Field of the Invention
The present invention relates to a hydraulic power unit for a refrigeration system,
particularly for use in a land transport vehicle such as a truck.
Background
Refrigeration systems are commonly used in all types of transport vehicles for
transporting perishable items, such as produce. As is typical in refrigeration and air
conditioning systems, such systems include a compressor for compressing a refrigerant
that is received by the compressor in a gaseous form and is compressed into a liquid form. This compression heats the refrigerant and the waste heat is convected away from the
system by passing the refrigerant through a radiator (condenser) downstream of the
compressor. The compressed refrigerant is then passed into an evaporator where it is
allowed to expand into the gaseous form. This expansion cools the fluid which draws
heat from the environment to produce the desired cooling. The gaseous refrigerant is then
returned to the compressor. The amount of cooling is controlled by controlling the speed
of the compressor. The refrigeration system attempts to provide and maintain a desired
temperature in a "box" or storage volume of the vehicle, which is typically a semi-trailer
pulled by a truck but may also be a railroad car pulled by a train engine.
Power for turning the compressor has typically been provided by a dedicated
internal combustion engine having its own dedicated fuel supply. The cooling output is
controlled by controlling the output of the engine. While providing a straight-forward
means for regulating cooling, the dedicated engine has the disadvantage that it adds cost
to the refrigeration system and is typically not as efficient as the engine used to power the
vehicle itself. It is also a drawback of such prior art systems that maintaining two
separate fuel supplies is inconvenient.
Alternatively, in the typical air conditioning system used in passenger vehicles,
power for the system is obtained from the vehicle engine. The power is typically taken
from the engine by belts and pullies and transmitted directly to the compressor.
However, the power provided to the compressor varies with engine speed, which in turn
varies with vehicle speed, so the amount of cooling cannot be controlled independently of
the desired operation of the vehicle. Heat from the vehicle's cooling system can be used
to compensate for over-cooling, but this is energy inefficient. Moreover, there is no mechanism for increasing the cooling if the engine output is too low.
In the context of a marine vehicle refrigeration system, the present inventor solved
the problems associated with both the prior art refrigeration and vehicle air conditioning
systems by powering a refrigeration system from the engine used for propelling the
vehicle through use of a hydraulic transmission system. The hydraulic transmission
system included a pump that was coupled directly to the engine. The engine turned the
pump which in turn pressurized hydraulic fluid in hydraulic fluid lines that carried the
pressurized hydraulic fluid to the remote location of the refrigeration system. A hydraulic
motor received the pressurized hydraulic fluid and was caused to turn as a consequence.
The system has not been known to function outside of the marine environment, however.
W
In particular, the system has not been known to function in a truck or other land transport vehicle.
It was a particular insight of the present inventor to employ a variable volume
pressure compensated pump to pump the hydraulic fluid. It is a characteristic of such
pumps that the pressure output of the pump can be optimized or controlled independent
of engine speed. As far as is known, the inventor's recognition of the advantage of this
type of pump for the purpose of powering a refrigeration system was and continues to be
unique.
Refrigeration systems also typically employ a blower for blowing air through the
evaporator, to increase the efficiency of conducting heat from the environment to the
expanding refrigerant at the evaporator and also for distributing the cooled air throughout the box. Typically, such blowers are directly connected to the compressor, although older units employed electrical power. When connected to the compressor, the blower speed
changes with compressor speed, while electrically powered blowers were typically
operated at a fixed speed.
Precise temperature control of the entire interior of the box can be critical. For
example, while it is necessary to maintain as low a temperature as possible for highly
perishable items, it may be critical that the items not be permitted to freeze. It has been
found that prior art refrigeration systems for truck use have not been entirely satisfactory in this regard.
Accordingly, there is a need for a hydraulic power unit for a refrigeration system
chat provides for improved cooling control without the need for a dedicated engine, particularly for use in trucks or other land transport vehicles.
Summary
The present invention provides for a hydraulic power unit for a refrigeration
system. According to one aspect of the invention, the power unit is provided for driving
the refrigeration system of a truck having an engine for propelling the truck and a power
take off from the engine. The refrigeration system has a compressor for compressing a
refrigerant and an evaporator which is cooled by the compressed refrigerant. The power
unit comprises a pump, a compressor motor, and a hydraulic circuit. The pump is
adapted for pumping hydraulic fluid and for connection to the power take off for driving
the pump. The compressor motor is adapted for driving the compressor in response to
receiving hydraulic fluid from the pump. The hydraulic circuit is adapted for conducting the hydraulic fluid from the pump to the compressor motor and for conducting the
hydraulic fluid from the compressor motor back to the pump. The hydraulic circuit
includes a temperature control portion having a heat exchanger and adapted for diverting
at least a portion of the hydraulic fluid through the heat exchanger in response to a
temperature indication indicating the temperature of the fluid.
