Tri-level multi-cylinder reciprocating compressor heat pump system

A multi-cylinder reciprocating compressor is automatically controlled in terms of two speed operation and selective utilization of the cylinders under single stage action to meet heating and cooling loads by way of a two step indoor thermostat and an outdoor thermostat. The compressor may supply energy to storage during heating and cooling or receive energy therefrom with the storage coil selectively loop connected to the outside or indoor coils. Subcooling return is directed to a specific cylinder and overrides refrigerant return vapor to that cylinder from other coils functioning as evaporators. Solenoid operated valves effect unloading of the compressor during start up and automatically effect removal or inclusion of selected cylinders to the single stage compressor operation. Automatic load responsive control of compressor drive motor speed is effected.

FIELD OF THE INVENTION 
This invention relates to air source heat pumps, and more particularly, to 
improved high efficiency heat pump systems employing a multi-cylinder 
reciprocating compressor. 
BACKGROUND OF THE INVENTION 
Reciprocating compressors are universally employed in heat pump systems for 
residential building structures and the like and operate in conjunction 
with outdoor and indoor coils which coils trade functions; the outdoor 
coil constituting an air source evaporator while under heating mode, for 
instance. While under cooling mode, the indoor coil becomes the system 
evaporator and the outdoor coil becomes the air source condenser. 
Depending upon the geographical location of the residence employing the 
heat pump, the loads during summer and winter operation vary. For 
instance, when in use in the northeastern states, the heat pump system is 
subjected to high heating loads in comparison to cooling loads, while in 
the southern states such as Florida, the heat pump system experiences 
heavy cooling loads during summer operation and light heating loads during 
the winter months. 
Further, to effect low cost construction, normally the compressor units, 
which may be of the hermetic design, employ single phase electric motors 
for driving the compressor. Where such compressors are under load during 
starting, the current loads on the motor are significantly large such that 
in most cases, the motor must be oversized for starting since the load is 
higher than normal operation high heating or cooling load conditions, 
after start up. Further, reciprocating compressors conventionally of the 
multi-cylinder type have the suction gas simply supplied to all cylinders 
in parallel under single stage compressor mode with little though to 
system efficiency both in terms of electrical loads imposed by starting 
the electric motor under load and loads imposed by refrigeration circuit 
operation conditions. 
Attempts have been made to improve system efficiency by operating the 
reciprocating multi-cylinder compressor in double stage operation, 
depending upon system conditions, this being the subject matter of the 
referred copending application. Further, it has been determined that 
system efficiency may be improved by incorporating a subcooler between the 
coils, which functions to subcool the liquid refrigerant downstream of 
that coil constituting the condenser prior to feeding the liquid 
refrigerant to the coil acting as the evaporator of the system for 
expansion within that evaporator coil. In such subcoolers, which also 
forms a part of the subject matter of the referred to copending 
application, a portion of the high pressure liquid refrigerant is bled 
from the system and vaporized in the presence of the total liquid 
refrigerant in a suitable subcooler heat exchanger to further reduce the 
temperature of that portion of the refrigerant delivered to the coil 
functioning as the evaporator under the particular mode, whether it be 
heating or cooling. The vapor generated in the subcooler, being at a 
pressure well above that of the vapor pressure from the coil or coils 
acting as the system evaporator and directed to the suction side of the 
reciprocating compressor, is permitted to return to the reciprocating 
compressor crank case for the multiple cylinders to maintain load reversal 
on the wrist pins of the reciprocating compressor piston and connecting 
rod assemblies of the multicylinder reciprocating compressor of the 
referred to copending application. 
It is, therefore, a primary object of the present invention to provide an 
improved, simplified automatic tri-level multiple cylinder reciprocating 
compressor heat pump system wherein compressor operation is matched to 
system heating and cooling loads regardless of the unequal load condition 
with three levels of compressor operation being readily achieved and 
automatically effected. 
It is a further object of the present invention to provide a simplified, 
automatic tri-level multiple cylinder reciprocating compressor heat pump 
system, in which, dependent upon indoor and outdoor conditions, the 
compressor three level operation may be effected by cutting out or adding 
compressor cylinders to the compressor compression process and/or shifting 
the compressor drive motor between low and high speed operation. 
