U.S. patent application number 10/490720 was filed with the patent office on 2004-12-16 for rotary machine.
Invention is credited to Adaniya, Taku, Arai, Tomoharu, Kanai, Akinobu, Kawaguchi, Masahiro, Kawata, Takeshi, Ota, Masaki.
Application Number | 20040253121 10/490720 |
Document ID | / |
Family ID | 19118449 |
Filed Date | 2004-12-16 |
United States Patent
Application |
20040253121 |
Kind Code |
A1 |
Ota, Masaki ; et
al. |
December 16, 2004 |
Rotary machine
Abstract
A torque receiving member is fixed to a rotary shaft of a
compressor. A pulley, which is rotated by an external drive source,
is rotatably supported by a housing of the compressor. The pulley
is substantially coaxial with the torque receiving member. A rubber
damper is located between the pulley and the torque receiving
member. The rubber damper absorbs rotational vibration transmitted
from the torque receiving member to the pulley. The pulley has a
dynamic damper. The dynamic damper includes rollers, which swing
like pendulums to reduce rotational vibration of the pulley. Stress
produced due to displacement between the axis of the pulley and the
torque receiving member is reduced by the rubber damper.
Inventors: |
Ota, Masaki; (Kariya-shi,
JP) ; Kawaguchi, Masahiro; (Kariya-shi, JP) ;
Adaniya, Taku; (Kariya-shi, JP) ; Kanai, Akinobu;
(Kariya-shi, JP) ; Arai, Tomoharu; (Kariya-shi,
JP) ; Kawata, Takeshi; (Kariya-shi, JP) |
Correspondence
Address: |
MORGAN & FINNEGAN, L.L.P.
3 WORLD FINANCIAL CENTER
NEW YORK
NY
10281-2101
US
|
Family ID: |
19118449 |
Appl. No.: |
10/490720 |
Filed: |
March 24, 2004 |
PCT Filed: |
September 27, 2002 |
PCT NO: |
PCT/JP02/10041 |
Current U.S.
Class: |
417/316 ;
417/222.2; 417/269 |
Current CPC
Class: |
F04B 27/0895 20130101;
F16D 1/06 20130101 |
Class at
Publication: |
417/316 ;
417/269; 417/222.2 |
International
Class: |
F04B 001/26; F04B
001/12; F04B 027/08 |
Foreign Application Data
Date |
Code |
Application Number |
Sep 27, 2001 |
JP |
2001-297358 |
Claims
1. A rotary machine comprising: a housing; a rotating shaft
rotatably supported by the housing; a first rotor rotatably fixed
to the rotating shaft and integrally rotatable therewith outside
the housing; a second rotor rotatably supported on the housing
outside the housing and operably connected to the first rotor in a
state substantially coaxial to the first rotor; and an absorption
mechanism, arranged on a connecting portion of the first rotor and
the second rotor, for absorbing misalignment between the axis of
the first rotor and the axis of the second rotor, the absorption
mechanism being formed from a rigid member.
2. The rotary machine as claimed in claim 1, wherein the absorption
mechanism includes: a support member rotatably supported on one of
the first rotor and the second rotor, the support member being
rotatably supported on that rotor about an axis parallel to a
rotation axis of the rotor supporting the support member; and an
eccentric pin, arranged relatively rotatable to at least one of the
other one of the first rotor and the second rotor and the support
member, for connecting the other one of the two rotors and the
support member, the eccentric pin being arranged eccentric to an
axis of the support member by an eccentricity amount that is
greater than an amount of axis misalignment between the two rotors
so that the axis of the other one of the two rotors does not
overlap the axis of eccentric pin.
3. The rotary machine as claimed in claim 1, wherein the absorption
mechanism includes: a connecting pin projecting from one of the two
rotors and having an axis parallel to a rotation axis of that
rotor; and an elongated hole-like or groove-like connecting portion
formed on the other one of the two rotors so as to generally extend
in a radial direction of that rotor and face the connecting pin,
wherein the connecting pin is inserted into the connecting portion
so as to be slidable along the connecting portion and rotatable
within the connecting portion.
4. The rotary machine as claimed in claim 1, further comprising a
compression mechanism for compressing refrigerant based on rotation
of the rotating shaft.
5. The rotary machine as claimed in claim 4, wherein the
compression mechanism is a piston-type compression mechanism.
6. The rotary machine as claimed in claim 4, wherein the
compression mechanism is formed so that the amount of refrigerant
discharges per one rotation of the rotating shaft is variable.
7. The rotary machine as claimed in claim 5, wherein the
compression mechanism is formed so that the amount of refrigerant
discharges per one rotation of the rotating shaft is variable.
