KR100401814B1 - Fluid compressor of rotary slant shaft type with multi - discharging system - Google Patents

Abstract

The present invention relates to a four-axis gas compressor having a multistage exhaust system.

According to the present invention, a cylinder head having six gas holes is integrally fixed to the drive shaft 10; and a gas guide member 20 that sucks gas from the outside and discharges the compressed gas to the outside; A case head member 30 having a suction port for supplying gas into the cylinder bore and three exhaust ports for discharging the compressed gas; and an outer surface of the cylinder head fixed to the inner side of the case head member and rotating; A valve plate member 50 having a gas intake valve groove and three gas exhaust valve grooves formed therein and six cylinder bores formed in a direction parallel to the drive shaft, integrally coupled with the cylinder head, and a piston being slidably inserted Cylinder block 60; And the cylinder block and the piston is connected, the swash plate member 80 for converting the rotational force transmitted from the drive shaft to a linear reciprocating motion to transmit to the piston; It consists of a cylinder bore to provide a bent axis gas compressor that can discharge the compressed gas in the interior.

Description

Slope type gas compressor with multi-stage exhaust system {FLUID COMPRESSOR OF ROTARY SLANT SHAFT TYPE WITH MULTI-DISCHARGING SYSTEM}

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a gas compressor, and more particularly, to a rotary slant shaft type gas compressor having a multi-stage exhaust system capable of selectively exhausting gas compressed in a cylinder according to the pressure of an exhaust passage. will be.

Compressors are mechanisms that force the medium to be compressed from the outside to increase the pressure and speed of the medium itself. A compressor is also called a fluid compressor because it targets a fluid regardless of the state of the medium to be compressed. Media that can be compressed by the compressor include air, gas such as nitrogen or oxygen, and liquid such as oil or refrigerant. The compressor described below may be used to compress a liquid such as oil, but is mainly described with a gas compressor used to compress a gas such as air.

What is widely known as a gas compressor is a reciprocating compressor which compresses a gas by using a piston for a simple reciprocating motion.

In general, a reciprocating compressor is composed of a cylinder in which a suction valve and an exhaust valve are installed in a head, such as an engine of an automobile, and a piston reciprocating linearly inside the cylinder. Such a reciprocating compressor performs a process of inhaling, compressing, and exhausting gas by opening and closing the intake valve and the exhaust valve according to the gas pressure in the cylinder when the piston makes a linear reciprocating motion in the cylinder. However, such a reciprocating compressor causes the intake valve and the exhaust valve installed in the cylinder head to face the piston head directly or with the piston head during the compression process of the gas. The collision of the valves primarily produces mechanical noise, and as the use time increases, the valve itself ruptures when the valve is bent or severely. In addition, the reciprocating compressor is provided with two valves in one cylinder, so the intake and exhaust of the gas occurs intermittently in the cylinder, so pulsation occurs when the gas is compressed. There is a problem that occurs.

In order to improve the noise problem of such a reciprocating compressor, it is usually equipped with a silencer for the intake and exhaust. However, when the silencer is installed in the reciprocating compressor, the compressor itself is mechanically complicated and the number of parts is increased. The silencer is installed to increase the gas resistance, causing another problem of degrading the performance of the compressor.

Another type of compressor capable of compressing gas is the bent axis compressor disclosed in Japanese Patent Laid-Open No. 61-65081 (published 1986,04,03).

65081 compressor relates to a device for transmitting a rotational force of the rotating shaft to the swash plate to convert the piston connected to the swash plate into a linear reciprocating motion. The compressor has a structure in which a cylinder block in which six cylinders are formed is fixed to a rotating shaft, and each cylinder formed in the cylinder block has an open face opposite to the piston. On the open cylinder side, a float valve having an intake and exhaust hole is blocked, and a compressor case head is in close contact with the back of the float valve. A rubber ring is interposed between the float valve and the case head to prevent leakage of compressed gas in each cylinder.

When the compressor rotates the drive shaft by the rotational force transmitted from the external power source, the cylinder block fixed to the drive shaft rotates together with the drive shaft, and the swash plate fixed to the end of the drive shaft rotates to rotate each piston sequentially in each cylinder. Have a straight reciprocating motion.

The feature of this compressor is that each cylinder rotates open, while the float valve and case head are stationary. Due to this feature of the compressor, each cylinder draws gas into the cylinder as it passes through the intake hole of the float valve, gradually compresses the gas while rotating, and compresses the gas formed in the case head as it passes the exhaust hole of the float valve. Will be discharged into the passageway. In this compression process, the float valve comes into close contact with the cylinder block due to the difference in force due to the gas pressure applied to the cylinder cross section and the valve cross section.

Comparing the 65081 compressor with the conventional reciprocating compressor as described above, the 65081 compressor has an advantage that the shape of the compressor can be made very compact because the piston reciprocates in a direction parallel to the direction of the drive shaft. In addition, the 65081 compressor does not use reciprocating intake and exhaust valves, but uses a fixed float valve so that the valve does not naturally generate mechanical noise due to direct collision with the piston head. In addition to these advantages, the 65081 compressor exhibits the same compression efficiency as conventional reciprocating compressors when continuously operated at rated load, and noise characteristics due to gas pressure difference are almost the same.

However, in spite of the advantages of the 65081 compressor as described above, this compressor has a fatal disadvantage that the cylinder block and the float valve must rub against each other in order to maintain airtightness between the rotating cylinder block and the stationary float valve. Thus, the compressor wears out due to the constant friction between the cylinder block and the float valve. In order to effectively remove the frictional heat generated by the friction, the target gas itself to be compressed must have lubricity. Therefore, the gas used in this compressor is inevitably limited to lubricious gas. In addition to the frictional heat, the compressor has a very high heat of compression generated during the compression of the gas, and thus a device for dissipating or absorbing heat inside or outside the compressor is required. However, since the 65081 compressor does not provide such a heat removing means, the durability of the compressor decreases due to various heat generated when the compressor is actually used, and there is a problem that the compression efficiency of the gas decreases as it is used.

On the other hand, when considering the structure of the compressor 65081 compressor aerodynamically, the 65081 compressor is characterized in that the pressure difference between the maximum pressure (Pm) of the compression section and the exhaust pressure (Pd) of the exhaust section is very large. As the pressure difference between the two zones increases, the aerodynamic noise generated when the high-pressure compressor body is ejected at a lower pressure becomes larger. In view of this point, when comparing the 65081 compressor with a conventional reciprocating compressor, the 65081 compressor exhibits a larger aerodynamic noise since this pressure difference is larger than that of the conventional reciprocating compressor.

In addition, when considering the compression load in the cylinder generated during operation of the 65081 compressor, the 65081 compressor exhibits a much greater variation per unit time of the compression load than the conventional reciprocating compressor. As described above, the larger the fluctuation range of the compression load in the cylinder, the greater the axial load on the drive shaft. Therefore, 65081 compressor has an axial force load proportional to the compression load is applied to the swash plate fixed to the end of the drive shaft, thereby directly affecting the ball bearing parts installed between the lower plate and the case, which reduces the durability of the compressor itself. It acts as a factor.

As described above, although the 65081 compressor has many advantages over the conventional reciprocating compressor, there is a limitation in commercializing it as a gas compressor due to problems caused by the structure of the 65081 compressor.

