BACKGROUND OF THE INVENTION
1. Field of the Invention
The present invention relates to a continuously variable
transmission and, in particular to a continuously variable transmission for a vehicle.
2. Description of the Related Art
Vehicles are changed in driving mode from manual gear shifting
to automatic gear shifting due to application of continuously variable transmissions
to the vehicles. Since vehicles with automatic gear shifting are convenient in operation,
fluent in acceleration, and safety in drive, they are increasingly appeal to car-purchasers.
In the prior art, the continuously variable transmissions for achieving automatic
gear shifting comprise hydraulic automatic transmissions and metal belt type continuously
variable transmissions. These two transmissions have many disadvantages which will
be described in detail as follows.
The hydraulic automatic transmissions comprise hydraulic
couplings and hydraulic torque converters. A hydraulic coupling operates based on
the following principle: a power source drives a pump impeller thereof to rotate
so that the pump impeller drives liquid among blades of the pump impeller to rotate.
As a result, energy from the power source is transmitted to the liquid so that the
liquid increases in kinetic energy. When the liquid increasing in kinetic energy
enters into spaces between respective blades of a runner, a part of kinetic energy
contained in the liquid is transmitted to the runner which rotates slower than the
pump impeller so that the runner outputs torque to the outside. In order to increase
the torque output from the runner, a reactor, which rotates in one direction, is
further disposed between the pump impeller and the runner, so that a hydraulic torque
converter is configured.
The hydraulic torque converter can counterbalance some
disadvantages of a mechanically controlled step transmission, and has some prominent
advantages. For example, a vehicle with a hydraulic torque converter is simple in
operation, and labor saving, as well as improves in safety and comfort of drive
and in overtaking acceleration, and lowers in harmful exhaust to comply with the
environmental protection requirement. Further, since the power source and a gear
train are coupled flexibly, a life time of the vehicle lengthens. Finally, the vehicle
has excellent adaptability to improve adaptability to various road surfaces. However,
the hydraulic torque converter has the following apparent disadvantages.
- 1. Since the hydraulic torque converter has a small range of change in speed
and a small range of change in torque so that it can not separately satisfy a usage
requirement of a vehicle, it is necessarily connected in parallel or series with
a mechanically controlled transmission cooperating therewith. In addition, in order
to achieve an automatic gear shifting of the mechanically controlled transmission,
a hydraulic or electric control system is needed to be provided. Therefore, the
corresponding structure is complicated, and the manufacturing cost thereof is expensive.
- 2. Since the energy is transmitted by a liquid as a transmitting medium, transmission
efficiency is low, and energy loss is great so that the vehicle is lowered in economic
usage of fuel oil.
- 3. Since the transmission is complicated in structure, it is difficult to service
the transmission. Therefore, a special service person that has a high service level
and malfunction inspecting and analyzing ability is required for the service. As
a result, service cost is high.
Chinese Patent
CN No. 1136108C published on January 28, 2004
discloses a hydraulic automatic transmitting system comprising a torque
converter for fluidly coupling a engine and a transmission; a primary gear shifting
part disposed on a first shaft; a secondary gear shifting part disposed on a second
shaft; a power transmitting device (consisting of four power transmitting gears)
for transmitting a rotary power from the primary gear shifting part to the secondary
gear shifting part; first, second, and third brakes and clutches for controlling
the primary gear shifting part and the secondary gear shifting part; and first and
second one-way clutches. The torque converter consists of a pump impeller, a runner,
a reactor disposed between the pump impeller and the runner. The pump impeller,
the runner, and a reactor constitute a hydraulic torque converter. Although the
above automatic transmitting system improves a lot over the prior art, it still
has many disadvantages as mentioned above.
A metal belt type continuously variable automatic transmission
operates based on the following principle: a power output from a power source is
transmitted to a driving working wheel of a stepless speed change device, and the
driving working wheel transmits the power to a driven wheel via a V-shaped metal
belt. After that, the power is transmitted to wheels of a vehicle through an intermediate
speed reducer, a main speed reducer, and a differential. A metal belt transmitting
device in the transmission is the core of the transmission and comprises the driving
working wheel, the driven working wheel, and the V-shaped metal belt connecting
the driving working wheel and the driven working wheel. Each of the driving working
wheel and the driven working wheel is composed of a stationary cone disc and a movable
cone disc assembled coaxially. The stationary cone disc and the movable cone disc
form a V-shaped groove with which the V-shaped metal belt engages, so that the V-shaped
belt transmits the power under pressing force of the stationary cone disc and the
movable cone disc. When the movable cone discs of the driving working wheel and
the driven working wheel moves axially during operation, a radius of the metal belt
varies, so that a transmission radio is changed. The movable cone disc moves axially
by adjusting hydraulic oil within cylinders for the driving working wheel and the
driven working wheel which are in turn controlled by means of a mechanical-hydraulic
or electric-hydraulic control system. Since pressure of the hydraulic oil can be
adjusted continuously, the transmission also can achieve stepless speed change.
With the metal belt type continuously variable automatic
transmission, a vehicle has advantages that it is excellent in kinetic characteristic,
convenient in operation, and high in transmission efficiency. In addition, the power
source can always operate within an economic rotary speed range thereof, so that
the vehicle is greatly increased in economic usage of fuel oil and is improved in
exhaust. However, the transmission still has the following unsurmountable disadvantages.
- 1. Since the power is transmitted by means of friction between the metal belt
and the driving and driven working wheels, slippage may occur between the metal
belt and the driving and driven working wheels so that motive power with a large
power or a large torque cannot be transmitted. Therefore, the transmission can not
match a power source with a large displacement. Currently, the transmission is only
used in cars with a medium or small displacement so that application range thereof
is limited.
- 2. A vehicle with the transmission is low in starting performance. If a driver
wants to abruptly accelerate, the transmission can not respond rapidly since it
will take time to change diameters of the working wheels. In addition, since the
vehicle needs a large starting torque, it is necessarily provided with starting
devices such as multiple-disc wet clutches, electromagnetic clutches, and hydraulic
torque converters. As a result, the structure of the transmission becomes very complicated.
- 3. It is difficult to manufacture the metal belt and the manufacturing cost
is expensive. Therefore, special equipment for the manufacturing is needed, and
an amount of replacement of equipment is large.
Chinese Laid-Open Application
CN No. 1442623 published on September 17, 2003
discloses such a metal belt type mechanically controlled frictional continuously
variable transmission which has the above disadvantages.
Furthermore,
US Patent No. 6,062,096
discloses a continuously variable transmission which transmits torque
by utilizing rocking arms oscillating. Eccentric masses are provided at ends of
the rocking arms. An input drives the eccentric masses to rotate around respective
axes thereof. Centrifugal forces generated when the eccentric masses rotate drive
the rocking arms to oscillate. Although the transmission overcomes some of the disadvantages
occurring in the above two transmissions, it still has the following disadvantages.
