Note: Descriptions are shown in the official language in which they were submitted.
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METHODS AND APPARATUS fOR OPERATING
A REFRIGERATION SYSTEM CHARACTERIZED BY
CONTROLLING ENGINE COOLANT
TECHNICAL FIELD
The invention relates in general to refrigeration
systems, and more specifically to refrigeration systems
which utilize a compressor having an intermediate pressure
port.
BACKGROUND ART
U.S. Patent 4,850,197, which is assigned to the
same assignee as the present application, discloses a vapor
compression refrigeration system based on an' economizer
cycle which utilizes a refrigerant compressor having an
intermediate pressure port, in addition to suction and
discharge ports. An economizer heat exchanger is used to
enhance hot gas cooling and heating cycles which are
initiated by associated electrical or electronic control to
achieve and maintain a predetermined temperature range
close to a selected set point temperature in a served space
to be conditioned.
U.S. Patent 5,174,123 issued December 29, 1992
entitled Methods and Apparatus for Operating a Refrigera
tion System, which is assigned to the same assignee as the
present application, discloses refrigeration methods and
apparatus which utilize a flash tank in a refrigeration
system which has an economizer cycle, in place of an
economizer heat exchanger. The refrigeration arrangement
disclosed in the aforesaid application eliminates the need
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for a float valve in the flash tank, enabling the flash
tank to be used in transport refrigeration applications.
It would be desirable, and it is an object of the
present application, to improve the reliability and effi
ciency, as well as the control methods and arrangements, of
refrigeration systems which have an economizer cycle, such
as the refrigeration systems disclosed in the hereinbefore
mentioned patent and patent application.
SUMMARY OF THE INVENTION
The invention includes methods and apparatus for
operating a refrigeration system which achieves and holds
a predetermined set point temperature in a conditioned
space via cooling and heating cycles. The refrigeration
system includes a refrigerant compressor having a suction
port, an intermediate pressure port., a discharge port, and
a compressor prime mover. The refrigeration system further
includes a hot gas compressor discharge line, first and
second hot gas lines, and first controllable valve means
having first and second positions which respectively
connect the hot gas compressor discharge line to the first
and second hot gas lines. A main condenser is connected to
the first hot gas line. An evaporator, which is associated
with the conditioned space, includes an evaporator expan-
sion valve. An auxiliary condenser is associated with the
conditioned space, with the auxiliary condenser being
connected to the second hot gas :Line. Economizer heat
exchanger means having first and second refrigerant flow
paths is provided, including an economizer expansion valve
which controls the rate of refrigerant flaw through the
second refrigerant flow path. A main liquid line connects
the main condenser to the evaporator expansion valve via
the first refrigerant flow path of the economizer heat
exchanger means, an auxiliary liquid line connects the
auxiliary condenser to the economizer heat exchanger means,
a main suction line connects the evaporator to the suction
port of the compressor, and an auxiliary suction line
connects the second flow path of the economizer heat
exchanger means to the intermediate. pressure port of the
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compressor. Second controllable valve means having first
and second positions is disposed to block and unblock the
main liquid line, and third controllable valve means is
disposed to selectively add heat to the economizer heat
exchanger means.
At least one predetermined null related pattern of
open/closed valve positions for the first, second and third
controllable valve means is provided, and a null cycle or
null operating mode is initiated when the temperature of
the conditioned space is in a predetermined null tempera
ture range adjacent to the predetermined set point tempera
ture while maintaining operation of the refrigerant com
pressor. The initiation of the mull operating mode in
cludes selecting the at least one predetermined null
associated pattern of open/closed valve positions.
In a preferred embodiment of the invention, a
plurality of predetermined patterns of valve positions are
provided, each implementing a slightly different null.
operating mode. At any given instant a null related
controllable valve position pattern is selected which will
result in the net heat gain or loss of the conditioned
space being substantially matched by the heat added to or
removed from the conditioned space by the evaporator and
auxiliary condenser.
Another embodiment of the invention includes
providing a refrigerant vent orifice which automatically
drains refrigerant from a heating related circuit to a low
pressure part of a cooling circuit, during each cooling
cycle. A preferred location for the vent orifice connects
one end at a junction between the auxiliary condenser and
a drain pan heating coil, and the other end is connected
either to a refrigerant distributor which distributes
refrigerant to the evaporator coil, or to a main suction
line downstream from an evaporator coil. These connection
points are preferred as only a short length of tubing is .
required, and the location provides defrosting of the vent
orifice tubing during a defrost cycle.
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In another embodiment of the invention, the
economizer heat exchanger means provides the second refrig-
erant path via an outer shell or :housing having an inlet
and an outlet, with the shell surrounding the first refrig-
erant path. Refrigerant is expanded into the shell via the
economizer expansion valve, to provide refrigerant expan-
sion of the flooded type, as no super heating exists with
this configuration. A refrigerant and compressor oil drain
line is connected from a low point of the shell to a
selected one of two locations. The first location is a
still lower point of the auxiliary suction line which
interconnects the outlet of the shell to tree intermediate
pressure port of the compressor. The second location is to
a higher point of the auxiliary suction line than the drain
elevation. A "lift" of the oil-liquid refrigerant mixture
is provided by running the drain line vertically along, and
in heat exchange relation with, a section of the liquid
line to create a percolator effect. This drain line
provides two advantages. By removing compressor oil which
2o has been carried into the shell of the economizer heat
exchanger means, which oil is at least partially miscible
with liquid refrigerant, the oil concentration in the
boiling refrigerant in the shell i~> reduced, resulting in
a dramatic increase in the heat transfer characteristic
between the first and second refrigerant flow paths.