Therefore, in order to maintain the basic characteristics of the bent gas compressor and to improve the problems of the 65081 compressor, the new bent axis gas compressor minimizes aerodynamic noise, improves the durability of the parts, increases the energy efficiency and increases the number of parts. It is necessary to develop a structure for minimizing the amount of gas, and to develop a structure for enabling a no-load operation and to diversify the compressible gas.

The present invention has been made in order to solve the problems of the conventional reciprocating compressor and the bent axis gas compressor, the object of which is to discharge the compressed gas in the cylinder bore at a time without a temporary discharge axis that can be selectively discharged The present invention provides a gas compressor.

Still another object of the present invention is to provide a four-axis gas compressor that does not generate mechanical or aerodynamic noise by designing the structure of the gas compressor to minimize noise sources aerodynamically.

Still another object of the present invention is to provide a bent axis gas compressor that can maximize energy efficiency by minimizing power required for gas compression.

It is still another object of the present invention to provide a bent-type gas compressor having improved durability by minimizing a variation in compression load per unit time applied to the gas compressor.

It is a further object of the present invention to provide a bent gas compressor which can circulate gas sucked into each cylinder bore into the crank chamber and then suck it into the cylinder bore.

Still another object of the present invention is to provide a bent axis gas compressor capable of high load operation with high efficiency.

Still another object of the present invention is to provide a bent gas compressor which can effectively release frictional heat generated inside the compressor and compressed heat generated by air compression.

1 is a cross-sectional view showing a bent axis gas compressor according to the present invention.

2 is a perspective view showing a gas guide member of the gas compressor according to the present invention.

3A is a perspective view showing a case head of a gas compressor according to the present invention;

B is sectional drawing which shows the state cut | disconnected by the I-I line in FIG. 3A,

C is sectional drawing which shows the state cut | disconnected by the II-II line in FIG. 3A.

Figure 4 is a perspective view showing a tension ring of the gas compressor according to the present invention.

5 is a perspective view showing a valve plate of the gas compressor according to the present invention.

6 is a cross-sectional view showing another embodiment of the gas compressor according to the present invention;

B is sectional drawing which shows the spring coupling cut | disconnected by the III-III line | wire in FIG.

C is sectional drawing which shows another embodiment of the spring coupling shown in FIG. 6A.

7 is a cross-sectional view showing yet another embodiment of a gas compressor according to the present invention.

8 is a cross-sectional view of FIG. 1 illustrating a gas suction process and an exhaust process of the gas compressor according to the present invention.

9A is an explanatory diagram illustrating a process of inhaling and compressing exhaust gas in the valve plate of the gas compressor according to the present invention;

B is explanatory drawing which shows the compression characteristic of gas at the time of one rotation of a cylinder in the valve plate of a gas compressor by this invention.

10A is an explanatory diagram showing the compression characteristics of a gas which appears during operation of the gas compressor according to the present invention,

B is explanatory drawing which shows the compression characteristic of the gas shown at the time of operation | movement of the conventional reciprocating compressor,

C is explanatory drawing which shows the compression characteristic of the gas shown at the time of operation | movement of the conventional bent axis gas compressor.

In order to achieve the object of the present invention as described above, the present invention is fixed to the case head and the valve plate which is in close contact with the rotating cylinder head and a plurality of suction grooves and a plurality of exhaust grooves formed in the compressed gas inside the cylinder bore Provided is a bent axis gas compressor that can be selectively discharged.

Referring to the compressor of the present invention in more detail, the bent axis gas compressor of the present invention includes a drive shaft in which a plurality of gas holes are formed on a concentric circle at regular intervals and integrally formed in a direction perpendicular to the direction of the drive shaft; A gas guide member having an intake manifold for sucking and an exhaust manifold for discharging the compressed gas in the cylinder bore to the outside; and rotatably supporting the drive shaft and supplying the gas sucked from the intake manifold into the cylinder bore; A case head member having at least two exhaust ports configured to discharge at least one suction port and compressed gas inside the cylinder bore to the exhaust manifold; and a cylinder head fixed and rotated to an inner surface of the case head member. It adheres to the outer side and the gas hole in the cylinder head A gas intake valve groove and at least two gas exhaust valve grooves are formed on the circumference, and the gas intake valve groove supplies the gas introduced from the suction port into the cylinder bore, and the gas exhaust valve groove is compressed inside the cylinder bore. A valve plate member for discharging the gas to the exhaust port; and a plurality of cylinder bores are formed in a direction parallel to the driving shaft, and one side is integrally coupled with the cylinder head, and the other side is inside each cylinder bore. A piston is slidably inserted, the cylinder block compressing the gas sucked into each cylinder bore; And a swash plate member connected to the central portion of the cylinder block with a bent axis, connected to the plurality of pistons and the piston rod, and converting the rotational force transmitted from the drive shaft into a linear reciprocating motion to be transmitted to the piston; and supporting the swash plate member. A case end plate having an inclined slope formed therein; and a case coupled to the case head and the case end plate to house the cylinder block and the swash plate member; Consists of

In the bent-type gas compressor of the present invention, each exhaust port of the case head member includes a check valve to selectively discharge the compressed gas discharged through each of the exhaust grooves of the valve plate member according to the internal pressure of the compression tank. .

In the bent axis gas compressor of the present invention, a circulation circuit is formed in the case head member and the drive shaft to introduce the gas introduced from the intake manifold into the cylinder bore via the sealed crank chamber inside the case. This can be minimized by limiting the aerodynamic noise generated when the compressed air remaining inside the cylinder bore joins the intake gas inside the case.

In addition, the bent-type gas compressor of the present invention inserts a tension ring and a ring-shaped leaf spring between the inner surface of the case head member and the valve plate member, so that the compressor may be used for a long time, resulting in a gap between the valve plate and the cylinder head. Do not occur at the source.

And the bent axis gas compressor of the present invention by connecting the rotating cylinder block and the swash plate member by a universal coupling or a spring coupling to minimize the mechanical noise generated during the operation of the compressor.

Hereinafter, exemplary embodiments of the present invention and modifications thereof will be described in detail with reference to the accompanying drawings.

As shown in FIG. 1, important parts of the bent axis gas compressor according to the present invention are housed inside the cylindrical case 1. Both sides of the case 1 are bolted by the case head 30 and the case end plate 3, and rubber rings 4 are respectively provided on the mating surfaces of the case parts 1, 30, and 3. The crank chamber 70 inside the case is sealed from the outside.

The gas compressor housed inside the case 1 includes a drive shaft 10 for transmitting a rotational force supplied from an external power source to the swash plate, a gas guide member 20 for sucking and discharging gas from the outside, and a cylinder for sucking the gas. A case head member 30 which supplies internally and selectively exhausts the compressed gas according to the pressure of the compression tank, and a valve which supplies gas into the rotating cylinder fixed to the inner side of the case head member and exhausts the compressed gas The plate member 50, the cylinder block 60 is built in a plurality of pistons to compress the gas, and a swivel plate member (80) for converting the rotational force of the drive shaft 10 into a linear reciprocating motion.

