EP3553315B1

EP3553315B1 - Discharge valve structure for compressor

Info

Publication number: EP3553315B1
Application number: EP17881648.4A
Authority: EP
Inventors: Hiroyuki Ishida; Hideyuki Takahashi; Takayuki Endo; Hao He
Original assignee: Valeo Japan Co Ltd
Current assignee: Valeo Japan Co Ltd
Priority date: 2016-12-12
Filing date: 2017-12-08
Publication date: 2021-07-07
Anticipated expiration: 2037-12-08
Also published as: EP3553315A1; JPWO2018110449A1; CN110073105A; CN110073105B; WO2018110449A1; EP3553315A4

Description

Technical Field

The present invention relates to a compressor valve structure, and particularly, to a discharge valve structure in which a valve plate is disposed between a cylinder block and a cylinder head, wherein ports formed in the valve plate are opened and closed by reed valves.

Background Art

As a reciprocating compressor, one that includes a cylinder block having formed therein cylinder bores; pistons which move reciprocally linearly in the cylinder bores; a cylinder head which, being provided on the opposite side of the cylinder block from the side on which the pistons are inserted, has formed demarcated therein a suction and a discharge chamber in which to temporarily store a working fluid; and a valve plate disposed between the cylinder block and the cylinder head, has been publicly known. In this kind of configuration, the cylinder bores and the suction and discharge chambers communicate with each other via ports provided in the valve plate, and the individual ports are opened and closed by valve elements (suction valves, discharge valves) each formed of a reed valve with elasticity. The leading end portion of the reed valve, being formed to be larger than each of the ports, comes into abutment with the valve plate, thereby blocking the circulation of a working fluid passing through the port (closed valve) . On the other hand, when the pressure on the upstream side of the port is higher than the pressure on the downstream side, the reed valve comes out of abutment with the valve plate due to the difference between the pressures of the working fluid which act on the leading ends of the reed valve, allowing the circulation of the working fluid (open valve).
The contact width in which the leading end of the reed valve comes into abutment with the valve plate preferably has a width enough to be sealable so as for the working fluid not to leak out, but if the area in which the leading end portion of the reed valve comes into contact with a valve seat portion is too large, there is a problem in that the reed valve is inhibited from opening due to the surface tension of a lubricant which is interposed between the leading end portion of the reed valve and the valve seat portion when the valve is closed, causing a decrease in performance or a vibration. For this reason, it is necessary to appropriately control the contact width in which the leading end of the reed valve comes into abutment with the valve plate.
For example, in a compressor shown in JP-A-2010-169077 as shown in Fig. 9, an annular groove 101 is formed around a port 100 in a valve plate 3 disposed between a cylinder block and a cylinder head, thereby forming an annular valve seat 102 on the periphery of the open end of the port 100, and a leading end portion 103a of a reed valve 103 is brought into abutment with the valve seat 102 elastically, thus opening and closing the port 100. Also, the outer edge of the leading end portion 103a of the reed valve 103 is brought into coincidence with the outer edge of the valve seat 102 (the inner edge of the annular groove 101), thus preventing the leading end portion 103a of the reed valve 103 from sticking out into the annular groove 101.
As the outer edge of the leading end portion 103a of the reed valve 103 is brought into coincidence with the outer edge of the valve seat 102 (the inner edge of the annular groove 101) in this way, with the reed valve 103 being closed, it is difficult for the outer edge of the reed valve 103 to come into contact with a lubricant accumulating in the annular groove 101, disrupting the supply of the lubricant to the space between a valve seat surface and the outer edge of the leading end portion of the reed valve 103. Because of this, the adhesion of the lubricant interposed between the valve seat 102 and the leading end portion of the reed valve 103 is reduced, and it is possible to suppress a disorder of valve opening timing due to the stiction of the reed valve 103 caused by the lubricant.
A compressor having a discharge valve structure according to the preambe of claim 1 is known from JP-A-2011-226464 .

