U.S. patent application number 10/183754 was filed with the patent office on 2003-01-16 for power transmission mechanism and compressor.
Invention is credited to Adaniya, Taku, Kanai, Akinobu, Kawaguchi, Masahiro, Kawata, Takeshi, Ota, Masaki, Suzuki, Takahiro.
Application Number | 20030012661 10/183754 |
Document ID | / |
Family ID | 19034403 |
Filed Date | 2003-01-16 |
United States Patent
Application |
20030012661 |
Kind Code |
A1 |
Kawata, Takeshi ; et
al. |
January 16, 2003 |
Power transmission mechanism and compressor
Abstract
A power transmission mechanism transmits power of an external
drive source to a rotary shaft. The transmission mechanism has a
first rotor rotated by the external power source. A second rotor is
connected to the rotary shaft. The second rotor is coaxial with the
first rotor and coupled to the first rotor transmit power to the
first rotor. A disconnecting body disconnects power transmission
from the first rotor to the second rotor when an excessive
transmission torque is generated between the rotors. A dynamic
damper is provided in at least one of the rotors. The dynamic
damper has a weight that swings like a pendulum. The axis of the
pendulum motion of the weight is separated by a predetermined
distance from and is substantially parallel to the rotation axis of
the corresponding rotor.
Inventors: |
Kawata, Takeshi;
(Kariya-shi, JP) ; Kawaguchi, Masahiro;
(Kariya-shi, JP) ; Ota, Masaki; (Kariya-shi,
JP) ; Adaniya, Taku; (Kariya-shi, JP) ; Kanai,
Akinobu; (Kariya-shi, JP) ; Suzuki, Takahiro;
(Kariya-shi, JP) |
Correspondence
Address: |
MORGAN & FINNEGAN, L.L.P.
345 Park Avenue
New York
NY
10154
US
|
Family ID: |
19034403 |
Appl. No.: |
10/183754 |
Filed: |
June 27, 2002 |
Current U.S.
Class: |
417/223 ;
417/222.1; 417/269 |
Current CPC
Class: |
F04B 27/0895 20130101;
F04B 27/1036 20130101; F04B 39/0088 20130101; F16F 15/145
20130101 |
Class at
Publication: |
417/223 ;
417/222.1; 417/269 |
International
Class: |
F04B 049/00 |
Foreign Application Data
Date |
Code |
Application Number |
Jun 28, 2001 |
JP |
2001-196633 |
Claims
1. A power transmission mechanism for transmitting power of an
external drive source to a rotary shaft, the transmission mechanism
comprising: a first rotor rotated by the external power source; a
second rotor connected to the rotary shaft, wherein the second
rotor is coaxial with the first rotor and coupled to the first
rotor to transmit power to the first rotor; a disconnecting body,
which disconnects power transmission from the first rotor to the
second rotor when an excessive transmission torque is generated
between the rotors; and a dynamic damper provided in at least one
of the rotors, wherein the dynamic damper has a weight that swings
like a pendulum, wherein the axis of the pendulum motion of the
weight is separated by a predetermined distance from and is
substantially parallel to the rotation axis of the corresponding
rotor.
2. The power transmission mechanism according to claim 1, wherein
the disconnecting body is a breakable member located in a power
transmission path between the rotors, wherein the breakable member
breaks when the excessive transmission torque is generated.
3. The power transmission mechanism according to claim 2, wherein
the breakable member is made of one of sintered metal and
low-carbon steel.
4. The power transmission mechanism according to claim 2, wherein
the first rotor is rotatably supported by a support member, which
rotatably supports the rotary shaft, and wherein the breakable
member is a pin and is connected to at least one of the rotors via
an elastic member fitted about the breakable member.
5. The power transmission mechanism according to claim 1, wherein
the disconnecting body has a coupling member, which couples the
rotors to each other to permit power to be transmitted between the
rotors and engages with at least one of the rotors, wherein, when
the coupling member is disengaged from the one of the rotors, the
power transmission between the rotors is disconnected.
6. The power transmission mechanism according to claim 5, wherein
the coupling member is a leaf spring.
7. The power transmission mechanism according to claim 1, wherein a
shock absorbing member is located in the power transmission path
between the rotors.
8. The power transmission mechanism according to claim 7, wherein
the second rotor is coupled to the rotary shaft and integrally
rotates with the rotary shaft, wherein the first rotor is rotatably
supported by a support member, which rotatably supports the rotary
shaft, and wherein the shock absorbing member is an elastic
member.
9. The power transmission mechanism according to claim 1, wherein
the weight is one of a plurality of weights, wherein the weights
are arranged or constructed to suppress a plurality of order
components in rotational vibration generated in the corresponding
rotor.
10. A compressor comprising: a rotary shaft; a compression
mechanism driven by rotation of the rotary shaft to compress fluid;
a first rotor supported by the rotary shaft to rotate integrally
with the rotary shaft; a second rotor fixed to the rotary shaft and
for rotating integrally with the rotary shaft, wherein the second
rotor is substantially coaxial with the first rotor and is rotated
by the external drive source; a disconnecting body located between
the rotors, wherein the disconnecting body disconnects power
transmission from the first rotor to the second rotor when an
excessive transmission torque is generated between the rotors; and
a dynamic damper provided in at least one of the rotors, wherein
the dynamic damper has a weight that swings like a pendulum,
wherein the axis of the pendulum motion of the weight is separated
by a predetermined distance from and is substantially parallel to
the rotation axis of the corresponding rotor.
11. The compressor according to claim 10, wherein the disconnecting
body is a breakable member located in a power transmission path
between the rotors, wherein the breakable member breaks when the
excessive transmission torque is generated.
12. The power transmission mechanism according to claim 11, wherein
the breakable member is made of one of sintered metal and
low-carbon steel.
13. The power transmission mechanism according to claim 10, wherein
a shock absorbing member is located in the power transmission path
between the rotors.
14. The power transmission mechanism according to claim 13, wherein
the first rotor is rotatably supported by a support member, which
rotatably supports the rotary shaft, and wherein the shock
absorbing member is an elastic member.
15. The power transmission mechanism according to claim 10, wherein
the weight is one of a plurality of weights, wherein the weights
are arranged or constructed to suppress a plurality of order
components in rotational vibration generated in the corresponding
rotor.
16. The compressor according to claim 10 further comprising a
piston type compression mechanism, which has piston accommodated in
a cylinder bore, and wherein the compression mechanism compresses
refrigerant in accordance with reciprocation of the piston.
17. The compressor according to claim 11, wherein the piston is one
a plurality of pistons, and the cylinder bore is one of a plurality
of cylinder bores, each of which accommodates one of the pistons,
and wherein the number of the cylinder bores is three, four, five,
six or seven.
18. The compressor according to claim 16, wherein the number of the
cylinder bores is three.
19. The compressor according to claim 18, wherein the compression
mechanism is driven by the rotary shaft, and wherein the
displacement per rotation of the rotary shaft is varied.
Description
BACKGROUND OF THE INVENTION
[0001] The present invention relates to a power transmission
mechanism having a power transmission means that is located between
two rotors and disconnects power transmission from one of the
rotors to the other when an excessive transmission torque is
generated between the rotors.
[0002] Japanese Utility Model Publication No. 6-39105 discloses a
compressor having such a power transmission mechanism.
[0003] The compressor disclosed in the publication includes a
pulley (first rotor), a transmission disk (second rotor), and a
shaft (driven rotary shaft). The pulley is rotatably supported by a
housing (support). The shaft is rotatably supported by the housing.
The transmission disk is secured to the shaft to rotate integrally
with the shaft. The rotors are coupled to each other by a breakable
member (power transmission disconnecting body). The breakable
member is located in a power transmission path and permits power to
be transmitted between the rotors. When receiving an excessive
load, the breakable member breaks.
[0004] When there is a malfunction in the compressor and a
transmission torque between the pulley and the transmission disk
exceeds a threshold level, the breakable member breaks and power
transmission from the pulley to the transmission disk is
disconnected.
[0005] The power transmission disconnecting body does not
disconnect the power transmission as long as the level of the
transmission torque is normal. When the transmission torque is
excessive due to an abnormality, the power transmission
disconnecting body disconnects the power transmission.