8. The rotary machine as claimed in claim 2, further comprising a
compression mechanism for compressing refrigerant based on rotation
of the rotating shaft.
9. The rotary machine as claimed in claim 3, further comprising a
compression mechanism for compressing refrigerant based on rotation
of the rotating shaft.
Description
BACKGROUND OF THE INVENTION
[0001] The present invention relates to a rotary machine that
includes a housing, a rotary shaft, which is rotatably supported by
the housing, a first rotor, which is fixed to the rotary shaft, and
a second rotor, which is rotatably supported by the housing.
Specifically, the present invention pertains to a rotary machine
that includes a first and second rotors, which are substantially
coaxial and coupled to each other.
[0002] For example, a typical vehicular compressor driven by an
external drive source such as a vehicle engine has a first rotor,
which is fixed to a rotary shaft for driving the compressor, and a
second rotor, which is coupled to the first rotor and to the engine
by a belt. Typically, the second rotor is rotatably supported by
the housing of the compressor with a bearing. This structure
reduces radial stress applied to the rotary shaft due to the
tension of the belt. Accordingly, the stress received by the
bearing supporting the rotary shaft is reduced.
[0003] Japanese Laid-Open Patent Publication No. 2000-297844
discloses such a structure. The structure of the publication
includes an inertial weight (second rotor) driven by an external
drive source and a hub (first rotor) fixed to a crankshaft (rotary
shaft). The inertial weight is coupled to the hub with an annular
elastomer (damping member).
[0004] The annular elastomer attenuates the rotational vibration
transmitted from the crankshaft to the inertial weight. Further,
the hub has a variable frequency vibration absorbing system. The
system includes centrifugal weights, which swing like pendulums to
reduce rotational vibration of the crankshaft. Accordingly, the
resonance between the first rotor and the second rotor is
suppressed.
[0005] If the first rotor is fixed to a rotary shaft and the second
rotor is rotatably supported by a housing, which supports the
rotary shaft, the displacement between the axes of the first and
second rotors needs to be absorbed (allowed). If there is no
structure for absorbing the axial displacement, the durability of
bearings supporting the rotary shaft and the durability of bearings
supporting the second rotor are degraded. The structure of the
publication No. 2000-297844 does not absorb such axial
displacement.
SUMMARY OF THE INVENTION
[0006] Accordingly, it is an objective of the present invention to
provide a rotary machine that suppresses resonance between a first
rotor and a second rotor and improves the durability of the rotary
machine.
[0007] To achieve the foregoing and other objectives and in
accordance with the purpose of the present invention, a rotary
machine driven by an external drive source is provided. The rotary
machine includes a housing, a rotary shaft, a first rotor, a second
rotor, a coupling member, and a dynamic damper. The rotary shaft is
rotatably supported by the housing. The first rotor is fixed to the
rotary shaft to rotate integrally with the rotary shaft. The second
rotor is rotatably supported by the housing. The second rotor is
substantially coaxial with the first rotor and is rotated by the
external drive source. The coupling member couples the rotors to
each other to transmit power from the second rotor to the first
rotor. The coupling member includes a damping member. 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.
[0008] 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
[0009] 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:
[0010] FIG. 1 is a cross-sectional view illustrating a compressor
having a power transmission mechanism according to a first
embodiment of the present invention;
[0011] FIG. 2(a) is a front view illustrating the power
transmission mechanism of the compressor shown in FIG. 1;
[0012] FIG. 2(b) is a cross-sectional view taken along line
2(b)-2(b) of FIG. 2(a);
[0013] 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);
[0014] FIG. 4(a) is a front view illustrating a power transmission
mechanism according to a second embodiment; FIG. 4(b) is a
cross-sectional view taken along line 4(b)-4(b) of FIG. 4(a);
[0015] FIG. 5 is a front view illustrating a power transmission
mechanism according to another embodiment;
[0016] FIG. 6 is a front view illustrating a power transmission
mechanism according to another embodiment (the hub is omitted for
purposes of illustration); and
[0017] FIG. 7 is a front view illustrating a power transmission
mechanism according to another embodiment (the hub is omitted for
purposes of illustration).
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0018] A compressor C according to one 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.
[0019] 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.
[0020] The cylinder block 11 and the front housing member 12 define
a crank chamber 15 in between.
[0021] A rotary shaft, which is a drive shaft 16 in this
embodiment, is housed in the compressor housing and extends through
the crank chamber 15. The front portion of the drive 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 drive shaft 16 is
supported by a radial bearing 11A located in the cylinder block
11.