Summary of Invention

Technical Problem

In the heretofore known configuration described above, however, when the relative position of the reed valve 103 and the valve plate 3 fluctuates due to production tolerance, the leading end portion 103a of the reed valve 103 easily sticks out into the annular groove 101. A portion of the leading end portion 103a of the reed valve 103, which sticks out of the valve seat 102, oscillates so as to thrust into the inside of the annular groove 101 due to the inertial force generated when the reed valve 103 is seated on the valve seat 102, and stretching stress and compression stress act repeatedly in the vicinity of the sticking out portion. Because of this, there is a concern about a problem in that the leading end portion of the reed valve 103 is subject to fatigue failure due to this bending stress.
Also, when a lubricant adheres to the outer edge of the reed valve 103 which sticks out into the annular groove 101, the adhering lubricant is led between the valve seat 102 and the leading end portion of the reed valve 103 along the outer edge, and there is also a problem in that a disorder of valve opening timing tends to occur due to the adhesion of the lubricant.
In order to avoid these problems, it is also considered to bring the outer edge of the leading end portion 103a of the reed valve 103 into abutment with the valve seat inside the inner edge of the annular groove 101, but in the heretofore known annular valve seat 102, the area of contact between the leading end portion 103a of the reed valve 103 and the valve seat 102 is small, and a high surface pressure acts momentarily on a portion of the valve seat 102 with which the reed valve 103 is in abutment.
Particularly when in high speed operation, the speed at which the leading end portion of the reed valve 103 when closed hits against the valve seat 102 is very high, and so when the area of abutment with the valve seat 102 is small, the reed valve 103 hits against the valve seat 102 at a high speed, so that a high surface pressure acts momentarily on the surface of the valve seat 102 with which the reed valve 103 is in abutment.
Because of this, there is a fear that the leading end portion 103a of the reed valve 103 or the valve seat 102 breaks, or the reed valve 103 breaks secondarily resulting from the damaged valve seat 102.
The invention, having been contrived taking into consideration these circumstances, has for its principal problem to provide a compressor valve structure wherein no break occurs in a reed valve or a valve seat even when in high speed operation, and a stiction of the reed valve caused by a lubricant can also be reduced, enabling a stable operation to be maintained.

Solution to Problem

In order to achieve the above-mentioned problem, the compressor valve structure according to the invention is a valve structure used in a compressor, which includes a cylinder block having formed therein cylinder bores; pistons which move reciprocally linearly in the cylinder bores; a cylinder head having formed therein a space in which to temporarily store a working fluid; a valve plate which, being provided between the cylinder block and the cylinder head, has formed therein ports which provide communication between the cylinder bores and the space; and reed valves which open and close the ports in the valve plate, wherein an annular valve seat with which each of the reed valves comes into abutment is provided on the periphery of the open end of each of the ports in the valve plate, the radial width of the valve seat is formed to be larger on the side corresponding to the leading end portion of the reed valve than on the side corresponding to the base end portion of the reed valve, and the outer edge of the leading end portion of the reed valve is positioned inside the outer edge of the valve seat in a state where the port is closed by the reed valve.
Consequently, in the annular valve seat provided on the periphery of the open end of the port with which the reed valve comes into abutment, as the leading end side radial width of the reed valve is set to be larger than the base end side radial width, the contact width of the reed valve on the leading end side which has a high power of impact is increased while suppressing a problem in that the area of abutment between the reed valve and the valve seat is too large, inhibiting the valve from opening due to surface tension, and it is thus possible to reduce the contact surface pressure generated when the reed valve when closed hits against the valve seat.
Also, as the outer edge of the leading end portion of the reed valve is positioned inside the outer edge of the valve seat, with the reed valve closing the port, it is difficult for a lubricant to flow into between the reed valve and the valve seat along the outer edge of the reed valve, reducing the problem of a stiction of the reed valve (the problem of a disorder of valve opening timing due to the adhesion of the lubricant), and at the same time, the whole of the valve seat surface does not serve as an adsorption surface even when the lubricant flows into between the reed valve and the valve seat, so that it is possible to suppress the disorder of valve opening timing due to the adhesion of the lubricant.
Herein, the valve seat may be formed by providing a plurality of inconsecutive recesses around the port in the valve plate, but the dimensional control of the valve plate is cumbersome, and so an annular valve seat may be formed on the periphery of the opening of the port by forming an annular groove around the port in the valve plate.
Also, in order to reduce the speed (valve closing speed) at which the reed valve is seated on the valve seat at valve closing and thus reduce the power of impact generated when the reed valve hits against the valve seat, the reed valve is formed so that a moment of inertia of area of the base end side thereof is higher than a moment of inertia of area of the leading end side.
The moment of inertia of area of the base end side is made higher by gradually increasing the width of the reed valve as it goes from the leading end portion to the base end portion.
By adopting this kind of configuration, it is possible to accelerate the valve closing timing of the reed valve and thus to close the reed valve before cylinder pressure falls below discharge chamber pressure, and it is possible to suppress the speed at which the reed valve hits against the valve seat being excessive.