Depending upon the current running condition this increase
is usually between 20% and 60%. By metering the flow of
liquid refrigerant from the shell back to the intermediate
pressure port of the compressor, an evaporation of refrig-
erant occurs in the compressor which cools the compressor
and thus limits the discharge temperature of the compres-
sor.
In still another embodiment of the invention
compressor discharge pressure and load on the compressor
prime mover are kept within desirable limits without
causing either a heating cycle or. a cooling cycle to
suffer, and without adding another restrictive valve to the
system, by using maximum operating pressure (MOP) expansion
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valves for both the evaporator expansion valve and for the
economizer expansion valve. The MOP evaporator expansion
valve controls the operating pressure during a cooling
cycle, and the MOP economizer expansion valve controls the
5 operating pressure during a heating/defrost cycle. Thus,
the maximum operating pressure setting for each MOP valve
is selected for the specific cycle it is associated with,
with the setting on one MOP expansion valve applying no
restriction on capacity during a cycle in which the other
MOP expansion valve is in control.
In another embodiment of the invention, an
economizer by-pass valve is connected between the auxiliary
and main suction lines, and in addition to operating the
valve in a conventional manner, eg., open during a heat-
ing/defrost cycle, and either closed during a cooling
cycle, or controlled to provide compressor unloading during
a cooling cycle for temperature control of a conditioned
space, it is used to provide engine load management, which
is especially required during high ambient temperatures.
2o The load on the prime mover is monitored, such as by
monitoring compressor discharge temperature, or by monitor-
ing the temperature of the prime mover, and the by-pass
valve is additionally controlled as a function of the
detected load. When a certain load is reached, the by-pass
valve is opened, and when the load drops to a predetermined
value, the valve is closed.
In still further embodiments of the invention,
which embodiments relate to when the compressor is driven
by a liquid cooled engine, and to a eompressor which has an
external oil cooler cooled by the engine coolant, the
connection of the oil cooler in the engine coolant circuit
is selected according to the type of thermostat used in the
engine coolant circuit. When a thermostat of the choke
type is used, ie . , a thermostat which has a single inlet
and a single outlet, and which is practically closed below
a predetermined temperature, the oil cooler is connected on
the upstream side of the input to the thermostat. Thus,
there is coolant flow to the oil cooler even when the
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thermostat is closed. When the thermostat is of the by-
pass type, ie., a thermostat with two inputs and a single
outlet, with the thermostat controlling the percentage of
flow into the two inlets, the oil cooler is located down-
s stream from the outlet, again assuring a constant flow of
coolant, regardless of the regulating action of the by-pass
thermostat at any'instant.
BRIEF DESCRIPTION OF THE DRAWINGS
The invention will become more apparent by reading
the following detailed description in conjunction with the
drawings, which are shown by way of example only, wherein:
Figure 1 illustrates a refrigeration system
constructed according to the teachings of the invention,
with the refrigeration system having an economizer cycle;
Figure 2 schematically illustrates a more detailed
arrangement for implementing the portion of the apparatus
shown in Figure 1 which is directly associated with the
space to be conditioned;
Figure 3 illustrates a modification of the
refrigeration system shown in Figure 1, related to the
cooling of a lubricant utilized by a refrigerant compres
sor, utilizing engine coolant and a single thermostat of
the by-pass type;
Figure 4 illustrates another modification of the
refrigeration shown in Figure 1, related to the cooling of
a lubricant utilized by a refrigerant compressor, utilizing
engine coolant and a single thermostat of the choke type;
Figure 5 illustrates a control algorithm having
heating, cooling and null operating modes which are imple
mented according to the teachings of the invention; and
Figure 6 illustrates other modifications of the
refrigeration system shown in Figure 1, according to the
teachings of the invention.
DESCRIPTION OF PREFERRED EMBODIMENTS
As used in the following description and claims,
the term "conditioned space" includes any space to be
temperature and/or humidity controlled, including station-
ary and transport applications, for the preservation of
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foods and other perishables, maintenance, of a proper
atmosphere for the shipment of industrial products, space
conditioning for human comfort, and the like. The term
"refrigeration system" is used to generically cover both
air conditioning systems for human comfort, and refrigera-
tion systems for preservation of perishables and shipment
of industrial products. When it is stated that the temper-
ature of a conditioned space is controlled to a selected
set point temperature, it is to be understood that the
l0 temperature of the conditioned space is controlled to a
predetermined temperature range adjacent to the selected
set point temperature. In Figure 1, controllable valves
which are normally open (n.o.) are illustrated with an
empty circle, and controllable valves which are normally
closed (n. c.) are illustrated with an "X" within a circle.
Of course, the associated electrical or electronic control,
hereinafter called "electrical control", may be changed to
reverse the de-energized states shown. An arrow ,pointed
toward a valve in Figure 1 indicates that the valve is
controlled by the associated electrical control.