The drive shaft 10 extends in parallel with the central axis of the case 1 and is rotatably fixed to the boss portion of the case head 30 by the ball bearing 11 and the tapered roller bearing 12. One end of the drive shaft 10 is fixed to the drive pulley 5 for transmitting the rotational force generated from an external power source (not shown) to the drive shaft, the other end for sealing the cylinder bore 61 of the cylinder block 60 The cylinder head 13 is integrally formed. The cylinder head 13 formed integrally with the drive shaft is formed in a disc shape, and six gas holes 14 are formed at regular intervals in the cylinder head 13 at positions concentric with the drive shaft. The shaft chamber 15 is partially drilled inside the drive shaft 10, and the shaft port 16 is drilled in a direction perpendicular to the axial direction, so that the gas introduced into the crank chamber 70 is supplied to the cylinder bore 61. It acts as a passageway. At the end of the shaft shaft 15 of the drive shaft 10, a liquid inflow prevention jaw 17 is formed, whereby scattered lubricating oil that can flow through the side surface of the block chamber 63 of the rotating cylinder block includes a cylinder bore ( 61) Prevent from entering inside.

As shown in Figs. 1 and 2, the gas guide member 20 has a cylindrical shape, and an intake pipe 21 and an exhaust pipe 22 are fixed on a cylinder. Here, the intake pipe 21 serves to inject gas to be compressed into the compressor, so a filter (not shown) is attached to the outside of the intake pipe, and the exhaust pipe 22 is a compression tank (not shown) storing the compressed gas from the compressor. It communicates with the compression tank because it plays a role of discharging. In addition, a check valve 27 is built in the exhaust pipe 22 to prevent the gas compressed in the compression tank from flowing back into the cylinder. On the bottom of the gas guide member 20, an intake manifold 23 communicating with the intake pipe 21 and an exhaust manifold 24 communicating with the exhaust pipe 22 are dug on the same circumference. In addition, the auxiliary intake pipe 25 and the auxiliary exhaust pipe 26 are attached to the side surfaces of the intake pipe 21 and the exhaust pipe 22, so that the gas compressor operates when it is not necessary to compress the gas during the operation of the gas compressor. The auxiliary intake pipe 25 and the auxiliary exhaust pipe 26 are interconnected without stopping the compressor so that the load of the compressor can be kept to a minimum. Proceeding the suction compression stroke in this state is called no-load operation. As such, in the no-load operation state, the pressure compressed in the compression tank does not flow back toward the exhaust pipe 22 due to the check valve 27 built in the exhaust pipe. Reference numeral 28, which is not described in FIG. 2, is used to fix the gas guide member 20 to the case head 30 described later by using bolts as bolt holes.

As shown in FIGS. 1 and 3, the case head member 30 is provided with a guide groove 32 for fixing the heat dissipation fin 31 and the gas guide member on the outer side thereof, and the cylinder head 13 on the inner side thereof. Insertion groove (33) for installing the thrust bearing (18) for maintaining a constant distance from the) and plate groove (34) to which the valve plate 50 is fixed, and the tension ring (40) for close contact with the valve plate is inserted A ring groove 35 for the purpose is formed. In addition, a boss 36 for accommodating the ball bearing 11 and the tapered roller bearing 12 for supporting the drive shaft 10 is formed at the center of the case head member 30. Here, the heat dissipation fan 31 formed on the outer surface of the case head member 30 is formed in a radial shape with respect to the driving shaft so that cooling air by an external air fan (not shown) flows smoothly to the outside of the case 1. In order to enhance the cooling effect, the heat dissipation fan 9 outside the case 1 is formed in a direction parallel to the drive shaft. At this time, the air fan for air cooling is preferably installed on the drive pulley (5) side.

The case head member 30 circulates the gas supplied from the gas guide member 20 into the crank chamber 70 to be sucked into the cylinder bore 61, and serves as a passage for selectively discharging the gas compressed in the cylinder. do. Therefore, the gas flow passage formed in the case head member 30 plays an important role for achieving the object of the present invention. The gas movement passage formed in the case head member 30 is mainly formed between the guide groove 32 and the plate groove 34 and is divided into a suction passage and an exhaust passage.

As shown in FIG. 3, the gas suction passage includes a first suction port 37a and a second suction port 38a. One side of the first suction port 37a (FIG. 3B) communicates with the guide groove 32 and is connected to the intake manifold 23 of the gas guide member, and the other side of the case head member is connected by the side port 37b. It is connected with the side of 30. Here, the side port 37b is formed below the installation groove 33 on which the trust bearing 18 is seated, so that the gas introduced does not pass through the thrust bearing 18. One side of the second suction port 38a (FIG. 3C) communicates with the plate groove 34 toward the cylinder bore 61, and the other side communicates with the drive shaft 10 by the side port 38b. It faces the shaft port 16 side of a drive shaft. As shown in FIG. 8, the gas supplied from the intake manifold 23 of the gas guide member is introduced into the first suction port 37a and the side port 37b by the gas suction passage formed as described above, and thus the crank chamber Via 70, the block chamber 63 and the shaft chamber 15, it is sucked into the cylinder bore 61 through the side port 38b and the second suction port 38a.

As shown in FIG. 3A, the gas exhaust passage consists of a primary exhaust port 41, a secondary exhaust port 42, and tertiary exhaust ports 43a and 43b, and these primary, secondary and tertiary exhausts The ports 41, 42, 43a and 43b pass through the case head member 30 and are connected to the exhaust manifold 24 of the gas guide member. In each of the exhaust ports 41, 42, 43a and 43b, a check valve 46 is incorporated to prevent backflow of compressed gas from the exhaust manifold 24 into the cylinder bore 61.

In FIG. 3A, reference numeral 44 denotes a drain port for discharging lubricating oil sucked into the shaft chamber 15 to the crank chamber, and 45 denotes a bolt hole for fixing the valve plate 50 to the case head member 30. .

4 shows the tension ring 40 inserted into the ring groove 35 of the case head member. The ring groove 35 is provided with a disc ring-shaped leaf spring 49 (FIG. 1), so that the leaf spring 49 exerts an elastic force on the tension ring 40, and the tension ring 40 is a valve plate (to be described later). 50 is pressed against the cylinder head 13. The tension ring 40 inserted into the ring groove 35 prevents the gas compressed during the gas compression stroke from leaking to the outside and prevents foreign substances such as cooling oil from flowing into the cylinder bore 61. In addition, the tension ring 40 is divided into the intake section 40a and the first and second exhaust sections 40b, 40c, and 40d so that the gas compressed during the gas compression stroke leaks from one section to another. prevent. The tension ring 40 is made of a material having elasticity and heat resistance, such as heat resistant rubber or urethane.

As shown in FIGS. 1 and 5, the valve plate member 50 is formed in a disc ring shape, and as shown in FIG. 5, the upper surface of the valve plate member 50 closely adheres to the outer surface of the cylinder head 13. In this plane, the cylinder head rotates and slides. On the upper surface of the valve plate member 50, one intake valve groove 51 and three isolated primary, secondary and tertiary exhaust valve grooves 52, 53 and 54 are excavated in an arc shape. Here, the groove widths of the suction valve groove 51 and the tertiary exhaust valve groove 54 are formed to be equal to or larger than the diameter of the gas hole 14 formed in the cylinder head, but the primary and secondary exhaust valve grooves 52 and 53 are formed. Is formed smaller than the diameter of the gas hole (14).