It is a further object of the present invention to provide an improved 
tri-level, multiple cylinder reciprocating compressor heat pump system 
which includes a subcooler for subcooling liquid refrigerant being fed to 
the heat pump system coil acting as the system evaporator, and wherein the 
vapor returned from the subcooler is passed over the motor windings in a 
hermetic reciprocating compressor package open to the compressor crank 
case and delivered to the low side of a given cylinder. 
It is a further object of the present invention to provide an improved air 
source heat pump three-step tri-level multiple cylinder reciprocating 
compressor heat pump system which permits thermal energy to be picked up 
by an outside air coil and supplied selectively to either one or all of an 
inside air coil, inside hydronic coil, and storage coil, depending upon 
system needs. 
It is a further object of the present invention to provide a simplified 
tri-level, multiple cylinder, reciprocating compressor heat pump system in 
which, under mild ambient conditions, thermal energy may be removed from 
the room being conditioned and stored by way of a storage coil during the 
day and may be supplied to the same room as usable heat from the storage 
coil during the night. 
It is a further object of the present invention to provide an improved, 
simplified tri-level multiple cylinder reciprocating compressor heat pump 
system, wherein heat may be removed from the room being conditioned during 
high ambient temperature conditions during the day and stored by way of 
the storage coil within the system for subsequent discharge by way of the 
outside air coil at night at lower ambient temperature for improved system 
thermal efficiency. 
SUMMARY OF THE INVENTION 
The present invention is principally directed to an improved air source 
heat pump system of the type having a first heat exchanger which forms an 
indoor coil, a second heat exchanger forming an outdoor coil, and a third 
intermediate pressure evaporator coil which may be a solar energy fed 
storage coil as an example. The invention involves a multi-cylinder 
reciprocating compressor and conduit means carrying refrigerant connects 
the coils and the compressor in a closed fluid circuit. Preferably, the 
conduit means includes a reversing valve for connecting the indoor and 
outdoor coils in a closed series loop with the reversing valve functioning 
to cause the indoor and outdoor coils to operate alternately as low 
pressure evaporator or high pressure system condenser. Further, means are 
provided for selectively supplying refrigerant to the intermediate 
pressure evaporator coil for evaporation therein and the improvement 
resides in the multi-cylinder reciprocating compressor constituting a 
hermetic compressor unit including a hermetic casing, at least three 
cylinders within the casing, pistons within the cylinders, a motor within 
the casing and operatively coupled to the pistons for driving the pistons 
of the reciprocating compressor. The conduit means includes first conduit 
means for supplying low pressure suction return vapor from the system 
evaporator to the hermetic casing for flow over the motor to cool the 
motor and thence to a first cylinder for recompression. Second conduit 
means are further provided for supplying intermediate pressure refrigerant 
vapor from the intermediate pressure evaporator coil to the first cylinder 
and includes means for cutting off the first cylinder to the suction 
return refrigerant vapor from the system evaporator when the intermediate 
pressure refrigerant vapor is being supplied to the first cylinder. 
Conduit means are provided for selectively directing low pressure suction 
return refrigerant vapor from the system evaporator to at least said third 
cylinder such that the compressor may be operated at partial load 
conditions with refrigerant vapor compressed at low pressure by said first 
and second cylinders and under increased load conditions with the low 
pressure suction return refrigerant vapor compressed by the second and 
third cylinders and the intermediate pressure refrigerant vapor from the 
intermediate pressure evaporator coil compressed by the first cylinder 
such that the third cylinder functions in a capacity control mode. 
Preferably, the compressor drive motor comprises a two speed motor and the 
system comprises control means for operating the motor at low speed with 
the first and second cylinders connected to the system evaporator under 
low system load conditions at low speed with the first, second and third 
cylinders connected to the system evaporator under intermediate load 
conditions and with the motor operating at high speed and the second and 
third cylinders connected to the system evaporator and the first cylinder 
connected to an intermediate pressure evaporator coil to thereby provide 
three step compressor loading. 