According to another aspect of the invention, the refrigeration system further
includes a blower for blowing air through the evaporator. The power unit comprises a
pump, a blower motor, and a hydraulic circuit. The pump is adapted for pumping
hydraulic fluid and for connection to the power take off for driving the pump. The
blower motor is adapted for driving the blower in response to receiving hydraulic fluid
from the pump. The hydraulic circuit is adapted for conducting the hydraulic fluid from
the pump to the blower motor and for conducting the hydraulic fluid from the blower motor back to the pump. The hydraulic circuit includes a temperature control portion
having a heat exchanger and adapted for diverting at least a portion of the hydraulic fluid
through the heat exchanger in response to a temperature indication indicating the
temperature of the fluid.
It is to be understood that this summary is provided as a means of generally
determining what follows in the drawings and detailed description of preferred
embodiments and is not intended to limit the scope of the invention. Moreover, the objects, features and advantages of the invention will be more readily understood upon consideration of the following detailed description taken in conjunction with the
accompanying drawings.
Description of the Drawings
Figure 1 is a schematic diagram of a hydraulic power unit for a refrigeration
system according to the present invention.
Figure 2 is a pictorial, partially cut-away view of a preferred land transport vehicle
for use with the present invention.
Figure 3 is a schematic diagram of a compressor motor control module according
to the present invention.
Figure 4 is a schematic diagram of a blower motor control module according to
the present invention for use with a pump having a substantially constant power output.
Figure 5 is a schematic diagram of a blower motor control module according to
the present invention with compensation for use with a pump subject to varying power
output.
Figure 6 is a schematic diagram of a generalized oil temperature control module
according to the present invention.
Figure 7 is a schematic diagram of the oil temperature control module of Figure 6
implemented with two thermostatic valves.
Figure 8 is a schematic diagram of a preferred hydraulic power unit according to
the present invention.
Figure 9 is a schematic diagram of a preferred oil temperature control module
according to the present invention.
Figure 10 is a schematic diagram of a means for coupling a hydraulic compressor
motor to a compressor according to the present invention.
Description of Preferred Embodiments
Figure 1 is a schematic view of a hydraulic power unit 10 for transmitting power
from an engine 12 to a refrigeration system 14. The present inventor had recognized the
desirability of providing a hydraulic power unit for a refrigeration system that is
particularly adapted for use in a truck and attempted to adapt the marine system described
above for that purpose. However, he discovered through these attempts that the hydraulic
fluid would boil under certain conditions, so that the system was not functional. The
present invention solves this problem.
Accordingly and with reference to Figure 2, the engine, hydraulic power unit, and
refrigeration system are all contained on a land transport vehicle 9, particularly in the
preferred embodiment of the invention a truck adapted for heavy or large cargo transport,
such as a standard semi-trailer truck. The truck has a cargo volume 11 which is referred
to herein as a "box."
The engine 12 is used for propelling the truck and is typically a large internal
combustion engine, most typically a diesel engine. The engine provides a torque output
over a range of engine speeds and is coupled to the driving wheels of the truck through a
transmission 13. The torque output of the engine is made available for powering
auxiliary devices through a power take off ("PTO") 15. As will be readily appreciated by
persons of ordinary mechanical skill, the PTO 15 may be coupled directly to the engine,
transmission, rear end, or other component of the truck's power train, or the PTO may be
coupled to an auxiliary device that is in turn coupled to the engine. The invention
provides the outstanding advantage, however, that the engine 12 is used as the ultimate
source of power provided to the refrigeration system 14.
Turning back to Figure 1, the hydraulic power unit 10 includes a hydraulic pump
24 that is coupled to the power output of the engine through the PTO 15. The hydraulic
pump is adapted to pump hydraulic fluid, typically (and hereinafter) oil, through a
hydraulic circuit 17 under pressure. The hydraulic pump 24 may be any standard type of
pump used in hydraulic systems such as earthmoving equipment. However, preferably,
the pump 24 is of the type known in the art of hydraulic systems as variable volume
pressure compensated ("VVPC")- The VVPC type of pump 24 compensates for both
load and engine speed so as to provide a substantially constant pumping pressure.