The position of the valve grooves 51, 52, 53, and 54 formed in the valve plate member 50 will be described in more detail with reference to FIGS. 5 and 9A. The suction valve groove 51 has a piston 64 having a top dead center. It is formed in the circumferential 180 ° section corresponding to the suction stroke section moving from (T) to the bottom dead center (B), the three exhaust valve grooves (52, 53, 54) is the piston 64 is the bottom dead center (B) It is formed in the remaining 180 ° section of the circumference corresponding to the compression stroke section moving from the top dead center (T). The radius (V _R ) of the circumference connecting the center lines of each of the valve grooves 51 (52, 53, 54) is the same as the radius (H _R ) of the circumference connecting the center lines of the six gas holes 14 formed in the cylinder head. Therefore, each gas hole 14 passes through each of the valve grooves 51, 52, 53, and 54 sequentially. In forming the valve grooves 51, 52, 53 and 54, an important point is that the distance between each valve groove (the length V _L of the compartment wall) must be spaced apart from the diameter of the gas hole 14 of the cylinder head. Another important point is that the length of each of the primary, secondary and tertiary exhaust valve grooves 52, 53, 54 is shorter than the distance between the gas holes 14, so that one exhaust valve groove 52, 53, 54 is provided. At least one gas hole 14 should not be located in the chamber.

In the suction and primary and secondary exhaust valve grooves 51 and 52 and 53, valve holes 51a and 52a and 53a penetrating the valve plate 50 are formed one by one, and the third exhaust valve groove ( In the 54, two valve holes 54a and 54b are formed. The suction valve hole 51a penetrating through the valve plate has a second suction port 38a at the bottom of the case head, and the primary exhaust valve hole 52a has a primary exhaust port 41 and a secondary exhaust valve hole 53a. The secondary exhaust port 42, the tertiary exhaust valve holes 54a, 54b are connected to the tertiary exhaust ports 43a, 43b, respectively. Accordingly, the gas introduced into the suction valve groove 51 of the valve plate from the second suction port 38a of the case head is sucked into each cylinder bore 61 through each of the gas holes 14 of the cylinder head to pivot, and the cylinder The block 60 is gradually compressed as it rotates, and then selectively discharged to the valve grooves 52, 53, and 54 of the valve plate where the gas holes 14 meet, depending on the pressure of the compression tank. Compressors introduced into the exhaust valve grooves 52, 53, and 54 are discharged through the exhaust valve holes 52a, 53a, 54a, and 54b and the exhaust ports 41, 42, 43a, and 43b. Join the exhaust manifold 24 formed in the guide.

As shown in FIG. 1, the cylinder block 60 is formed in a cylindrical shape as a whole, and the center of the cylinder axis is drilled to form the block chamber 63, and the same diameter is formed on the same circumference of the cylinder of the block chamber 63. Six cylinder bores 61 having are drilled in a direction parallel to the drive shaft. One end of the cylinder block 60 is bolted to the cylinder head 13 of the drive shaft and sealed, and six pistons 64 are slidably inserted into each cylinder bore 61. In addition, a coupling 65 for coupling the universal coupling 66 to be described later is fixed to the block chamber 63 with bolts. Here, the heat dissipation fins 67 are formed in a spiral shape on the outer circumferential surface of the cylinder block 60, and this heat dissipation fins accelerates the circulation of gas flowing in the inner space of the crank chamber 70 when the cylinder block 60 is rotated. The heat radiation fins 67 may be formed of a plurality of circular fins having a constant spacing in addition to the spiral. In addition, the block chamber 63 is drilled in a two-stage cylinder having a different diameter to prevent the lubricated oil from flowing to the shaft chamber 15.

The cylinder block 60 is connected to the swash plate member 80, which will be described later, and the universal coupling 66 along the rotation axis center line of each component. The universal coupling 66 is composed of a drive joint 68 and a driven joint 69, the drive joint 68 and the driven joint 69 being interconnected by cross shafts 71 at respective fork arms. do. The drive joint 68 of the universal coupling has a hollow cylindrical shape and has splines formed on its outer circumference so that the drive joint 68 is splined with the coupling 65 fixed to the cylinder block 60, and the driven joint 69 is flanged at one end. Is formed is coupled to the swash plate member 80 with a bolt.

The swash plate member 80 is axially coupled by a tapered roller bearing 81 to a driven shaft 7 whose central axis of rotation is fixed to the inclined surface 6 of the case end plate 3. Six pairs of brackets 82 are formed on the outer circumference of the upper surface so that one end of the piston rod 73 can be coupled, and the lower surface has a jaw so that the thrust bearing 83 can be installed between the inclined surface 6. Is formed.

Six pistons 64 respectively inserted in the cylinder bore 61 are rotatably fastened to the bracket 82 of the swash plate member by respective piston rods 73. The piston rod 73 fastened between each piston 64 and the swash plate member 80 may selectively use a universal coupling or a two-section crank. The piston rod 73 in the form of a two-section crank consists of a first rod 74 and a second rod 76. One side of the first rod 74 is connected to the piston 64 by the connecting pin 75 and the fork arm is formed on the other side. One side of the second rod 76 is connected to the bracket 82 of the swash plate by the connecting pin 77 and the other side is formed with a coupling ring. The coupling of the first rod 74 and the second rod 76 is rotatably fixed to the connecting pin 78 by the coupling ring of the second rod 76 is inserted into the fork arm of the first rod 74. And a bearing is coupled to the outside of the connecting pin to the portion where each connecting pin of the two-section crank piston rod 73 is fixed. The piston rod 73 of the universal coupling type, which is not shown in the figure, is similar to the piston rod of the two-crank crank type coupling structure, and both rods are rotatably fixed by the cross shaft.

6 and 7 show another embodiment of the present invention.

In another embodiment of the present invention, the functions of the main components for determining the driving mechanism of the present invention are the same, except that the structure of the gas suction passage and the connection structure connecting the cylinder block 60 and the swash plate member 80 and the gas compressor. The cooling structure for the cooling of the various changes.

Another embodiment of the present invention shown in FIG. 6 is basically the same as the embodiment shown in FIG. 1 and the difference is a connection structure connecting the cylinder block 60 and the swash plate member 80. As shown in FIG. 6A, the piston rod 73 connecting each piston 64 and the swash plate member 80 is in the form of a two-section crank, which is the same as the embodiment shown in FIG. However, the cylinder block 60 and the swash plate member 80 are connected by a spring coupling 72 rather than a universal coupling. Therefore, the power of the drive shaft 10 is transmitted to the swash plate member 80 by the piston rod 73 rather than the universal coupling.

As shown in FIG. 6, the spring coupling 72 connects the cylinder block 60 and the swash plate member 80 in parallel with two springs 72a and 72b. The cylinder block 60 is formed with two block rings 68a and 68b in the diagonal direction, and the swash plate member 80 is also formed with the swash plate rings 69a and 69b in the diagonal direction, so that the first spring 72a is formed. ) Is fastened to the first block ring (68a) and the first swash plate ring (69a), the second spring (72b) is fastened to the second block ring (86b) and the second swash plate ring (69b). By the fastening structure of the spring coupling 72, the cylinder block 60 and the swash plate member 80 are mutually in the same direction as the rotation direction of the cylinder block 60 and the piston rod 73 (arrow direction in Fig. 6B). Compressive force will act. This compression force cancels the reaction force generated by the swash plate member 80 when the cylinder block 60 starts to rotate. Here, the springs 72a and 72b of the spring coupling 72 have their elastic modulus determined in consideration of the reaction force of the swash plate member 80. And the block ring (68a, 68b) and the swash plate (69a, 69b) is a holder part to which the fixing pin and the spring is connected to each ring is fixed in consideration of the fact that the swash plate member (80) is a swing movement when the gas compressor operates Are connected by ball joints.