Advantageous Effects of Invention

As described above, according to the compressor valve structure according to the invention, in the annular valve seat, the leading end side radial width of the reed valve is formed to be larger than the base end side radial width, and at the same time, the outer edge of the leading end portion of the reed valve is positioned inside the outer edge of the valve seat, so that it is possible, by effectively increasing the contact area of the reed valve on the leading end side which has a high power of impact when the valve is closed, to reduce the surface pressure between the reed valve and the valve seat when the valve is closed. Because of this, it is possible to avoid a problem in that the reed valve is broken, or the valve seat is damaged, by the power of impact generated when the reed valve hits against the valve seat.
Also, as a configuration is such that the outer edge of the leading end portion of the reed valve does not stick out of the outer edge of the valve seat, there is less fear that a lubricant flows into between the reed valve and the valve seat when the valve is closed, and also, the whole of the valve seat surface does not serve as an adsorption surface even when the lubricant flows into between the reed valve and the valve seat, so that it is possible to eliminate the problem of the disorder of valve opening timing caused by the reed valve sticking to the valve seat due to the adhesion of the lubricant. Brief Description of Drawings

[Fig. 1] Fig. 1(a) is a sectional view showing one portion of a compressor including a valve structure according to the present invention, and Fig. 1(b) is a sectional view showing the valve structure according to the present invention.
[Fig. 2] Fig. 2(a) is a diagram showing a cylinder block side end surface of a valve plate, and Fig. 2 (b) is a diagram showing a suction valve seat to be superimposed on the end surface.
[Fig. 3] Fig. 3 is a diagram showing the state in which the suction valve seat is superimposed on the valve plate.
[Fig. 4] Fig. 4(a) is a diagram showing a cylinder head side end surface of the valve plate, and Fig. 4(b) is a diagram showing a discharge valve seat to be superimposed on the end surface.
[Fig. 5] Fig. 5 is a diagram showing the state in which the discharge valve seat is superimposed on the valve plate.
[Fig. 6] Fig. 6 is an enlarged view describing the valve structure, wherein (a) is a plan view describing the shape of the valve seat and the positional relationship between a discharge valve and the valve seat, and (b) is a sectional side view of the valve structure in (a).
[Fig. 7] Fig. 7 is a diagram wherein the cylinder pressure - valve lift characteristic diagram of a heretofore conventional valve structure is superimposed on that of the valve structure according to the present invention when in high speed operation.
[Fig. 8] Fig. 8(a) is a diagram showing the cylinder pressure - valve lift characteristics of the heretofore conventional valve structure, and Fig. 8(b) is a diagram showing the cylinder pressure - valve lift characteristics of the valve structure of the invention.
[Fig. 9] Fig. 9 is a diagram showing the heretofore conventional valve structure, wherein (a) is a plan view describing the relationship between a valve seat of the heretofore conventional valve structure and a reed valve seated on the valve seat, and (b) is a sectional side view of the valve structure in (a) .