Referring now to the drawings, and to Figure 1 in
particular, there is shown a refrigeration system 10
constructed according to the teachings of the invention.
Refrigeration system 10 is of the type having an economizer
cycle, including a refrigerant compressor 12 having a
suction port S, a discharge port D, and an intermediate
pressure port IP. Compressor 12 is driven by a prime mover
14, which, in a preferred embodiment of the invention,
includes a liquid cooled internal combustion engine, such
as a diesel engine, linked to compressor 12 as indicated
generally by broken line 16. Prime mover 14 may also
include an electric motor, as the sole prime mover, or as
a stand-by prime mover.
A compressor hot gas discharge line 18 connects
the discharge port D of compressor 12 to first controllable
valve means 20 via a discharge service valve 22. The first
controllable valve means 20 connects the compressor hot gas
discharge line 18 to a selected one of first and second hot
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gas lines 24 and 26. As illustrated in Figure 1, the first
controllable valve means 20 may include a n.c. pilot
solenoid valve 28 and a three-way valve 30. Pilot solenoid
valve 28 selectively connects the; low pressure side of
compressor 12 to the three-way valve 30, such as by tapping
a main suction line 32 via a tee 34, with the main suction
line 32 being connected to the suction port S of compressor
12 via a suction line service valve 36. Pilot solenoid
valve 28 is operably controlled by electrical control 38
via means indicated generally by arrow 29. When pilot
solenoid valve 28 is de-energized <~nd thus closed, three-
way valve 30 interconnects the compressor hot gas discharge
line 18 to the first hot gas line :Z4, and when electrical
control 38 energizes and opens pilot solenoid valve 28,
three-way valve 30 is operated by compressor pressure~to
interconnect compressor hot gas discharge line 18 to the
second hot gas line 26.
The first and second hot gas lines 24 and 26
respectively direct hot compressor discharge gas to cooling
2o and heating circuits 40 and 42. The cooling circuit 40
includes main refrigeran L condenser means 44 which includes
a condenser coil 46 and condenser a:ir mover means 48. The
first hot gas line 24 is connected to an inlet side of
condenser coil 46, and an outlet side is connected to an
inlet 51 of a refrigerant receiver 50 via a main liquid
line 52 which includes a check valve 54. The cooling
circuit 40 and main liquid line 52 continues from an outlet
53 of receiver 50 to an inlet side of an evaporator expan-
sion valve 56, via a refrigerant dehydrator or dryer 58,
economizer heat exchanger means 60, and second controllable
valve means 62, such as a n.o. solenoid valve operably
controlled by electrical control 38 via means indicated
generally by arrow 63.
Economizer heat exchanger means 60 includes first
and second refrigerant flow paths 64 and 66, respectively,
with the first refrigerant flow path 64 including a heat
exchanger coil 68 in the liquid .Line 52. The second
refrigerant flow path 66 includes a shell or housing 70
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disposed to surround heat exchanger coil 68, with shell 70
having a refrigerant inlet 72 and a refrigerant outlet 74.
The second flow path 66 taps the main liquid line 52 via a
tee 76 and a conduit 77, with an economizer expansion valve
78 being connected in conduit 77 between tee 76 and shell
inlet 72. Thus, a portion of t:he liquid refrigerant
flowing through the main liquid line 52 is diverted through
the economizer expansion valve 78 into the second refriger-
ant flow path 66, expanding refrigerant into shell 70 and
providing an economizer cycle by subcooling liquid refrig-
erant flowing through heat exchanger coil 68. Shell outlet
74 is connected to the intermediate pressure port IP of
compressor 12 via an auxiliary suction line 80 and a
service valve 82. Refrigerant in shell 70 is at a.higher
pressure than refrigerant returning to suction port S of
compressor 12, and is thus returned to the higher pressure
intermediate port IP.
Economizer heat exchanger means 60 also includes
heating means 84 for selectively adding heat to the refrig
erant flowing through economizer heat exchanger means 60.
Heating means 84, in a preferred embodiment of the inven-
tion in which the prime mover 14 includes a liquid cooled
internal combustion engine, includes a heating or water
jacket 86 connected to receive liquid coolant from prime
mover 14 via third controllable valve means 88, which may
be a n.c. solenoid valve operably controlled by electrical
control 38 via means indicated generally by arrow 89.
Liquid coolant from a liquid coolant circuit associated
with prime mover 14 enters an inlet snide of water jacket 86
via a first liquid flow conduit 90, and liquid coolant is
returned from water jacket 86 to a water pump 92 via a
second liquid flow conduit 94. Valve 88 and conduit 90 tap
the liquid circuit of the prime mover 14 without going
through a thermostat T associated with prime mover 14.
Refrigerant flow rate through the second refrigerant flow
path 66 is controlled by the economizer expansion valve 78
as a function of the refrigerant temperature at the outlet
74, as indicated by thermal bulb 96.
