Note : Les descriptions sont présentées dans la langue officielle dans laquelle elles ont été soumises.
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1 ~YDRAULIC SYSTEM FOR AN INDUSTRIAL MACHINE
Background of the Invention
~his invention relates to a hydraulic power and control
system for an industrial machine, such as a hydraulically
powered excavator or the like.
Hydraulic excavators conventionally utilize hydraulic
cylinders to control the position of the main excavator boom, to
control the position of the arm at the end of the boom, and to
control the position of the bucket at the end of the arm. Also,
a hydraulic motor is conventionally utilized to rotate the
excavator body that carries the boom on the undercarriage and it
is also known to use a pair of independently controlled
hydraulic motors to drive the opposite tracks of the machine's
undercarriage to propel and steer the machine. Typically, such
hydraulic systems have been open center type systems.
Since more than one of the hydraulically powered functions
on such excavators are normally being operated at the same time,
it is necessary that the hydraulic systems have a relatively
high flow capacity. For that purpose, it is known to provide a
pair of hydraulic pumps to supply fluid pressure for the system,
since the smaller pumps are more readily available and normally
are cheaper per unit of capacity than a single large pump.
Moreover, the smaller pumps can be independently replaced at a
much cheaper cost.
While high flow capacity is a desirable feature generally,
it can be a detriment when an excavator is being used in an
operation that requires relatively slow, accurately controlled
movement of the various functions, such as when an excavator is
being used to lay pipe, or unload trucks, or operating in close
proximity to a structure or workers.
Summary of the Invention
According to the present invention, a multi-pump, closed
center, pressure on demand system is provided for an excavator
or the like.
An important feature in the invention resides in the
provision of means for reducing the flow rate of the system at
the option of the operator. More specifically, a valve is
provided to optionally shut off the load sensing circuit for at
least one of the pumps, thereby causing the pump to destroke at
the option of the operator.
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1 Brief Description of the Drawin~s
Fig. 1 is a side elevation view of an excavator embodying
the invention.
Fig. 2 is a schemetic view of the hydraulic system for
controlling the various hydraulically powered functions on the
excavator, the figure being divided into sections 2a and 2b on
separate sheets of drawing in order to afford enough space for
the hydraulic system.
Description o~ the Preferred Embodiment
The invention is embodied in a hydraulic excavator that is
shown generally in Fig. 1. The excavator includes a main body
10, having an operator's cab 12 at one end and mounted on an
undercarriage 14 by means of a swing pivot 16. The swing pivot
is also schematically shown in Fig. 2, and the body 10 is
rotatable through a full circle relative to the undercarriage 14
on the pivot 16, the swinging of the body being accomplished by
a hydraulic motor 18 that drives a gear train 20 engageable with
a large rir.g gear in the pivot mechanism to rotate the body 10.
The undercarriage 14 includes a pair of tracks 22 on
opposite sides of the undercarriage, and the respective tracks
are driven by hydraulic motors 24 and 25 through respective
clutches 26 and 27 and planetary reduction gearing (not shown)
in the well known manner.
The excavator includes a large boom 28 that extends from the
body 10 and is swingable in a vertical arc by actuation of a
pair of boom cylinders 30. A dipper stick or arm 32 is
swingably mounted on the outer end of the boom and its position
is controlled by a hydraulic cylinder 34. At the lower end of
the dipper stick or arm 32, there is mounted a conventional
excavator bucket 36 that is swingable relative to the arm 32 by
means of a hydraulic cylinder 38. All of the above represents
more or less conventional construction, and the various
hydraulic functions are all schematically shown on Fig. 2.
An internal combustion engine 40 is mounted in the body 10
and has an output shaft 42 that drives a pair of identical
hydraulic pumps 44 and 46. The pumps 44 and 46 are identical
variable displacement pumps and respectively have stroke control
systems 48 and 50. While only two pumps are illustrated in the
preferred embodiment, it is to be understood that three or more
commonly driven and connected pumps could he included in the
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l system. Since the stroke control system for the two pumps is
essentially the same, only the stroke control system 48 for the
pump 44 will be described in detail.