[0022] A cylindrical support 40 is formed at the front end of the
front housing member 12. The front end portion of the drive shaft
16 extends through the front wall of the front housing member 12
and is located in the cylindrical support 40. A power transmission
mechanism PT is fixed to the front end of the drive shaft 16. The
power transmission mechanism PT includes a pulley 17. The front end
of the drive 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 engaged with the
pulley 17. The power transmission mechanism PT and the compressor
form a rotary machine.
[0023] A lug plate 19 is coupled to the drive shaft 16 and is
located in the crank chamber 15. The lug plate 19 rotates
integrally with the drive 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 drive
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 drive shaft 16 and to incline with
respect to the drive shaft 16 while sliding along the rotation axis
of the drive shaft 16.
[0024] A snap ring 22 is fitted about the drive 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 drive shaft 16 is closest to ninety degrees.
[0025] 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 drive 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 drive shaft 16, the shoes 26 converts
the rotation into reciprocation of each piston 25.
[0026] The cylinder block 11 (cylinder bores 24), the drive 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.
[0027] 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 32
correspond to one of the cylinder bores 24 (compression
chambers).
[0028] 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.
[0029] 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.
[0030] 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.
[0031] 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 drive shaft 16, is
varied.
[0032] As shown in FIGS. 1 to 2(b), the cylindrical support 40
protrudes from the front wall of the front housing member 12 and
surrounds the front portion of the drive shaft 16. The axis of the
circumference of the support cylinder 40 substantially coincides
with the axis of the drive shaft 16.
[0033] A lip seal 41 is located in the support cylinder 40 to fill
the space between the support cylinder 40 and the drive shaft 16.
The lip seal 41 prevents refrigerant from escaping the crank
chamber 15 through the space between the support cylinder 40 and
the drive shaft 16.
[0034] A first rotor, which is a torque receiving member 42 in this
embodiment, is secured to the front end of the drive shaft 16 to
rotate integrally with the drive shaft 16. The torque receiving
portion 42 includes a boss 42A and a circular hub 42B. The boss 42A
is fitted in the support cylinder 40 and is located forward of the
lip seal 41. The hub 42B is integrally formed with the boss 42A and
is located forward of the support cylinder 40.
[0035] A second rotor, which is the pulley 17 in this embodiment,
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. 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 inner cylinder 17B. That is, the pulley 17 is
rotatably supported by the housing. Also, the pulley 17 rotates
relative to the drive shaft 16 and the torque receiving member 42
with the rotation axis of the pulley 17 is coaxial with those of
the drive shaft 16 and the torque receiving member 42.
[0036] The inner surface of the inner cylinder 17B and the outer
surface of the hub 42B are connected to each other by an annular
elastic member (damping member), which is a rubber damper 43 in
this embodiment. The rubber damper 43 is located in the power
transmission path between the pulley 17 and the torque receiving
member 42. The rubber damper 43 functions as a coupling member for
coupling the pulley 17 and the torque receiving member 42 to each
other. The rubber damper 43 is elastically deformed to allow the
pulley 17 to move circumferentially and radially relative to the
torque receiving member 42.
[0037] Six weight receptacles 45 (only one is shown in FIG. 1) are
formed in the pulley 17 between the belt receiving portion 17A and
the inner cylinder 17B. The receptacles 45 function as receiving
portions. The weight receptacles 45 are angularly spaced at the
constant intervals.
[0038] Each weight receptacle 45 has a weight guiding surface 45A.
The cross-section of each of the guiding surfaces 45A is arcuate
along a plane perpendicular to the rotation axis of the pulley 17.
Each weight guiding 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.
[0039] A weight, which is a rigid roller 46 in this embodiment, is
accommodated in each weight receptacle 45. The diameter and the
weight of each roller 46 are referred to as d.sub.1 and m.sub.1,
respectively. Each roller 46 rolls in the circumferential direction
along the weight guiding surface 45A of the corresponding weight
receptacle 45. An annular lid 47 is fixed to the front face of the
pulley 17 by bolts. The lid 47 covers the weight receptacles 45 to
prevent the rollers 46 from falling off the receptacles 45.
[0040] When the compressor C is being driven by the engine E, or
when the drive shaft 16 is rotating, centrifugal force causes each
roller 46 to contact the corresponding guiding surface 45A (see
FIGS. 1 to 2(b)). If torque fluctuation is generated due to, for
example, torsional vibrations of the drive shaft 16, each roller 46
starts reciprocating along the guiding surface 45A of the
corresponding receptacle 45. In other words, each roller 46 moves
along the circumferential direction of the guiding surface 45A.