Description of Embodiments

Hereafter, a description will be given, while referring to the accompanying drawings, of a valve structure according to the invention and a compressor using the valve structure.
Fig. 1 shows a piston type compressor 1 using the valve structure according to the present invention. The piston type compressor 1 is configured having a cylinder block 2, a cylinder head 4 which is assembled to the rear side of the cylinder block 2 via a valve plate 3, and a front housing 6 which, being assembled so as to cover the front side of the cylinder block 2, defines a crankcase 5 on the front side of the cylinder block 2. The front housing 6, cylinder block 2, valve plate 3, and cylinder head 4 are axially fastened by not-shown fastening bolts, configuring a housing 7 of the compressor.
A drive shaft 8 disposed in the crankcase 5 is rotatably retained in the front housing 6 and the cylinder block 2 via a bearing 9 (only the cylinder block side thereof is shown) . The drive shaft 8 protrudes from the front housing 6 and is connected to a not-shown travelling engine via a belt and a pulley, and the power of the travelling engine is transmitted to the drive shaft 8, causing the drive shaft 8 to rotate.
A support hole 11, in which the bearing 9 is housed, and a plurality of cylinder bores 12, which are disposed at equally spaced intervals on the circumference of a circle centered on the support hole 11, are formed in the cylinder block 2. Single head pistons 13 are reciprocally slidably inserted in their respective cylinder bores 12.
In the crankcase 5, a swash plate 14 which rotates in synchronism with the rotation of the drive shaft 8 is provided on the drive shaft via a hinge ball 15. Engaging portions 13a of the single head pistons 13 are retained in engagement with the peripheral portion of the swash plate 14 via a pair of shoes 16 provided one on each of the front and back of the swash plate peripheral portion.
Consequently, when the drive shaft 8 rotates, the swash plate 14 rotates correspondingly, and the rotational motion of the swash plate 14 is converted to a reciprocal linear motion of the single head pistons 13 via the shoes 16, changing the volume of compression chambers 17 which are formed between the single head pistons 13 and the valve plate 3 in the cylinder bores 12.
A suction port 20 and a discharge port 30 are formed in the valve plate 3 so as to correspond to each of the cylinder bores 12. Also, a suction chamber 18, in which to store a working fluid to be supplied to the compression chambers 17, and a discharge chamber 19, in which to store the working fluid discharged from the compression chambers 17, are demarcated in the cylinder head 4. In this example, the suction chamber 18 is formed in the central portion of the cylinder head 4, and the discharge chamber 19 is annularly formed around the suction chamber 18.
The suction chamber 18 can communicate with the compression chambers 17 via the suction ports 20 which are opened and closed by respective suction valves 21 to be described hereafter. Also, the discharge chamber 19 can communicate with the compression chambers 17 via the discharge ports 30 which are opened and closed by respective discharge valves 31 to be described hereafter.
A suction valve seat 22, which is superimposed on and attached to the cylinder block side end surface of the valve plate 3 and has formed therein the suction valves 21, and a gasket 23, which is superimposed on the suction valve seat 22 and is sandwiched and fixed between the valve plate 3 and the cylinder block 2, are provided between the valve plate 3 and the cylinder block 2.
Also, a discharge valve seat 32, which is superimposed on and attached to the cylinder head side end surface of the valve plate 3 and has formed therein the discharge valves 31, and a gasket 34, which is superimposed on the discharge valve seat 32 and is sandwiched and fixed between the valve plate 3 and the cylinder head 4 and with portions of which opposite to the discharge valves 31 retainers 33 are formed integrally, are provided between the valve plate 3 and the cylinder head 4.
The cylinder block 2, gasket 23, suction valve seat 22, valve plate 3, discharge valve seat 32, gasket 34, and cylinder head 4 are positioned by not-shown positioning pins and fixed pressed against each other by the fastening bolts which fasten the component members of the housing 7.
The suction chamber 18 communicates with a not-shown suction opening which is connected to the low-pressure side (the outlet side of an evaporator) of an external refrigerant circuit via a suction passage which is radially extended so as to pass through the discharge chamber 19. Also, the discharge chamber 19 communicates with a discharge space 41, which is formed in the peripheral wall portion of the cylinder block 2, via a passage formed in the gasket 34, valve plate 3, suction valve seat 22, gasket 23, and cylinder block 2. The discharge space 41 is defined by the cylinder block 2 and a covering 42 attached thereto and is connected to the high-pressure side (the inlet side of a radiator) of the external refrigerant circuit via a discharge opening 43 formed in the covering 42.
The suction valve seat 22, being superimposed on and attached to the cylinder block side end surface of the valve plate 3 shown in Fig. 2 (a), is configured having an aggregation of the plurality of suction valves 21 which open and close the respective suction ports 20, as shown in Fig. 2 (b) . The suction valves 21 are circumferentially formed at predetermined spaced intervals so as to correspond to the number of cylinder bores 12, and through holes 28 through which to insert the fastening bolts, through holes through which to insert the not-shown positioning pins, and the like, are formed, in the suction valve seat 22. Also, a through hole 24 which avoids interference with the suction port 30 is formed in the base end portion of each of the suction valves 21.
Each of the suction valves 21 is configured of a partial portion of the suction valve seat 22, and by forming a U-shaped punched-out slit 25 in the vicinity of the periphery of the suction valve seat 22, is integrally extended from radially outside to inside. That is, a leading end portion 21a of each of the suction valves 21 is disposed radially inside a base end portion 21b (the leading end portion 21a is disposed so as to be closer to the center of the suction valve seat 22 than the base end portion 21b).