Preferably, the refrigerant vapor from either the low pressure suction 
return from the evaporator or the intermediate pressure vapor returning 
from the intermediate pressure evaporator is directed over the motor 
windings to the crank case of the multiple cylinder reciprocating 
compressor to assure wrist pin load reversal for all of the compressor 
cylinders and to the first cylinder for recompression. The conduit means 
may include a suction line leading from the reversing valve to at least 
one of the cylinders, a discharge line leading from the discharge side of 
all of the compressor cylinders to the reversing valve for supplying 
compressed refrigerant vapor to the system condenser regardless of system 
mode, and with the system further including a shorting line connecting the 
discharge line to the suction line with the shorting line carrying a 
solenoid operated control valve for selectively opening the shorting line 
to permit the compressor drive motor to be energized with all compressor 
cylinders fully unloaded. 
Preferably, the intermediate pressure evaporator coil comprises a subcooler 
and the means for selectively cutting off the first cylinder comprises a 
check valve within the means leading from the reversing valve to the 
hermetic casing such that the evaporator suction return refrigerant vapor 
flows to the hermetic casing for cooling of the motor and recompression by 
the first cylinder only in the absence of subcooler operation. 
A storage coil may be provided in heat transfer relation with a thermal 
energy storage media and the conduit means may include additional conduit 
means for selectively connecting the storage coil in parallel with the 
outside air coil for supplying heat to the refrigerant simultaneously with 
that supplied by the outside air coil under heat pump system heating mode 
or for removing heat from the system either simultaneously with the 
outside air coil or exclusive of the outside air coil when the heat pump 
system is operating under cooling mode. Further, the invention is directed 
to a hermetic multi-cylinder reciprocating compressor unit to achieve this 
purpose and, preferably, comprises four cylinders within two cylinder 
heads. The first and second cylinders are located within the first 
cylinder head and the third and fourth cylinders are located within the 
second cylinder head. Manifold means for the cylinder heads define 
separate inlets for the first and second cylinders and a common outlet for 
those cylinders and a common inlet and a common outlet for said third and 
fourth cylinders within the second cylinder head. A first discharge line 
leads from the common outlet for the first and second cylinders and a 
second discharge line leads from the common outlet of the third and fourth 
cylinders, and a common discharge manifold connects to the first and 
second discharge lines. A check valve within one of the lines, such as the 
discharge line from the outlet of the first and second cylinders, permits 
compressor discharge flow from cylinders 1 and 2 through the common 
discharge manifold but prevents reverse flow. A heat rejector storage line 
connects to the common outlet for cylinders 1 and 2 such that the 
discharge from cylinders 1 and 2 may be supplied to a low pressure 
condenser while the compressor discharge from cylinders 3 and 4 may be 
directed to the coil functioning as the system high pressure condenser. 
Further, under such arrangement, the first cylinder is of less 
displacement than the second, third and fourth cylinders, the second 
cylinder is of a displacement larger than the first but smaller than the 
third and fourth cylinders combined, such that the third and fourth 
cylinders provide capacity control and the first cylinder functions to 
compress the intermediate pressure evaporator return vapor to a common 
discharge pressure with that of said second cylinder. The displacement of 
the cylinders may be achieved by providing cylinders and pistons of given 
diameter such that the second cylinder has a given diameter, the first 
cylinder has a diameter smaller than that of the second cylinder, and the 
third and fourth cylinders, each have a diameter equal to that of the 
second cylinder with the pistons having equal strokes.

DESCRIPTION OF THE PREFERRED EMBODIMENT 
Referring to FIGS. 1 and 2, there is illustrated a compressor package 10 
which is shown schematically as an outer housing 11 and purposesly shown 
in dotted lines as indicative only of the fact that the hermetic 
compressor unit indicated generally at 12 and interior of the housing 
constitute a portion of the compressor package 10 along with a plurality 
of lines, couplings and various control elements. The hermetic compressor 
unit 12 constitutes a compressor casing 14 having sealed end caps or end 
bells 16 and 18 at respective ends which are sealably, fixedly coupled to 
the casing 14 by way of cooperating flanges. The hermetic unit 12 is 
formed principally of a hermetic, single phase electric motor indicated 
generally at 20 on the left side of the unit and the reciprocating 
compressor on the right hand side, and indicated generally at 22. The 
compressor 22 constitutes in the illustrated embodiment four cylinders 1, 
2, 3 and 4, although the invention has application to a multi-cylinder 
reciprocating compressor of more or fewer than four cylinders; three 
cylinders for instance. 