[0006] However, in a typical compressor, the transmission torque
between the first rotor and the second rotor fluctuates relatively
greatly in a normal state. The values of the peaks of the torque
fluctuation are close to the threshold torque level at which the
power transmission is disconnected. This reduces the design margin
of the power transmission disconnecting body. The reduced design
margin adds to difficulty in designing.
SUMMARY OF THE INVENTION
[0007] Accordingly, it is a first objective of the present
invention to facilitate designing of a power transmission mechanism
that does not disconnect power transmission while a power
transmission disconnecting body is in a normal power transmission
state and disconnects power transmission when transmission torque
is excessive.
[0008] A second objective of the present invention is to provide a
compressor that uses the power transmission mechanism.
[0009] To achieve the foregoing and other objectives and in
accordance with the purpose of the present invention, a power
transmission mechanism for transmitting power of an external drive
source to a rotary shaft is provided. The transmission mechanism
includes a first rotor, a second rotor, a disconnecting body, a
dynamic damper. The first rotor is rotated by the external power
source. The second rotor is connected to the rotary shaft. The
second rotor is coaxial with the first rotor and coupled to the
first rotor transmit power to the first rotor. The disconnecting
body disconnects power transmission from the first rotor to the
second rotor when an excessive transmission torque is generated
between the rotors. The dynamic damper is provided in at least one
of the rotors. The dynamic damper has a weight that swings like a
pendulum. The axis of the pendulum motion of the weight is
separated by a predetermined distance from and is substantially
parallel to the rotation axis of the corresponding rotor.
[0010] Other aspects and advantages of the invention will become
apparent from the following description, taken in conjunction with
the accompanying drawings, illustrating by way of example the
principles of the invention.
BRIEF DESCRIPTION OF THE DRAWINGS
[0011] The invention, together with objects and advantages thereof,
may best be understood by reference to the following description of
the presently preferred embodiments together with the accompanying
drawings in which:
[0012] FIG. 1 is a cross-sectional view illustrating a compressor
having a power transmission mechanism according to a first
embodiment of the present invention;
[0013] FIG. 2(a) is a front view illustrating the power
transmission mechanism of the compressor shown in FIG. 1;
[0014] FIG. 2(b) is a cross-sectional view taken along line
2(b)-2(b) of FIG. 2(a);
[0015] FIG. 3(a) is a front view illustrating a power transmission
mechanism according to a second embodiment; FIG. 3(b) is a
cross-sectional view taken along line 3(b)-3(b) of FIG. 3(a);
[0016] FIG. 4(a) is a front view illustrating a power transmission
mechanism according to a third embodiment; FIG. 4(b) is a
cross-sectional view taken along line 4(b)-4(b) of FIG. 4(a);
[0017] FIG. 5(a) is a front view illustrating a power transmission
mechanism according to a fourth embodiment; FIG. 5(b) is a
cross-sectional view taken along line 5(b)-5(b) of FIG. 5(a);
[0018] FIG. 6 is a front view showing a state of the power
transmission mechanism of FIG. 5(a); and
[0019] FIG. 7 is a partial front view showing the power
transmission mechanism of FIG. 5(a).
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0020] A compressor C according to a first embodiment of the
present invention will now be described with reference to FIGS. 1
to 2(b). The left end of the compressor C in FIG. 1 is defined as
the front of the compressor, and the right end is defined as the
rear of the compressor C.
[0021] The compressor C forms a part of a vehicular air
conditioner. As shown in FIG. 1, the compressor C includes a
cylinder block 11, a front housing member 12, a valve plate
assembly 13, and a rear housing member 14. The front housing member
12 is secured to the front end of the cylinder block 11. The rear
housing member 14 is secured to the rear end of the cylinder block
11 with the valve plate assembly 13 in between. The cylinder block
11, the front housing 12, the valve plate assembly 13, and the rear
housing member 14 form the housing of the compressor C.
[0022] The cylinder block 11 and the front housing member 12 define
a crank chamber 15 in between.
[0023] A rotary shaft 16 in this embodiment, is housed in the
compressor housing (supporting body) and extends through the crank
chamber 15. The front portion of the rotary shaft 16 is supported
by a radial bearing 12A located in the front wall of the front
housing member 12. The rear portion of the rotary shaft 16 is
supported by a radial bearing 11A located in the cylinder block
11.
[0024] A support cylinder 40 is formed at the front end of the
front housing member 12. The front end portion of the rotary shaft
16 extends through the front wall of the front housing member 12
and is located in the support cylinder 40. A power transmission
mechanism PT is fixed to the front end of the rotary shaft 16. The
power transmission mechanism PT includes a pulley 17 (first rotor).
The front end of the rotary shaft 16 is coupled to an external
drive source, which is a vehicular engine E in this embodiment, by
the power transmission mechanism PT and a belt 18, which is wound
around the pulley 17. The power transmission mechanism PT and the
compressor form a rotary machine.
[0025] A lug plate 19 is coupled to the rotary shaft 16 and is
located in the crank chamber 15. The lug plate 19 rotates
integrally with the rotary shaft 16. A cam plate, which is a swash
plate 20 in this embodiment, is housed in the crank chamber 15. The
swash plate 20 slides along and inclines with respect to the rotary
shaft 16. The swash plate 20 is coupled to the lug plate 19 by the
hinge mechanism 21. The lug plate 19 permits the swash plate 20 to
rotate integrally with the rotary shaft 16 and to incline with
respect to the rotary shaft 16 while sliding along the rotation
axis of the rotary shaft 16.
[0026] A snap ring 22 is fitted about the rotary shaft 16. A spring
23 extends between the snap ring 22 and the swash plate 20. The
snap ring 22 and the spring 23 limit the minimum inclination angle
of the swash plate 20. At the minimum inclination angle of the
swash plate 20, the angle defined by the swash plate 20 and the
axis of the rotary shaft 16 is closest to ninety degrees.
[0027] Cylinder bores 24 (only one is shown in FIG. 1) are formed
in the cylinder block 11. The cylinder bores 24 are located about
the rotation axis of the rotary shaft 16. A single-headed piston 25
is reciprocally housed in each cylinder bore 24. The front and rear
openings of each cylinder bore 24 are closed by the associated
piston 25 and the valve plate assembly 13. A compression chamber is
defined in each cylinder bore 24. The volume of the compression
chamber changes according to the reciprocation of the corresponding
piston 24. Each piston 25 is coupled to the peripheral portion of
the swash plate 20 by a pair of shoes 26. When the swash plate 20
is rotated by rotation of the rotary shaft 16, the shoes 26 convert
the rotation into reciprocation of each piston 25.
[0028] The cylinder block 11 (cylinder bores 24), the rotary shaft
16, the lug plate 19, the swash plate 20, the hinge mechanism 21,
the pistons 25, and the shoes 26 form a piston type variable
displacement compression mechanism.
[0029] Sets of suction ports 29 and suction valve flaps 30 and sets
of discharge ports 31 and discharge valve flaps 32 are formed in
the valve plate assembly 13. Each set of the suction port 29 and
the corresponding suction valve flap 30 and each set of the
discharge port 31 and the corresponding discharge valve flap 30
correspond to one of the cylinder bores 24 (compression
chambers).
[0030] A suction chamber 27 and a discharge chamber 28 are defined
in the rear housing member 14. The front ends of the suction
chamber 27 and the discharge chamber 28 are closed by the valve
plate assembly 13. As each piston 25 moves from the top dead center
position to the bottom dead center position, refrigerant gas is
drawn into the corresponding cylinder bore 24 (compression chamber)
through the corresponding suction port 29 while flexing the suction
valve flap 30 to an open position. Low pressure refrigerant gas
that is drawn into the cylinder bore 24 is compressed to a
predetermined pressure as the piston 25 is moved from the bottom
dead center position to the top dead center position. Then, the gas
is discharged to the discharge chamber 28 through the corresponding
discharge port 31 while flexing the discharge valve flap 32 to an
open position.
[0031] The suction chamber 27 is connected to the discharge chamber
28 by an external refrigerant circuit (not shown). Refrigerant that
is discharged from the discharge chamber 28 flows into the external
refrigerant circuit. The external refrigerant circuit performs heat
exchange using the refrigerant. The refrigerant is drawn into the
suction chamber 27 from the external refrigerant circuit. Then, the
refrigerant is drawn into each cylinder bore 24 to be compressed
again.