- 1. Since the rocking arms oscillate during operation, in order to assure that
an output shaft always outputs a speed in one direction, it is required that two
one-way clutches which have respective lockup directions opposite to each other
are disposed on sleeves directly connected with the rocking arms, and a set of reversing
gears are disposed to reverse a rotary speed output from one of the one-way clutches,
so as to assure that the one-way clutches output the speed in the same direction.
Therefore, the output torque is output alternately through the two one-way clutches.
As a result, the transmission is bulky and complicated in structure, and when the
output speed is high, the oscillating rocking arms are subject to a large inertia
force, thereby making the requirement for material and manufacturing accuracy of
the rocking arms and bearings thereof high.
- 2. Since a one-way clutch is not disposed between the torque converter and a
prime mover, a part of kinetic energy stored in the eccentric masses is transmitted
back to the prime mover so as to cause circulation of the power, thereby affecting
effective exertion of performance of the transmission.
- 3. Since the masses of the eccentric masses vary by changing phases of the eccentric
masses so as to adjust output torque and rotary speed, additional manual or automatic
controlling mechanism is required. If the manual controlling mechanism is adopted,
the phases of the eccentric masses necessarily are manually adjusted, but can not
be automatically adjusted. If the automatic controlling mechanism is added, a complicated
feed back system is required, so that the manufacturing cost greatly increases.
- 4. The one way clutch at output part of the transmission is a radially engaging
line-contact sprag clutch, the one way clutch having such structure not only is
not sufficiently sensitive in response since the sprags are affected by centrifugal
forces generated by rotation thereof, respectively, but also a large moment is required
during disengagement and a large amount of energy is consumed so as to decrease
the transmission efficiency of the transmission. In addition, since the sprags are
in line contact with both an inter ring and an outer ring, the clutch is low in
wear resistance so that it can not transmit a large torque and is short in lift
time. Therefore, the lift time of the entire transmission is shortened.
US Patent No. 6044718
, which is a continuation-in-par of
US Patent No. 6062096
, further discloses a new solution in addition to the original ones. However,
the further disclosed solution is substantially similar to the others. In the further
solution, the rocking arm mechanism is substantially not changed, and two one-way
clutches which have respective lockup directions opposite to each other are still
needed to be disposed on sleeves directly connected with the rocking arms. The further
solution differs from the others in that one clutch of the one-way clutches is fitted
between the sleeves of the rocking arms and a base frame. The one clutch thus limits
swing of the rocking arms towards one direction so that the rocking arms only can
swing intermittently in another direction. The transmission not only still has the
above disadvantages, but also makes operating condition of the one-way clutch fitted
between the rocking arms and the base frame worse, since the base frame is stationary,
and the one-way clutch bears a large load so as to limit the swing of the rocking
arms towards the one direction. As a result, the sprags of the clutch tightly wedge
into the sleeves of the rocking arms and a corresponding hole of the frame, so that
it is more difficult to disengage the clutch, more energy is consumed, the wear
of the clutch is more severe, and thus the life time is severely affected.
SUMMARY OF THE INVENTION
It is an object of the present application to overcome
some of the above disadvantages existing in the prior art. The present application
provides a mechanically controlled continuously variable automatic transmission
which can automatically adjust the output speed and moment depending upon the magnitude
of load applied thereto so as to well achieve a function of stepless automatic change
of the speed. Moreover, the transmission is simple in structure, high in transmission
efficiency, and long in life time. Furthermore, the transmission can be applied
to an operating condition of high rotary speed and can effectively transmit large
power and motive power with large torque, so that it can be used in various vehicles
and has a wide application range.
The mechanically controlled continuously variable automatic
transmission comprises a housing and a torque converting mechanism mounted in the
housing. The torque converting mechanism comprises an input part, a rotatable or
rotary carrier, at least one eccentric assembly which is rotatablely mounted on
the carrier, and an output part. The input part and the carrier can rotate independently
with respect to each other, and have respective rotation axes which are collinear.
Each of the at least of one eccentric assembly comprises an eccentric mass which
is driven to rotate around its rotation axis by the input part. The output part
is provided with only one one-way clutch directly connected therewith, and the one-way
clutch is a first one-way clutch. The first one-way clutch is an axially pressing
surface-contact one-way clutch with a driving part and a driven part arranged axially.
When the driving part and the driven part engage with each other, engagement surfaces
of engagement elements of the driving part and the driven part abut against each
other so as to transmit moment by means of frictional force therebetween.
The mechanically controlled continuously variable automatic
transmission according to the present invention further comprises the following
additional features.
In a further improved embodiment of the present invention,
the transmission further comprises a second one-way clutch located at a front end
of the torque converting mechanism which has an input part coupled with a power
source and an output part coupled with the input part of the torque converting mechanism.
In a still further improved embodiment of the present invention,
the transmission further comprises a third one-way clutch disposed at a rear end
of the first one-way clutch, and the third one-way clutch has a lockup direction
opposite to that of the first one-way clutch and has a movable part coupled with
an output part of the first one-way clutch and a stationary part fixed in the housing.
In a preferable embodiment of the present invention, the
torque converting mechanism comprises the carrier, and the at least one eccentric
assembly includes two eccentric assemblies symmetrically mounted at two ends of
the carrier. Each of the two eccentric assemblies comprises a driven shaft, and
the eccentric mass and a driven gear both mounted on the driven shaft. The eccentric
masses and the driven gears are hinged to the two ends of the carrier by means of
the driven shafts. The input part of the torque converting mechanism comprises a
driving shaft and a driving gear mounted on the driving shaft. The driving shaft
is coupled with the power source or the second one-way clutch, and the driving gear
is in mesh with the driven gears. The output part is an output shaft fixed at a
center of the carrier, and the driving part of the first one-way clutch is coupled
with the output shaft.
In another preferable embodiment of the present invention,
the torque converting mechanism comprises the carrier, and the at least one eccentric
assembly includes three eccentric assemblies arranged on the carrier at equal intervals
in a circumferential direction thereof. Each of the three eccentric assemblies comprises
a driven shaft, and the eccentric mass and a driven gear both mounted on the driven
shaft. The carrier is a disc-like body, and the eccentric masses and the driven
gears are hinged to an edge of the carrier by means of the driven shafts. The input
part of the torque converting mechanism comprises a driving shaft and a driving
gear mounted on the driving shaft. The driving shaft is coupled with the power source
or the second one-way clutch, and the driving gear is in mesh with the driven gears.
The output part is an output shaft fixed at a center of the carrier, and the driving
part of the first one-way clutch is coupled with the output shaft.
In a preferable embodiment of the present invention, the
first one-way clutch is a screw-pressing one-way clutch which is engaged by screw-pressing.
The screw-pressing one-way clutch comprises a clutch drum and first and second clutch
discs disposed in the clutch drum and parallel to each other, at least one drum
friction plate and at least one spring. The first and second clutch discs clamp
the at least one drum friction plate under the action of the at least one spring,
and the at least one drum friction plate is fitted over a sleeve and coupled with
the clutch drum in such a manner that a torque can be transmitted. The sleeve has
an inner thread, and a transmitting shaft projects into the sleeve and has an end
projecting into the sleeve which is formed with an outer thread engageable with
the inner thread.