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When prime mover 14 is an electric motor, heating
jacket 86, instead of being a water jacket, may be an
electrical resistance coil, with the third controllable
valve means 88 being replaced by an on/off switch. Also,
5 while heat is preferably added t« the external side of
shell 70, it is to be understood that liquid coolant may be
directed to a heat exchanger coil disposed within shell 70,
and electrical resistors, instead of heating the external
side of shell 70, may be disposed within shell
10 The cooling circuit 40 continues from evaporator
expansion valve 56, which separates high arid low pressure
sides of the cooling circuit 40, via a refrigerant distrib-
utor 98 which distributes refrigerant to evaporator means
100. Evaporator means 100 includes an evaporator coil 102,
which has a plurality of flow paths receiving refrigerant
from distributor 98, and evaporator air mover means 104.
Air mover means 104 circulates aii- between a conditioned
space, indicated generally at 106, and the evaporator coil
102. An outlet side of evaporator coil 102 is connected to
the hereinbefore mentioned main suction line 32, to return
refrigerant to suction port S of compressor 12. The flow
through the first flow path 64 of economizer heat exchanger
means 60 is thus controlled by the evaporator expansion
valve 56, which controls flow rate according to the degree
of superheat in the refrigerant vapors leaving evaporator
coil 102, as indicated by thermal bulb 107.
The heating circuit 42 includes the second hot gas
line 26, an auxiliary condenser 108, and an auxiliary
liquid line 110. Auxiliary condenser 108 is associated
with evaporator means 100 and is thus also in heat exchange
relation with conditioned space 106. The second hot gas
line 26 is connected to an inlet side of auxiliary condens-
er 108, and an outlet side of auxiliary condenser 108 is
connected to the auxiliary liquid line 110. Auxiliary
liquid line 110 taps the.main liquid line 52 via a tee 112,
with a check valve 114 being disposed in auxiliary liquid
. line 110 to prevent flow from the main liquid line 52 to
the auxiliary condenser 108.
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In a preferred embodiment of the invention the
auxiliary condenser 108 is divided into first and second
serially connected sections 116 and 118 which respectively
function as a defrost pan heater coil and a heating coil
for adding heat to conditioned space 106. Figure 2 is a
schematic representation of a suitable implementation of
the evaporator means 100 and auxiliary condenser means 108
in which the heating coil 118 is implemented by using one
row or refrigerant flow path of a pT~urality of rows or flow
paths which make up the evaporator' coil 102. Return air
from conditioned space 106, indicated by arrow 120, is
drawn into a plenum 122 by air mover means 104, and air is
forced to flow through a plurality of refrigerant flow
paths which include flow paths of evaporator coil 102 and
one or more flow paths associated with auxiliary condenser
108, with heating coil 118 being one or move of the rows of
heat exchanger tubes in a structure which makes up evapora-
tor coil 102, as hereinbefore stated. The location of
heating coil 118 relative to the air flow direction through
plenum 122 depends upon the specific application of refrig-
eration system 10. If de-humidifying is a requirement of
the application, a tube location or row close to the
entering air would be selected, as illustrated in Figure 2.
If de-humidifying is not a requirement, the selected row
may be centered to enhance a defrosting cycle of evaporator
coil 102. Even when heating coil 118 is close to the
entering side of the air flow, however, defrosting is
rapid, as a controllable defrost damper 124, controlled by
electrical control 38, is closed during defrost, which
circulates air rapidly about all of the rows of the tube
bundle which makes up the evaporator coil 102, spreading
heat from the heating coil 118 rapidly to all rows of the
structure. The discharge or conditioned air, indicated by
arrow 126, is forced to flow back into conditioned space
106 by air mover means 104. Return air and discharge air
temperature sensors 128 and 130 provide control signals for
electrical control 38. As shown in Figure 1, an ambient
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air temperature sensor 132 may also provide an input to
electrical control 38.
In a desirable embodiment of the invention a
refrigerant vent line 133 is provided, witr~ vent line 133
having a predetermined orifice size, as indicated at 134.
The vent line 133 is connected to apply suction pressure to
the heating circuit 42 during a cooling cycle, to enhance
the cooling cycle without adding to the overall refrigerant
requirements of the system, by forcing refrigerant trapped
in the heating circuit 42 into the cooling circuit 40.
Refrigerant vent line 133 is connected between the heating
circuit 42, which includes the circuit between three-way
valve 30 and check valve 114, ie., the second hot gas line
26, auxiliary condenser 108, and auxiliary liquid line 110,
and the low pressure side of the cooling circuit 40, ie.,
between the outlet side of evaporator expansion valve 56
and suction port S of compressor 12. In a preferred
embodiment of the invention the defrost pan coil 116 is
connected in series with the heating coil 118, and the
refrigerant vent line 133 is connected from a junction or
tee 136 between coif 116 and 118 to one of two predeter-
mined points. In the embodiment of the invention shown in
Figure 1, the vent line 133 is connected to the refrigerant
distributor 98. Figure 6, to be hereinafter explained,
illustrates the other predetermined point. These preferred
arrangements have the advantages of minimizing the length
of the vent line 133, and of providing defrosting of the
vent line 133 during a defrost cycle. Since during a
heating/defrost cycle the vent line 133 will create a
capacity loss, the vent orifice 134 is preferably selected
to be in a range of about .03 to .1 inch (.8 - 2.5 mm), to
minimize this capacity loss during a heating/defrost cycle.
In another desirable embodiment of the invention
a compressor oil drain line 138 is connected from a low
point 140 of shell 70 to one of two predetermined points.