[0032] A bleed passage 33 is formed in the housing to connect the
crank chamber 15 with the suction chamber 27. A supply passage 34
is formed in the housing to connect the discharge chamber 28 with
the crank chamber 15. A control valve 35 is located in the supply
passage 34 to regulate the opening degree of the supply passage
34.
[0033] The opening of the control valve 35 is adjusted to control
the flow rate of highly pressurized gas supplied to the crank
chamber 15 through the supply passage 34. The pressure in the crank
chamber 15 (crank chamber pressure Pc) is determined by the ratio
of the gas supplied to the crank chamber 15 through the supply
passage 34 and the flow rate of refrigerant gas conducted out from
the crank chamber 15 through the bleed passage 33. As the crank
chamber pressure Pc varies, the difference between the crank
chamber pressure Pc and the pressure in the compression chambers
varies, which changes the inclination angle of the swash plate 20.
Accordingly, the stroke of each piston 25, or the compressor
displacement during one rotation of the rotary shaft 16, is
varied.
[0034] As shown in FIGS. 1 to 2(b), the support cylinder 40
protrudes from the front wall of the front housing member 12 and
surrounds the front portion of the rotary shaft 16.
[0035] A lip seal 41 is located in the support cylinder 40 to fill
the space between the support cylinder 40 and the rotary shaft 16.
The lip seal 41 prevents refrigerant from escaping the crank
chamber 15 through the space between the support cylinder 40 and
the rotary shaft 16.
[0036] A torque receiving member 42 is fixed to the front end of
the rotary shaft 16. The torque receiving member 42 rotates
integrally with the rotary shaft 16. The torque receiving member 42
includes a boss 42A, which functions as a second rotor. The boss
42A is fitted in the support cylinder 40 and is forward of the lip
seal 41. The torque receiving member 42 is coupled to the rotary
shaft 16 at the boss 42A. The torque receiving member 42 also
includes a pulley seat 42C, which surrounds the boss 42A and is
coupled to the boss 42A by spokes 42B. The boss 42A, the pulley
seat 42C, and the spokes 42B are formed integrally. The spokes 42B
function as power transmission disconnecting body (breakable
members). A radial bearing 40A is fitted about the support cylinder
40. The outer ring of the radial bearing 40A is secured to the
inner surface of the pulley seat 42C.
[0037] The pulley 17 is fixed to the outer surface of the pulley
seat 42C. The pulley 17 has a belt receiving portion 17A, about
which the belt 18 is wound. The belt 18 transmits power (torque) of
the output shaft of the engine E to the pulley 17. The pulley 17
also has an inner cylinder 17B. The pulley 17 is coupled to the
torque receiving member 42 by fitting the inner cylinder 17B about
the pulley seat 42C so that the pulley 17 rotate integrally with
the torque receiving member 42.
[0038] The spokes 42B, the number of which is three in this
embodiment, radially extends from the boss 42A to the pulley seat
42C. The spokes 42B transmit power between the boss 42A and the
pulley seat 42C. In other words, the spokes 42B are located in the
power transmission path between the pulley 17 and the boss 42A.
[0039] In this embodiment, the torque receiving member 42, which
includes the integrally formed boss 42A, the spokes 42B, and the
pulley seat 42C, is formed with sintered metal. The fatigue ratio
.sigma..sub.W/.sigma..su- b.B of the sintered metal is about 0.5.
The sign .sigma..sub.W represents the fatigue limit and the sign
.sigma..sub.B represents the tensile strength.
[0040] Four weight receptacles 45 and two weight receptacles 46
(one of each is shown in FIGS. 1 and 2(b)) are formed in the pulley
17 between the belt receiving portion 17A and the inner cylinder
17B. The weight receptacles 45, 46 function as weight guides. The
weight receptacles 45, 46 are angularly spaced by substantially
equal intervals to minimize the displacement of the center of
gravity of the pulley 17 from the rotation axis of the pulley
17.
[0041] Each weight receptacle 45 has a weight guide surface 45A,
and each weight receptacle 46 has a weight guide surface 46A. The
cross-section of each of the guide surfaces 45A and 46A is arcuate
along a plane perpendicular to the rotation axis of the pulley 17.
Each weight guide surface 45A forms a part of an imaginary
cylinder, the axis of which is parallel to the rotation axis of the
pulley 17. The radius of the imaginary cylinder is represented by
r.sub.1, and the axis of the imaginary cylinder is spaced from the
rotation axis of the pulley 17 by a distance R.sub.1. Each weight
guide surface 46A forms a part of an imaginary cylinder, the axis
of which is parallel to the rotation axis of the pulley 17. The
radius of the imaginary cylinder is represented by r.sub.2, and the
axis of the imaginary cylinder is spaced from the rotation axis of
the pulley 17 by a distance R.sub.2.
[0042] A weight, which is a roller 47, is accommodated in the
weight receptacle 45. Also, a weight, which is a roller 48, is
accommodated in the weight receptacle 46. The diameter and the mass
of the roller 47 are represented by d.sub.1, and m.sub.1. The
diameter and the mass of the roller 48 are represented by d.sub.2
and m2. Each roller 47 rolls in the circumferential direction along
the weight guide surface 45A of the corresponding weight receptacle
45. Each roller 48 rolls in the circumferential direction along the
weight guide surface 46A of the corresponding weight receptacle 46.
An annular lid 49 is fixed to the front face of the pulley 17 by
bolts. The lid 49 covers the weight receptacles 45, 46 to prevent
the rollers 47, 48 from falling off the weight receptacles 45,
46.
[0043] When the compressor C is being driven by the engine E, or
when the rotary shaft 16 is rotating, centrifugal force causes each
roller 47, 48 to contact the corresponding guide surface 45A, 46A
(see FIGS. 1 to 2(b)). If torque fluctuation is generated due to,
for example, torsional vibrations of the rotary shaft 16, each
roller 47, 48 starts reciprocating along the guide surface 45A, 46A
of the corresponding weight receptacle 45, 46. In other words, each
roller 47, 48 moves along the circumferential direction of the
guide surface 45A, 46A. That is, each roller 47, 48, or the center
of gravity of each roller 47, 48, swings like a pendulum about the
axis of an imaginary cylinder that includes the corresponding guide
surface 45A, 46A. That is, each roller 47, 48 acts as a centrifugal
pendulum when the compressor C is being driven by the engine E. The
size and mass of the rollers 47, 48 and the locations of the
rollers 47, 48 in the pulley 17 are determined such that the torque
fluctuation is suppressed by pendulum motion of the rollers 47,
48.
[0044] The pulley 17 (the weight receptacles 45, 46) and the
rollers 47, 48 form a dynamic damper.
[0045] The settings of the rollers 47, 48, which function as
centrifugal pendulums, will now be described.
[0046] Each roller 47, 48 (centrifugal pendulum) suppresses the
amplitude of torque fluctuation when the frequency of the
fluctuation is equal to the characteristic frequency of the roller
47, 48. Therefore, the location, the size, and the mass of each
roller 47, 48 are determined such that the characteristic frequency
of the roller 47, 48 is set equal to the frequency of a peak of the
torque fluctuation. Accordingly, the peak of the torque fluctuation
is suppressed, and the influence of the torque fluctuation is
effectively reduced. A peak of the torque fluctuation represents a
peak of the fluctuation range, or a rotation order component.
[0047] The frequency of the torque fluctuation and the
characteristic frequency of each roller 47, 48 are proportional to
the angular velocity .omega..sub.1 of the rotary shaft 16, which
corresponds to the speed of the rotary shaft 16. The frequency of
the torque fluctuation when its amplitude is the greatest is
represented by the product of the rotation speed of the rotary
shaft 16 per unit time (.omega..sub.1/2.pi.) and the number N of
the cylinder bores 24. That is, the frequency is represented by the
formula (.omega..sub.1/2.pi.).multidot.N. Through experiments, it
was confirmed that an nth greatest peak (n is a natural number) of
the torque fluctuation has a value equal to a product
n.multidot.(.omega..sub- .1/2.pi.).multidot.N.