As is typical, the refrigeration system 14 includes a compressor 16, a condenser
18, and an evaporator 20 having the usual functions. A refrigerant flows through a
refrigerant circuit 18 through refrigerant carrying lines ISa. The hydraulic power unit 10
drives the compressor; more particularly, the hydraulic power unit 10 includes a hydraulic
compressor motor 22 for this purpose.
The hydraulic circuit 17 includes hydraulic oil carrying lines 17a that carry and
route the hydraulic oil that is pressurized by the pump 24. The hydraulic circuit routes
the pressurized hydraulic oil to the compressor motor 22 as well as to a compressor motor
control module 26 for controlling the amount of the hydraulic oil that is provided to the
compressor motor. The compressor motor 22 and the control module 26 are coupled in parallel.
Particularly, both the compressor motor 22 and the control module 26 receive hydraulic
oil from the circuit 17 at "A," and both the compressor motor and the control module 26
output hydraulic oil at "B." The control module 26 controls the amount of oil provided to
the compressor motor 22 by accepting (shunting) more or less of the oil through the
control module. In a preferred embodiment of the invention, the control module 26
provides for just two operating modes of control of the compressor motor, "high cool"
and "low cool."
Turning to Figure 3, the control module 26 includes a signal input "ICOMP" for
receiving a signal "SCOMP" indicating either "high cool" or "low cool" modes of operation.
The signal "SC0MP " may be generated electrically, mechanically, hydraulically, or
pneumatically and is selected by a user of the system such as by use of a toggle or rotary
switch.
A binary state, flow control valve 23 of the control module 26 is either "open" or
"closed." When the signal indicates "high cool" mode, the valve 23 is closed so that
substantially no hydraulic oil is shunted away from the compressor motor 22; substantially all of the hydraulic oil flowing in the line 17a passes through the compressor
motor. When the signal indicates "low cool" mode, the valve 23 is opened so that a set amount of the hydraulic oil is shunted away from the compressor motor 22. Preferably, a
flow-set valve 25 is used to set the proportion of the oil that is accepted through the control module 26 rather than being provided to the compressor motor 22. The valve 25
may provide for a fixed or adjustable flow rate, and if the latter may easily be manually
pre-set to determine the flow in low cool mode. The valve 25 may also provide for
additional cooling modes, and may provide for a continuous range of adjustment, and
therefore a continuous range of cooling output, either manually or automatically, remotely
or locally.
As an example of setting the valve 25 for two cooling modes, the compressor
motor 22 may turn 1800 rpm in high cool mode and only 1400 rpm in low cool mode.
Where, for example, 10.5 gallons are required to turn the motor 1800 revolutions, to a
first approximation about 1400/1800 gallons (0.78) would be required to turn the motor
1400 revolutions. Thence, (1 - 0.78) X 10.5 gallons (2.3 gallons) would be shunted
through the valve 23, or about 22% of the total flow. The actual amount of flow set by
the valve 25 is best determined empirically.
Preferably, the hydraulic power unit 10 also includes a fan or- blower for blowing
air through the evaporator 20 and thereby increasing the efficiency of heat transfer
between the air and the evaporator as well as distributing the cooled air throughout the
box 11 (Figure 2). More particularly, referring back to Figure 1 , the hydraulic power unit
10 includes a hydraulic blower motor 28 for mechanically driving a blower 29. The
hydraulic circuit 17 routes the pressurized hydraulic oil to the blower motor 28 as well as
to a blower motor control module 30 for controlling the amount of the hydraulic oil that is
provided to the blower motor.
As for the compressor motor and its associated control module, the blower motor 28 and the blower control module 30 are coupled in parallel. Particularly, both the blower motor 28 and the blower control module 30 receive hydraulic oil from the circuit
17 at "B," and both the blower motor and the blower control module output hydraulic oil at "C." The blower control module 30 controls the amount of oil provided to the blower
motor 28 by accepting more or less of the oil through the control module.
It is recognized herein that it is desirable to maintain the speed of the blower
motor 28 to be substantially constant, or at least independent of the speed of the engine 12
or the load of the hydraulic circuit 17. It is further recognized that it is desirable to
employ a VVPC type pump 24 to accomplish this purpose.