In the above preferable embodiment, the at least one drum
friction plate includes a plurality of drum friction plates, and a disc friction
plate is interposed between every two adjacent drum friction plates. The disc friction
plate is fitted over the sleeve and coupled with the sleeve in such a manner that
a torque can be transmitted. The clutch drum, the first and second friction discs,
the drum friction plates, the sleeve and the disc friction plate have respective
rotation axes which are collinear. The drum friction plates are splined to the clutch
drum and the disc friction plate is also splined to the sleeve. The clutch drum
has one open end and the other end formed with an outwards extending hollow shaft
at a center portion thereof. The first clutch disc is fixed on the transmitting
shaft, and mounted at the open end of the clutch drum by means of a first snap ring.
The second clutch disc is integrally formed with the sleeve, and mounted on the
transmitting shaft by means of a second snap ring. Only one spring is provided in
the clutch drum. The one spring is a plate-like compression spring fitted over the
transmitting shaft, and is disposed between the second snap ring and the second
clutch disc. The end, which projects into the sleeve, of the transmitting shaft
is supported in the hollow shaft via a bearing.
In another preferable embodiment of the present invention,
the first one-way clutch is a four-bar-linkage-mechanism-pressing one-way clutch
which is engaged by pressing by a four-bar linkage mechanism. The four-bar-linkage-mechanism-pressing
one-way clutch comprises a housing composed of a clutch disc and a clutch cover
both combined or covered together, and a clutch hub, a plurality of hub friction
plates, at least one set of disc friction plates, and at least one four-bar linkage
pressing mechanism are disposed in the housing. A through hole is formed in a center
portion of the clutch cover, and an end of the clutch hub is exposed from the through
hole. The hub friction plates have a shape of a circular ring, and are fitted over
the clutch hub and coupled therewith in such a manner that a torque can be transmitted.
A number of sets of the least one set of disc friction plates and a number of the
at least one four-bar linkage pressing mechanism are the same, and each set of the
at least one set of disc friction plates comprise a plurality of friction plates
having a shape of a part of a circular ring which are arranged so as to be alternated
with the hub friction plates. Each of the at least one four-bar linkage pressing
mechanism has a traverse bar, each set of the at least one set of disc friction
plates have a through hole at a same position of respective disc friction plates,
and the traverse bars pass through the through holes, respectively. The four-bar
linkage mechanisms press the disc friction plates and the hub friction plates against
a friction surface of the clutch disc.
In the above preferable embodiment, each of the at least
one four-bar linkage pressing mechanism comprise two supporting arms parallel to
each other and the traverse bar jointing the two supporting arms. The two supporting
arms have ends hinged to the housing by means of jointing pins, respectively, and
the other ends hinged to jointing blocks by means of jointing pins, respectively,
and the traverse bar has two ends which are fixed to the jointing blocks, respectively.
A spring is fitted over the end of the traverse bar and a pressing block is disposed
at the other end of the traverse bar, and the pressing block is hinged to the jointing
pin. Each set of the at least one set of the disc friction plates are further provided
with two through holes through which two cylindrical pins pass, respectively, so
that the respective disc friction plates are connected together in series. The clutch
disc and the clutch cover are correspondingly formed with two elongate grooves extending
in a circumferential direction, respectively. Two ends of each of the two cylindrical
pins are inserted in the elongate grooves, each of the ends of the two cylindrical
pins is provided with two planes parallel to each other, each of the elongate grooves
is provided with two planes parallel to each other, and the two planes of each of
the ends fit with the corresponding two planes of each of the elongate grooves.
The hub friction plates are splined to the clutch hub. A transmitting shaft is mounted
at a center of the clutch disc, and has an end projecting outwards and the other
end supported in a center hole of the clutch hub through a bearing. The at least
one four-bar linkage pressing mechanism includes three four-bar linkage pressing
mechanisms disposed on the housing, the at least one set of disc friction plate
includes three sets of disc friction plates correspondingly disposed in the housing,
the three four-bar linkage pressing mechanisms and the three sets of friction plates
are arranged in equal intervals in a circumferential direction.
The mechanically controlled continuously variable automatic
transmission according to the present application has following advantages with
respect to the prior art.
- 1. Since the mechanically controlled transmission transmits torque by changing
a moment of momentum of the eccentric mass, output rotation speed and moment can
be automatically adjusted depending on magnitude of a load without other manual
control or electronic feedback system. Not only a function of automatic stepless
speed change is achieved really, but also the structure of the transmission is simplified
and the manufacturing cost of the transmission is decreased.
- 2. The transmission of the present invention is not only high in transmission
efficiency, but also can effectively transmit large power and motive power with
large torque. Since the energy is transmitted by means of the eccentric mass as
a transmitting medium, which differs from the case where the energy is transmitted
by means of the fluid as a transmitting medium, not only the transmission efficiency
is high, but also the large power and the motive power with large torque can be
transmitted. Therefore, the transmission can be applied to cars with a medium or
small displacement as well as heavy vehicles with a large displacement. The transmission
can be applied widely. A hydraulic torque converter and a metal belt type continuously
variable automatic transmission cannot compare with the transmission of the present
invention in terms of the advantages.
- 3. The torque converting mechanism of the transmission of the present invention
is more reasonable in structure and working conditions. The output of the carrier
of the transmission according to the present invention is provided with only one
one-way clutch directly coupled therewith in order to achieve unidirectional torque
and speed output without the carrier of the torque converting being necessarily
limited to oscillate and two one-way clutches and a reversing mechanism disposed
at an output of the carrier. In this way, not only the working conditions of the
carrier of the torque converting mechanism are improved so that the function of
automatic stepless speed change is achieved better, but inertia force to which the
carrier of the torque converting mechanism is subjected greatly decreases, so that
strength requirement for components of the transmission greatly lowers. Therefore,
the transmission of the present invention can be suitable for a high speed engine
of a speed of up to 6000 rpm.
- 4. A greater torque ratio can be obtained with the transmission of the present
invention, and the transmission of the present invention has larger transmission
efficiency. Since a one-way clutch is also disposed between an engine and the input
of the torque converting mechanism, it can be assured that kinetic energy stored
in the eccentric assembly can not be transmitted back to the engine so that more
energy can be transmitted to a load. Therefore, the torque ratio of the entire transmission
becomes greater, the transmission efficiency of the entire transmission is higher,
and the structure of the entire transmission is finer.
- 5. An axially pressing surface-contact one-way clutch is adopted in the present
invention. The one-way clutch not only is sensitive in response so as to satisfy
requirement of an engine operating at a high speed, but also can transmit a large
torque and has high transmission efficiency and long life time. Therefore, the entire
transmission becomes stable and reliable in performance, low in failure rate, and
long in life time.