[0048] The characteristic frequency of each roller 47, 48 is
obtained by multiplying the rotation speed of the rotary shaft 16
per unit time (.omega..sub.1/2.pi.) with the square root of the
ratio R/r. The sign R represents the distance between the rotation
axis of the pulley 17 (a rotor having weights that swing like
pendulums) and the axis of the pendulum motion of each roller 47,
48. The sign r represents the distance between the axis of the
pendulum motion of each roller 47, 48 and the center of gravity of
the roller 47, 48.
[0049] Accordingly, by equalizing the square root of the ratio R/r
with the product n.multidot.N, the characteristic frequency of each
roller 47, 48 is equalized with the frequency of the nth greatest
peak of the torque fluctuation. Accordingly, the torque fluctuation
at the nth greatest peak is suppressed.
[0050] To suppress the greatest peak of the torque fluctuation, the
values of the distances R and r of the rollers 47 are determined
such that the square root of the ratio R/r of the rollers 47 is
equal to N, or the value of the product n.multidot.N when n is one.
To suppress the second greatest peak of the torque fluctuation, the
values of the distances R and r of the rollers 48 are determined
such that the square root of the ratio R/r of the rollers 48 is
equal to 2N, or the value of the product n.multidot.N when n is
two.
[0051] The torque produced about the rotation axis of the pulley 17
by the rollers 47 is represented by a sign T.sub.1. The torque
produced about the rotation axis of the pulley 17 by the rollers 48
is represented by a sign T.sub.2. To effectively reduce peaks of
the torque fluctuation by the pendulum motion of the rollers 47,
48, the torques T.sub.1, T.sub.2 need to counter the torque
fluctuation and the amplitudes of the torques T.sub.1, T.sub.2 need
to be equal to the amplitude of the peaks of the fluctuation. When
the frequency of a peak of the torque fluctuation is equal to the
characteristic frequencies of the rollers 47, 48, the torques
T.sub.1, T.sub.2 are represented by the following equations.
T.sub.1=m.sub.t1.multidot.(.omega..sub.a1).sup.2.multidot.(R+r).multidot.R-
.multidot..phi..sub.1 (Equation 1)
T.sub.2=m.sub.t2.multidot.(.omega..sub.a2).sup.2.multidot.(R+r).multidot.R-
.multidot..phi..sub.2 (Equation 2)
[0052] The sign m.sub.t1 represents the total weight of the rollers
47 (m.sub.t1=4m.sub.1), and the .phi..sub.1. The sign m.sub.t2
represents the total weight of the rollers 48 (m.sub.t2=2m2), and
the sign .omega..sub.a2 is the average angular velocity of the
rollers 48 when the rollers 48 swing in a minute angle
.phi..sub.2.
[0053] The masses m.sup.t1, m.sub.t2 of the rollers 47, 48 are
maximized to minimize the values R, r, and .phi..sub.1,
.phi..sub.2, so that the size of the pulley 17 is minimized, and
the torques T.sub.1, T.sub.2 are maximized.
[0054] The axis of each imaginary cylinder that includes one of the
weight guide surfaces 45A coincides with the axis, or fulcrum, of
the pendulum motion of the corresponding roller 47. Likewise, the
axis of each imaginary cylinder that includes one of the weight
guide surfaces 46A coincides with the axis, or fulcrum, of the
pendulum motion of the corresponding roller 48. That is, the
distance R.sub.1 between the rotation axis of the pulley 17 and the
axis of each imaginary cylinder that includes one of the weight
guide surfaces 45A corresponds to the distance R for each roller
47. Likewise, the distance R.sub.2 between the rotation axis of the
pulley 17 and the axis of each imaginary cylinder that includes one
of the weight guide surfaces 46A corresponds to the distance R for
each roller 48.
[0055] The distance between the axis of the pendulum motion of each
roller 47 and the center of gravity of the roller 47 is equal to
the value obtained by subtracting the half of the diameter d.sub.1
of the roller 47 from the radius r.sub.1 of the corresponding
imaginary cylinder. Likewise, the distance between the axis of the
pendulum motion of each roller 48 and the center of gravity of the
roller 48 is equal to the value obtained by subtracting the half of
the diameter d.sub.2 of the roller 48 from the radius r.sub.2 of
the corresponding imaginary cylinder. That is, the difference
(r.sub.1-(d.sub.1/2)) corresponds to the distance r for each roller
47. Likewise, the difference (r.sub.2-(d.sub.2/2)) corresponds to
the distance r for each roller 48.
[0056] To suppress the greatest peak of the torque fluctuation, the
values of the distances R.sub.1, r.sub.1, and the diameter d.sub.1
of each roller 47 are determined such that the square root of
R.sub.1/(r.sub.1-(d.sub.1/2), which corresponds to the square root
of the ratio R/r, is equal to N, or the value of the product
n.multidot.N when n is one.
[0057] To suppress the second greatest peak of the torque
fluctuation, the values of the distances R.sub.2, r.sub.2, and the
diameter d.sub.2 of each roller 48 are determined such that the
square root of R.sub.2/(r.sub.2-(d.sub.2/2), which corresponds to
the square root of the ratio R/r, is equal to 2N, or the value of
the product n.multidot.N when n is two.
[0058] The settings are determined by regarding each roller 47, 48
as a particle at the center of gravity.
[0059] The operation of the compressor C and the power transmission
mechanism PT will now be described.
[0060] When the power of the engine E is supplied to the rotary
shaft 16 through the pulley 17, the swash plate 20 rotates
integrally with the rotary shaft 16. As the swash plate 20 rotates,
each piston 25 reciprocates in the associated cylinder bore 24 by a
stroke corresponding to the inclination angle of the swash plate
20. As a result, suction, compression and discharge of refrigerant
gas are repeated in the cylinder bores 24.
[0061] If the opening degree of the control valve 35 is decreased,
the flow rate of highly pressurized gas supplied to the crank
chamber 15 from the discharge chamber 28 through the supply passage
34 is decreased. Accordingly, the crank chamber pressure Pc is
lowered and the inclination angle of the swash plate 20 is
increased. As a result, the displacement of the compressor C is
increased. If the opening degree of the control valve 35 is
increased, the flow rate of highly pressurized gas supplied to the
crank chamber 15 from the discharge chamber 28 through the supply
passage 34 is increased. Accordingly, the crank chamber pressure Pc
is raised and the inclination angle of the swash plate 20 is
decreased. As a result, the displacement of the compressor C is
decreased.
[0062] During rotation of the rotary shaft 16, the compression
reaction force of refrigerant and reaction force of reciprocation
of the pistons 25 are transmitted to the rotary shaft 16 through
the swash plate 20 and the hinge mechanism 21, which torsionally
(rotationally) vibrates the rotary shaft 16. The torsional
vibrations generate torque fluctuation. The torque fluctuation
causes the compressor C to resonate. The torque fluctuations also
produce resonance between the compressor C and external devices
(the engine E and auxiliary devices), which are connected to the
pulley 17 by the belt 18.
[0063] When the torque fluctuation is produced, the rollers 47,48
start swinging like pendulums. The pendulum motion of the rollers
47, 48 produces torques about the rotation axis of the pulley 17.
The produced torques suppress the torque fluctuation. The
characteristic frequencies of the rollers 47, 48 are equal to the
frequencies of the first and second greatest peaks of the torque
fluctuation. Therefore, the peaks of the torque fluctuation are
suppressed, which effectively reduce the torque fluctuations of the
pulley 17.
[0064] As long as the magnitude of the transmission torque between
the pulley 17 and the boss 42A does not adversely affect the engine
E, that is, as long as the transmission torque is within a normal
power transmission state, the power is transmitted from the engine
E to the rotary shaft 16.
[0065] However, if there is an abnormality in the compressor C, for
example, if the compressor C is locked, and the transmission torque
is excessive, the spokes 42B break due to the excessive load.
Accordingly, the power transmission from the pulley 17 to the boss
42A is disconnected. This prevents the engine E from being
adversely affected by the excessive torque.
[0066] The present embodiment has the following advantages.