Cylinders 1, 2 and 4 are shown as being of the same diameter, and larger 
than cylinder 3. Cylinder 3 may have 1/12 to 1/4 the total compressor 
displacement. As will be appreciated hereinafter, cylinders 1 and 2 
function as capacity control cylinders for the compressor unit for the air 
source heat pump system with which it is preferably employed, cylinder 3 
functioning as the cylinder in which subcooler return, intermediate 
pressure refrigerant vapor is compressed by the compressor for either 
common discharge or selectively directed to a rejection/storage unit. In 
this respect, the pistons, which are sized to the cylinders, reciprocate 
with the same stroke and cylinder 3 thus has a different displacement from 
those of cylinders 1, 2 and 4 which have equal displacement. The cylinders 
and pistons could be identically sized. Further, capacity could be varied 
by providing shorter or longer strokes for selected pistons, as at 
P.sub.1, P.sub.2, P.sub.3 and P.sub.4, FIG. 1. 
By reference to FIG. 3, it may be appreciated that the compressor casing 14 
houses the stator 68 which is fixed at one end of the housing and which 
concentrically surrounds the rotor 70 of the single phase hermetic drive 
motor 20, the rotor being fixedly mounted to a shaft 71 and supported by 
way of bearings 69 within opposed end bells, the rotor 70 thus being 
mounted for rotation about its axis. Conventionally, the cylinders are 
connected by means of their connecting rods to the shaft 71 such that the 
motor directly drives the pistons. In that regard, the sectional view 
shows cylinders 3 and 4, the pistons P.sub.3 and P.sub.4 within respective 
cylinders and connected to the shaft by means of connecting rods R.sub.3, 
R.sub.4 and crank arms C.sub.3, C.sub.4, respectively. The partition 76 
defines with the hermetic casing 14 and the end bell 18 to the right, the 
cylinders and the cylinder heads 24, 26, a crank case 73 which is 
pressurized by the return vapor passing over the rotor 70 and stator 68, 
the crank case 73 being open to the portion of the hermetic compressor 
casing 14 housing the stator 68 and rotor 70 by way of passage 74 within 
the transverse partition or wall 76. Further, as may be appreciated by 
reference to FIG. 1, by way of tube or conduit 78, the vapor normally 
returned from the subcooler after it passes over the motor rotor 70 and 
stator 68 and cools the windings of the single phase two speed motor, 
enters the low pressure, low side or inlet 38 of cylinder head 26 for 
compression by the smaller displacement cylinder, and in this case, 
smaller diameter cylinder 3. At the same time, the crank case 73 is 
pressurized at a pressure which is normally in excess of the suction 
pressures being applied to the faces of the pistons within respective 
cylinders, with the exception of cylinder 3. 
In this case, the compressor is provided with cylinders 1 and 2 within 
cylinder head 24 and cylinders 3 and 4 within cylinder head 26. The 
cylinder head 24 includes manifold means such as 28 to divide the cylinder 
head 24 into a low pressure side, low side or inlet 30 and a high pressure 
side, high side or outlet 32. For compressor cylinders 3 and 4, a manifold 
34 divides the low pressure or low side of the compressor cylinder head 26 
from that of the high pressure or high side, and in addition, the cylinder 
head 26 incorporates additional manifold means as at 36, whereby the low 
pressure sides of both cylinders 3 and 4 are cut off from each other. In 
that regard, the low pressure side, low side or inlet to cylinder 3 is 
shown at 38, the low pressure, low side or inlet for cylinder 4 is shown 
at 40, and there is a common high pressure, high side or discharge 42 for 
both cylinders 3 and 4. The compressor, the cylinders, the manifolding and 
the heads are shown schematically as well as the outlet or discharge 
connections for the compressor cylinder heads. In that regard, a discharge 
port 44 for the common discharge 32 for cylinders 1 and 2 is connected by 
way of conduit or passage 46 shown in dotted line to a compressor 
discharge manifold 48 to which is also connected the common discharge or 
high side 42 of cylinders 3 and 4 through a discharge or outlet port 50. 