That is, each roller 46, or the center of gravity of each roller
46, swings like a pendulum about the axis of an imaginary cylinder
that includes the corresponding guiding surface 45A. That is, each
roller 46 acts as a centrifugal pendulum when the compressor C is
being driven by the engine E. The size and mass of the rollers 46
and the locations of the rollers 46 in the pulley 17 are determined
such that the torque fluctuation is suppressed by pendulum motion
of the rollers 46.
[0041] The pulley 17 (the weight receptacles 45) and the rollers 46
form a dynamic damper.
[0042] The settings of the rollers 46, which function as
centrifugal pendulums, will now be described.
[0043] The rollers 46 suppress torque fluctuation when the
frequency of the fluctuation is equal to the characteristic
frequency of the roller 46 (centrifugal pendulum). Therefore, the
location, the size, and the mass of the rollers 46 are determined
such that the characteristic frequency of the rollers 46 is set
equal to the frequency of a peak component of the torque
fluctuation. Accordingly, the amplitude of the peak component 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.
[0044] The frequency of the torque fluctuation and the
characteristic frequency of the rollers 46 are proportional to the
angular velocity .omega..sub.1 of the drive shaft 16, which
corresponds to the speed of the drive shaft 16. The frequency of
the torque fluctuation when its range is the greatest is
represented by the product of the rotation speed of the drive 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- .) N.
[0045] The characteristic frequency of the rollers 46 is obtained
by multiplying the rotation speed of the drive 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 46 (weight). The
sign r represents the distance between the center of the pendulum
motion of each roller 46 and the center of gravity of the roller
46.
[0046] Therefore, by equalizing the square root of the ratio R/r
with the product n.multidot.N, the characteristic frequency of each
roller 46 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.
[0047] Accordingly, to suppress the greatest peak of the torque
fluctuations, the value of the signs R and r are determined such
that 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.
[0048] The torque produced about the rotation axis of the pulley 17
by the rollers 46 is represented by a sign T. To effectively reduce
peaks of the torque fluctuation by the pendulum motion of the
rollers 46, the torques T need to counter the torque fluctuation
and the amplitudes of the torques T need to be equal to the
amplitude of the peaks of the fluctuation. When the frequency of
the peak of the torque fluctuations is equal to the characteristic
frequency of the rollers 46, the torque T is represented by the
following equation.
T=m.multidot.(.omega..sub.a).sup.2.multidot.(R+r).multidot.R.multidot..phi-
. (Equation 1)
[0049] In the equation 1, the sign m represents the total mass of
the rollers 46 (m=6 m.sub.1), and the sign .omega..sub.a represent
the average angular velocity of the rollers 46 when the rollers 48
swing in a minute angle .phi..
[0050] In this embodiment, the mass m is maximized to minimize the
values R, r, and .phi., so that the size of the pulley 17 is
minimized, and the torque T is maximized.
[0051] The axis of each imaginary cylinder, which includes one of
the guiding surfaces 45A, coincides with the axis, or the fulcrum,
of the pendulum motion of the corresponding roller 46. That is, the
distance R.sub.1 between the rotation axis of the pulley 17 and the
axis of each imaginary cylinder corresponds to the distance R.
[0052] The distance between the axis of the pendulum motion of each
roller 46 and the center of gravity of the roller 46 is equal to
the value obtained by subtracting the half of the diameter d.sub.1
of the roller 46 from the radius r.sub.1 of the corresponding
imaginary cylinder. That is, the difference (r.sub.1-(d.sub.1/2))
corresponds to the distance r.
[0053] 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
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.
[0054] The settings are determined by regarding each roller 46 as a
particle at the center of gravity.
[0055] The operation of the compressor C will now be described.
[0056] When the power of the engine E is supplied to the drive
shaft 16 through the pulley 17, the swash plate 20 rotates
integrally with the drive 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.
[0057] 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.
[0058] During rotation of the drive shaft 16, the compression
reaction force of refrigerant and reaction force of reciprocation
of the pistons 25 are transmitted to the drive shaft 16 through the
swash plate 20 and the hinge mechanism 21, which torsionally
(rotationally) vibrates the drive 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.
[0059] When the torque fluctuations are generated, the rollers 46
start swinging like pendulums. The pendulum motion of the rollers
46 produces torques about the rotation axis of the pulley 17. The
produced torques suppress the torque fluctuation. The
characteristic frequency of the rollers 46 is equal to the
frequency of the greatest peak of the torque fluctuations.
Therefore, the peak of the torque fluctuations is suppressed, which
effectively reduce the torque fluctuations of the pulley 17.
[0060] Further, since the pulley 17 is coupled to the drive shaft
16 (the torque receiving member 42) by the rubber damper 43, 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.