And another embodiment of the spring coupling 72 is shown in Fig. 6C. The spring coupling 72 shown in FIG. 6C has both ends of one cylindrical spring 72c fastened to the block flange 68c and the swash plate flange 69c, and the block flange 68c and the swash plate flange 69c are Bolted to the cylinder block 60 and the swash plate member 80, respectively. When the cylindrical spring (72c) is fastened here, the cylinder block (60) and the piston rod (73) are fastened in a state of twisting by a predetermined amount in the same direction as the rotational direction. Accordingly, the reaction force generated by the swash plate member 80 is canceled when the cylinder block 60 starts to rotate by the twisting stress of the spring 72c.

Another embodiment of the present invention shown in Figure 7 compared to the embodiment shown in Figure 1 the structure of the gas suction passage and the connecting structure for connecting the cylinder block 60 and the swash plate member 80 and the cooling of the gas compressor The cooling structure for this is changed. Hereinafter, the main components of the embodiment thus modified will be described in order.

7 is different from the embodiment shown in FIG. 1. First, a through hole 15a through which the inside of the drive shaft 10 penetrates completely in the axial direction is formed. As such, when the through hole 15a is formed in the driving shaft, the shaft port 16 as shown in FIG. 1 is not formed. The through hole 15a serves as a passage through which the lubricating oil mist (oil vapor) generated in the crank chamber 70 is discharged to the outside of the compressor when the gas compressor operates in the gas compressor using the lubricating oil. Air flows into the crank chamber 70 serves as a cooling tube.

The second difference is that the suction port 39 of the case head member 30 penetrates between the guide groove 32 and the plate groove 34. When comparing the embodiment of FIG. 1 with the embodiment of FIG. 7, in the embodiment of FIG. 7, the first suction port 37, the second suction port 38, The side ports 37b and 38b and the separate drain port 44 are not drilled. The direct inlet of the gas suction port 39 is to supply the gas introduced from the gas guide member 20 directly into the cylinder bore 61 without circulating the gas into the crank chamber 70.

The third difference is that the cylinder block 60 and the swash plate member 80 are biaxially coupled to the bevel gear 86 to improve assembly.

The fourth difference is that the piston 64 and the swash plate member 80 are connected to the piston rod 73 in the form of a two-strand extension rod. This is because the piston rod is divided into two and manufactured in the form of male and female bolts, and then finely adjusts the stroke gap of the piston when assembling the piston using the lock nut 84. When the piston rod 73 is used in the form of an elastic rod as shown in Fig. 7, a leaf spring 85 is provided on the outer side of each piston rod 73 and the outer circumferential surface of the swash plate member 80. The leaf spring 85 as described above serves to cancel the centrifugal force applied to the piston 64 when the cylinder block 60 rotates together with the swash plate member 80.

When using the form of the piston rod 73 in the form of a flexible rod, as shown in Figure 7, it is preferable to use a compressor by injecting lubricant to the bottom of the crank chamber 70. When the compressor is operated by injecting lubricating oil into the bottom of the crank chamber 70 as described above, the lubricating oil is scattered into the chamber during the operation of the compressor and is generated between the ball 87 and the ball joint 88 of the flexible rod. The frictional heat can be cooled.

The fifth difference is that the cooling case 6 surrounding the case 1 is attached to the outside of the cylindrical case 1. Therefore, a cooling chamber 8 is formed between the case 1 and the cooling case 6. The cooling case 6 has a cooling water inlet 6a and a cooling water discharge port 6b formed therein, and the heat dissipation fan 9a formed outside the case 1 is formed in a spiral shape and circulated in the cooling chamber 8. The coolant is discharged to the outlet 6b after circulating the outer circumferential surface of the case 1. The heat dissipation fins 9 and 9a are manufactured in a spiral form 9a as shown in FIG. 7 in the case of water cooling, and in a direction parallel to the drive shaft 7 as shown in FIG.

Finally, in the embodiment of FIG. 7, a plurality of blades 89 are formed on the outer circumferential surface of the swash plate member 80. Such a blade 89 serves to scatter the lubricating oil accumulated on the bottom of the crank chamber 70 in the compressor using the lubricating oil into the chamber.

Although the embodiments of the present invention and its modified examples have been described with respect to the components necessary for the operation of the present invention and the coupling relationship thereof, some of the components, which are not described, for example, case sealing parts and rolling ball piston rods are conventional The detailed description thereof is omitted since it is the same as the mechanical parts generally used for the phosphorus machine.

In addition, the shape of the shaft coupling connecting the cylinder block 60 and the swash plate member 80, the piston rod 73, and the cooling structure described in the embodiments of the present invention and its modifications are not dedicated to one embodiment. According to the intended use of the gas compressor can be used in combination selectively.

Hereinafter, the operation and operation characteristics of the gas compressor according to the embodiment of the present invention and its modification will be described with reference to FIGS. 8 to 10.

As shown in FIG. 8, the rotational force generated from an external power source such as a motor (not shown) is transmitted to the pulley 5 through a power transmission means such as a belt (not shown). When the rotational force transmitted to the pulley 5 rotates the drive shaft 10, the cylinder head 13 and the cylinder block 60, which are integrally coupled with the drive shaft, rotate together with the drive shaft 10. At the same time, the swash plate member 80 coupled with the cylinder block 60 and the universal coupling 73 (or the spring coupling of FIG. 6 or the bevel gear of FIG. 7) is rotated about the driven shaft 7. When the swash plate 80 rotates while swinging in a direction inclined to the drive shaft, each piston rod 73 coupled to the swash plate member 80 is linearly reciprocated in the drive shaft direction.

6, the power of the drive shaft 10 is transmitted to the swash plate member 80 by the piston rod 73 instead of the spring coupling 72. At this time, since the compressive or twisting stress is applied to each of the springs 72a, 72b, and 72c of the spring coupling 72, the swash plate member 80 when the cylinder block 60 starts to rotate due to such compressive or twisting stress. This counteracts the reaction forces generated by).

The movement of the compressor parts as described above is made at the same time as the power is input, the six pistons 64 are rotated together with the cylinder block 60 and at the same time reciprocating movement, respectively, to selectively perform the compression exhaust stroke. At this time, the stroke distance of the piston 64 is the same as the distance moved between the swash plate member 80 and the piston rod 73 in the drive shaft direction during the swing movement of the swash plate, it can be represented by 2Rsin (K0). (Here; R: distance from the center of the driven shaft 7 to the connection point of the swash plate 80 and the piston rod 73, K0: inclination angle of the drive shaft 10 and the driven shaft 7).

On the other hand, the flow of gas in the compressor during the operation of the compressor of the present invention is performed as follows.