BRIEF DESCRIPTION OF THE DRAWINGS
The features and the advantages of the present invention
will be further explained in detail by embodiments taken in combination with accompanying
drawings, in which:
- Fig. 1A is a schematic block diagram of a configuration of a drive system of
a vehicle showing a general mounting position of a continuously variable automatic
transmission according to the present invention;
- Fig. 1B is a perspective view of a mechanical configuration of a continuously
variable automatic transmission according to the present invention;
- Fig. 2 is a schematic block diagram of a preferable embodiment of the present
invention showing a basic configuration;
- Fig. 3 is a schematic block diagram of another preferable embodiment of the
present invention showing a basic configuration, in which a second one-way clutch
is added based on the preferable embodiment shown in Fig. 2;
- Fig. 4 is a schematic block diagram of a further preferable embodiment of the
present invention showing a basic configuration, in which a third one-way clutch
is added based on the preferable embodiment shown in Fig. 2;
- Fig. 5 is a schematic block diagram of a still further preferable embodiment
of the present invention showing a basic configuration, in which a third one-way
clutch is added based on the preferable embodiment shown in Fig. 3;
- Fig. 6A is a perspective view of a torque converting mechanism in the embodiments
shown in Figs. 2 through 5, in which the torque converting mechanism comprises two
eccentric assemblies;
- Fig. 6B is an exploded perspective view of the torque converting mechanism of
Fig. 6A showing specific configurations and mounting manners of components thereof;
- Fig. 6C is a configuration of a torque converting mechanism with which the torque
converting mechanism shown in Fig. 6A is replaced;
- Fig. 6D is an exploded perspective view of the torque converting mechanism of
Fig. 6C showing specific configurations and mounting manners of components thereof;
- Fig. 7A is a perspective view of another torque converting mechanism in the
embodiments shown in Figs. 2 through 5, in which the torque converting mechanism
comprises three eccentric assemblies;
- Fig. 7B is an exploded perspective view of the torque converting mechanism of
Fig. 7A showing specific configurations and mounting manners of components thereof;
- Fig. 7C is a configuration of a torque converting mechanism with which the torque
converting mechanism shown in Fig. 7A is replaced;
- Fig. 7D is an exploded perspective view of the torque converting mechanism of
Fig. 7C showing specific configurations and mounting manners of components thereof;
- Fig. 8A is a perspective view of a configuration type of first and second one-way
clutches in the embodiments shown in Figs. 2 through 5;
- Fig. 8B is an exploded perspective view of the clutches shown in Fig. 8A;
- Fig. 8C includes four block diagrams which all together show operation principle
of the one-way clutches of Fig. 8A;
- Fig. 8D is a view showing an entire mechanical configuration of a transmission
of the present invention in the case where both the first and second one-way clutches
adopt those shown in Figs. 8A and 8B;
- Fig. 9A is a perspective view of another configuration type of first and second
one-way clutches in the embodiments shown in Figs. 2 through 5;
- Fig. 9B is an exploded perspective view of the clutches shown in Fig. 9A;
- Fig. 9C is a block diagram showing operation principle of the one-way clutches
of Fig. 9A;
- Fig. 9D is a view showing an entire mechanical configuration of a transmission
of the present invention in the case where both the first and second one-way clutches
adopt those shown in Figs. 9A and 9B;
- Fig. 10A is a schematic view showing an operating condition of the continuously
variable transmission according to the present invention, in which the first one-way
clutch is in a lockup state; and
- Fig. 10B is a schematic view showing another operating condition of the continuously
variable transmission according to the present invention, in which the first one-way
clutch is in an overrunning state.
DETAILED DESCRIPTION OF THE EMBODIMENTS
Referring to Fig. 1A, a configuration of a drive system
of a vehicle is schematically shown. The drive system comprises a power source 100,
a transmission 200, and a load 300 (i.e., wheels and the like). The power source
100 comprises an engine, a motor, or other prime movers. The present invention is
explained with respect to the engine for convenient description. The transmission
200 is mounted between the power source 100 and the load 300, and power output from
the power source 100 is transmitted to the wheels 300 of a vehicle by torque and
speed changes by the transmission 200, so that the wheels rotate, thereby driving
the vehicle to move forward or backward. The prevent invention improves the transmission
200 in the system.
With reference to Fig. 1B, a mechanically controlled continuously
variable automatic transmission 200 comprises a housing 8 and a torque converting
mechanism 1 mounted in the housing 8. The torque converting mechanism 1 includes
an input part 11, a rotatable or rotary carrier 12, at least one eccentric assembly
13 (two eccentric assemblies are shown in Fig. 2) which is rotatablely mounted on
the carrier 12, and an output part 14, as shown in Fig. 2. As shown, the input part
11 and the carrier 12 can rotate independently with respect to each other, and have
respective rotation axes which are collinear. Each of the at least of one eccentric
assembly 13 comprises an eccentric mass 131 which is driven to rotate around its
rotation axis by the input part 11, and a specific configuration of which will be
described in detail in the following.
The output part 14 is provided with only one one-way clutch
2 directly connected therewith. The power subjected to torque change and speed change
by the torque converting mechanism 1 is output through the one one-way clutch 2.
That is, a rotation speed can be output in a single direction. Therefore, the torque
converting mechanism is simple in structure, convenient in assembly, and low in
manufacturing cost.
The one-way clutch 2 is a first one-way clutch. The first
one-way clutch 2 is an axially pressing surface-contact one-way clutch with an engagement
element 23 of a driving part 21 and an engagement element 24 of a driven part 22
arranged axially, as shown in Fig. 2. When the driving part and the driven part
engage with each other, engagement surfaces of engagement elements 23 and 24 abut
against each other so as to transmit moment by means of frictional force therebetween.
Not only the first one-way clutch is high in transmission efficiency and can satisfy
requirements of operating conditions of high speed and high frequency in running
of a vehicle, but also excellent in wear resistance and long in life time, as compared
with a radially pressing point-contact clutch in the prior art. An output shaft
25 of the one-way clutch 2 is coupled with the load 300 directly or through an auxiliary
device 7 to be described in the following to transmit the moment to the load 300.
Referring to Fig. 3, as a further improvement of the above
preferable embodiment, the transmission 200 further comprises a second one-way clutch
3 located at a front end of the torque converting mechanism 1. The second one-way
clutch 3 has an input part 31 coupled with the power source 100 and an output part
32 coupled with the input part 11 of the torque converting mechanism 1. A lockup
direction of the second one-way clutch 3 is the same as a forward rotation direction
of the engine and a lockup direction of the first one-way clutch 2. With the second
one-way clutch 3, the power of the engine is transmitted to the torque converting
mechanism 1 only in a single direction. Even if a rotation speed of the output part
11 of the torque converting mechanism 1 is higher than that of the engine, since
the second one-way clutch 3 is in an overrunning state when it rotates in a reverse
direction, a moment in the reverse direction can not be transmitted back to the
engine so that energy stored in the eccentric assembly 13 can be transmitted to
the first one-way clutch 2 even more. Therefore, the transmission of the present
invention can obtain a large torque ratio and circulation of power to lower the
transmission efficiency does not occurs.