In the embodiment of the invention shown in Figure 1, the
oil drain line is connected to a still lower point, eleva-
tion-wise, on auxiliary suction line 80, with the lower
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elevation connection to auxiliary suction line 80 being
indicated by tee 142. Figure 6, to be hereinafter ex-
plained, illustrates the other predetermined point, which
is a higher point, elevation-wise, on the auxiliary suction
line 80 than the drain point 140. Compressor oil that is
carried out into the system with the hot gas discharge from
compressor 12 is at least partially miscible with liquid
refrigerant in shell 70. Compressor oil which collects in
shell 70 decreases the heat transfer efficiency between the
flooded type evaporation taking place in shell 70 and heat
exchanger coil 68. In the Figure 1 embodiment of the drain
line 138, drain line 138 was found to function well when
constructed using tubing having an outside diameter (OD) of
.25 inch (6.35 mm) and an orifice of .09 inch (2.3 mm).
Drain line 138 thus provides the advantage of reducing the
concentration of compressor oil in shell, increasing the
heat transfer efficiency by 20% to 60%, depending upon the
current running condition. Drain line 138 also returns a
metered flow of liquid refrigerant to compressor 12,
injecting the oil and liquid refrigerant into the interme-
diate pressure port IP. The metered amount of liquid
refrigerant evaporates and cools the compressor, maintain-
ing the discharge temperature of compressor 12 within a
desirable limit.
As is common with compressors which have an
intermediate pressure port IP, a n.c. controllable valve
144, called an economizer by-pass valve, is provided, which
by-passes economizer refrigerant vapors to the suction port
P when open. By-pass valve 144 is operably controlled by
electrical control 38 via means indicated generally by
arrow 147. Valve 144 may be internal to compressor 12, or
external, as illustrated, with valve 144 being connected
between tees 146 and 148 which respectively tap the auxil-
iary and main suction lines e0 and 32. A normal duty for
economizer by-pass valve 144 is to be open during a heat- ,
ing/defrost cycle, to preclude any limitation on compressor
pumping capability. During a heating/ defrost cycle the
normal flow to suction port S is closed. If compressor 12
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pumps only through the intermediate pressure port IP the
pumping capability may be limited, and it also pulls a
vacuum on the main suction line. An open line between the
auxiliary and main suction lines, via the open by-pass
valve 144 thus eliminates these problems. By-pass valve
144 may also be opened during a cooling cycle as part of a
temperature control algorithm, to unload compressor 12 for
temperature control in the conditioned space 106 as the
selected set point temperature is approached. The set
to point temperature of conditioned space 106 is selected on
a set point temperature selector :145, which provides an
input to electrical control 38.
In a desirable embodiment of the invention
economizer by-pass valve 144 provides another function,
engine load management, when prime mover 14 is an interrial
combustion engine. It is desirable that the temperature of
the engine coolant and the exhaust temperature be main-
tained within reasonable limits. 'With excessive load on
engine 14, especially during high ambient temperatures, it
would be desirable to unload the engine 14 to maintain the
desired limits. Thus, according to the teachings of the
invention, load on engine 14 is monitored, and when it
exceeds a predetermined value, by-pass valve 144 is opened
by electrical control 38, and valve 144 remains open until
the monitored load falls below a predetermined smaller
value. Load on engine 14, for example, may be monitored by
monitoring the compressor discharge pressure. A discharge
pressure sensor 150 provides an indication of the compres-
sor discharge pressure to electrical. control 38. When the
discharge pressure reaches a predetermined value, for
example a value of 360 psig (2482 kPa gauge) for R22
refrigerant, electrical control 38 energizes economizer by-
pass valve 144 to open it and unload engine 14. When the
discharge pressure drops to a predetermined value, such as
314 psig (2165 kPa gauge) for R22, electrical control 38
de-energizes by-pass valve 144, closing it. Other indica-
tions of engine load may be used, for example engine
coolant temperature, as sensed by a temperature sensor 152
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associated with an engine coolant circuit 154. An engine
coolant -temperature rise to 215 °F (101 °C) , for example,
may be used to initiate opening of valve 144, while a
temperature drop to 200 °F (93 °C) may initiate closing.
Engine exhaust temperature may also be used to indicate
engine load, as sensed by a temperature sensor 156 associ
ated with an exhaust conduit 158. An exhaust temperature
rise to 850 °F (454 °C), for example, may be used to
initiate opening of valve 144, while a temperature drop to
800 °F (426 °C) may initiate closing.
Engine coolant is used in, another embodiment of
the invention to cool the compressor oil. When compressor
12 is compressing at high pressure :ratios and the specific
heat ratio of the refrigerant is high, compressor 12 needs
some cooling to limit the discharge temperature so neoprene
or similar O-ring seals may be u~:ed with the discharge
service valve 22. Compressor cooling is achieved by taking
oil from the compressor 12, cooling the oil in an oil.
cooler 160, and injecting the oil back into compressor 12
at an intermediate point, which operation also lubricates
the shaft seal. The engine coolant is preferably a solu-
tion of ethylene glycol and water. It would be desirable
to cool both the engine and compressor oil with a single
thermostat, even though the engine and compressor have
different cooling needs. Neither the compressor 12 nor the
engine 14 should be too hot or too cold, with the compres-
sor 12 generally heating up more quickly than engine 14
during most operating conditions.