[0061] The rotation axes of the pulley 17 and the torque receiving
member 42 may be displaced from each other. However, since the
rubber damper 43 is located between the pulley 17 and the torque
receiving member 42 (the drive shaft 16), stress applied to the
radial bearings 12A, 40A due to the displacement of the axes is
reduced.
[0062] The rubber damper 43 functions effectively when the
frequency of the torque fluctuation is relatively high. The rollers
46 function effectively when the frequency of the torque
fluctuation is relatively low.
[0063] The present embodiment has the following advantages.
[0064] (1) The rollers 46 are provided in the pulley 17. Each
roller 46 swings like a pendulum about its axis, which is spaced
from the rotation axis of the pulley 17 by the predetermined
distance R.sub.1 and is parallel to the rotation axis of the pulley
17. The pendulum motion of the rollers 46 suppresses the torsional
vibration (the torque fluctuation), which suppresses resonance
produced in the power transmission mechanism PT and the compressor
C. Further, the pendulum motion suppresses resonance produced
between the compressor C and the external devices that are coupled
to the pulley 17 by the belt 18.
[0065] The rubber damper 43 is located in the power transmission
path between the pulley 17 and the torque receiving member 42. The
rubber damper 43 is elastically deformed to allow the pulley 17 to
move circumferentially relative to 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 46, the rubber damper 43 functions as a damper. Therefore,
the resonance is effectively suppressed.
[0066] Since the structure for suppressing resonance is provided in
the power transmission mechanism PT, the drive shaft 16 need not
have any means for suppressing resonance. This reduces the weight
and the size of the compressor C.
[0067] (2) Radial stress is applied to the drive shaft 16 due to
the tension of the belt 18 coupling the pulley 17 with the engine
E. However, since the pulley 17 is supported by the housing, radial
stress applied to the drive shaft 16 is reduced.
[0068] (3) The damper 43 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 (the drive shaft 16) can be displaced from each
other. However, since the rubber damper 43 is elastically deformed
to allow the pulley 17 to move radially relative to the torque
receiving member 42, deformation of the rubber damper 43 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.
[0069] (4) The rollers 46, which are rigid cylinders, move along
the arcuate weight guiding surface 45A formed in the weight
receptacles 45 of the pulley 17. Since the rollers 46 are not fixed
to the fulcrum of the pendulum motion, the structure is simplified
compared to a structure in which weights are fixed to the fulcrums.
In a structure in which the weights are fixed to the fulcrums, the
distance between each weight and the corresponding pendulum axis
(fulcrum) varies due to space created between the fulcrum and the
hole formed in the weight for receiving the fulcrum. The structure
of the above embodiment has no such drawback. Therefore, the
resonance is effectively suppressed.
[0070] 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.
[0071] As shown in FIGS. 3(a) and 3(b), a torque receiving member
42 of the second embodiment has a hub 42B. Compared to the first
embodiment, the hub 42B has a greater diameter and substantially
covers the entire opening of each receptacle 45. In the second
embodiment, the lid 47 of the first embodiment is omitted. Instead,
the hub 42B prevents the rollers 46 from falling off the
receptacles 45.
[0072] In the second embodiment, the rubber damper 43, which is
located between the outer surface of the hub 42B and the inner
surface of the inner cylinder 17B of the first embodiment, is
omitted.
[0073] A damper receptacle 51 is formed between each adjacent pair
of the weight receptacles 45. The front end of each damper
receptacle 51 is open. That is, the pulley 17 has the six damper
receptacles 51.
[0074] A tubular elastic member (damping member), which is a rubber
damper 52, is fitted in each damper receptacle 51. The rubber
dampers 52 have circular cross-sections. The outer surface of each
rubber damper 52 closely contacts the inner surface of the
corresponding damper recess 51.
[0075] Each rubber damper 52 has a through hole 52A, the
cross-section of which is circular. The hub 42B has power
transmission pins 53 projecting rearward. Each pin 53 corresponds
to one of the rubber dampers 52. The rear end (right end as viewed
in the drawings) of each pin 53 is fitted in the through hole 52A
of the corresponding damper 52. Each pin 53 is press fitted in a
hole formed in the peripheral portion of the hub 42B, and extends
in the axial direction of the torque receiving member 42. The
number of the pins 53 is six in this embodiment.
[0076] Power transmitted from the engine E to the pulley 17 is
transmitted to the torque receiving member 42 through the rubber
dampers 52 and the power transmission pins 53. The rubber dampers
52 and the power transmission pins 53 are located in the power
transmission path between the pulley 17 and the torque receiving
member 42. The rubber dampers 52 attenuate the torque fluctuation
transmitted from the torque receiving member 42 to the pulley
17.