With reference to Figs. 4 and 5, as a further improvement
of the above two preferable embodiments, the transmission 200 further comprises
a third one-way clutch 4 disposed at a rear end of the first one-way clutch 2. The
third one-way clutch 4 has a lockup direction opposite to that of the first one-way
clutch 2 and has a movable part coupled with the driven part 22 of the first one-way
clutch 2 and a stationary part fixed in the housing 8. The third one-way clutch
4 may comprise a roller or sprag one-way clutch in the prior art, which is adapted
to limit a reverse rotation of the output part 25 of the first one-way clutch 2.
Therefore, the third one-way clutch 4 does not constrain a moving state of the carrier
12. For example, when a vehicle is in a forward or driving shift while stopping
on an upwardly-inclined road, even if a parking brake is not actuated, the vehicle
can not move backwards due to constraining action of the third one-clutch, which
is greatly convenient for a driver to drive. The third one-way clutch has a function
different from that of a one-way clutch disposed between a sleeve of a rocking arm
and a base frame in
US Patent No. 6,044,718
which directly limits a rotation of the rocking arm towards a direction.
As an example of the preferable embodiments shown in Figs.
2 through 5, the torque converting mechanism 1 is configured as shown in Figs. 6A
and 6B. The torque converting mechanism 1 comprises the carrier 12, and the at least
one eccentric assembly includes two eccentric assemblies 13 symmetrically mounted
at two ends of the carrier 12. Each of the two eccentric assemblies 13 comprises
a driven shaft 132, and the eccentric mass 131 and a driven gear 133 both mounted
on the driven shaft 132. The eccentric mass 131 and the driven gear 133 are hinged
to the two ends of the carrier 12 by means of the driven shafts 132. As shown, the
carrier 12 has a H-shaped longitudinal section, and the driven shafts 132 pass through
two side walls of the carrier 12. Each of the driven shafts 132 has one end fixed
with the driven gear 133, and the other end fixed by means of a snap ring 137. Spacers
135 are interposed between the snap rings 137 and the side wall of the carrier 12
as well as between the driven gears 133 and the side wall of the carrier 12, respectively,
so as to reduce wear of the carrier 12. The eccentric masses 131 is held between
the two side walls of the carrier 12, so as to be axially positioned conveniently
and so as to be safe and reliable.
As shown in Figs. 6A and 6B, the input part 11 of the torque
converting mechanism 1 comprises a driving shaft 111 and a driving gear 112 mounted
on the driving shaft 111. The driving shaft 111 is coupled with an output shaft
of the engine (in the embodiments shown in Figs. 2 and 4), or the output part of
the second one-way clutch 3 (in the embodiments shown in Figs. 3 and 5). The driving
gear 112 is in mesh with the driven gears 133. The output part 14 is an output shaft
fixed at a center of the carrier 12, and the driving part 21 of the first one-way
clutch 2 is coupled with the output shaft 14. The eccentric masses 131 are connected
with the driven shafts 132 by means of keys 136, respectively. Therefore, when the
driving gear 112 drives the driven gear 133 to rotate, the eccentric masses 131
rotate around the driven shaft 132 as an axis over an entire circumference and thus
generate centrifugal forces to drive the carrier 12 to rotate. The centrifugal forces
generated by the eccentric masses 131 drive the carrier 12 to rotate.
Figs. 6C and 6D show an alternative embodiment of the torque
converting mechanism which differs from that shown in Figs. 6A and 6B in that the
carrier 12 has a flat plate shape. The carrier 12 is simpler in structure and the
torque converting mechanism 1 and the entire transmission 200 decrease in weight,
as compared with the above carrier 12.
As another example of the preferable embodiments shown
in Figs. 2 through 5, the torque converting mechanism 1 is configured as shown in
Figs. 7A and 7B. The torque converting mechanism 1 comprises the carrier 12, and
the at least one eccentric assembly includes three eccentric assemblies 13 arranged
on the carrier 12 at equal intervals in a circumferential direction thereof. Each
of the three eccentric assemblies 13 comprises a driven shaft 132, and the eccentric
mass 131 and a driven gear 133 both mounted on the driven shaft 132. As shown, the
carrier 12 is a disc-like body, and the eccentric masses 131 and the driven gears
133 are hinged to an edge of the carrier 12 by means of the driven shaft 132. This
torque converting mechanism can generate greater torque ratio so as to be more suitable
for heavy vehicles with a large load as compared with the torque converting mechanism
shown in Figs. 6A and 6D. Since the three eccentric assemblies 13 are distributed
at equal intervals in the circumferential direction, it can be assured that radial
forces acting on the carrier 12 in a radial direction with respect to a rotation
axis thereof always are balanced, and circumferential forces acting on the carrier
12 are in a same circumferential or rotational direction so as to produce a moment
of couple to drive the carrier 12 rotate the rotation axis thereof.
The input part 11 and the output part 12 of the torque
converting mechanism 1 have the same configuration as those of the above embodiments.
The input part 11 comprises a driving shaft 111 and a driving gear 112 mounted on
the driving shaft 111. The driving shaft 111 is coupled with the output shaft of
the engine (in the embodiments shown in Figs. 2 and 4), or the second one-way clutch
3 (in the embodiments shown in Figs. 3 and 5). This embodiment is different from
the above embodiments in that the driving gear 112 is in mesh with the three driven
gears 133. The output part 14 is an output shaft fixed at a center of the carrier
12, and the driving part 21 of the first one-way clutch 2 is coupled with the output
shaft 14. Similar to the above embodiment, the Figs. 7C and 7D show an alternative
embodiment of the torque converting mechanism which differs from that shown in Figs.
7A and 7B in that the carrier 12 has a flat plate shape. Thus, the carrier 12 is
simpler in structure and the torque converting mechanism 1 and the entire transmission
200 decrease in weight.
With reference to Figs. 8A and 8B, as an example of the
preferable embodiments shown in Figs. 2 through 5, the first one-way clutch 2 is
a screw-pressing one-way clutch which is engaged by screw-pressing. The screw-pressing
one-way clutch comprises a clutch drum 51 and first and second clutch discs 52 and
53 disposed in the clutch drum 51 and parallel to each other, at least one drum
friction plate 54 and at least one spring 55. The spring 55 can apply a pre-pressing
force to the first and second clutch discs 52 and 53. The first and second clutch
discs 52 and 53 clamp the at least one drum friction plate 54 under the action of
the spring 55, and the at least one drum friction plate 54 is fitted over a sleeve
56 and coupled with the clutch drum 51 in such a manner that a torque can be transmitted.
The sleeve 56 has an inner thread, and a transmitting shaft 57 projects into the
sleeve 56 and has an end projecting into the sleeve which is formed with an outer
thread engageable with the inner thread.