Port 50 through conduit indicated by dotted lines 52 opens to the common 
discharge manifold 48. Incorporated within conduit or passage 52, is a 
check valve 54 for permitting flow from the discharge or high side of 
cylinders 3 and 4, that is, from outlet 42 for cylinders 3 and 4 to the 
common discharge manifold 48 but prevents reverse flow. 
With respect to the above portion of the description of the compressor 14, 
its make-up and manifolding is somewhat similar to that of the copending 
application noted above. However, there are a substantial number of 
differences between the compressor of that copending application and the 
instant invention. 
First of all, with respect to the hermetic unit 12, the normal return to 
compressor suction from the various coils functioning as evaporators 
within the system is via a first conduit 60 at suction return C of the 
compressor housing 10. Line 60 connects to the inlet or low side of 
compressor head 24, through line 60a common to cylinders 1 and 2, line 60a 
including a solenoid operated shut-off valve V2c. This permits the 
refrigerant vapor being returned selectively by energization of the 
solenoid operated valve V2c to two of the cylinders 1 and 2 of the 
compressor 14. Further, line 60 connects to end bell 16 of the hermetic 
compressor unit 12 via conduit 62 from connection point 64 downstream of 
check valve 66. Line 60b defines another flow path for suction return C to 
cylinder 4. 
One aspect of the present invention is to employ the refrigerant vapor 
returning to the suction side of the compressor as the means for cooling 
the hermetic motor by flowing over the stator 60 and rotor 70 of that unit 
of the motor 20. 
Further, vapor within the chamber housing the rotor and stator enters the 
crank case indicated generally at 73 via passage 74 within wall 76 which 
divides a motor section from the compressor section of the hermetic 
compressor unit 12. 
Further, as shown in dotted line at 78, cylinder 3 is fed with the same 
refrigerant vapor as is directed to the crank case 73 vapor passing 
through inlet port 80 which opens up into the inlet or low side 38 to 
cylinder 3 of cylinder head 26 via tube 78 from motor chamber 75 
downstream of motor 20. The refrigerant vapor is compressed by cylinder 3 
for discharge to the common discharge or high side 42 of that head and for 
normal delivery to the common discharge manifold 48 of the compressor. 
Further, by way of line 60b, refrigerant vapor can pass through check valve 
82, to the inlet or low side 40 of cylinder 4 for compression by that 
cylinder. The check valve 66, within line 60 and 82, within line 60b 
prevent refrigerant vapor flow towards cylinders 1 and 2 should 
refrigerant vapor be returned to the compressor for feeding cylinders 3 
and 4 at pressure levels in excess of that available to the multiple 
cylinders of the compressor through suction line 60. 
In that respect, it is contemplated that in addition to normal suction line 
supply to cylinders 3 and 4, refrigerant vapor may be made available to 
selected cylinders of the compressor as at 3 and 4 from a storage source, 
through line 84 and/or from a system subcooler or other intermediate 
pressure evaporator, through line 86 opening to line 84 at junction 88. In 
that regard, the line 84 from storage includes a solenoid operated control 
or shut-off valve V7, within line 84 upstream of check valve 90, line 84 
being coupled to line 62 at point 64 and entering compressor housing at 
point D, line 84 permitting refrigerant flow simultaneously to cylinder 3 
through line 62 and also to cylinder 4 through line 92 and intersecting 
line 60b at point 94. Line 86 from the subcooler intersects lines 84 at 
point 88 upstream of the connection of line 84 to lines 62 and 92 such 
that refrigerant vapor either from storage or from subcooler may flow to 
cylinders 3 and 4 as desired and permitted by the system, while at the 
same time the check valve 90 permits the storage evaporator to generate a 
higher pressure level refrigerant vapor than the subcooler evaporator, 
permitting both flows in a direction toward cylinders 3 and 4 via lines 62 
and 92, but prevents flow from the subcooler to the storage due to the 
presence of that check valve 90. At the same time, since normally the 
refrigerant vapor from storage or from the subcooler is at a higher 
pressure than that returning to the compressor unit via suction line 60, 
the check valves 66 and 82 prevent the storage or subcooler refrigerant 
vapor from passing towards suction line 60 and cylinders 1 and 2 due to 
the presence of those check valves 66 and 82 within lines 60 and 60b, 
respectively. 