[0077] In addition to the advantages (1) to (4), the second
embodiment has the following advantages.
[0078] (5) Each rubber damper 52 is located between one of the
adjacent pairs of the weight receptacles 45. The spaces between the
receptacles 45 are effectively used for providing rubber dampers.
The structure of the second embodiment reduces the axial size of
the power transmission mechanism PT compared to the first
embodiment.
[0079] (6) The hub 42B of the torque receiving member 42 prevents
the rollers 46 from falling off the receptacles 45. Therefore,
there is no need for providing an additional member for preventing
the rollers 46 from falling, such as the lid 47 in the first
embodiment. This reduces the number of the parts and thus reduces
the costs.
[0080] A third embodiment of the present invention will now be
described. The third embodiment is the same as the second
embodiment except for the structures of the receptacles and the
rollers and the location of the rubber dampers. 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.
[0081] As shown in FIGS. 4(a) and 4(b), the power transmission
mechanism PT of the third embodiment has six receiving portions,
which are weight receptacles 55. An arcuate weight guiding surface
55A is formed in each receptacle 55. Each guiding surface 55A is a
part of an imaginary cylinder the radius of which is larger than
the radius r.sub.1 of the guiding surface 45A in the second
embodiment.
[0082] An auxiliary guiding surface 55B is formed in each weight
receptacle 55. The auxiliary guiding surface 55B is separated from
the guiding surface 55A toward the axis of the pulley 17 by a
predetermined distance and has an arcuate cross-section. As viewed
from the front side of the compressor C, each weight receptacle 55
appears as an arc having a constant width with its middle portion
located closer to the periphery of the pulley 17 than its ends.
Also, as viewed from the front side of the compressor C, each
weight receptacle 55 is symmetrical with respect to an imaginary
line that contains the rotation axis of the pulley 17 and the
center of the corresponding imaginary cylinder.
[0083] A weight, which is a rigid roller 56 in this embodiment, is
accommodated in each weight receptacle 55. The diameter of each
roller 56 is slightly less than the distance between the guiding
surfaces 55A and the auxiliary guiding surface 55B. The axial
dimension of the rollers 56 is slightly less than the depth of the
receptacles 55, or the dimension along the axis of the pulley 17.
That is, each roller 56 can move along, or can swing like a
pendulum in, the guiding surface 55A of the corresponding
receptacle 55.
[0084] The rubber dampers 52 is located in the power transmission
path between the pulley 17 and the torque receiving member 42.
Also, each adjacent pair of the dampers 52 are located at the ends
of the pendulum motion of one of the rollers 56, or at the ends of
the corresponding receptacle 55. Part of each rubber damper 52 is
exposed in the corresponding receptacles 55. The rubber dampers 52
and the receptacles 55 form receiving portions. When the rollers 56
are moved by an excessive degree, the rollers 56 contact the rubber
dampers 52. That is, the rubber dampers 52 attenuate the torque
fluctuation transmitted from the torque receiving member 42 to the
pulley 17 and function as shock absorbing members for absorbing the
shock due to collision of each roller 56 with the corresponding
receiving portion.
[0085] Each rubber damper 52 is located between an adjacent pair of
the receptacles 55. The outer surface of each damper 52 is exposed
in the corresponding receptacles 55. That is, each damper 52 can
contact the rollers 56 in the corresponding pair of the receptacles
55.
[0086] The pulley 17 (the weight receptacles 55), the rubber
dampers 52, and the rollers 56 form a dynamic damper.
[0087] In addition to the advantages (1) to (6), the third
embodiment has the following advantages.
[0088] (7) The rubber dampers 52 are located at the ends of the
pendulum motion of each roller 56 in the receiving portions.
Therefore, when each roller 56 is moved by an excessive degree, the
shock due to the collision of the roller 56 and the corresponding
receiving portion is absorbed. This prevents the receiving portion
and the roller 56 from being deformed or broken and suppresses
noise.
[0089] (8) The rubber dampers 52 attenuate the torque fluctuation
transmitted from the torque receiving member 42 to the pulley 17
and function as shock absorbing members for absorbing the shock due
to collision of each roller 56 with the corresponding receiving
portion. This structure facilitates creating of the spaces for the
shock absorbing members and the damping members and reduces the
number of the parts, which reduces the costs.
[0090] (9) The shock absorbing member (each shock absorbing damper
52) is used for an adjacent pair of the receiving portions. This
structure facilitates creating of the spaces for the shock
absorbing member (52) and reduces the number of the parts, which
reduces the costs.