[0067] (1) The rollers 47, 48 are provided in the pulley 17. The
axis of each roller 47, 48 is spaced from the rotation axis of the
pulley 17 by a predetermined distance and is parallel to the
rotation axis of the pulley 17. Each roller 47, 48 swings like a
pendulum about its axis. The pendulum motions of the rollers 47, 48
suppress peaks of the torque fluctuation. Accordingly, the
magnitude of the transmission torque at which the power
transmission is disconnected is set greatly different from peaks of
the torque fluctuation in the normal power transmission state.
Since the spokes 42B are not broken by the magnitude of
transmission torque in the normal power transmission state but are
broken when the transmission torque is excessive, a large margin is
obtained. This facilitates designing.
[0068] Since the greatest peak of the torque fluctuation is
suppressed, resonance produced in the power transmission mechanism
PT and resonance produced between the compressor C and the external
devices coupled to the pulley 17 by the belt 18 are suppressed.
[0069] Since the structure for suppressing resonance is provided in
the power transmission mechanism PT, the rotary shaft 16 need not
have any means for suppressing resonance. This reduces the weight
and the size of the compressor C.
[0070] (2) The spokes 42B are located in the power transmission
path between the pulley 17 and the boss 42A. When the transmission
torque is excessive, the spokes 42B are broken so that the power
transmission is disconnected. That is, in this embodiment, the
power transmission between the first rotor (the pulley 17) and the
second rotor (the boss 42A) is disconnected when the breakable
members (the spokes 42B) are broken. Compared to a case where a
coupler is located in the power transmission path between a first
rotor and a second rotor to couple to the rotors and the coupler
can be detached from at least one of the rotors, the present
embodiment reduces the number of the parts.
[0071] (3) The torque receiving member 42 (the spokes 42B) is
formed with sintered metal. Since the ductility of the sintered
metal is relatively low, the threshold level of the transmission
torque at which the spokes 42B are broken is easily determined.
Also, the fatigue ratio .sigma..sub.w/.sigma..sub.B of the sintered
metal is easily set high. Therefore, the durability of the spokes
42B to withstand repetitive stress in the normal power transmission
state is set relatively high. Also, the balance between the
durability of the spokes 42B and the level of the transmission
torque at which the spokes 42B are broken is easily optimized.
Accordingly, it is easy to design the mechanism such that the
spokes 42B have a satisfactory durability and do not break for the
transmission torque in the normal transmission state, and break
when the transmission torque is excessive.
[0072] (4) The pulley 17 has multiple types of weights (the rollers
47, 48) so that a plurality of components of rotation order in the
rotation vibration produced in the pulley 17 are suppressed.
Specifically, the greatest peak and the second greatest peak of the
torque fluctuation are suppressed. Compared to a structure that
suppresses only one of the components of the rotation order, the
torque fluctuation is effectively suppressed in the normal power
transmission state.
[0073] A second embodiment of the present invention will now be
described. The second embodiment is the same as the first
embodiment except for the structure of the power transmission
mechanism PT. Mainly, the differences from the first embodiment
will be discussed below, and same or like reference numerals are
given to parts that are the same as or like corresponding parts of
the first embodiment.
[0074] In the power transmission mechanism PT of the second
embodiment, which is shown in FIGS. 3(a) and 3(b), the pulley 17
rotates relative to the torque receiving member 42. An annular
elastic member (shock absorbing member), or a rubber damper 51, is
located in the power transmission path between the pulley 17 and
the torque receiving member 42.
[0075] The torque receiving member 42 of the second embodiment has
a boss 42A, a cylindrical damper seat 42D, and spokes 42B. The
torque receiving member 42 is attached to the rotary shaft 16 at
the boss 42A. The damper seat 42D is integrally formed with the
boss 42A and the spokes 42B and is located forward of the support
cylinder 40.
[0076] A radial bearing 40A is fitted about the support cylinder
40. The outer ring of the radial bearing 40A is directly attached
to the inner surface of the inner cylinder 17B of the pulley 17
without any member corresponding to the pulley seat 42C of the
first embodiment. That is, the pulley 17 is rotatably supported by
the housing. Also, the pulley 17 rotates relative to the rotary
shaft 16 and the torque receiving member 42 with the rotation axis
of the pulley 17 is coaxial with those of the rotary shaft 16 and
the torque receiving member 42.
[0077] The inner cylinder 17B of the pulley 17 includes an extended
portion 17C, which is integrally formed with the inner cylinder 17B
and protrudes forward. The rubber damper 51 couples the inner
surface of the extended portion 17C and the outer surface of the
damper seat 42D.
[0078] Torque fluctuation is suppressed in a manner similar to the
first embodiment. That is, when torque fluctuation is produced
while the rotary shaft 16 is rotating, the rollers 47, 48 swing
like pendulums and produce torques. The produced torques suppress
the torque fluctuation.
[0079] Further, since the pulley 17 is coupled to the rotary shaft
16 (the torque receiving member 42) by the rubber damper 51, torque
fluctuation transmitted from the torque receiving member 42 to the
pulley 17 is attenuated. As a result, the resonance produced by the
torque fluctuations is effectively suppressed.
[0080] The rotation axes of the pulley 17 and the torque receiving
member 42 may be displaced from each other due to errors. However,
since the rubber damper 51 is located between the pulley and the
torque receiving member 42 (the rotary shaft 16), stress applied to
the radial bearings 12A, 40A due to the displacement of the axes is
reduced.
[0081] The rubber damper 51 functions effectively when the
frequency of the torque fluctuation is relatively high. The rollers
47, 48 function effectively when the frequency of the torque
fluctuation is relatively low.
[0082] In addition to the advantages (1), (2), (3), and (4), the
second embodiment has the following advantages.
[0083] (5) The rubber damper (damper member) 51 is located in the
power transmission path between the pulley 17 and the torque
receiving member 42, which attenuates the torque fluctuation
transmitted from the torque receiving member 42 to the pulley 17.
That is, in addition to the rollers 47, 48, the rubber damper 51
functions as a damper. Therefore, the resonance is effectively
suppressed.
[0084] (6) The damper 51 (elastic member) is located between the
pulley 17 and the torque receiving member 42, or in the power
transmission path in between. The rotation axes of the pulley 17
and the torque receiving member 42 may be displaced from each other
due to errors. However, deformation of the rubber damper 51 reduces
stress applied to the radial bearings 12A, 40A due to the
displacement of the axes. Therefore, the durability of the rotary
machine, which includes the power transmission mechanism PT and the
compressor C, is improved.
[0085] A third embodiment of the present invention will now be
described. The third embodiment is the same as the second
embodiment except for the structure of the power transmission
mechanism PT. Mainly, the differences from the second embodiment
will be discussed below, and same or like reference numerals are
given to parts that are the same as or like corresponding parts of
the second embodiment.
[0086] As shown in FIGS. 4(a) and 4(b), the second rotor, which is
a torque receiving member 55 in the third embodiment, is secured to
the front end of the rotary shaft 16 to rotate integrally with the
rotary shaft 16. The torque receiving portion 55 includes a boss
55A and a circular hub 55B. The boss 55A is fitted in the support
cylinder 40 and is located forward of the lip seal 41. The hub 55B
is integrally formed with the boss 55A and is located forward of
the support cylinder 40.
[0087] Unlike the torque receiving member 42 of the first and
second embodiments, the torque receiving member 55 of the third
embodiment is not made of sintered metal but is made of a common
iron material.
[0088] The hub 55B substantially covers the opening of each weight
receptacle 45, 46. In the third embodiment, the lid 49 of the first
and second embodiments is omitted. Instead, the hub 55B prevents
the rollers 47, 48 from falling off the weight receptacles 45,
46.
[0089] A damper receptacle 56 is formed between each adjacent pair
of the weight receptacles 45, 46. That is, the pulley 17 has the
six damper receptacles 56.
[0090] In the third embodiment, the rubber damper 51 of the second
embodiment, which is located between the outer surface of the
damper seat 42D and the extended portion 17C, is omitted. Instead,
tubular rubber dampers 57 each having a circular cross-section are
provided. That is, each rubber damper 57 is fitted into one of the
damper receptacles 56 and functions as an elastic member (damper
member). The outer surface of each rubber damper 57 closely
contacts the inner surface of the corresponding damper receptacle
56.