First, when gas passing through an external filter (not shown) is introduced through the inlet pipe 21 of the gas guide, in the embodiments of FIGS. 1 and 6, the crank chamber 70 is circulated to enter the cylinder bore 61. 7 is directly introduced into the cylinder bore 61 in the embodiment of FIG. Referring to the gas circulation path in the case where the gas is circulated inflow as in the embodiment of Figure 1 as follows. That is, the gas introduced into the intake pipe 21 of the gas guide is the first suction port 37a, the side port 37b of the case head, the crank chamber 70, the block chamber 63, the shaft chamber 15, It passes through the shaft port 16, the side port 38b, the suction valve hole 51a of the valve plate, and the gas hole 14 of the cylinder head, and flows into the cylinder bore 61 in order. The purpose of circulating and introducing the sucked gas into the crank chamber 70 without directly entering the inside of the cylinder bore 61 is a residual pressure that may remain inside the cylinder bore 61 immediately after the compression exhaust stroke is completed. Is buffered by the low-pressure gas introduced into the crank chamber 70 to cause the explosion occurring at the moment when different gases are mixed inside the crank chamber 70, which is an enclosed space, resulting from the confluence of two gases having different pressures. This is to reduce noise.

Next, the gas introduced into the cylinder bore 61 is compressed while the cylinder block 60 and the piston 64 are rotated, so that the gas hole 14 of the cylinder head is the primary exhaust groove 52 of the valve plate, At the moments of sequentially meeting the secondary exhaust grooves 53 and the tertiary exhaust grooves 54, the respective exhaust holes 52a, 53a, 54a and 54b of the valve plate and the exhaust ports 41 of the case head ( 42) (43a, 43b) is selectively discharged according to the pressure of the compression tank, and then joined in the exhaust manifold 24 of the gas guide member is discharged to the exhaust pipe (22).

On the other hand, the frictional heat generated as the components rub against each other during the operation of the compressor of the present invention is cooled by the following method.

When operating using the lubricant in the embodiment of Figure 7, the compressor is operated in a state where the lubricant is filled to the level of the blade 84 of the swash plate member to the bottom of the crank chamber 70. In the case of such a compressor, the blade 84 of the rotating swash plate scatters the lubricating oil to the inner wall of the crank chamber 70, and at the same time, the lubricating oil of the radiating heat 67 of the rotating cylinder block is stirred. Therefore, the lubricated oil is lubricated to each of the moving parts while cooling the frictional heat generated by the mutual friction of the parts. At this time, when a part of the lubricating oil is evaporated to the oil vapor state, such oil vapor is discharged to the outside of the compressor through the hollow portion 15a of the drive shaft. The heat dissipated toward the block chamber 63 among the compressed heat generated in the cylinder bore 61 is also cooled by the vortex of gas formed inside the crank chamber 70.

When operating without lubricating oil, the compressor is combined with a power transmission structure that minimizes friction. In the case of such a compressor, since the gas flowing into the crank chamber 70 circulates while forming a vortex in the chamber, the circulating gas itself serves as a cooling source.

Although the above description has been described with respect to the cooling action inside the crank chamber 70, the present invention cools the outside of the case (1) by the air cooling or water cooling method together with the compressor internal cooling. Air cooling is performed by the heat dissipation fan 31 of the case head formed in a radial shape with the drive shaft outside the case and the heat dissipation fan 9 of the case formed in a direction parallel to the drive shaft. That is, when an air fan is installed on a motor (not shown) located in front of the drive shaft 10 to cause wind toward the case 7, the compressor flows along the heat dissipation fans 31 and 9 outside the case to cool the outside of the compressor. . In the water cooling type, when the refrigerant is supplied to the intake opening 6a between the cooling chamber 8 formed between the case 1 and the cooling case 6, the outer circumferential surface of the case 1 is circulated along the spiral heat dissipation fan 9a, and then the outlet The outside of the compressor is cooled while being discharged to 6b.

The centrifugal force acts on the piston 64 which rotates together with the cylinder block 60 in a direction in which the radius increases with respect to the drive shaft 10. In order to offset the centrifugal force applied to such a piston, as shown in FIG. 7, in the embodiment of the present invention, a plate spring 87 is installed on the piston rod 73 so that the centrifugal force applied to the piston 64 during the movement is prevented. Offset.

The following describes in more detail the process of the compression and exhaust stroke in the compressor according to an embodiment of the present invention. 9 shows the moment when the valve grooves 51, 52, 53 and 54 of the valve plate 50 and the gas hole 14 of the cylinder head 13 meet each other during the reciprocation of the piston 64. Suction and compressed exhaust characteristics.

In FIG. 9A, symbol T denotes a point where the piston 64 is located at the top dead center, and symbol B denotes a point where the piston 64 is located at the bottom dead center. The symbol K means the angle at which the gas hole 14 of the cylinder head moves the valve plate 50. In FIG. 9A, when the gas hole 14 rotates counterclockwise along the valve plate 50, a section of K = 0 ° to 180 ° corresponds to a suction stroke section of the piston, and K = 180 ° to 360 °. The section of corresponds to the compressed exhaust stroke section of the piston.

At the top dead center T, the section S1 until the gas hole 14 rotates to meet the intake valve groove 51 and immediately before K1 expands some of the remaining compressed gas remaining without being exhausted inside the cylinder bore 61. It is a section. Section S2 is a suction section through which the gas hole 14 sucks gas while passing through the suction valve groove 51. The valve plate 50 blocks the gas hole 14 from K2 to K3, which is the bottom dead center B, to prepare for compression. In addition, R1 represents a section in which the gas hole 14 is blocked at the bottom dead center B and moves to K4, and the gas sucked in this section is primarily compressed. E1 is a section in which the gas hole 14 primarily exhausts the first compressed gas while passing through the primary exhaust valve groove 52. R2 represents a section in which the gas hole 14 is moved again to K6 in the state where the gas hole 14 is again blocked, and the first compressed gas is secondarily compressed. E2 is a section in which the gas hole 14 passes the secondary exhaust valve groove 53 to exhaust the secondary compressed gas secondaryly. R3 represents a section in which the gas hole 14 is moved again to K8 in a state where the gas hole 14 is again blocked, and in this section, the secondary compressed gas is again tertiarily compressed. E3 is a section in which the gas hole 14 passes the tertiary exhaust valve groove 54 and tertiaryly exhausts the tertiary compressed gas. In addition, K9 to K10 is a section in which the valve plate 50 closes the gas hole 14 and compresses the gas until the piston reaches the top dead center T to prepare for the next suction stroke.

The characteristics of the present invention are in the compression exhaust stroke, and the exhaust valve grooves act as the exhaust stroke section when the pressure of the exhaust manifold 24 is low during the exhaust stroke, but the compression is performed when the pressure of the exhaust manifold 24 is high. It acts as an administrative section. That is, even if the gas hole 14 meets the primary, secondary and tertiary exhaust valve grooves 52, 53 and 54 in the exhaust sections E1, E2 and E3, they are compressed in the cylinder bore 61. Only when the pressure of the compressed gas is higher than the pressure in the exhaust manifold 24, the compressed gas is discharged from each of the exhaust valve grooves 52, 53, and 54, and the compressed gas is discharged in the cylinder bore 61. If the pressure is lower than the pressure in the exhaust manifold 24, the check valve 46 installed in each exhaust port is closed to prevent the flow back from the exhaust manifold 24 into the cylinder bore 61. It acts as a compressed administrative section rather than an administrative section.

FIG. 9B shows the pressure of the gas that can be obtained in the cylinder bore 61 corresponding to each rotation angle when one cylinder bore 61 rotates once along the valve plate 50 with all the check valves closed. (P) is shown.