Turning to Figure 4, a detail of the blower control module 30 is shown configured
for the simple case where the VVPC pump 24 is used. In that case, the blower control
module 30 may simply provide for a "blower on" and a "blower off mode of operation,
the blower motor speed being governed by the pump 24. The blower control module 30
includes a signal input "IBLOWER" f°r receiving a signal "SBL0WER" indicating either "blower
on" or "blower off modes of operation. The signal "SBLOWER" maY be generated electrically, mechanically, hydraulically, or pneumatically and is selected by a user of the
system.
A binary state, flow control valve 33 of the control module 30 is either "open" or
"closed." When the signal indicates "blower on" mode, the valve 33 is closed so that
substantially no hydraulic oil is shunted away from the blower motor 28; substantially all
of the hydraulic oil flowing in the line 17a passes through the blower motor. When the signal indicates "blower off mode, the valve 33 is opened so that substantially all the
hydraulic oil is shunted away from the blower motor 22.
Similar to the compressor control module 26, the blower control module 30 may
be modified to provide for two blower speeds, or additional blower speeds, and may
provide for a continuous range of adjustment of blower speed, and therefore a continuous
range of blower output, either manually or automatically, remotely or locally.
Turning to Figure 5, where the output of the pump 24 is variable, the blower
control module may include a variable flow-set valve 35 that is automatically controlled
to compensate for variations in the pressure of the hydraulic fluid. Since the power
provided to the blower motor is defined by the rate of flow of the oil to the blower motor
multiplied by the pressure of the oil at the blower motor, the control module 30 may
provide a transducer 36 tor measuring the ml pressure and. a compensating controller ό$
for receiving the output of the transducer 36 and automatically adjusting the flow rate of
the valve to compensate for changes in the pressure. The desired speed of the blower
may be provided as a set-point with the signal "SBL0WER." Changes in pressure may also
be deduced, for example, by monitoring the speed of the engine 12. The compressor
control module 26 can be similarly adapted to compensate for variable pump output.
Turning back to Figure 1, the compressor motor 22 and the compressor motor
control module 26 may be considered to define a compressor portion 40 (shown in Figure
1 between "A" and "B") of the hydraulic circuit 17, where the blower motor 28 and the
blower motor control module 30 define a blower portion 42 (shown in Figure 1 between
"B" and "C") of the hydraulic circuit. While the compressor and blower portions of the
circuit 17 are shown in series in Figure 1, it should be understood that they may be provided in parallel with no loss of generality.
Regardless, the two circuit portions are together coupled in series with a
temperature control portion 44 of the circuit 17. The temperature control portion 44
provides for controlling the temperature of the oil to protect the compressor and blower
motors and to ensure that these components operate at peak efficiency.
Referring to Figure 6, the temperature control portion 44 of the hydraulic circuit
17 includes an oil temperature control module 46, a heat exchanger 48 and an oil
reservoir 50. A hydraulic line 17a, (Figure 1) routes the hydraulic oil from the blower
portion 42 of the hydraulic circuit 17 to the oil reservoir 50.
The heat exchanger is provided for cooling oil that is too hot, however the heat
exchanger could be used for heating oil that is too cold, and two heat exchangers could be
used to both cool oil that is too hot and heat oil that is too cold with slight modification to
the temperature control module 46 as will be readily apparent to persons of ordinary skill.
The heat exchanger can exchange heat with the air cooled by the refrigeration system 14
or may be cooled by air, water, oil or other fluid provided from an external source.
The oil is preferably always passed through the reservoir 50, however this is not
essential to the invention. The reservoir 50 provides room for the oil to expand as it is
heated, and it provides for the removal of bubbles in the oil.
The control module 46 receives oil from the pump 24 and senses the oil
temperature, or receives an indication thereof from another source, the sensing being
indicated generally at 54. The temperature control module 46 provides a controller 56 including three valves V1, V2, and V3 that together define three different flow configurations, or patterns of oil flow F1, F2, and F3, depending on the sensed temperature
of the oil. If the oil is too cold, i.e., less than a predetermined minimum TL (not shown),
the controller defines a warm-up flow configuration whereby the valve V1 is closed to
prevent the oil from reaching the point "A" in Figure 1 and thereby to prevent the oil
from reaching the compressor or blower motors. The valve V2 is also closed to prevent
flow to the heat exchanger. The valve V3 is open to recirculate the oil to the pump 24, in
this case by passing it through the reservoir 50 which in turn returns the oil to the pump.