Since line 62 directs refrigerant vapor to the crank case via passage 74 
within wall 76, it is obvious that at all times the pressure beneath the 
pistons in the crank case is no lower than the suction pressure at the top 
of the pistons. 
As stated previously, by operation of the solenoid operated valve V2c 
within line 60a, both cylinders 1 and 2 of cylinder head 24 may be 
selectively cut in and cut out of the compression process. In that respect 
and throughout the application, the solenoid operated valves are intended 
to be shut-off valves and to be normally closed in the de-energized 
condition. Thus, with solenoid valve V2c de-energized, cylinders 1 and 2 
are cut out and suction return via suction line 60 is directed only 
through lines 60 and 62 to cylinder 3 and through line 60b to cylinder 4. 
Further, the discharge manifold 48 permits the compressor to discharge 
refrigerant vapor fully compressed in a single stage compression process 
via discharge line 96, the discharge line 96 including a check valve 98 
and discharging refrigerant vapor from the machine at discharge point A, 
FIG. 1. 
In order to provide the capability of the rejection of heat in cases where 
there is more heat being generated than may be used, for instance, in 
either supplying one or more indoor coils functioning as condensers or 
heat supply coils for the space or room being conditioned and for 
supplying heat to a storage coil within the system, the common discharge 
from cylinders 3 and 4 may be directed to a heat rejection coil external 
of the compressor housing environment through line 100 which exits from 
the compressor housing 10 at point B and is connected directly to the 
common compressor discharge or outlet side 42 of the compressor cylinder 
head 26. 
Further, the discharge line 96 for the compressor is provided with a bypass 
or shorting line as at 102 between point 104 of line 96 and point 106 of 
suction line 60, line 102 including a solenoid operated control valve V1c 
which selectively connects, when energized, the suction and discharge 
sides of the compressor cylinders together so as to completely unload the 
machine during start up. The solenoid control valve V1c being a normally 
closed valve causes, when energized, the connection of the high side of 
the compressor directly to the low side, preventing pressure build up and 
allowing totally unloaded start conditions. This permits the hermetic 
motor 20 to be of relatively small size and improving the efficiency of 
the system. 
An explanation of the hermetic compressor unit and the control elements and 
fluid connections of FIG. 1 may be best appreciated by reference to FIGS. 
2 through 2e which illustrate a preferred embodiment of the tri-level 
multiple cylinder reciprocating compressor heat pump system of the present 
invention under various stages of operation. Like elements are given like 
numerical designations, and FIG. 2 is essentially a complete multiple coil 
heat pump system utilizing the hermetic compressor package 10 of FIG. 1. 
It should be apparent that while the system of FIG. 2 makes use of a 
storage coil and an inside hydronic coil in addition to conventional 
outside air and inside air coils, the system of the present invention may 
advantageously employ only an outside air coil and an inside air coil, 
preferably with the subcooler for three level multiple cylinder 
reciprocating compression operation for improved system efficiency. 
The electric motor 20 may be of the single phase two speed type wherein 
conventionally suitable controls such as control unit 107 acts to change 
the number of poles as from four poles to two poles, or eight poles to 
four poles, etc., to double the speed of rotation of rotor 70 by applied 
electrical control signals through leads 109. Current is delivered to the 
motor via the control unit 107 by way of leads 105 leading to an 
electrical source (not shown). Control unit 107 may be connected to all 
solenoid operated control valves of the system, the reversing or four way 
valve 114, FIG. 2, may be programmed for system operation as hereinafter 
described and receive inputs from outdoor thermostat OT adjacent outside 
air coil 110 sensing the temperature of the air passing over that coil, a 
two step or two position indoor or room thermostat IT, within space 146 
being conditioned, and storage thermostat ST sensing the temperature of 
storage media M, FIG. 2. 