[0091] Each rubber damper 57 has a through hole 57A, the
cross-section of which is circular. The hub 55B has power
transmission pins 58 projecting rearward. The pings 55B function as
power transmission disconnecting body (breakable members). Each pin
58 corresponds to one of the rubber dampers 57. The rear end (right
end as viewed in the drawings) of each pin 58 is fitted in the
through hole 57A of the corresponding damper 57. Each pin 58 is
press fitted in a hole formed in the peripheral portion of the hub
55B, and extends in the axial direction of the torque receiving
member 55. The number of the pins 58 is six in this embodiment.
[0092] The power transmission pins 58 are made of sintered metal.
The fatigue ratio .sigma..sub.W/.sigma..sub.B of the sintered metal
is about 0.5. The sign .sigma..sub.w represents the fatigue limit
and the sign .sigma..sub.B represents the tensile strength.
[0093] Power transmitted from the engine E to the pulley 17 is
transmitted to the torque receiving member 55 through the rubber
dampers 57 and the power transmission pins 58. The rubber dampers
57 and the power transmission pins 58 are located in the power
transmission path between the pulley 17 and the torque receiving
member 55. The rubber dampers 57 attenuate the torque fluctuation
transmitted from the torque receiving member 55 to the pulley
17.
[0094] As long as the magnitude of the transmission torque between
the pulley 17 and the torque receiving member 55 does not adversely
affect the engine E, that is, as long as the transmission torque is
within a normal power transmission state, the power is transmitted
from the engine E to the rotary shaft 16.
[0095] However, if there is an abnormality in the compressor C, for
example, if the compressor C is locked, and the transmission torque
is excessive, the power transmission pins 58 break due to the
excessive load. Accordingly, the power transmission from the pulley
17 to the hub 55B is disconnected. This prevents the engine E from
being adversely affected by the excessive torque.
[0096] In addition to the advantages (1) to (5) and (6), the third
embodiment has the following advantages.
[0097] (7) Each rubber damper 57 is located between each adjacent
pair of the weight receptacles 45, 46. The spaces between the
weight receptacles (45, 46) are effectively used for providing
rubber dampers. The structure of the third embodiment reduces the
axial size of the power transmission mechanism PT compared to the
second embodiment.
[0098] (8) The hub 55B of the torque receiving member 55 prevents
the rollers 47, 48 from falling off the weight receptacles (45,
46). Therefore, there is no need for providing an additional member
for preventing the rollers (47, 48) from falling, such as the lid
49 in the first and second embodiments. This reduces the number of
the parts and thus reduces the costs.
[0099] (9) Since the breakable members are the power transmission
pins 58, the structure of the breakable members and the structure
of the elastic members (rubber dampers 57) fitted about the
breakable members are simplified.
[0100] A fourth embodiment will now be described. The fourth
embodiment is greatly different from the third embodiment in the
power transmission mechanism PT. Particularly, the power
transmission mechanism PT of the fourth embodiment has no rubber
member between the pulley and the torque receiving member. Instead,
coupling members including leaf springs are provided. Mainly, the
differences from the third embodiment will be discussed below, and
same or like reference numerals are given to parts that are the
same as or like corresponding parts of the third embodiment.
[0101] As shown in FIGS. 5(a) and 5(b), the pulley 17 (the first
rotor) of the fourth embodiment includes a belt receiving portion
17A and an annular base portion 17D. The base portion 17D projects
inward from the inner surface of the belt receiving portion 17A.
The base portion 17D is coupled to the support cylinder 40 with a
radial bearing 40A such that the pulley 17 rotates relative to the
support cylinder 40. The pulley 17 is coaxial with the rotary shaft
16 and rotates relative to the rotary shaft 16.
[0102] The second rotor, which is a torque receiving member 61 in
the fourth embodiment, is secured to the front end of the rotary
shaft 16 to rotate integrally with the rotary shaft 16. The torque
receiving member 61 has a substantially circular hub 61A.
[0103] Support pins 61B, the number of which is four in this
embodiment, are fixed to the peripheral portion of the rear face of
the hub 61A. The support pins 61B are spaced by equal angular
intervals (ninety degrees in this embodiment). The proximal end of
a power transmission arm 62, which is a leaf spring, is wound about
each support pin 61B at a cylindrical sleeve. The power
transmission arms 62 function as coupling members. The proximal end
of each power transmission arm 62 is rotated with the sleeve
relative to the corresponding support pin 61B when rotational force
about the pine 61B having a certain level is applied to the
proximal end.
[0104] Engaging pins 17E, the number of which is four in this
embodiment, are fixed to the peripheral portion of the front face
of the pulley base portion 17D. The engaging pins 17E are spaced by
equal angular intervals (ninety degrees in this embodiment). A
cylindrical sleeve 17F is rotatably supported by each engaging pin
17E. The sleeves 17F are located closer to the peripheral portion
of the pulley 17 than to the outer surface of the hub 61A.
[0105] The distal end of each power transmission arm 62 is engaged
with the outer surface of one of the sleeves 17F. The distal
portion of each power transmission arm 62 is hooked on the
corresponding sleeve 17F. That is, each power transmission arm 62
extends through a position radially outside of the corresponding
sleeve 17F and then is curved radially inward toward the rotation
axis of the pulley 17. Specifically, each power transmission arm 62
is hooked on the corresponding cylindrical portion 17F at a hook
62A formed at the distal end. Each power transmission arm 62 is
engaged with the corresponding sleeve 17F of the pulley 17 at the
hook 62A. The torque receiving member 61 and the pulley 17 can
rotate relative to each other in a predetermined angular range. As
long as the power transmission arms 62 are engaged with the sleeves
17F, power (torque) can be transmitted from the pulley 17 to the
torque receiving member 61.
[0106] At least the power transmission arms 62 and the sleeves 17F
form the power transmission disconnecting body.
[0107] Three weight guides, which are weight receptacles 63, are
located between each adjacent pair of the engaging pins 17E. The
weight receptacles 63 are formed in the pulley base portion 17D.
Only one of the weight receptacles 63 is shown in FIG. 5(b). The
weight receptacles 63 in each set are angularly spaced by
substantially equal intervals between the corresponding pair of the
engaging pins 17E to minimize the displacement of the center of
gravity of the pulley 17 from the rotation axis of the pulley
17.
[0108] Each weight receptacle 63 has a weight guide surface 63A.
The cross-section of each of the guide surfaces 63A is arcuate
along a plane perpendicular to the rotation axis of the pulley 17.
Each guide surface 63A forms a part of an imaginary surface.
[0109] A weight, which is a rigid roller 64 in this embodiment, is
accommodated in each weight receptacle 63. Each roller 64 rolls in
the circumferential direction along the weight guide surface 63A of
the corresponding weight receptacle 63. Four arcuate lids 65 are
fixed to the front face of the pulley 17 by bolts. Each lid 65
covers a set of the weight receptacles 63 to prevent the
corresponding 64 from falling off the weight receptacles 63.
[0110] When the torque fluctuation is produced while the rotary
shaft 16 is rotating, each roller 64 reciprocates along the weight
guide surface 63A in the corresponding weight receptacle 63. In
other words, the roller 64 swings like a pendulum the fulcrum of
which coincides with the axis of the imaginary cylinder containing
the guide surface 63A.
[0111] The size and mass of the rollers 64 and the locations of the
rollers 64 in the pulley 17 are determined such that the torque
fluctuation is suppressed by pendulum motion of the rollers 64.
Since the settings of the rollers 64 are similar to those of the
rollers in the first to third embodiments, the description is
omitted.
[0112] The pulley 17 (the weight receptacles 63) and the rollers 64
form a dynamic damper.
[0113] In FIGS. 5(a) and 5(b), centrifugal force produced by
rotation of the rotary shaft 16 causes each roller 64 to contact
the corresponding guide surface 63A.
[0114] When the torque fluctuation is produced, each roller 64
starts swinging like a pendulum. The pendulum motion of each roller
64 generates torque about the rotation axis of the pulley 17. The
produced torques suppress the torque fluctuation.