Each of the suction valves 21 is formed as a cantilevered reed valve, and as shown in Fig. 3, too, the leading end portion 21a is made to serve as a seat portion which is seated on a valve seat 26 formed around the suction port 20 in the valve plate 3.
Through holes which avoid interference with the cylinder bores 12 are circumferentially formed at predetermined spaced intervals so as to correspond to the number of cylinder bores 12, and through holes through which to insert the fastening bolts, through holes through which to insert the positioning pins, and the like, are formed, in the gasket 23 interposed between the suction valve seat 22 and the cylinder block 2.
The discharge valve seat 32, being superimposed on and attached to the cylinder head side end surface of the valve plate 3 shown in Fig 4(a), is configured having an aggregation of the plurality of discharge valves 31 which open and close the respective discharge ports 30, as shown in Fig. 4(b). The discharge valves 31 are circumferentially formed at predetermined spaced intervals so as to correspond to the number of cylinder bores 12. Also, through holes 35 which avoid interference with the suction ports 20, not-shown through holes through which to insert the positioning pins, and the like, are formed in the discharge valve seat 32.
Each of the discharge valves 31, being configured of a partial portion of the discharge valve seat 32, is integrally radially extended from the seat central portion. That is, a leading end portion 31a of each of the discharge valves 31 is disposed radially outside a base end portion 31b (the leading end portion 31a is disposed so as to be farther away from the center of the discharge valve seat 32 than the base end portion 31b) .
Each of the discharge valves 31 is formed as a cantilevered reed valve, and as shown in Fig. 5, too, the leading end portion 31a is made to serve as a seat portion which is seated on a valve seat 36 formed around the discharge port 30 in the valve plate 3.
Through holes which avoid interference with the suction ports 20 are circumferentially formed at predetermined spaced intervals so as to correspond to the number of cylinder bores, and through holes through which to insert the fastening bolts, through holes through which to insert the positioning pins, and the like, are formed, in the gasket 34 interposed between the discharge valve seat 32 and the cylinder head 4, and each of the retainers 33 is formed integrally with the portion of the gasket 34 opposite to the discharge valve 31 so as to separate gradually from the discharge valve 31 as it goes from the base end portion 31b to the leading end portion 31a of the discharge valve 31.
Consequently, when in a suction process, a refrigerant is sucked into the compression chambers 17 from the suction chamber 18 via the suction ports 20 which are opened and closed by the suction valves 21, and when in a compression process, the compressed refrigerant is discharged into the discharge chamber 19 from the compression chambers 17 via the discharge ports 30 which are opened and closed by the discharge valves 31.
In this kind of compressor 1, the valve seat 26, on which the suction valve 21 is seated, and the valve seat 36, on which the discharge valve 31 is seated, are formed integrally with the periphery of the opening of the suction port 20 and with the periphery of the opening of the discharge port 30, respectively, in the valve plate 3.
Fig. 6 shows a valve structure wherein the valve seat 36 on which the discharge valve 31, out of the two valves, is seated is formed integrally with the valve plate 3, and hereafter, a description will be given, focusing on the valve structure on this discharge side.
The valve seat 36, by forming an annular groove 37 around the discharge port 30 in the valve plate 3, is annularly formed on the periphery of the opening of the discharge port 30, and is formed flush with the cylinder head side end surface of the valve plate 3. The valve seat 36 is not formed having a uniform width (radial width) all over the circumference, and the radial width thereof is set to increase toward the leading end side of the discharge valve (reed valve) 31 (the radial width of the valve seat 36 corresponding to the leading end side of the discharge valve 31 is set to be larger than the radial width corresponding to the base end side).
In this example, as the discharge valve 31 is extended from radially inside to outside the discharge valve seat 32, and the leading end portion 31a is positioned radially outside the base end portion 31b, the radial width of the valve seat 36 is formed to increase gradually as it goes from radially inside to outside the valve plate 3. Also, in this example, an outer edge of the valve seat 36 is formed in a continuous circular arc curve.
Furthermore, in this example, with the discharge valve 31 closing the discharge port 30, the circular-arc outer edge of the leading end portion of the discharge valve 31 is formed so as to be positioned inside the outer edge of the valve seat 36, that is, inside the inner edge of the annular groove 37. Because of this, the outer edge of the leading end portion of the discharge valve 31 is brought into abutment with a wide valve seat surface, which corresponds to the leading end side of the discharge valve 31, without sticking out into the annular groove 37.
Consequently, by adopting this kind of configuration, it is possible, in spite that the outer edge of the leading end portion of the discharge valve 31 does not stick out of the outer edge of the valve seat 36, to increase the area of contact between the leading end side of the leading end portion 31a of the discharge valve 31, which is high in the power of impact when the valve is closed, and the valve seat 36, as compared with in a heretofore conventional configuration wherein the radial width of the valve seat 36 is uniformly formed all over the circumference, and it is possible to increase a region on which a high contact pressure acts and thus reduce surface pressure.
Consequently, it is possible to avoid a problem in that the leading end portion 31a of the discharge valve 31 or the valve seat 36 breaks due to stress fluctuation or contact pressure, leading to a deterioration in the compression efficiency of the compressor.