For convenience of explanation, it is assumed that the pressure in the exhaust pipe 22 and the pressure in the compression tank are the same, ignoring the pressure loss caused by the check valve and the exhaust passage, and the pressure in the intake pipe 21 is P ₀₀ , and the cylinder bore 61 is used. When the pressure at the point of inhalation at the rotation angle K1 is P ₀ , it is assumed that P ₀₀ > P ₀ is due to air friction loss during suction. And Ptr means the rated pressure in the compression tank, Pmax means the maximum reaching pressure that can be reached in the cylinder bore 61 with all the check valve closed. When the compressor is actually designed, as shown in Fig. 9B, the set pressure Ptr in the compression tank is set, and then the stroke length of the piston is adjusted so that the maximum reaching pressure Pmax in the cylinder bore 61 is higher than the set pressure ptr. Be sure to When the rotation angle K of the gas hole 14 becomes K4, the gas pressure in the cylinder bore 61 at point 4 becomes P4, and when K5, K6, K7, K8, and K9 are sequentially, P5, P6, and P7, respectively. , P8, P9. Therefore, when designing the compressor according to the present invention, the design pressure Ptr of the compression tank is set to be located between the pressure reached when the gas hole 14 is located in the last exhaust section E3, that is, between P8 and P9, As indicated by the dotted line in Fig. 9B, the position of K1 is determined so that the pressure of the gas remaining in the cylinder bore 61 after the compression exhaust stroke is completed is equal to the suction pressure P ₀ .

When the compressor of the present invention is continuously operated based on the characteristics of the suction and compressed exhaust stroke of the present invention described above and the pressure change in the compressor, the compression characteristics of the gas will be described with reference to FIG.

10 shows the compression characteristics of the compressor obtained by one cylinder bore 61, Figure 10A shows the compression characteristics of the compressor according to the present invention, Figure 10B shows the compression characteristics of the conventional reciprocating compressor 10, C shows the compression characteristics of a conventional bent axis compressor. In FIG. 10, the horizontal axis represents the number of reciprocating strokes of the piston, which is equal to the rotation speed N of the cylinder block 60.

First, A of FIG. 10 will be described. When the pressure Pt of the compression tank indicated by point D corresponds to the pressure between points 4 and 5, which represent rotational positions K4 and K5 in Fig. 9A, the pressure in the cylinder bore 61 is equal to points 3, The process changes from point 4, point 4D, point 5, point 6, point 7, point 8 and point 9. That is, the compression proceeds from the initial pressure P0 to the point 4, and the compression proceeds to the point 4D with the check valve 46 in the primary exhaust port 41 closed. However, passing through point 4D, the check valve 46 in the primary exhaust port 41 opens, reaching point 5 with the same pressure. Subsequently, from point 5 to point 6, the gas hole 14 is closed and the R2 section of Fig. 9A is passed, so compression is performed. Since the pressure in the cylinder bore 61 is higher than the pressure at the point D, which is the pressure of the compression tank, at the point 6 at which the secondary exhaust section E2 starts, the check valve 46 in the secondary exhaust port 42 is opened. The pressure is then temporarily reduced until point 7 of the secondary exhaust process (E2). Subsequently, from point 7 to point 8, the gas hole 14 is again closed, so that it passes through the section R3 of Fig. 9A, and the compression proceeds again. Then, when the point 8 reaches the start point of the tertiary exhaust section E3, the pressure in the cylinder bore 61 is higher than the pressure in the compression tank, so that the check valve 46 in the tertiary exhaust port 43 is opened. The pressure drops to point 9, the end point of the exhaust process E2, resulting in the pressure at point D.

This process requires that the check valve 46 is always closed if the pressure P in the cylinder bore is lower than the pressure tank pt of the compression tank indicated by point D. If the pressure P is higher than the pressure of the compression tank, the check valve 46 is always open. do. The symbol D in FIG. 10A indicates the position where the check valve 46 is opened. Therefore, as the number of revolutions of the compressor increases, the pressure D of the compression tank increases and the pressure discharged from the cylinder bore 61 toward the compression tank moves from point 4 to point 9 (i.e., dotted line in FIG. 10A). .

On the other hand, the auxiliary exhaust pipe 26 is connected to the auxiliary suction pipe 25 in a state in which the compressor of the present invention continues to operate, that is, in a state in which the compression load is not applied to the inside of the cylinder bore 61 (ie, no-load operation state). The compression characteristics when operating the compressor are shown on the right side of Fig. 10A. In this case, considering that a constant pressure loss occurs when the pressure P ₀₀ of the external intake gas is sucked into the cylinder bore 61, the pressure in the cylinder bore 61 becomes P ₀ , and this pressure is the point. It becomes the pressure at 3. When the compression stroke is performed under no load, the pressure curve in the cylinder bore 61 generates some compression sections (points 3 to 4, points 5 to 6, and points 7 to 8), but eventually the external suction It is equal to the gas pressure (P ₀₀ ) This is because the compression exhaust stroke is performed with all the check valves 46 all open.

In order to compare with the compression characteristics of the compressor of the present invention described above, the compression characteristics of a conventional reciprocating compressor and a conventional bent axis compressor will be described.

Referring to the conventional reciprocating compressor shown in FIG. 10B is as follows. When the compressor is operated for the first time with the pressure at the point D of the compression tank (pt) lower than the set pressure (ptr), the pressure (point B) of the compressed gas in the cylinder before the piston reaches top dead center is As soon as the pressure is higher than the pressure of the compression tank, the exhaust valve opens and gas is released. That is, during the compression process of the gas, if the pressure P of the cylinder chamber is lower than the pressure D of the compression tank, the pressure is continuously compressed. If the pressure P is the same, the exhaust valve for exhaust starts to open and exhausts. Therefore, as the number of times of compression increases, that is, the pressure of the compression tank increases, the position of the point B which terminates the compression process for each rotation increases. In the case of no load operation, the exhaust valve is opened at the H point at which the pressure of the exhaust pipe becomes P ₀₀ , thereby exhibiting compression characteristics as shown in the right curve of Fig. 10B.

In the conventional bent axis compressor shown in FIG. 10C, the gas in the cylinder chamber is always compressed to point B, and then immediately exhausted when this pressure is higher than point D, which is the pressure of the compression tank, and further compressed if lower than point D. There is a characteristic. In the case of no-load operation, the gas is compressed to point H and immediately dropped to the external inlet gas pressure (P ₀₀ ).

As shown in FIG. 10, when the pressure of the compression tank reaches the set pressure Ptr by operating the compressor, the energy efficiency is improved by changing to no-load operation immediately. Therefore, the total load of the compressor required to operate the compressor to compress the gas can be expressed as the sum of the area of the polygon consisting of points 3 to 9 or points A, B, C, and F for each revolution. The total load of such a compressor is proportional to the total amount of energy required to drive the compressor.

In the case of the compressor according to the present invention, the total energy required for compression is similar to that of the conventional reciprocating compressor shown in FIG. 10B, but it can be seen that it is very small compared to the conventional bent-type compressor shown in FIG. 10C.

Therefore, it can be seen that the compressor according to the present invention has a very high energy efficiency as compared to the conventional four-axis compressor. In particular, even in the case of no load operation, the compressor according to the present invention shows a significantly smaller amount of energy required than a conventional bent axis compressor.