[0115] As long as the magnitude of the transmission torque between
the pulley 17 and the torque receiving member 61 does not adversely
affect the engine E, that is, as long as the transmission torque is
within a normal power transmission state, the engagement between
the hooks 62A and the sleeves 17F is maintained and the power is
transmitted from the engine E to the rotary shaft 16.
[0116] However, if there is an abnormality in the compressor C, for
example, if the compressor C is locked, and the transmission torque
is excessive, the elastic force of the power transmission arms 62
cannot maintain the engagement between the arms 62 and the sleeves
17F. Thus, each sleeve 17F comes off the corresponding hook 62A,
which disengages the sleeve 17F from the arm 62. Accordingly, the
power transmission from the pulley 17 to the torque receiving
member 61 is disconnected (see FIG. 6). This prevents the engine E
from being adversely affected by the excessive torque.
[0117] When the sleeves 17F are disengaged from the hooks 62A, the
pulley 17 freely rotates relative to the torque receiving member
61. This causes each sleeve 17F to contact the back of one of the
power transmission arms 62. Accordingly, each power transmission
arm 62 receives a pivotal force about the corresponding support pin
61B due to the contact with the sleeve 17F. Then, each power
transmission arm 62, together with the sleeve about which the
proximal end of the arm 62 is wound, is rotated clockwise about the
support pin 61B as viewed in FIG. 7. This moves the power
transmission arm 62 to a position where the arm 62 does not contact
the sleeves 17F.
[0118] Each power transmission arm 62 and the corresponding sleeve
are rotated about the corresponding support pin 61B when receiving
a rotational force that is greater than a predetermined level.
Therefore, even if receiving an external force (for example, force
based on vibration of the vehicle), the power transmission arms 62
are held at the position to avoid contacting the sleeves 17F.
Therefore, each power transmission arm 62 does not contact the
sleeves 17F, which consecutively pass over the arm 62. This
prevents noise and vibration due to interference of the sleeves 17F
with the arm 62.
[0119] When the torque fluctuation is produced while power is being
transmitted between the pulley 17 and the torque receiving member
61, the magnitude of the force applied to each power transmission
arm 62 by the corresponding sleeve 17F is repeatedly increased and
decreased. Also, the contact position of the sleeve 17F on the hook
62A repeatedly moved. The degree of the elastic deformation of each
power transmission arm 62 is changed in accordance with the changes
of the force applied by the corresponding sleeve 17F. This reduces
the fluctuation of the transmission torque between the pulley 17
and the torque receiving member 61. Also, friction between each
engaging pin 17E and the corresponding sleeve 17F and friction
between each power transmission arm 62 and the corresponding sleeve
17F attenuate the fluctuation of the transmission torque between
the pulley 17 and the torque receiving member 61. That is, the
power transmission arms 62 function as shock absorbing member
(elastic members).
[0120] In addition to the advantage (1), the third embodiment has
the following advantages.
[0121] (10) Friction between each engaging pin 17E and the
corresponding sleeve 17F and friction between each power
transmission arm 62 and the corresponding sleeve 17F attenuates the
torque fluctuation transmitted from the torque receiving member 61
to the pulley 17. That is, in addition to the rollers 64, the
frictions attenuate the torque fluctuation. Therefore, the
resonance is effectively suppressed.
[0122] (11) The power transmission arms 62 (elastic members) are
located between the pulley 17 and the torque receiving member 61,
or in the power transmission path in between. The rotation axes of
the pulley 17 and the torque receiving member 61 may be displaced
from each other due to errors. However, deformation of the power
transmission arms 62 reduces stress applied to the radial bearings
12A, 40A due to the displacement of the axes. Therefore, the
durability of the rotary machine, which includes the power
transmission mechanism PT and the compressor C, is improved.
[0123] (12) Power is transmitted between the pulley 17 and the
torque receiving member 61 when the power transmission arms 62 are
engaged with the sleeves 17F. The power transmission is
disconnected when the power transmission arms 62 are disengaged
from the sleeves 17F. Therefore, the power transmission
disconnecting body can be reused without replacing any parts.
Therefore, compared to a structure in which power transmission
between the pulley 17 and the rotary shaft 16 is disconnected by
breaking breakable members, the structure of the fourth embodiment
reduces the costs for reusing the power transmission disconnecting
body.
[0124] (13) The coupling members (the power transmission arms 62)
are formed with leaf springs. This simplifies the structure of the
coupling members. Also, the leaf springs reduces the fluctuation of
the transmission torque between the pulley 17 and the torque
receiving member 61.
[0125] It should be apparent to those skilled in the art that the
present invention may be embodied in many other specific forms
without departing from the spirit or scope of the invention.
Particularly, it should be understood that the invention may be
embodied in the following forms.
[0126] In the first embodiment, a portion of the boss 42A that is
located forward of the support cylinder 40 may be radially
enlarged, and a dynamic damper having weights may be provided in
the enlarged portion. In this case, the dynamic damper need not be
provided in the pulley 17.
[0127] In the second, third and fourth embodiments, the dynamic
damper is provided in the pulley 17. However, the dynamic damper
may be provided in the torque receiving member (42, 55, 61).
Alternatively, both of the pulley 17 and the torque receiving
member (42, 55, 61) may have the dynamic damper.
[0128] In the illustrated embodiments, spherical weights may be
used.
[0129] In the illustrated embodiments, the number of weight
receptacles (45, 46, 63), in which the rollers (47, 48, 64) are
provided, may be changed. The number of the weight receptacles need
not correspond to the cylinder bores of the compressor C.
[0130] In the illustrated embodiments, the cross-sectional shape of
each weight receptacle (45, 46, 63) along a plane perpendicular to
the rotation axis of the pulley 17 may be circular. This
facilitates machining of the weight receptacles (45, 46, 63).
[0131] In the illustrated embodiments, the ratios R/r (the square
root of the ratios R/r) may be equalized for all the weights. That
is, the square root of the ratio R/r for all the rollers may be
equalized at N, which is the product of n.multidot.N when n is
one.
[0132] In the illustrated embodiments, the ratios R/r (the square
root of the ratios R/r) may be three or more different values. In
this case, since there are three or more values that correspond to
the ratios R/r, three or more peaks of the torque fluctuation are
suppressed. In this case, the values of n are preferably selected
from three of more numbers in order from one. For example, when
three numbers are selected, one, two and three are preferably used.
Accordingly, the square roots of the ratios R/r correspond to the
numbers represented by the products n.multidot.N, in which n is
one, two and three. Therefore, the three or more greatest peaks of
the torque fluctuation are suppressed. That is, the resonance is
effectively suppressed.
[0133] In the illustrated embodiments, the weight guide surface
(45A, 46A, 63A) is formed in each of the weight receptacles (45,
46, 63) formed in the pulley 17, and each roller (47, 48, 64)
swings like a pendulum along the corresponding guide surface.
However, the pulley may have weights each of which is coupled to a
fulcrum pin fixed to the pulley and swings like a pendulum.
Alternatively, each weight may have a fulcrum pin, which is engaged
with a hole formed in the pulley. In this case, each weight swings
like a pendulum about the pin.
[0134] In the illustrated embodiments, the settings are determined
by regarding each weight as a particle at the center of gravity.
However, the settings are preferably determined by taking the
inertial mass of each weight into consideration. For example, in
the case of the rollers (47, 48, 64), the settings are preferably
made based on the ratio 2R/3r instead on the ratio R/r to take the
inertial mass into consideration. In this case, the equations 1 and
2 are replaced with the following equations 3 and 4.
T.sub.1=(3/2).multidot.m.sub.t1.multidot.(.omega..sub.a1)2.multidot.(R+r).-
multidot.R.multidot..phi..sub.1 (Equation 3)
T.sub.2=(3/2).multidot.m.sub.t2.multidot.(.omega..sub.a2).sup.2.multidot.(-
R+r).multidot.R.multidot..phi..sub.2 (Equation 4)
[0135] If spherical weights are used instead of the rollers (47,
48, 64), the settings are preferably made based on the ratio 5R/7r
to take the inertial mass into consideration. When the frequency of
a peak of the torque fluctuation is equal to the characteristic
frequency of each spherical weight, the torque T is represented by
the following equation.