Herein, even though the valve seat 36 has all over the circumference a uniform width which is the same as the leading end side width of the discharge valve 31 (even though the radial width of the valve seat 36 corresponding to the base end side of the discharge valve 31 has as large a width as the width corresponding to the leading end side), the same advantageous effect can be obtained in that the leading end side contact area of the leading end portion 31a of the discharge valve 31 is increased, reducing stress. As the area of contact between the leading end portion 31a of the discharge valve 31 and the valve seat 36 increases in excess when the valve is closed, however, there is a concern about a problem in that the discharge valve 31 is inhibited from opening due to the surface tension of a lubricant interposed between the leading end portion 31a of the discharge valve 31 and the valve seat 36, causing a decrease in performance or a vibration. In the previously described configuration example, as the width of the valve seat 36 corresponding to the base end portion of the discharge valve 31 is formed to be smaller than the width corresponding to the leading end side, coupled with the fact that the outer edge of the leading end portion of the discharge valve 31 is positioned inside the outer edge of the valve seat 36, it does not happen that the area of contact between the discharge valve 31 and the valve seat 36 increases in excess, and thus the above-described concern does not arise.
Furthermore, in this example, the discharge valve 31, by the width thereof being formed to increase gradually as it goes from the leading end portion to the base end portion, is configured so that a moment of inertia of area of the base end side is made higher than a moment of inertia of area of the leading end side.
By adopting this kind of configuration, as hereinafter described in detail, it is possible to advance the closing timing of the discharge valve, as compared with the heretofore known discharge valve (reed valve) whose width is equal from the leading end portion to the base end portion, and to close the discharge valve before cylinder pressure falls below discharge pressure, and it is thus possible to restrain the speed at which the discharge valve 31 hits against the valve seat 36 from being excessive.
Figs. 7 and 8 are diagrams wherein the behaviors of the pressure (cylinder pressure) in the compression chamber 17, and of the opening height (valve lift) of the discharge valve, with respect to the rotation angle of the shaft, at high speed operation, are analytically calculated. These examples use as analysis conditions the rotation speed: 9000 rpm, the discharge chamber pressure: 15 bar, the suction pressure: 2 bar, and the maximum opening height (maximum opening degree) of the discharge valve: 1 mm.
A description will hereafter be given, while referring to Fig. 8(b), of the behavior of the heretofore known discharge valve (reed valve) whose width is equal from the leading end portion to the base end portion (whose second moment of area is equal from the leading end portion to the base end portion) .
In the compression process (the section in which the angle of rotation of the shaft is 0° to 180°), when the cylinder pressure exceeds the discharge chamber pressure and the discharge valve 31 starts to open (in the drawing, when the angle of rotation shown by I is passed), a refrigerant gas in the cylinder starts to be discharged into the discharge chamber 19, but the refrigerant in the cylinder is not immediately discharged into the discharge chamber 19 due to a delay in opening of the discharge valve 31 or to the discharge valve's own resistance, and the pressure in the cylinder becomes higher than the discharge chamber pressure (in this example, 1.5 Bar).
The opening degree of the discharge valve 31 changes under the influence of the inertial force deriving from the discharge valve's own mass in addition to the balance between the force based on the difference between the pressures acting on the front and back of the valve (the difference between the cylinder pressure and the discharge chamber pressure) and the spring force of the discharge valve 31, and when the discharge valve 31 reaches the retainer 33, the opening degree of the discharge valve 31 is maintained at the maximum opening degree (in the drawing, at the rotation angle shown by II).
After that, the speed of the piston 13 decreases toward the top dead center, and so the cylinder pressure starts to drop. Then, when the force based on the difference between the pressures acting on the front and back of the discharge valve 31 cannot exceed the spring force of the discharge valve 31 when at the maximum opening degree, the opening degree of the discharge valve 31 starts to decrease (in the drawing, at or after the rotation angle shown by III).
Then, when the cylinder pressure falls below the discharge chamber pressure, the force based on the difference between the discharge chamber pressure and the cylinder pressure acts on the discharge valve 31 in the direction of closing the valve, so that the speed of closing of the valve, coupled with the discharge valve's own spring force, is accelerated (in the drawing, at or after the rotation angle shown by IV), and the discharge valve 31 hits strongly against the valve seat 36. For this reason, there is a concern that the discharge valve 31 or the valve seat 36 may break due to the power of impact with which the leading end portion of the discharge valve 31 hits against the valve seat 36.
The above-described concern, being the event which cannot be seen when in low to medium speed operation wherein the closing of the discharge valve is completed before the cylinder pressure falls below the discharge chamber pressure, stems from the fact that a natural valve-closing response with a heretofore conventional discharge valve's own spring force cannot follow the change of pressure in high speed operation.
Next, a description will hereafter be given, while referring to Fig. 8(b), of the behavior of the discharge valve 31 formed so that the width thereof increases gradually as it goes from the leading end portion to the base end portion (a moment of inertia of area of the base end side is made higher than a moment of inertia of area of the leading end side).