On the other hand, the noise generated in the compressor mainly occurs as the pressure difference between the cylinder and the compression tank increases. Considering this point, in the case of a conventional bent axis compressor, the pressure difference between the point B and the point D is very large as shown in FIG. 10C, while in the compressor of the present invention, as shown in A of FIG. The pressure difference from D is very small since it is represented by the pressure difference from point 5 to point 6 and from point 7 to point 8. This result means that the pressure difference between the compressed gas in the cylinder and the compressed gas in the compression tank is very small, which means that there is almost no explosion when the two gases are combined. Therefore, the gas compressor according to the embodiment shown in FIGS. 1, 6, and 7 of the present invention has an advantage that little noise is generated.

In consideration of the fact that the compression load applied to the cylinder bore 61 is the same as the axial load applied to the drive shaft 10, the compressor according to the present invention, as shown in FIG. It can be seen that it is much smaller than this conventional bent axis compressor. Therefore, the compressor of the present invention has an advantage that the durability of the bearings supporting the swash plate 80 directly affected by such a variable load and the bearings connected to the drive shaft 10 can be greatly increased.

As described above, the compressor of the present invention is designed to allow the exhaust stroke to be selectively made according to the pressure of the pressure tank, and is designed to circulate or directly suck the gas introduced into the cylinder bore toward the crank chamber. To effect.

First: the aerodynamically reduced noise source allows quiet operation of the compressor.

Secondly, energy efficiency can be maximized by minimizing the power required to compress the gas.

Third, the durability of the compressor is improved by reducing the fluctuation of the compression load per unit time.

Fourth: Provide a compressor with high efficiency and no-load operation.

Fifth: The refrigerant can be circulated around the compressor, and the heat generated by mechanical friction and air compression inside the compressor can be discharged smoothly to the outside using cooling lubricating oil to improve the compression efficiency of the gas and the durability of the compressor. .

Sixth: By offsetting the centrifugal force generated by the rotation of the piston, by reducing the relative friction force generated in the contact surface inside the cylinder, heat generation can be suppressed, and the life of the compressor can be extended.

Seventh: The piston rod is made of two-section, making the assembly of the piston easier, ultimately improving the assembly productivity of the compressor.

Claims (22)

A driving shaft in which a plurality of gas holes are formed on a concentric circle at regular intervals and integrally formed perpendicular to the direction of the driving shaft; and A gas guide member having an intake manifold for sucking gas from the outside and an exhaust manifold for discharging the gas compressed in the cylinder bore to the outside; and At least one suction port rotatably supporting the drive shaft and supplying the gas sucked from the intake manifold into the cylinder bore and at least two exhaust ports for discharging the gas compressed inside the cylinder bore to the exhaust manifold. Case head member is formed; And A gas inlet valve groove and at least two or more gas exhaust valve grooves are formed on a circumference of the cylinder head, which is fixed to the inner surface of the case head member and is in close contact with the outer surface of the cylinder head. The groove supplies the gas introduced from the suction port into the cylinder bore, and the gas exhaust valve groove includes a valve plate member for discharging the gas compressed in the cylinder bore into the exhaust port. A plurality of cylinder bores are formed in a direction parallel to the drive shaft, one side is integrally coupled with the cylinder head, the other side is slidably inserted into each cylinder bore, so that each cylinder bore A cylinder block for compressing the sucked gas; and A swash plate member connected to the central portion of the cylinder block by a bent axis and connected to the plurality of pistons by a piston rod, and converting the rotational force transmitted from the drive shaft into a linear reciprocating motion and transmitting it to the piston; and A case end plate having an inclined slope for supporting the swash plate member; and A case coupled to the case head and the case end plate to house the cylinder block and the swash plate member; A four-axis gas compressor comprising a. The bent-type gas compressor of claim 1, wherein a check valve is built in each exhaust port of the case head member. The gas inlet valve groove of the valve plate member is formed in a circumferential 180 ° section corresponding to a suction stroke section in which a piston moves from a top dead center to a bottom dead center. A four-axis gas compressor characterized in that it is formed in the remaining 180 ° section of the circumference corresponding to the compression stroke section moving from point to top dead center. The method of claim 3, wherein the gas intake valve groove and the gas exhaust valve grooves of the valve plate member is a screw shaft, characterized in that the (L V _L of the partition wall), a distance between the respective valve, largely formed more gas hole diameter of the cylinder head Gas compressors. 5. The bent axis gas compressor of claim 4, wherein a length of each of the gas exhaust valve grooves of the valve plate member is shorter than a distance between gas holes. The bent gas compressor of claim 5, wherein the groove width of each valve groove of the valve plate member is equal to or larger than the diameter of the gas hole. The bent gas compressor of claim 6, wherein the groove width of at least one of the gas exhaust valve grooves of the valve plate member is smaller than a diameter of the gas hole. 6. The circuit according to claim 5, wherein the case head member and the drive shaft are provided with a circulation circuit for introducing gas introduced from the intake manifold into the cylinder bore via a sealed crank chamber formed inside the case. 4-axis gas compressors. 9. The circuit of claim 8, wherein the circulation circuit comprises at least one suction passage communicating with the case head member from the intake manifold to the crank chamber and at least one suction passage communicating with the cylinder bore via the cylinder block chamber from the crank chamber. And a hollow portion is formed in a portion of the drive shaft in an axial direction, and a shaft port is formed in a vertical direction in the hollow portion. 10. The bent axis gas compressor according to claim 9, wherein at least one liquid inlet bump is formed in the hollow portion or the cylinder block chamber in the direction of the drive shaft of the circulation circuit. 11. The bent gas compressor of claim 10, further comprising a tension ring and a ring-shaped leaf spring inserted between the inner surface of the case head member and the valve plate member. 12. The method of claim 11, wherein a cooling case surrounding the case is attached to the outer surface of the case, a cooling water inlet and a cooling water outlet are formed in the cooling case, and a spiral heat dissipation fan is formed outside the case. A four-axis gas compressor. 13. The bent gas compressor of claim 12, wherein a plurality of blades are formed on an outer circumferential surface of the swash plate member. The heat dissipation fan of claim 13, wherein the heat dissipation fan is formed on the outer side of the case head member with respect to the driving shaft, and the heat dissipation fan is formed on the outer side of the case head in a direction parallel to the drive shaft. 4-axis gas compressors. 15. The bent gas compressor of claim 14, wherein the gas guide member has an auxiliary inlet pipe and an auxiliary exhaust pipe connecting the intake pipe communicating with the intake manifold and the exhaust pipe communicating with the exhaust manifold. . 16. The bent axis gas compressor according to any one of claims 1 to 15, wherein the bent axis connecting the cylinder block and the swash plate member is connected to any one of a universal coupling and a bevel gear. 17. The bent gas compressor of claim 16, wherein the piston rod connecting the piston and the swash plate member is connected to any one of a universal coupling, a two-section crank, and an extension rod. 18. The bent gas compressor of claim 17, wherein the expansion rod has a shape of two male and female bolts, and the stroke gap of the piston is adjusted using a lock nut coupled to an outer side of the male bolt. 19. The bent axis gas compressor according to claim 18, wherein a leaf spring is provided on an outer side of the expansion rod. 16. The bent axis gas compressor according to any one of claims 1 to 15, wherein the bent axis connecting the cylinder block and the swash plate member is connected by a spring coupling. 21. The bent axis gas compressor of claim 20, wherein the piston rod connecting the piston and the swash plate member is connected by a two-section crank. 22. A bent gas compressor as set forth in claim 21, wherein said spring coupling is connected so that compressive or twisting stress acts in the same direction as the rotation direction of said cylinder block and piston rod.