The displacement of the pump 44 is controlled by a stroke
control cylinder 52 that is actuated by a change of pressure in
a destroke line 54, the pressure in the line 54 acting against a
spring in the stroke control cylinder 52 to destroke the pump.
When the pressure is exhausted from the line 54, the spring
moves the piston of the stroke control cylinder 52 toward full
displacement of the pump.
The pump destroke line 54 is connected to a horsepower
control valve 56, which connects the line 54 to a load sensing
spool valve 58 when the horsepower control valve 56 is in the
position illustrated in Fig. 2. The load sensing spool valve 58
in turn connects the destroke line 54 to a drain line 60 when
the valve 58 is in the illustrated position, and the drain line
60 is connected to a line 61 that dumps into the hydraulic
systems reservoir 62.
The pumps 44 and 46 include inlet lines 64 that draw fluid
from the reservoir 62 and also leakage lines 65 that dump pump
leakage back to the reservoir 62. The pump 44 has an outlet
line 66 and the pump 46 has an outlet line 67 that are connected
to a common outlet line 68, which is illustrated on Fig. 2b,
while the pumps and their respective controls are illustrated in
Fig. 2a. Referring back to Fig. 2a, the outlet line 66 is
connected to one end of the horsepower control valve 56 via a
pilot line 70, and movement of the control valve 56 is caused by
said pilot line pressure working against springs 72. As is
apparent, if pressure in the outlet line 66 starts to exceed a
certain va'ue that would overload the engine, the horsepower
control valve shifts downwardly in Fig. 2 to meter pressurized
fluid from the outlet line 66 back to the destroke line 54,
thereby reducing the pump displacement and thereby reducing the
load on the engine.
A load sensing line 74 is connected to the load sensing
spool valve 58 for the motor 44 through an on-off type solenoid
actuated valve 78, that is controlled by the operator by a
conventional electric switch at the operator's cab 12. The line
74 is connected to the load sensing spool valve for the control
system 50 directly by a hydraulic line 80, so that the pressure
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1 in the load sensing line 74 is always communicated to one end of
the load sensing spool valve for the motor 46 while it is
optionally connected to the valve 58 for the other pump 44. As
is apparent, when the pressure in the load sensing line 74 is
communicated to the valve 58 through the solenoid valve 78, it
works against the pressure in the pump outlet line 66 that is
communicated to the other end of the valve 58 by means of a
pilot line 82. The pressure in the line 76 plus the force of a
spring 84 are sufficient to shift the valve to the position
shown in Fig. 2a when the pressure in the line 74 reaches a
certain value, and, as is apparent, when the system is
sufficiently loaded to open the valve 58 as shown but not
overloaded so that the valve 56 stays open, the pump destroke
line 54 is connected to the reservoir to shift the pump 44 to
its maximum displacement.
Since the line 80 from the load sensing line 74 to the load
sensing valve for the motor 46 is always connected to the valve,
the pump responds to pressure in the load sensing line 74 to
place the pump in stroke when the pressure in the load sensing
line 74 indicates a pressure demand. The same is normally true
for the control system for the other motor 44, except that the
operator by shifting the valves 78 to its closed condition, can
shut off the connection of the load sensing line 74 to the valve
s 58, so that pressure in the line 66 shifts the valve 58 to
connect the outlet line 66 to the pump destroke line 54, which
thereby destrokes the pump. Thus, the operator can optionally
maintain the pump 44 in its destroked condition regardless of
the demand indicated by the pressure in the load sensing line
74.
The common outlet line 68 for the two pumps serves as a
common input for a bank of control valves 85, 86, 87, 88, 89,
and 90 that respectively control the flow of pressurized fluid
to and exhaust fluid from the boom cylinders 30, the arm
cylinder 34, the bucket cylinder 38, the drive motor 24, the
drive motor 25, and the swing control motor 18. The valves 85-
90 are essentially identical, and, therefore, only the valve 87
for the bucket control cylinder 38 will be described in detail.