In the compression process (the section in which the angle of rotation of the shaft is 0° to 180°), when the cylinder pressure exceeds the discharge chamber pressure and the discharge valve 31 starts to open (in the drawing, when the angle of rotation shown by I' is passed), a refrigerant gas in the cylinder, by the discharge valve 31 opening, is discharged into the discharge chamber 19, but the refrigerant in the cylinder is not immediately discharged into the discharge chamber 19 due to a delay in opening of the discharge valve 31 or to the discharge valve's own resistance, and the pressure in the cylinder becomes higher than the discharge chamber pressure. Moreover, as the base end side second moment of area of the discharge valve 31 is made higher than the leading end side one, the pressure in the cylinder becomes slightly higher than heretofore known.
The opening degree of the discharge valve 31 changes under the influence of the inertial force deriving from the discharge valve's own mass in addition to the balance between the force based on the difference between the pressures acting on the front and back of the valve (the difference between the cylinder pressure and the discharge chamber pressure) and the spring force of the discharge valve 31, and when the discharge valve 31 reaches the retainer 33, the opening degree of the discharge valve is confined at the maximum opening degree (in the drawing, at the angle of rotation shown by II'). Even in this kind of discharge valve 31, as the moment of inertia of area of the base end side is higher than the leading end side one, the time at which a maximum lift is reached is more delayed than heretofore known.
After that, the speed of the piston decreases toward the top dead center, so that the cylinder pressure starts to drop, and when the force based on the difference between the pressures acting on the front and back of the discharge valve 31 cannot exceed the spring force of the discharge valve when at the maximum opening degree, the opening degree of the discharge valve 31 starts to decrease (in the drawing, at or after the angle of rotation shown by III') . As the discharge valve 21 is formed so that the moment of inertia of area of the base end side is higher than the leading end side one, the discharge valve 31 starts to be closed earlier than the heretofore conventional discharge valve. Also, as the leading end portion side width of the valve is smaller than the base end portion side width, the leading end portion side mass of the valve is not so large, and the response speed of the discharge valve can be effectively increased.
For this reason, as the discharge valve 31 is seated on the valve seat 36 before the rotation angle shown in IV' at which the cylinder pressure falls below the discharge chamber pressure, the problem is eliminated that the force based on the difference between the discharge chamber pressure and the cylinder pressure acts on the discharge valve 31 in the direction of closing the valve, accelerating the speed of closing of the discharge valve 31, and it is possible to avoid the discharge valve 31 hitting strongly against the valve seat 36. Because of this, it is possible to avoid the situation in which the discharge valve 31 or the valve seat 36 breaks due to the power of impact with which the leading end portion 31a of the discharge valve 31 hits against the valve seat 36.
The discharge side valve structure has heretofore been described, but as for the radial width of the valve seat, the suction side valve structure also adopts the same configuration, and thereby it is possible to exert the same working effect.
That is, the valve seat 26 on which the leading end portion 21a of the suction valve 21 is seated is annually formed on the periphery of the opening of the suction port 20 by forming the annular groove 27 around the suction port 20 in the valve plate 3, as shown in Fig. 2, while the suction valve 21 is extended from radially outside to inside the suction valve seat 22, and the leading end portion 22a is positioned radially inside the base end portion 22b, so that the radial width of the valve seat 26 is formed to increase gradually as it goes from the base end side to the leading end side of the suction valve 21 (as it goes from radially outside to inside the valve plate 3). Also, with the suction port 20 being closed by the suction valve 21, the outer edge of the leading end portion 21a of the suction valve 21 may be positioned inside the outer edge of the valve seat 26.
By adopting this kind of configuration, it is possible, in the suction valve 21, too, to increase the surface area of the valve seat 26 with which the leading end portion 21a of the suction valve 21 comes into abutment, relaxing a high stress resulting from a small surface area. Because of this, it is possible to avoid a problem in that the leading end portion 21a of the discharge valve 21 or the valve seat 26 breaks due to stress fluctuation or contact pressure, leading to a deterioration in the compression efficiency of the compressor.
Also, as the radial width of the valve seat 26 corresponding to the base end side of the suction valve 21 is formed to be smaller than that on the side corresponding to the leading end portion of the suction valve 21, it is possible to reduce the stiction of a lubricant, and thus possible to reduce a disorder of valve opening timing.
The above-described examples show an example in which the radial width of the valve seat 36, 26 is gradually increased toward the leading end side of the reed valve (the discharge valve 31, the suction valve 21), but the radial width of the valve seat 36, 26 may be locally increased only on the leading end side of the reed valve (the discharge valve 31, the suction valve 21).
Also, in the above-described examples, the width is formed to increase gradually as it goes from the leading end portion to the base end portion of the reed valve (discharge valve 31), thereby making the moment of inertia of area of the base end side of the reed valve (discharge valve 31) higher than the moment of inertia of area of the leading end, but the mode of increasing the moment of inertia of area of the base end side not being limited to this, for example, even though not covered by the present invention, the thickness of the reed valve may be gradually increased toward the base end side.