T=(7/5).multidot.m.multidot.(.omega..sub.a).sup.2.multidot.(R+r).multidot.-
R.multidot..phi. (Equation 5)
[0136] In the equation 5, the sign m represents the total mass of
the spherical weights, and the sign .omega..sub.a represents the
average angular velocity of the spherical weights when the weights
swing in a minute angle .phi..
[0137] If the weights are not formed cylindrical or spherical, the
settings are preferably made by taking the inertial mass of the
weights into consideration so that the resonance is effectively
suppressed.
[0138] In the first and second embodiments, the fatigue ratio
.sigma..sub.W/.sigma..sub.B of the sintered metal forming the
breakable members (42B, 58) is substantially 0.5. However, the
fatigue ratio .sigma..sub.W/.sigma..sub.B may be changed.
Specifically, the fatigue ratio .sigma..sub.W/.sigma..sub.B may be
any value as long as the breakable members are broken when
receiving excessive torque.
[0139] In the first to third embodiments, the breakable members
(42B, 58) are made of sintered metal. However, the breakable
members may be made of other material. For example, the breakable
members may be made of low-carbon steel. The fatigue ratio
.sigma..sub.W/.sigma..sub.B of low-carbon steel is easily set high
(approximately 0.5). Therefore, the durability of the breakable
members (42B, 58) to withstand repetitive stress in the normal
power transmission state is set relatively high. Also, the balance
between the durability of the breakable members (42B, 58) and the
level of the transmission torque at which the breakable members
(42B, 58) are broken is easily optimized. Accordingly, it is easy
to design the apparatus such that the breakable members (42B, 58)
have a satisfactory durability and do not break for the
transmission torque in the normal transmission state, and break
when the transmission torque is excessive.
[0140] In the first to third embodiments, the breakable members
(42B, 58) are made of metal. However, the breakable members may be
made of other material. Specifically, as long as the breakable
members are broken when receiving a torque that exceeds a threshold
level, any material such as resin or ceramic may be used for the
breakable members.
[0141] In the third embodiment, the power transmission pins 58 may
be integrally formed with the hub 55B. In this case, if the torque
receiving member 55, with which the power transmission pins 58 are
integrally formed, is made of breakable material such as sintered
metal, the power transmission is disconnected when the pins 58 are
broken.
[0142] In the third embodiment, the number of the power
transmission pins 58 is six and equal to the number of the weight
receptacles 45, 46. However, the number of the pins 58 may be
different from the number of the weight receptacles 45, 46. The
number of the pins 58 located between an adjacent pair of weight
receptacles may be changed. No pins 58 may be provided between one
of the adjacent pairs of the weight receptacles.
[0143] In the third embodiment, each power transmission pin 58 is
coupled to the pulley 17 with the corresponding rubber damper 57.
However, the power transmission pins 58 may be directly coupled to
the pulley 17 without the rubber dampers 57.
[0144] In the third embodiment, each power transmission pin 58 is
fixed to the hub 55B and coupled to the pulley 17 with the
corresponding rubber damper 57. However, the pins 58 may be fixed
to the pulley 17 and coupled to the hub 55B with rubber
dampers.
[0145] In the third embodiment, each power transmission pin 58 may
be coupled to the hub 55B with a tubular shock absorbing member
(rubber damper). In other words, the power transmission pins 58 may
be coupled to both of the hub 55B and the pulley 17 with shock
absorbing members such as rubber dampers.
[0146] In the third embodiment, the rubber dampers 57 are tubular
and have circular cross-section. However, the cross-section of the
rubber dampers 57 may be changed.
[0147] In the third embodiment, a space may exist between the inner
surface of each damper receptacle 56 and the corresponding rubber
damper 57 or between the inner surface of each rubber damper 57 and
the outer surface of the corresponding power transmission pin 58.
That is, narrow spaces may exist as long as the power transmitting
performance and the durability of the power transmission mechanism
PT are not adversely affected.
[0148] In the second and third embodiments, the rubber dampers (51,
57) are used. However, dampers made of elastomer may be used.
[0149] In the third embodiment, the rubber dampers 57 are used as
shock absorbing members. However, sealed gel containers may be used
as shock absorbing members and located in the power transmission
path between the pulley 17 and the torque receiving member 42. For
example, a cylindrical gel container may be located in each damper
receptacle 56 and the rear end of the corresponding power
transmission pin 58 may be fitted in a hole formed in the gel
container. In this structure, gel sealed in the gel containers
attenuates the torque fluctuation transmitted from the torque
receiving member 55 to the pulley 17 through the power transmission
pins 58. The rotation axes of the pulley 17 and the torque
receiving member 55 may be displaced from each other due to errors.
However, deformation of the gel containers reduces stress applied
to the radial bearings 12A, 40A due to the displacement of the
axes. Instead of the gel containers, containers having fluid such
as liquid or gas may be used as shock absorbing members. If fluid
is used, the same advantages as the case where gel is used are
achieved. However, gel attenuates vibration by a greater degree
compared to fluid.
[0150] In the fourth embodiment, the power transmission arms 62 are
formed with leaf springs. However, rigid arms may be used as
coupling members. In this case, the proximal end of each rigid arm
is rotatably coupled to the corresponding support pin 61B and a
hook is formed at the distal end of the arm. The hook is engaged
with the corresponding sleeve 17F. An urging member such as a
spring is provided for each rigid arm to urge the hook to the
sleeve 17F.
[0151] In the fourth embodiment, the power transmission arms 62 are
formed with leaf springs. However, the power transmission springs
62 may be formed with wire springs.
[0152] In the fourth embodiment, the proximal end of each coupling
member (62) may be supported at the pulley 17 and engaging pins may
project from the hub 61A. In this case, a cylindrical member is
fitted about each engaging pin, and the hook (62A) of each coupling
member (62) is engaged with the corresponding cylindrical
member.
[0153] The number of cylinder bores 24 in the compressor C may be
changed. A typical compressor for a vehicular air conditioner has
three to seven cylinder bores. If the number of the cylinder bores
24 is three, the fluctuation of torque transmitted between the
pulley 17 and the torque receiving member 42 due to rotational
vibration produced in the rotary shaft 16 is greater compared to a
case where the number of the cylinders 24 is four or greater. That
is, in a rotary machine that has three cylinder bores, the dynamic
dampers and the shock absorbing members of the illustrated
embodiment effectively suppress resonance.
[0154] In the illustrated embodiment, the power transmission
mechanism PT is used for the compressor C, which has the single
headed pistons 25. However, the mechanism PT may be used for a
compressor that has double-headed pistons. In this type of
compressor, cylinder bores are formed on either side of a crank
chamber and each piston compresses gas in one of the pairs of the
cylinder bores.
[0155] In the illustrated embodiment, the cam plate (the swash
plate 20) rotates integrally with the rotary shaft 16. However, the
present invention may be applied to a compressor in which a cam
plate rotates relative to a rotary shaft. For example, the present
invention may be applied to a wobble type compressor.
[0156] The pulley 17 may be used in a fixed displacement type
compressor, in which the stroke of the pistons is constant.
[0157] In the illustrated embodiments, the present invention is
applied to a reciprocal piston type compressor. However, the
present invention may be applied to rotary compressors such as a
scroll type compressor
[0158] In the illustrated embodiment, the first rotor is the pulley
17. However, a sprocket or a gear may be used as the first
rotor.
[0159] In the illustrated embodiments, the bearing (40A) for
rotatably support the first rotor (17) is directly attached to the
housing, which support the rotary shaft 16. However, the bearing
(40A) may be attached to the housing with another structure in
between.
[0160] In the illustrated embodiments, the present invention is
applied to the compressor C. However, the present invention may be
applied to any apparatus that has a rotary shaft coupled to the
power transmission mechanism PT, and torsional vibration is
produced in the rotary shaft.
[0161] The axis of the pendulum motion of each weight need not be
parallel to the rotation axis of the rotor in which the weights are
provided. Specifically, the axis of the pendulum motion may be
inclined relative to the rotation axis of the rotor in a range in
which the maximum torque fluctuation is suppressed by a desirable
degree.
[0162] Therefore, the present examples and embodiments are to be
considered as illustrative and not restrictive and the invention is
not to be limited to the details given herein, but may be modified
within the scope and equivalence of the appended claims.
* * * * *