[0091] 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.
[0092] A dynamic damper having weights that swing like pendulums
may be provided in the torque receiving member 42 instead of the
pulley 17. Alternatively, both of the pulley 17 and the torque
receiving member 42 may have the dynamic damper.
[0093] In the illustrated embodiments, spherical weights may be
used.
[0094] In the illustrated embodiments, the number of weight
receptacles 45, in which the rollers 46 are provided, may be
changed. The number of the recesses need not correspond to the
cylinder bores of the compressor C.
[0095] In the illustrated embodiments, the cross-sectional shape of
each receptacle 45 along a plane perpendicular to the rotation axis
of the pulley 17 may be circular. This facilitates machining of the
receptacles 45.
[0096] In the illustrated embodiments, the square root of the ratio
R/r is equal to N, which is the value of n.multidot.N when the n is
one. However, the square root of the ratio R/r may be the value of
n.multidot.N when n is a natural number that is equal to or more
than two (for example, two or three).
[0097] In the illustrated embodiments, the ratio R/r, or the square
root of the ratio R/r, of the weights (rollers) may be different.
In this case, since there are two or more values that correspond to
the ratios R/r, two or more peaks of the torque fluctuations are
suppressed. In this case, the values of n are preferably selected
from 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-N, in which n is one, two and three.
Therefore, the two or more greatest peaks of the torque fluctuation
are suppressed. That is, the resonance is effectively suppressed.
FIG. 5 illustrates a power transmission mechanism PT having a
pulley 17 according to another embodiment. The pulley 17 has a
dynamic damper having two values of the ratio R/r. Specifically,
two receptacles 61 and four receptacles 45 are formed in the pulley
17. The cross-section of the receptacles 61 along a plane
perpendicular to the axis of the pulley 17 is different from that
of the receptacles 45. The weight guiding surface 61A of each
receptacle 61 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. The values of R.sub.2 and r.sub.2
and the diameter d.sub.2 of a roller 62 accommodated in each
receptacle 61 are determined such that the value corresponding to
the ratio R/r is different between the receptacle 61 and the
receptacles 45.
[0098] In the illustrated embodiments, the weight guiding surface
45A is formed in each of the weight receptacle 45 in the pulley 17,
and each roller 46 swings like a pendulum along the corresponding
guiding surface 45A. 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.
[0099] 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 cylindrical rollers 46, the settings are preferably
made based on the ratio 2R/3r instead on the ratio R/r to take the
inertial mass into consideration. When the frequency of the peak of
the torque fluctuation is equal to the characteristic frequency of
the rollers 46, the torque T is represented by the following
equation.
T=(3/2).multidot.m.multidot.(.omega..sub.a).sup.2.multidot.(R+r).multidot.-
R.multidot..phi. (Equation 2)
[0100] If spherical weights are used, 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 3)
[0101] 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.
[0102] In the second embodiment, each power transmission pin 53 may
be coupled to the hub 42B with a tubular rubber damper. In other
words, the power transmission pins 53 may be coupled to both of the
hub 42B and the pulley 17 with damping members such as rubber
dampers.
[0103] In the second embodiment, each power transmission pin 53 is
fixed to the hub 42B and coupled to the pulley 17 with the
corresponding rubber damper 52. However, the pins 53 may be fixed
to the pulley 17 and coupled to the hub 42B with rubber
dampers.
[0104] In the second embodiment, the rubber dampers 52 are tubular
and have circular cross-section. However, the cross-section of the
rubber dampers 52 may be changed.
[0105] In the second embodiment, a space may exist between the
inner surface of each receptacle 51 and the corresponding rubber
damper 52 or between the inner surface of each rubber damper 52 and
the outer surface of the corresponding power transmission pin 53.
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.
[0106] In the illustrated embodiments, the rubber dampers (43, 52)
are used. However, dampers made of elastomer may be used.
[0107] In the illustrated embodiment, rubber dampers (43, 52) are
used as damping members. However, the pulley 17 may be coupled to
the torque receiving member 42 by springs (damping members (elastic
members)) such as metal springs. FIG. 6 describes such an
embodiment. In the embodiment of FIG. 6, the rubber dampers 52 of
the second embodiment are replaced with leaf springs 71. The leaf
springs 71 transmit power from the pulley 17 to the power
transmission pins 53. In FIG. 6, the hub 42B is omitted for
purposes of illustration. In the structure of FIG. 6, a spring
receptacle 72 is formed between each adjacent pair of the weight
receptacles 45. A plurality of leaf springs 71 are located in each
spring receptacle 72. The rear portion of each power transmission
pin 53 is inserted in one of the spring receptacle 72. In each
spring receptacle 72, two of the rectangular springs 71 are
laminated and located at each side of the power transmission pin 53
in the circumferential direction of the pulley 17. The ends of the
springs 71 (the ends in the radial direction of the pulley 17) are
engaged with steps 72A formed in the inner wall of the spring
receptacle 72, which prevents the power transmission pin 53 from
moving toward the opposite side in the circumferential direction.