Reference Signs List

1: Piston type compressor
2: Cylinder block
3: Valve plate
4: Cylinder head
12: Cylinder bore
18: Suction chamber
19: Discharge chamber
20: Suction port
21: Suction valve
26: Valve seat
27: Annular groove
30: Discharge port
31: Discharge valve
36: Valve seat
37: Annular groove

Claims

A discharge valve structure configured to be used in a compressor (1) comprising a cylinder block (2) having formed therein cylinder bores (12), pistons (13) which move reciprocally linearly in the cylinder bores (12), and a cylinder head (4) having formed therein a space (18, 19) in which to temporarily store a working fluid; the discharge valve structure comprising:
a valve plate (3) which, configured to be provided between the cylinder block (2) and the cylinder head (4), has formed therein ports (30) which provide communication between the cylinder bores (12) and the space (19); and reed valves (31) which open and close the ports (30) in the valve plate (3);

wherein an annular valve seat (36) with which each of the reed valves (31) comes into abutment is provided on the periphery of the open end of each of the ports (30) in the valve plate (3), the radial width of the valve seat (36) is formed to be larger on the side corresponding to a leading end portion (31a) of the reed valve (31) than on the side corresponding to a base end portion (31b) of the reed valve (31), and the outer edge of the leading end portion (31a) of the reed valve (31) is positioned inside the outer edge of the valve seat (36) in a state where the port (30) is closed by the reed valve (31);

characterized in that the discharge reed valve (31) is formed so that a moment of inertia of area of a base end side is higher than a moment of inertia of area of a leading end side, wherein the reed valve is formed so that the width thereof increases gradually as it goes from the leading end portion (31a) to the base end portion (31b).
The compressor discharge valve structure according to any one of claims 1, wherein the annular valve seat (36) is formed on the periphery of the opening of the port (30) by forming an annular groove (37) around the port in the valve plate (3).
A compressor comprising a discharge valve structure according to any one of the preceding claims.