When the oil reaches a desired operating temperature, i.e., the temperature
exceeds Tu an operating flow configuration is defined whereby the valve V1 is opened to
permit oil to flow to the compressor and blower portions 40 and 42 of the hydraulic
circuit 17 through point A (Figure 1). The valve V3 is closed to cease recirculating oil to
the pump and the valve V2 leading to the heat exchanger remains closed.
When the oil is about to become too hot, i.e., the temperature reaches a pre-set
higher temperature limit TH (not shown), an over-temperature flow configuration is
defined whereby the valve V2 is opened to permit oil to flow through the heat exchanger
48, to cool the oil. If the oil becomes dangerously hot, the valve "V," permitting flow to
the compressor and blower portions 40 and 42 of the circuit 17 may also be closed.
The valves "V" may be solenoid controlled in response to electrical signals issued
by an electrical controller 56, where the electrical controller receives an electrical signal
from a sensor 54 having an electrical signal output for indicating the temperature.
However, in the preferred embodiment of the invention, the controller 56 and the valves
V are provided in the form of "three-way thermostatic control valves" that provide the advantage of automatic control without the need for any electrical or other source of
power. Such valves are commercially available, e.g., from Fluid Power Energy, Inc. of
Waukesha, Wisconsin.
Three-way thermostatic control valves (hereinafter "diverter valves") employ a
semi-liquid wax that undergoes large expansion within a relatively narrow temperature
range. The expansion of the wax provides for movement of a slider sleeve which
provides positive three-way valve action. The valves are factory set at predetermined
temperatures. A single diverter valve provides for a "straight-through" fluid flow path
and a "bypass" fluid flow path. If the fluid temperature is below a threshold, the valve
fully closes the bypass fluid flow path and the straight-through path is fully open. When
the temperature reaches the threshold, the valve partially opens the bypass path and
partially closes the straight-through path. As the temperature continues to rise, the valve
more completely opens the bypass path and more completely closes the straight-through
path until the bypass path is fully open and the straight-through path is fully closed.
In the simplest embodiment of the temperature control portion 44 of the hydraulic
circuit 17 as described above, the valve V, is closed when the oil temperature is below TL,
the valve V2 is closed when the oil temperature is below TH, and the valve V3 is closed
when then the temperature is above TL. Two diverter valves VD, and V02 may be employed to be responsive to the two different temperatures as shown in Figure 7.
The diverter valve VD, has a wax set-point temperature of TH and defines a straight-through fluid flow path "STRAIGHT-THROUGH," and a bypass fluid flow path "BYPASS,." Similarly, the diverter valve VD2 has a wax set-point temperature of TL and defines a straight-through fluid flow path "STRAIGHT-THROUGH2" and a bypass fluid
flow path "BYPASS2." To the extent that the temperature at the valve VDI increases
beyond its set-point Tn, more of the flow received from the pump 24 is diverted to the
heat exchanger 48 through the path BYPASS, and less of the flow is transmitted straight
through to the valve VD2 through the path STRAIGHT-THROUGH,. Conversely, to the
extent that the temperature at the valve V02 exceeds its set-point Tu less of the flow
received from the valve V01 is diverted to the reservoir 50 through the path STRAIGHT-
THROUGH2 and more of the flow is transmitted through the path BYPASS3 to the
compressor and blower circuit portions 40 and 42 through the point A (Figure 1).
Figure 8 shows a preferred hydraulic power unit 100 for the refrigeration system
14 of Figure 1. The power unit 100 is substantially the same as the power unit 10 of
Figure 1 (and therefore retains the same reference designators) except that a hydraulic line
Ha2 routes the hydraulic oil from the blower portion 42 of the hydraulic circuit 17 to a
temperature control module 46a rather than to the reservoir 50.