As shown in Figs. 8A and 8B, in a preferable embodiment
of the present invention, the at least one drum friction plate includes a plurality
of drum friction plates 54, and a disc friction plate 58 is interposed between every
two adjacent drum friction plates 54. The disc friction plate is fitted over the
sleeve 56 and coupled with the sleeve in such a manner that a torque can be transmitted.
When a moment is transmitted from a left side to a right side (which are described
with respect to Figs. 8A and 8B), the clutch drum 51 and the drum friction plates
54 constitute the driving part 21 of the first one-way clutch 2, the disc friction
plate 58, the sleeve 56, the first and second clutch discs 52 and 53, and the transmitting
shaft 57 constitute the driven part 22 of the first one-way clutch 2, and the drum
friction plates 54 and the disc friction plate 58 correspond to the engagement elements
23 and 24 of the driving part and the driven part, respectively. When one drum friction
plate 54 is used, the first and second clutch discs 52 and 53 correspond to the
engagement element 24 of the driven part 22, that is, the disc friction plates.
Therefore, it is not needed to provide the disc friction plate 58.
As shown in Figs. 8A and 8B, the clutch drum 51, the first
and second friction discs 52 and 53, the drum friction plates 54, the sleeve 56
and the disc friction plate 58 have respective rotation axes which are collinear.
The drum friction plates are splined to the clutch drum 51 and the disc friction
plate 58 is also splined to the sleeve 56. With the spline coupling, not only it
can be assure that input torque can be effectively transmitted, but also the torque
converting mechanism is convenient in manufacture and manufacture cost thereof is
reduced. The clutch drum 51 has one open end (right end in Figs. 8A and 8B) and
the other end formed with an outwards extending hollow shaft 511 at a center portion
thereof. The end, which projects into the sleeve, of the transmitting shaft 57 is
supported in the hollow shaft 511 via a bearing 50. Therefore, it is convenient
to mount and position the transmitting shaft 57, and it can be assured that the
axes of the transmitting shaft 57 and the clutch drum 51 are collinear. In a preferable
embodiment of the present invention, the bearing 50 comprises needle roller bearing
which facilitates to assure accurate positioning of the transmitting shaft.
As shown in Figs. 8A and 8B, the first clutch disc 52 is
fixed on the transmitting shaft 57, and mounted in the open end of the clutch drum
51 by means of a first snap ring 59A. The second clutch disc 53 is integrally formed
with the sleeve 56, and mounted on the transmitting shaft 57 by means of a second
snap ring 59B. As shown in Figs. 8A and 8B, in a preferable embodiment of the present
application, only one spring 55 is provided in the clutch drum 51. The spring 55
is a plate-like compression spring fitted over the transmitting shaft 57, and is
disposed between the second snap ring 59B and the second clutch disc 53. With the
plate-like compression spring, not only a pre-pressing force can be applied, but
the clutch is small in axial size, so that a space occupied by the spring 55 in
the one-way clutch drum and in turn the volume of the entire one-way clutch are
reduced.
In regard to the first one-way clutch 2, in use, the hollow
shaft 511 of the clutch drum 51 is coupled with the output shaft 14 of the torque
converting mechanism 1 through a key, and the transmitting shaft 57 is also coupled
with an outer part by a key. Referring to Figs. 8A and 8C, operating principle and
operation process of the screw-pressing one-way clutch are explained as follows.
When the output shaft 14 rotates in a direction L1 (a lockup direction of the one-way
clutch) in the case where a resistance direction of a load is opposite to the direction
L1, the output 14 drives the clutch drum 51 and the drum friction plates 54 to rotate
in the direction L1. Because of action of the pre-pressing force of the compression
spring 55, there is frictional force between the drum friction plates 54 and the
disc friction plate 58. The drum friction plates 54 drive the disc friction plate
58 to rotate likewise in the direction L1 by means of the frictional force. The
disc friction plate 58 drives the sleeve 56 to rotate in the direction L1 through
the spline. Since the transmitting shaft 57 and the sleeve 56 are coupled with each
other by right-hand thread, the sleeve 56 and the second clutch disc 53 move alone
the transmitting shaft 57 towards a right side so as to further press the drum friction
plates 54 and the disc friction plate 58. At that time, the one-way clutch is in
a lockup state so that the transmitting shaft 57 also rotates in the direction L1.
Otherwise, the one-way clutch is in a overrunning state. An equivalent coefficient
of friction can effectively increase by providing a plurality of drum friction plates
54 and a plurality of disc friction plates 58. In the case where the drum friction
plates 54 and the disc friction plates 58 do not vary in number, self-locking can
be achieved between the driving part 21 and the driven part 22 by reasonably designing
size of a helix angle &bgr;, i.e. assuring that the input moment is in the direction
L1. Otherwise, when the input moment is in a direction opposite to the direction
L1, the clutch is in an overrunning state so that it can not transmit moment. When
the transmitting shaft 57 and the sleeve 56 are coupled with each other by left-hand
thread, the lockup direction is opposite to that mentioned above.
Fig. 8C is a schematic block diagram showing operation
principle of the screw-pressing one-way clutches of Fig. 8A. A basic principle of
the spiral pressing is identical to a basic principle of inclined plane pressing.
A thread corresponds to a formation of a hypotenuse of a right triangle by winding
the right triangle around a cylinder in the case where a right-angle side of the
right triangle is perpendicular to an axis of the cylinder, and an inclination angle
&bgr; of the hypotenuse corresponds to a helix angle of the thread. The operation
principle of the screw-pressing one-way clutches is shown in Fig. 8C(b). Fit between
a wedge and an inclined plane corresponds to fit between an inner thread of the
sleeve 56 and an outer thread of the transmitting shaft 57, the wedge corresponds
to a combination of the second clutch disc 53 and the sleeve 56, and the inclined
plane and a portion integral therewith correspond to a combination of the transmitting
shaft 57 and the first clutch disc 52.
Referring to Fig. 8D, as a preferable embodiment of the
present invention, the second one-way clutch 3 located at the front end of the torque
converting mechanism 1 is the screw-pressing one-way clutches of Figs. 8A and 8B,
too. Therefore, the second one-way clutch 3 can also transmit large torque and motive
power with large power and is long in life time. The transmission with the above
configuration can be applied to not only small cars, but also heavy vehicles.
With reference to Figs. 9A and 9B, as another example of
the preferable embodiments shown in Figs. 2 through 5, the first one-way clutch
2 is a four-bar-linkage-mechanism-pressing one-way clutch which is engaged by pressing
by a four-bar linkage mechanism. The four-bar-linkage-mechanism-pressing one-way
clutch comprises a housing 60 composed of a clutch disc 61 and a clutch cover 62
both combined together. As shown in Figs. 9A and 9B, in a preferable embodiment
of the present invention, the clutch disc 61 and the clutch cover 62 are connected
to each other by means of bolts. A clutch hub 63, a plurality of hub friction plates
65, at least one set of disc friction plates 64, and at least one four-bar linkage
pressing mechanism 66 are disposed in the housing 60. As shown in Figs. 9A and 9B,
the clutch hub 63 is located at a center of the housing 60, a through hole 622 is
formed in a center portion of the clutch cover 62, and an end (right end in Figs.