The valves are of the type known as load sensing control valves,
and such valves are available from the Linde Hydraulics Company
of Canfield, Ohio, as well as others. The
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1 valve 87 is connected to the common pump outlet line 68 by a
hydraulic line 92 and a pair of load sensing lines 93 and 94
connect the valve to the main load sensing line 74. Similarly,
a pair of drain lines 95 and 96 connect the valve to the main
drain line 97 that connects the entire valve bank to the
hydraulic reservoir 62 through an oil cooler 98 and an oil
filter system 99, both of which are of known construction.
A relief valve system 100, also of conventional
construction, is connected to the load sensing line 74 by a line
102 and to the common outlet line 68 by a line 104, the relief
valve package being operable to dump the lines 102 and 104 to
the reservoir line 97 when certain pressures are exceeded.
The valve 87 is connected to the bucket control cylinder by
lines 106 and 108 and is shiftable in opposite directions by
means of pilot operated control systems 110 and 112, of known
construction. As is apparent, when the valve is shifted in one
direction from the neutral position as shown, the pressurized
inlet line 92 is connected to the valve outlet line 106, and the
pressure directed to the cylinder 38 is also communicated to the
load sensing line 93. The other end of the cylinder is
connected to the reservoir through the line 108 and the valve
87. Conversely, when the valve is shifted in the opposite
direction, it connects the pressure line 92 to the valve outlet
line 108 while simultaneously communicating that pressure to the
load sensing line 94, and the line 106 is connected to the
reservoir via the drain line 95.
The swing motor 18 is provided with conventional check and
pressure relief valves, and in addition to the control valves 88
and 89 for the track motors 24 and 25, metering valves 114 are
provided to meter fluid to the respective motors. Since the
present invention is not specifically directed to the hydraulic
circuitry associated with the respective motors, the various
relief and check valves are not described in detail.
The load sensing line 74 is also connected to a sensing line
116 that provides the input to a pressure actuated switch 118 of
conventional construction. A predetermined pressure in the load
sensing system and consequently the pressure in line 116
operates to hold the switch 118 in an open condition. However,
when the pressure in the line 116 falls below a predetermined
point as a result of a lack of load in the system, a spring in
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1 the switch 118 closes the switch to connect the electrical
system of the excavator to an automatic timer 120. The timer is
of known construction and can be set to provide varying time
delays, the preferable delay being in the range of 5 to 6
seconds. After the preset time delay has occurred, the timer
actuates to complete the electrical circuit to a solenoid 122.
The solenoid 122 is a dual coil type, the first or holding coil
being immediately energized when the switch 118 closes, while
the second coil, which causes the solenoid to shift against the
bias of a spring 124, is energized only through the timer 120.
Actuation of the solenoid 122 shifts a fuel injection pump lever
126 to throttle back the engine. A manually actuated switch 128
is preferably provided in the electrical system to disconnect
the system from the power source.
In operation, the excavator operator normally sets the
engine throttle at a relatively high setting to provide adequate
power to drive the pumps 44 and 46. If, however, the operator
interrupts the operation of the machine, such as between truck
loads, the entire control valve bank would be de-activated and
the pressure in the load sensing line 74 would fall off. The
reduced pressure would cause the pressure actuated switch 118 to
close, and presuming that the switch 128 is closed to actuate
the automatic idle system, electric energy would be directed to
the timer 120. After the preset time delay, the power would be
directed to the solenoid 122, to shift the solenoid and thereby
throttle back the engine. Thus, after an appropriate interval
when the hydraulic functions are not being operated, the engine
is automatically throttled back to save on engine wear and
fuel. Obviously, the operator can de-activate the automatic
idle system by simply opening the switch 128.
Also, in operation, the operator normally maintains the
valve 78 in its open condition, as shown. As previously
described, this allows load sensing pressure to be supplied to
both pump displacement control systems 48 and 50 so that the
displacement of both pumps 44 and 46 increases in response to
load. However, if the operator wishes to reduce the flow rate
of the system and thereby provide finer control of the various
motors, he simply closes the valve 78 so that the load sensing
pressure is not provided to the control system 48 for the motor
44, whereby the motor 44 is maintained at minimum displacement.
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