Figure 9 shows the oil temperature control module 46a of the preferred
embodiment in more detail. The temperature control module 46a includes a valve Vla
and a diverter valve V03. If the oil is determined to be at or above a desired operating
temperature, the valve V |a routes the oil through a flow path Fla leading to the compressor
and blower portions 40 and 42 of the hydraulic circuit 17 through point A (Figure 8). At
the same time and to the same extent, oil is prevented from flowing through the flow path
F pleading to the heat exchanger 48 and the reservoir 50. The valve may variably apportion the flow between the two paths but is preferably a binary state valve that provides for full flow through a selected one of the flow paths while completely
preventing flow through the other of the flow paths. The valve is preferably simply
operated by hand, but it may be adapted for electrical control for remote manual operation, or may be part of an automatic temperature control system that measures or
otherwise responds to the oil temperature and adjusts the valve accordingly.
Where the valve V]a is set to route oil to either the heat exchanger 48 or the
reservoir, the oil is caused to flow through the path Fjb to a diverter valve V03. The
diverter valve V03 has a wax set-point temperature of TH and defines a straight-through
fluid flow path "STRA1GHT-THROUGH3" and a bypass fluid flow path "BYPASS3."
To the extent that the temperature at the valve VD3 increases beyond its set-point TH, more
of the flow received from the pump 24 is diverted to the heat exchanger 48 through the
path BYPASS3 and less of the flow is transmitted straight through to the reservoir 50
through the path STRAIGHT-THROUGH3. Oil received from point C (Figure 1) through the hydraulic line 17a2 is also provided to the input "I" of the diverter valve VD3 for processing through the diverter valve.
Turning to Figure 10, the compressor motor 22 typically has a motor shaft 22a and
the compressor 16 has a compressor shaft 16a. Typically, prior art compressors that are
not coupled directly to an internal combustion engine include a pulley adapted to receive
a belt for driving the shaft 16a. For example, an electric compressor motor would
typically include a shaft having a first pulley and the compressor shaft 16a would include
a second pulley. A belt couples the first pulley to the second pulley. The pulley has been provided for the purpose of adjusting the gearing ratio between the two shafts.
Alternatively, where the compressor 16 is coupled directly to an internal combustion engine, an axial coupler is typically used to coaxially couple the shaft of the
internal combustion engine to the shaft 16a.
The present inventor has recognized that in the case of coupling directly to an
internal combustion engine, the rotating mass of the internal combustion engine provides
a flywheel effect that is important for smoothing vibrations emanating from the
compressor, and that this function was provided by the pulley when a belt drive system was used.
It was further recognized that the hydraulic compressor motor 22 operates more
like an internal combustion engine in terms of the variation in engine speeds that it can
1?
provide, so that employing a pulley system for changing gear ratios is unnecessary.
Thence, according to the present invention, a coupler 62 is preferably employed that
coaxially couples the shaft 22a to the shaft 16a, and a vibration dampener 64 is preferably
added to the system to smooth the vibrations. The vibration dampener is preferably a
metal disk or flywheel that is mounted to either the shaft 16a or the shaft 22a but may
have other configurations. While a flywheel or other vibration dampener is not essential
to the invention, the hydraulic motor 22 has a relatively
low mass and the compressor 16 typically produces a high level of vibration, so that the
vibration dampener is highly desirable in practice. The vibration dampener is also
preferably dynamically balanced, and is further preferably dynamically balanced on the
shaft with the power unit and refrigeration system in full operation.
The compressor motor 22 and the compressor 16 are preferably both mounted, e.g., by bolting or welding, to a rigid mount 60 so that alignment between the compressor
motor shaft 22a and the compressor shaft 16a can be reliably maintained. To minimize
the effect of any misalignment, the coupler 62 is preferably flexible, such as by having at
least a joint portion 62a formed of rubber. Further, to provide for operator safety, an
enclosure 66 is provided to prevent inadvertent access to rotating parts.
It is to be recognized that, while a hydraulic power unit for a refrigeration system
has been shown and described as preferred, other configurations and methods could be
utilized, in addition to those already mentioned, without departing from the principles of
the invention. For example, the logic described above for providing the oil temperature
control, compressor control, and blower control portions of the hydraulic circuit 17 could
18
SUBSTITUTE SHEET (JRULE 26)
be implemented by various means, automatic, semi-automatic, or manual, distributed or
integrated, in any combination of electrical, mechanical, hydraulic, and pneumatic
elements and circuits, as will be readily appreciated by persons of ordinary skill.
The terms and expressions which have been employed in the foregoing
specification are used therein as terms of description and not of limitation, and there is no
intention in the use of such terms and expressions to exclude equivalents of the features
shown and described or portions thereof, it being recognized that the scope of the
invention is defined and limited only by the claims which follow.