve, and an evaporator.