Note: Descriptions are shown in the official language in which they were submitted.
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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 environrnent 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
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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.
Altematively, 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.
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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 pennitted 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
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~hat provides for improved cooling control without the need for a dedicated
engine,
particularly for use in trucks or other land transport vehicles.
SumnnarV
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
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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.
Descrintion 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.
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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.
Descrintion 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
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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
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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 18a. 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."
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Turning to Figure 3, the control module 26 includes a signal input "IcoW" for
receiving a signal "ScoMp" indicating either "high cool" or "low cooi" modes
of operation.
The signal "ScoMp " 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 conipressor
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
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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 deterrnined 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
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
10 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
ernploy a VVPC type pump 24 to accomplish this purpose.
Turning to Figure 4, a detail of the blower control module 30 is shown
configured
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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 "IHioWER" for receiving a signal "SB,A,,ER" indicating
either "blower
on" or "blower off ' modes of operation. The signal "SamWER" 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 blbwer motor, the control module
30 may
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provide a transducer 36 for measuring the oil pressure and a compensating
controlter jts
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 "SB,AwER." 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
I 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
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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 generaily at 54. The temperature control module 46 provides a
controller 56
including three valves V,, V2, and V3 that together define three different
flow
configurations, or patterns of oil flow Fõ 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 V, 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 TL, an operating flow configuration is defined whereby the valve V, is
opened to
perrnit oil to flow to the compressor and blower portions 40 and 42 of the
hydraulic
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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 T. (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
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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 Tt, Two diverter valves Vp, and VD2 may be
employed to be responsive to the two different temperatures as shown in Figure
7.
The diverter valve Vpt has a wax set-point temperature of TH and defines a
straight-through fluid flow path "STRAIGHT-THROUGHI" and a bypass fluid flow
path
"BYPASS,.' Similarly, the diverter valve VDZ has a wax set-point temperature
of TL and
defines a straight-through fluid flow path "STRAIGHT-THROUGH2" and a bypass
fluid
flow path "BYPASSz " To the extent that the temperature at the valve Vp,
increases
beyond its set-point Tl1, 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 Vp2 through the path STRAIGHT-THROUGHi. Conversely, to
the
extent that the temperature at the valve VD2 exceeds its set-point TL, less
'of the flow
received from the valve Vn, 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
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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
17a2 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
V,a
and a diverter valve VD3. 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 Fi, 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
F1eleading 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 Fti6 to a diverter valve
VD,. The
diverter valve VD3 has a wax set-point temperature of Ta and defines a
straight-through
fluid flow path "STRAIGHT-THROUGH3' and a bypass fluid flow path "BYPASS3."
To the extent that the temperature at the valve Vp3 increases beyond its set-
point TH, more
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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.
Tuming to Figure 10, the compressor motor 22 typically has a motor shafl. 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.
Altennatively, 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 intemal 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 fiuther 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
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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
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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.
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