More specifically, a compressor oil cooler 160
having an inlet 161 and an outlet 163 is provided which has
a heat exchanger coil 162 connected to compressor oil sump
164 via conduits 166 and 168. A water jacket 170 surrounds
heat exchanger coil 162, with water jacket 170 being
connected to the engine coolant circuit 154. Engine
coolant circuit 154 includes a thermostat 172, a radiator
174, and an expansion tank 176, as well as the hereinbefore
mentioned coolant pump 92. Engine coolant is indicated at
177 in expansion tank 176. As illustrated, water jacket
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170 may be connected to receive coolant from thermostat 172
via a conduit 178, and to return coolant to pump 92 via a
conduit 180.
Figures 3 and 4 illustrate desirable embodiments
of the invention related to connecting oil cooler 160 into
the engine coolant circuit 154. Figure 3 relates to the
use of a thermostat 182 of the by-pass type. By-pass
thermostat 182 has first and second inlets 184 and 186 and
an outlet 188. By-pass thermostat initially blocks inlet
l0 186, causing all of the coolant to by-pass radiator 174
until the temperature of the coolant rises to a predeter-
mined value, at which point inlet 186 starts to open and
inlet 184 starts to close. At a predetermined higher
temperature thermostat inlet 184 will be substantially
closed and inlet 186 will be substantially completely open,
and all of the coolant will circulate through radiator 174.
In order to insure that there is always a constant flow of
coolant through the oil cooler 160, independent of the
position of thermostat 182 at any instant, water jacket 170
is connected to the outlet 188 of thermostat 186, down-
stream from the thermostat 182 and radiator 174.
Figure 4 illustrates an arrangement which utilizes
a thermostat 190 of the choke type, having a single inlet
192 and a single outlet 194. Choke type thermostat 190 is
substantially totally closed below a predetermined tempera-
ture, and when the predetermined temperature is reached, it
starts to open, reaching a fully open position at a prede-
termined higher temperature. Instead of connecting oil
cooler 160 downstream from radiator 174 and thermostat 182,
as in the Figure 3 embodiment, in the Figure 4 embodiment
oil cooler 160 is connected on the upstream side of thermo
stat 190, ie., at a tee 196 which taps the liquid coolant
circuit 154 prior to inlet 192 of thermostat 190. Thus,
oil cooler 160 receives coolant flow regardless of the
internal flow position of thermostat 190.
In order to construct and operate refrigeration
system 10 with the features hereinbefore described, with
economical sizing of the various heat exchangers and prime
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mover 14 relative to the compressor 12, and at the same
time keep compressor discharge pressure and temperature,
and engine load under control, some type of capacity
control is desirable, in addition to the hereinbefore
described optional engine load management use of the
economizer by-pass valve 144. T'he most simple way to
accomplish this is to introduce a ~>ressure drop on the low
pressure side of refrigeration system, ie., on the suction
side, such as with either a suction line throttling valve
l0 or a maximum operating pressure (MOP) evaporator expansion
valve. However, to keep compressor' discharge pressure and
temperature and engine load under control with a suction
line throttling valve or with a MOP evaporator expansion
valve, one of the modes, cooling or heating/defrost, has to
suffer with too large a restriction, as the desirable
pressure drops are different for the two modes.
In a desirable embodiment of the invention a
compromise in suction pressure control does not have to be
made, without adding an additional valve, by providing MOP
expansion valves for both the evaporator expansion valve 56
and the economizer expansion valve '78, each with a maximum
operating pressure setting which is optimum for the associ-
ated operating made. The evaporator' MOP expansion valve 56
thus has a relative low setting, compared with the setting
of economizer MOP expansion valve, with the evaporator MOP
expansion valve 56 controlling the maximum compressor
operating pressure during a cooling cycle, and with the
economizer MOP expansion valve 78 controlling the maximum
compressor operating pressure during a heating/defrosting
cycle. With R22 refrigerant, for example, the main MOP
expansion valve 56 would normally be set to provide a
maximum pressure somewhere in a range of 10 psia to 50 Asia
(68.96 kPa absolute to 344.7 kPa absolute), while the
economizer MOP expansion valve 78 would normally be set to
provide a maximum pressure somewhere in a range of 60 psia
to 100 psia (413.7 kPa absolute to E~89.5 kPa absolute.
Figure 5 illustrates a control algorithm 198
having operating modes which are implemented according to
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the teachings of the invention, including a plurality of
selectable null operating modes which smoothly maintain the
temperature of conditioned space 106 in a null temperature
range close to the selected set point temperature without
shutting down the prime mover 14 or compressor 12. This
arrangement assures constant air flow by evaporator air
mover means 104 a't all times, maintaining a substantially
uniform temperature throughout conditioned space 106.
Thus, the temperature of conditioned space 106 may be
controlled very close to the selected set point temperature
without danger of top freezing of a perishable cargo stored
therein.
The left hand side of control algorithm 198 of
Figure 5 illustrates the control error change points
between operating modes with a falling temperature in
conditioned space 106, while the i~ight hand side illus-
trates the control error change points for a rising temper-
ature in conditioned space 106. Electrical control 38
computes the control error as a function of the difference
between the temperature of the conditioned space 106, as
sensed by either, or both, of the temperature sensors 128
and 130, and the selected set point temperature SP.