The middle portions of the springs 71 are elastically deformed by
the power transmission pin 53, which absorbs the torque fluctuation
transmitted from the torque receiving member 42 to the pulley 17.
The friction between the laminated springs 71 attenuates the torque
fluctuation. The rotation axes of the pulley 17 and the torque
receiving member 42 may be displaced from each other. However,
deformation of the springs 71 and the changes of the contact point
between the springs 71 and the pins 53 reduce stress applied to the
radial bearings 12A, 40A due to the displacement of the axes.
[0108] In the illustrated embodiments, rubber dampers (43, 52) are
used as damping members. However, the pulley 17 may be coupled to
the torque receiving member 42 by damping members having gel
containers. FIG. 7 describes such an embodiment. In the embodiment
of FIG. 7, the power transmission pins 53 of the second embodiment,
which are located between the pulley 17 and the hub 42B, are
replaced with power transmission projections 73 and containers 74.
The power transmission projections 73 project rearward from the hub
42B. The pulley 17 is coupled to the torque receiving member 42 by
the projections 73 and the containers 74. In FIG. 7, the hub 42B is
omitted for purposes of illustration. In the structure of FIG. 7, a
container receptacle 75 is formed between each adjacent pair of the
weight receptacles 45. A plurality of containers 74 are located in
each container receptacle 75. The cross-section of each power
transmission projection 73 along a plane perpendicular to the
rotation axis of the torque receiving member 42 is substantially
rectangular. The rear portion of each projection 73 is inserted in
the corresponding container recess 75. In each container receptacle
75, a container 74 is located at each side of the projection 73 in
the circumferential direction of the pulley 17. Each container 74
includes a bag-like member and gel, which is included in the
bag-like member. In this structure, gel sealed in the gel
containers 74 attenuates the torque fluctuation transmitted from
the torque receiving member 42 to the pulley 17 through the
projections 73. 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 gel containers 74 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 damping 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.
[0109] 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 drive 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 damping members of the illustrated embodiment
effectively suppress resonance.
[0110] In the third embodiment, each shock absorbing members (52)
need not be used for an adjacent pair of the guides. Each shock
absorbing members may correspond to one of the guides.
[0111] In the third embodiment, the rubber dampers 52 attenuate the
torque fluctuation transmitted from the torque receiving member 42
to the pulley 17 and also function as shock absorbing members for
absorbing the shock due to collision of each roller 56 with the
corresponding receiving portion. However, the damping members may
be provided separately from the shock absorbing members.
[0112] In the illustrated embodiments, the shock absorbing members
for damping shock due to collision between the guides and the
weights may be provided at a position other than the ends of
pendulum motion of each weight. For example, the shock absorbing
members may be provided on any surface forming each receptacle (45,
55, 61). In the first embodiment, the shock absorbing members may
be provided on a side of the lid 47 that faces the receptacles. In
the second and third embodiments, the shock absorbing members may
be located on a side of the hub 42B that faces the receptacles. In
these cases, rubber sheets or elastic coating may be used as the
shock absorbing members. The shock absorbing members dampen the
shock due to collision between the weights and the guides. The
weights collide with the guiding surface (45A, 55A, 61A) of the
receptacles when, for example, the pulley 17 suddenly starts
rotating at a high speed. When the rotation speed of the pulley 17
is decreased while the weight contacts and is swinging like a
pendulum along the guiding surface, the weight is separated from
the guiding surface. At this time, the weight collides with the
radially inner surface of the receptacle (for example, the
auxiliary guiding surface 55B).
[0113] 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.
[0114] In the illustrated embodiment, the cam plate (the swash
plate 20) rotates integrally with the drive shaft 16. However, the
pulley 17 may be used in a compressor, in which a cam plate rotates
relative to a drive shaft. For example, the pulley 17 may be used
in a wobble type compressor.
[0115] The pulley 17 may be used in a fixed displacement type
compressor, in which the stroke of the pistons are not
variable.
[0116] 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.
[0117] In the illustrated embodiment, the second rotor is the
pulley 17. However, a sprocket or a gear may be used as the first
rotor.
[0118] 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.
[0119] 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
located. 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.
[0120] 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.
* * * * *