9A and 9B) of the clutch hub 63 is exposed from the through hole 622 so as to be
coupled with an outer input shaft or an output shaft. The hub friction plates 65
have a shape of a circular ring, and are fitted over the clutch hub 63 and coupled
therewith in such a manner that a torque can be transmitted. In a preferable embodiment
shown in Figs. 9A and 9B, the clutch hub 63 and the hub friction plates 65 are coupled
with each other by a spline. A transmitting shaft 612 is mounted at a center of
the clutch disc 61, and has an end projecting outwards and coupled with an outer
input or output shaft and the other end supported in a center hole of the clutch
hub 63 through a bearing 67. The bearing 67 is a needle roller bearing. Therefore,
the transmitting shaft not only is positioned with accuracy, but also can rotate
agilely.
A number of sets of the least one set of disc friction
plates 64 and a number of the at least one four-bar linkage pressing mechanisms
66 are the same. As shown in Figs. 9A and 9B, in a preferable embodiment of the
present invention, the at least one four-bar linkage pressing mechanism includes
three four-bar linkage pressing mechanisms 66 disposed on the housing 60, the at
least one set of disc friction plate includes three sets of disc friction plates
64 correspondingly disposed in the housing 60, the three four-bar linkage pressing
mechanisms 66 and the three sets of friction plates 64 are arranged in equal intervals
in a circumferential direction.
As shown in Figs. 9A and 9B, each set of the at least one
set of disc friction plates comprise a plurality of friction plates having a shape
of a part of a circular ring which are arranged so as to be alternated with the
hub friction plates 65. Each of the at least one four-bar linkage pressing mechanism
66 has a traverse bar 661, each set of the at least one set of disc friction plates
66 have a through hole 641 at a same position of respective disc friction plates,
and the traverse bars 661 pass through the through holes 641. The four-bar linkage
mechanisms 66 press the disc friction plates 64 and the hub friction plates 65 against
a friction surface of the clutch disc 61. With reference to Fig. 9B, specifically,
each of the at least one four-bar linkage pressing mechanism comprise two supporting
arms 662 and 663 parallel to each other and the traverse bar 661 jointing the two
supporting arms 662 and 663. The supporting arms 662 and 663 have a same length.
The two supporting arms 662 and 663 have ends hinged, by means of jointing pins
664, to the housing 60, and specifically to the clutch disc 61 and the clutch cover
62, respectively, as shown in Fig. 9B, and the other ends hinged to jointing blocks
667 and 668 by means of jointing pins 665 and 666, respectively, and the traverse
bar 661 has two ends which are fixed to the jointing blocks 667 and 668, respectively.
The housing 60, the two supporting arms 662 and 663, and the traverse bar 661 constitute
a four-bar linkage mechanism, of which opposite links are parallel to each other.
A spring 660 is fitted over the end of the traverse bar 661 and a pressing block
669 is disposed at the other end of the traverse bar 661, and the pressing block
669 is hinged to the jointing pin 666.
As shown in Fig. 9B, the spring 660 is a compression spring.
The compression spring 660 and the pressing block 669 act on two sides of each set
of the at least one set of the disc friction plates 64, respectively. The compression
spring 660 generates a pre-pressing force to press the disc friction plates 64 and
the hub friction plates 65, so that the disc friction plates 64 and the hub friction
plates 65 abut against each other.
As shown in Figs. 9A and 9B, each set of the at least one
set of the disc friction plates 64 are further provided with two through holes 642
and 643 through which two cylindrical pins 68 and 69 pass, respectively, so that
the respective disc friction plates are connected together in series. The clutch
disc 61 and the clutch cover 62 are correspondingly formed with two elongate grooves
611 and 621 extending in a circumferential direction thereof, respectively. Two
ends of each of the two cylindrical pins 68 and 69 are inserted in the elongate
grooves 611 and 621, each of the ends of the two cylindrical pins is provided with
two planes parallel to each other, each of the elongate grooves is provided with
two planes parallel to each other, and the two planes of each of the ends fit with
the corresponding two planes of each of the elongate grooves. Therefore, each set
of the at least one set of the disc friction plates 64 are supported to the clutch
cover 61 and the clutch disc 62 by means of the cylinder pins 68 and 69, and the
cylinder pins 68 and 69 bear centrifugal forces generated by the respective disc
friction plates 64 when they rotate.
With reference to 9C, operation process and operation principle
of four-bar linkage pressing mechanisms shown in Figs. 9A and 9B are described as
follows. When a moment is input from the clutch hub 63, the clutch hub 63 and the
hub friction plates 65 constitute the driving part 21, and the disc friction plate
64, the housing 60 and the transmitting shaft 612 constitute the driven part 22.
When the input moment drives the clutch hub 63 to rotate in the direction L1, the
clutch hub 63 drives the hub friction plates 65 to rotate in the direction L1 through
the spline. The hub friction plates 65 drive the disc friction plates 64 to rotate
likewise in the direction L1 with frictional force therebetween. The hub friction
plates 65 drive the housing 60 to rotate in the direction L1 through the four-bar
linkage mechanisms 60. At that time, the supporting arms 663 and the pressing blocks
669 further press the hub friction plates 65 and the disc friction plates 64, so
as to achieve self-locking. Otherwise, when the input moment drives the clutch hub
63 to rotate in a direction opposite to the direction L1, the four-bar linkage mechanisms
66 is automatically set to an unlocking state, the pressing force applied to the
disc friction plates 64 and the hub friction plates 65 by the pressing blocks 669
decreases, so that the driving part 21 and the driven part 22 can rotate relative
to each other. As a result, the clutch is in an overrunning state.
With reference to Fig. 9D, as a preferable embodiment of
the present invention, the second one-way clutch 3 is also the above four-bar-linkage-mechanism-pressing
one-way clutch, which is sensitive in response and excellent in wear resistance,
and can transmit moment with a large power, so that it can be used in various heavy
vehicles.
Operation principle and operation process of the transmission
of the present invention are described as follows. With reference to Figs. 10A and
10B, a cutaway perspective view of a combination of a torque converting mechanism
having two eccentric assemblies and a four-bar-linkage-mechanism-pressing one-way
clutch.
Referring to Figs. 10A and 10B, when an engine drives the
driving shaft 111 to rotate in a clockwise direction (viewed from a left side, the
following is the same as this), so long as there is a speed difference between the
input part 11 and the carrier 12, the driving gear 112 drives the two drivengears
133 to rotate in directions shown in Figs 10A and 10B, respectively. Therefore,
the eccentric masses 131 rotate around the driven shafts 132 over an entire circumference.