Figure 5 also illustrates t:he open/closed patterns
of controllable valves 28, 62, 88 and 144 which implement
the different operating modes of the: control algorithm. A
"C" indicates the associated valve is closed, an "O"
indicates the valve is open, and an "X" for by-pass valve
144 indicates that valve 144 may be: opened or closed for
additional fine tuning temperature control by loading and
unloading compressor 12. Internal unloading of compressor
12, ie., a reduction in displacement, such as with a slide
valve, slot valve, or a lift valve, may also be used to
obtain fine temperature control, as is well known in the
art.
It will be assumed that. the temperature of
conditioned space 106 is in the stage of initial pull-down,
and thus refrigeration system 10 will. be in full or maximum
cool. When prime mover 14 is an internal combustion
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engine, the engine speed is usually controlled by electri-
cal control 38 between two speeds, called high speed and
low speed, with temperature pull-down being initiated with
a high speed cool mode 200, to obtain maximum cooling.
Pilot solenoid valve 28 will be closed, causing three-way
valve 30 to select the cooling circuit 40, liquid line
valve 62 will be open, to enable evaporator coil 102 to
function in a cooling mode, engine coolant valve 88 will be
closed, preventing heat from being' applied to economizer
heat exchanger 60, and economizer by-pass valve 144 will be
closed. Thus, liquid, high pressure refrigerant will be
subcooled in heat exchanger coil 68 by the expanding,
flooded evaporating state of the refrigerant in the second
refrigerant flow path defined by shell 70. Refrigerant
returns to compressor 12 via both the suction port S and
the intermediate pressure port IP.
When the control error drops to a point indicated
at 202, engine 14 is switched to the lower of its two
standard operating speeds, without change in the controlla-
ble valve open/closed pattern, enterring a low speed cool
operating mode 204.
At a still smaller control error, indicated at
point 206, a low speed partial or reduced cooling mode 208
is initiated by opening engine coolant valve 88. Thus, the
subcooling of the high pressure liquid refrigerant in heat
exchanger coil 68 is reduced, reducing the cooling rate of
conditioned space 106 so the set point temperature SP is
approached at a slower, more contro7Lled rate.
When the set point temperature SP is reached, a
null temperature range adjacent to the set point tempera
ture SP is entered, which, in a preferred embodiment of the
invention is divided into a plurality of different null
operating modes, such as first, second and third operating
modes 210, 212, and 214, with each null operating mode
being respectively implemented by different open/closed
patterns 211, 213 and 215 of controllable valve positions.
The first null mode 210 is initiated at set point SP, the
second null operating mode 212 is initiated at a slightly
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larger control error indicated at point 216, and the third
null operating mode 214 is initiated at a still larger
control error indicated at point 2;18. The prime mover 14
and compressor 12 remain operational during all three null
operating modes, with the engine 14 remaining at the low.
speed setting.
In the first null mode 210, which is closest to
set point SP, both heating and cooling takes place in
evaporator means 100, with the emphasis being on cooling to
prevent a quick return to the low speed partial cool mode
208. The emphasis on cooling also e=nables same dehumidify-
ing to take place. The first nul:L operating mode 210 is
implemented by opening pilot solenoid valve 28 to switch
the flow of hot compressor discharge gas to the heating
circuit 42, while maintaining liquid line valve 62 in an
open position to allow cooling to take place in evaporator
coil 102. In other words, the :Flow path includes the
second hot gas line 26, the au:!ciliary condenser 108,
receiver 50, both flow paths 64 and 66 through economizer
heat exchanger 60, subcooling the liquid refrigerant
flowing through heat exchanger coil 68, expansion valve 56,
and evaporator coil 102, with refrigerant being returned to
both the suction port S and the intermediate pressure port
IP.
In the second null mode 212, which is midway
between the control errors which will terminate the null
operating modes, no cooling or heating takes place in
evaporator means 100, while engine coolant 177 is circu-
lated through the water jacket 86 to keep the refrigerant
in shell 70 fully evaporated for return to compressor 12,
while simultaneously providing a desirable cooling of the
engine coolant. The second null operating mode 212 is
implemented by closing pilot solenoid valve 28, to switch
the hot compressor discharge gas back to the first hot gas
line 24, which prevents auxiliary condenser 108 from adding
heat to conditioned space 106, by closing liquid line valve
. 62, which prevents evaporator coil 102 from removing heat
from conditioned space 106, and by opening engine coolant
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valve 88, to enable engine coolant to give up heat to the
refrigerant in shell 70. By-pass valve 144 may also be
opened to prevent the suction side of refrigeration system
from being pulled down into a vacuum.
5 Thus, in the second null operating mode 212 the
refrigerant flow circuit includes hot gas lines 18 and 24,
main condenser 46, receiver 50, the second flow path 66
through economizer heat exchanger means 60, and the auxil-
iary and main suction lines 80 and 32.