In FIG. 2, in addition to the compressor unit 12, the main components of 
the heat pump system in a preferred embodiment include conventionally an 
outside air coil 110, an inside air coil 112, a four way valve 114, a 
receiver 116, a subcooler 118 and an accumulator 120, these being 
essentially minimal components for a heat pump system incorporating the 
present invention. Additionally, however, there is provided an inside 
hydronic coil 122 which is in parallel with the inside air coil 112 and 
which conventionally supplies heat to a circulating liquid such as water 
to effect, for instance, soft heating of limited areas of a room or space 
148 to be conditioned such as those adjacent to an outside wall of an 
enclosure 144, while the inside air coil 122 permits heat delivery and 
under heating cycle or heating mode operation to space 146 being 
conditioned, or extracts heat therefrom when the inside air coil 112 acts 
as an evaporator during reverse flow cooling mode operation. Further, the 
system employs a storage coil 124 whose function is to selectively store 
thermal energy or remove thermal energy from a storage media M within a 
storage container 125, during either cooling or heating mode of the heat 
pump system, depending upon outdoor ambient conditions and indoor 
temperature conditions of the space being conditioned, or when the system 
has no demands from enclosure 144. 
The solenoid operated four way valve 114 is of conventional construction 
and simply reverses suction line 60 and discharge line 96 with respect to 
lines 126 and 128, line 126 being connected to the outside air coil 110 
through check valve 129 and to the storage coil through line 130, which 
line also includes a check valve 132. The check valves 129 and 132 
provide, respectively, for refrigerant flow through outside air coil 110 
and storage coil 124 in the direction of their common juncture point 134 
but not reverse flow. Line 128 is connected commonly at point 136 to the 
inside air coil 112 and the inside hydronic coil 122, through paired lines 
138 and 140, respectively. A check valve 142 is provided within line 140 
to permit the inside hydronic coil 122 to function as a condenser but 
prevent its operation as an evaporator by reverse refrigerant flow through 
that coil. No such check valve is provided for the inside air coil 112 
which may function alternately as a condenser and evaporator coil 
depending upon the necessity to heat or cool the space 146 being 
conditioned. In that respect, the total space to be conditioned within 
enclosure 144 is divided into an interior space 146 subjected to heating 
or cooling by the heat transfer via the inside air coil 112 and a second 
space 148 which is subjected, only to heating by controlled flow of 
refrigerant through the inside hydronic coil 122 when the heat pump system 
is operating under heating mode. The space 148 may constitute a room 
specifically being heated or hot water heaters under retrofit application 
of the present invention to an existing hot water heating system. 
The refrigerant R within line 140 is directed to receiver 116 through line 
150 which intersects line 140 at point 152. Further, refrigerant within 
line 138 and the inside air coil 112 may also be returned to the receiver 
through line 154 which intersects line 138 at point 156 and which is 
connected to line 150 at point 158, permitting commonly, refrigerant flow 
to the receiver from both coils 112 and 122. Line 154 includes a check 
valve 160 which permits flow of refrigerant towards the receiver from the 
inside air coil 112 but not in the reverse direction from the receiver 
116. The refrigerant R, which is provided to the system, accumulates as a 
liquid within the receiver 116 and is directed from the receiver through 
liquid refrigerant supply line 162 to the accumulator 120 where that 
liquid refrigerant is subcooled to some extent, prior to reaching 
subcooler 118, by way of accumulator coil 164. Line 162 extends to the 
subcooler and bears the subcooler coil 166 such that the liquid 
refrigerant can be subcooled both at the accumulator and at the subcooler, 
prior to its being directed selectively to either the storage coil 124 or 
the outside air coil 110, when the heat pump system is operating under 
heating mode, or to indoor air coil 146 during cooling mode. 