When the eccentric masses 131 rotate, centrifugal forces F are generated. Each of
the forces F is directed in a direction from a center of the corresponding driven
shaft 132 towards a center of mass of the corresponding the eccentric mass 131 as
shown in Figs. 10A and 10B. The forces F act on the carrier 12 through the driven
shafts 132. Each of the forces F decomposes in a circumferential direction and a
radial direction of rotation of the carrier 12 into a radial force Fr and a tangential
force Ft. Since the two eccentric assemblies 13 is symmetric about the center of
the carrier 12 at any moment, the two radial forces Fr are equal in magnitude to
each other, are directed to opposite directions, respectively, and are located in
a same line, so as to cancel out. The two tangential forces Ft are equal in magnitude
to each other, are directed to opposite directions, respectively, and are parallel
to each other with a distance therebetween (a distance d between axes of the two
driven shafts), so as to constitute a moment of couple M. The moment of couple M
drives the carrier 12 to rotate.
The centrifugal forces F generated by the eccentric masses
131 when they rotate have respective directions and respective magnitudes both periodically
varying with rotation positions of the eccentric masses 131 relative to the carrier
12. Therefore, the moment of couple M has a direction and a magnitude which are
both periodically varying. When the eccentric masses 131 rotate to the positions
shown in Fig. 10A, the moment of couple M acts in the clockwise direction and the
carrier 12 and the output shaft 14 rotate in the clockwise direction. As the eccentric
masses 131 further rotate, the moment of couple M varies in direction. When the
eccentric masses 131 rotate to the positions shown in Fig. 10B, the centrifugal
forces F generated by the eccentric masses 131 are directed in directions shown
in Fig. 10B, respectively. The tangential forces Ft decomposed from the forces F
act in directions opposite to those shown in Fig. 10A, respectively. Therefore,
a moment of couple M in a counter-clockwise direction is generated. The moment of
couple M makes the carrier 12 rotate in the counter-clockwise direction or decreases
a rotation speed of the carrier 12 in the clockwise direction. Therefore, every
moment of an entire period of rotation of the eccentric masses 131 is considered
theoretically, a moment M applied to the carrier 12 varies in magnitude in every
moment, and is alternately changed in direction. The carrier 12 rotates in various
motion states depending upon the rotation speed of the one-way clutch 12. In the
case where the input part 11 rotates normally, the carrier 12 takes on discontinuous
fluctuation when a rotation speed of the output part of the one-way clutch 2 is
zero, the carrier 12 takes on back-and-forth swing when the output part of the one-way
2 has a certain rotation speed but the carrier 12 and components rotating therewith
have moment of momentum smaller than amount of change of the moment of momentum
of the eccentric masses 131, and the carrier 12 rotates continuously towards the
lockup direction of the one-way clutch when the output part of the one-way 2 has
a large rotation speed and the carrier 12 and the components rotating therewith
have moment of momentum greater than amount of change of the moment of momentum
of the eccentric masses 131. The rotation of the eccentric masses 131 makes the
rotation of the carrier 12 periodically quicken and slow.
It can be found from the above analysis, the speed and
moment outputted from the output shaft 14 which have undergone moment and speed
changes by the torque converting mechanism 1 is fluctuated. The operation process
is described in combination with the one-way clutch mounted at the rear end of the
torque converting mechanism 1 as follows. With reference to Fig. 10A, when the moment
of couple M acts in the clockwise direction, the carrier 12 and the output shaft
14 rotate in the clockwise direction, the output shaft 14 drives the clutch hub
63 and the hub friction plates 65 of the first one-way clutch 2 located at the rear
end to rotate in the clockwise direction, and the first one-way clutch 2 is automatically
set to a self-locking state. Therefore, the hub friction plates 65 drive the disc
friction plates 64, the housing 60 and the transmitting shaft 612 to rotate in the
clockwise direction, and a moment in the direction L1 shown in Fig. 10A is outputted.
With reference to Fig. 10B, when the moment of couple M acts in the counter-clockwise
direction, the carrier 12 rotates in the counter-clockwise direction or a speed
of rotation of the carrier 12 in the clockwise direction slows down, and the first
one-way clutch 2 is automatically set to a overrunning state, so that the torque
is not transmitted, but a large rotation speed difference exists between the driving
gear 112 and the carrier 12. Therefore, the driving gear 112 drives the eccentric
assemblies 13 to rotate in a higher rotation speed &ohgr;, and the eccentric assemblies
13 store more kinetic energy (according to E=0.5J&ohgr;2, where J represents
moment of inertia of the eccentric assembly 13 with respect to its axis). When the
direction of the moment of couple M is changed into the clockwise direction from
the counter-clockwise direction, the moment of couple M in the clockwise direction
becomes greater. That is, when the moment of couple M acts in the counter-clockwise
direction, the eccentric masses 131 store energy, and when the moment of couple
M acts in the clockwise direction, the eccentric masses 131 output energy through
the carrier 12. Therefore, the output shaft 612 of the one-way clutch can continuously
output outwards moment in the clockwise direction (a direction L3 shown in Figs.
10A and 10B).
As described above, the carrier 12 is lower in rotation
speed than the driving gear 112 so that a rotation speed difference exists between
the driving gear 112 and the carrier 12 in operation. With the rotation speed difference,
the speed and the moment outputted from the output shaft 14 and the transmitting
shaft 612 of the one-way clutch can be automatically adjusted according to change
of magnitude of the load from an outside part. Specific processes and principle
of the adjustment are described as follows. When a vehicle runs upwards on an upwards
inclined road, resistance to the output shaft 25 of the first one-way clutch increases
and the rotation speed of the output shaft 25 of the first one-way clutch slows
down. When the first one-way clutch 2 is in the lockup state, resistance to the
output shaft 14 of the torque converting mechanism 1 also increases, and the rotation
speed of the carrier 12 automatically lowers, so that the rotation speed difference
between the driving gear 112 and the carrier 12 increases. As described above, at
that time, the eccentric masses 131 rotate at a higher speed, and store more kinetic
energy, so that greater moment of couple M can be outputted at a next period. When
a load decreases, change of variables is just contrary to that of the above described
operation condition, the output speed of the torque converting mechanism automatically
increases, the output torque of the torque converting mechanism correspondingly
decrease. Characteristic of this type complies with requirement of stepless speed
change of a vehicle. Therefore, the vehicle can achieve stepless speed change.
As described above, in order to prevent the kinetic energy
stored in the eccentric masses 131 from being transmitted back to the engine, in
the embodiments shown in Figs. 3 and 5, the second one-way clutch 3 is additionally
mounted between the engine and the torque converting mechanism, and thus the energy
from the engine can be transmitted only to the torque converting mechanism 1 in
a single direction, so as to prevent kinetic energy of the eccentric assemblies
13 from being transmitted back to the engine to decrease the speed of the of the
eccentric assemblies 13 and thus to affect magnitude of moment of couple M, as shown
in detail in Figs. 8D and 9D.
The above embodiments are illustrative. It would be appreciated
by those skilled in the art that modifications and changes may be made in these
embodiments without departing from the principles and spirit of the invention. For
example, one or four sets of the eccentric assemblies 13 can be disposed on the
carrier 12. All the equivalents fall within the scope of the present invention.