10 The third null operating mode 214 again provides
both heating and cooling in evaporator means 100, similar
to the first null operating mode 210, with more heat being
added to the refrigerant than in the first null operating
mode 204, to attempt to maintain the. temperature of condi-
tinned space 106 in the null temperature zone, by allowing
engine coolant valve 88 to remain open as the operating
mode changes .form null mode 212 to null mode 214. Thus,
the third null operating mode 214 is implemented by opening
pilot solenoid valve 28, to select the heating circuit 42,
by opening liquid line solenoid valve 62, and by allowing
engine coolant valve 88 to remain open. The refrigerant
flow path is the same as described relative to the first
null operating mode 204, with less subcooling of the liquid
refrigerant in heat exchanger coil 68. Since some cooling
takes place in the evaporator means 100, some dehumidifying
also takes place.
Thus, at any given instant when the control error
is close to the set point temperature, a null related
operating mode is selected which will attempt to match the
heat loss or gain of conditioned space 106 with the heat
being added to, or removed from, the conditioned space by
the evaporator coil 102 and the auxiliary condenser 108.
If the third null operating mode 214 does not keep
the control error from increasing, indicating still more
heat is required than is being provided in the third null
operating mode 214, a control error value indicated at 220
initiates a low speed partial heating mode 22z which allows
pilot solenoid valve 28 to remain open while liquid line
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valve 62 and engine coolant valve 88 are closed. Economiz-
er by-pass valve 144 may also be opened to prevent limiting
compressor pumping capacity and prevent a vacuum from being
pulled in the main suction line 32.. The refrigerant flow
path includes hot gas lines l8 and 26, auxiliary condenser
108, auxiliary liquid line 110, receiver 50, the second
refrigerant path 66 through economizer heat exchanger 60,
and both the auxiliary and main suction lines 80 and 32.
If the control error continues to~ increase,
reaching a value indicated at point: 224, a higher heating
rate low speed heat mode 225 is entered which adds addi
tional heat by opening engine coolant valve 88. Pilot
solenoid valve 28 and by-pass valve 144 remain open and
liquid line valve 62 remains closed.. The refrigerant flow
path is the same as the partial heat operating mode 222.
If the control error continues to increase,
reaching a value indicated at point 228, maximum heating is
achieved by switching engine 14 to the higher of the two
operating speeds, ie., to a high speed heat operating mode
230. The valve open/closed pattern remains the same as in
the low speed heat operating mode 226.
With a rising temperature in conditioned space
106, the operating modes just described are entered i.n
reverse order, at slightly different control errors, ie.,
higher up the control algorithm, to provide a hysteresis
which prevents quickly switching back to the immediately
prior operating mode.
Figure 6 illustrates two desirable modifications
of the refrigeration system 10 shown in Figure 1 which may
3o be used. Like reference numbers in Figures 6 and 1 indi
cate like components, with similar but modified components
being given a prime mark in Figure 6. A first modification
relates to vent line 133. Instead of connecting the second
end of vent line 133 to the refrigerant distributor 98, it
may be connected to a tee 197 in the main suction line 32,
downstream from evaporator coil 102, between evaporator
coil 102 and thermal bulb 107. This arrangement has an
advantage over the Figure 1 embodiment in that it avoids
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the pressure drop associated with the distribution tubes in
distributor 98.
A second modification relates to oil drain line
138. During transient testing of refrigeration system l0
as set forth in Figure 1, system 10 was operated in a low
speed cool mode with a 70 °F (21.1 °C) box, and with an
ambient temperature of 120 °F (48.9 °C) . Compressor 12 was
then shut down. While compressor 12 was off the ambient
was changed over a period of several hours to -25 °F (-31.67
°C) while the box was maintained at a temperature of 35 °F
(1.67 °C). During such an operation, the refrigerant
migrates to the cool ambient, and thus the condenser coil
46 usually cools the fastest of any component. This did
not happen, however, as the oil return drain line 138,
being connected to a point on auxiliary suction line, below
outlet 140, allowed the economizer liquid to go to compres-
sor 12. Thus, compressor 12 cooled faster than condenser
coil 46, and most of the refrigerant liquid ended up in
compressor 12. This severe a change in conditions would
not be likely to happen during actual operating conditions.
However, this undesirable result can be prevented, even
during such a severe test by an oil drain arrangement shown
in Figure 6. Drain line 138' is directed to run in an
upward direction, above the level of drain point 140, while
in heat exchange relation with the liquid line 77, which
causes the oil return line to function as an oil-refriger-
ant liquid lift or percolator. The tapping point tee 142'
is located on auxiliary suction line 80 at an elevation
above drain point 140. The Figure 6 embodiment of drain
line 138' will keep the oil concentration down in the
economizer heat exchanger 60, and when compressor 12 is
shut down, drain line 138' will not drain the liquid
refrigerant into compressor 12. The high pressure, con-
densing temperature, liquid line 77 is subcooled by the
partially boiling liquid refrigerant-oil solution. The
vertical oil lift portion of drain line 138' may be provid-
ed by one or more ~ 2 5 inch ( 6 . 3 5 mrn) OD tubes , with the
horizontal portion of the oil return line, or lines, being
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.375 inch (9.5 mm) OD tubing. The Figure 6 embodiment of
drain line 138' also has the temperature control advantage
of the Figure 1 embodiment, directly limiting the economiz-
er suction temperature and indirectly limiting the dis-
charge temperature.
