Technical Field
[0001] The present invention relates to a compressor valve structure, and particularly,
to a 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
[0002] 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).
[0003] 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.
[0004] For example, in a compressor shown in PTL 1, 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.
[0005] 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.
Citation List
Patent Literature
Summary of Invention
Technical Problem
[0007] 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.
[0008] 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.
[0009] 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.
[0010] 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.
[0011] 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.
[0012] 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
[0013] 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.
[0014] 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.
[0015] 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.
[0016] 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.
[0017] 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, it is preferable to form the reed
valve 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.
[0018] For example, the moment of inertia of area of the base end side may be made higher
by gradually increasing the width of the reed valve as it goes from the leading end
portion to the base end portion.
[0019] 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
[0020] 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.
[0021] 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
[0022]
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(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 is a diagram showing the state in which the suction valve seat is superimposed
on the valve plate.
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 is a diagram showing the state in which the discharge valve seat is superimposed
on the valve plate.
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 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(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 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
[0023] 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.
[0024] 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.
[0025] 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.
[0026] 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.
[0027] 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.
[0028] 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.
[0029] 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.
[0030] 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.
[0031] 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.
[0032] 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.
[0033] 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.
[0034] 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.
[0035] 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.
[0036] 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).
[0037] 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.
[0038] 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.
[0039] 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.
[0040] 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) .
[0041] 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.
[0042] 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.
[0043] 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.
[0044] 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.
[0045] 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.
[0046] 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).
[0047] 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.
[0048] 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.
[0049] 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.
[0050] 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.
[0051] 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.
[0052] 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.
[0053] 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.
[0054] 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.
[0055] 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) .
[0056] 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).
[0057] 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).
[0058] 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).
[0059] 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.
[0060] 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.
[0061] 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).
[0062] 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.
[0063] 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.
[0064] 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.
[0065] 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.
[0066] 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.
[0067] 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.
[0068] 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.
[0069] 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.
[0070] 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).
[0071] 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, the thickness of the reed valve
may be gradually increased toward the base end side.
Reference Signs List
[0072]
- 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