In that regard, line 162 includes a further check valve 168 spaced from its 
connection to the outside air coil at point 170. Line 162 carries solenoid 
operated control valve V2 such that line 162 can be shut off as desired, 
preventing liquid refrigerant R from the receiver 116 to reach the outside 
air coil 110 except upon energization of the solenoid operated control 
valve V2. Similarly, line 172 connects to line 162 at point 174 and to the 
storage coil 124 at point 176, the line 172 bearing a check valve 178 and 
a solenoid operated control valve V4. Thus, refrigerant flow from the 
receiver through the subcooler can flow only to the storage coil upon 
energization of the solenoid operated control valve V4 and in a direction 
to permit the storage coil to act as an evaporator; reverse flow being 
prevented by the check valve 178. The outside air coil 110, the storage 
coil 124 and the inside air coil 122 are provided with expansion devices 
such as capillary tubes, thermal expansion valves or the like (not shown), 
to effect expansion and vaporization of the liquid refrigerant within 
these coils selectively and to thereby permit those coils to operate as 
evaporators under given system conditions. When the outside air coil 110 
and the storage coil 124 are functioning as evaporators, the vaporized 
refrigerant after picking up heat is returned to the compressor through 
check valves 129 and 132 respectively and four way valve 114. In this 
case, a solenoid operated control valve V6 within line 130 is energized as 
well as solenoid operated control valve V4 within line 172 and solenoid 
operated control valve V2 for outside air coil 110. A bypass line 178 
bypasses the check valve 129, line 178 being connected to line 126 at 
points 134 and 180. Line 178 bears solenoid operated control valve V5, 
permitting by energization of that solenoid operated control valve V5, 
compressed refrigerant vapor flow from the four way valve 114 through line 
126 to the outside air coil 110. In that situation, the outside air coil 
acts as a condenser. A return line 182 is provided with a solenoid 
operated control valve V12, thus with coil 110 as a high pressure 
condenser, condensed refrigerant flows through line 182 and check valve 
184 to the receiver, line 182 intersecting line 150 at point 152. This 
point is also a common connection for line 186 which bears check valve 188 
and a solenoid operated control valve V10 and is connected to the storage 
coil 124 at point 176 such that under certain conditions where the storage 
coil is acting as a condenser, condensed refrigerant can flow from the 
storage coil 124, after being received from the compressor, through check 
valve 188 and line 186 to the receiver via line 150. 
The subcooler is conventional. The refrigerant line 162 is tapped at 190 
via line 192 leading to the subcooler and bearing a solenoid operated 
control valve V3, such that upon energization of the control valve V3, 
liquid refrigerant enters the subcooler and expands to subcool the liquid 
refrigerant within coil 166 upstream of tap point 190. The vaporized 
refrigerant at an intermediate pressure (between compressor suction and 
discharge) is directed to connection point 64 and line 62, via line 86 
which constitutes the subcooler return, this refrigerant vapor being 
compressed by cylinder 3 since the vapor not only passes over the motor 
stator and rotor to cool the same, but also reaches the crank case to 
pressurize the crank case, entering the inlet or low side 38 of compressor 
head 26 prior to recompression by cylinder 3. 
Further, when the inside air coil 112 is functioning as a low pressure 
evaporator coil, it receives refrigerant from liquid refrigerant line 162, 
downstream of the subcooler 118, through a conduit 194 which connects to 
line 162 at point 198, the intersects line 138 at point 156. The line 194 
bears a check valve 200 which permits liquid refrigerant flow from the 
liquid refrigerant line 162, downstream of the subcooler, to the inside 
air coil 112 but prevents reverse flow; the flow through this line 194 
being further controlled by solenoid operated control valve V1 located 
between the check valve 200 and connection point 156. 
Further, while the solenoid operated control valves V2 and V6 permit the 
outside air coil 110 and the storage coil 124 to return refrigerant vapor 
from these coils selectively to the compressor, the present system permits 
heating requirements to be achieved by feeding refrigerant from a high 
evaporating pressure rather than a low evaporating pressure such that 
depending upon the evaporating pressure within the outside air coil 110 or 
the storage coil 124, the flow can be controlled in a suitable manner. For 
instance, between the storage coil 124 and check valve 132, at point 202 
within line 130, there is connected one end of line 84 which bears 
solenoid operated control valve V7, this line 84 permitting by 
energization of the solenoid operated control valve V7, refrigerant vapor 
flow to the inlet or low pressure side 38 of head 26 for compression of 
that vapor by cylinder 3, and by way of line 92 to the low pressure or low 
side 40 of the head 26 for compression by cylinder 4. 
Additionally, line 206 is connected at an end to line 84, at 204, and thus 
storage coil 124, between the storage coil 124 and the solenoid operated 
control valve V7, and at its opposite end, at point 208, to compressor 
discharge line 96 such that by energization of the solenoid operated 
control valve V8 within line 206, compressed refrigerant vapor discharged 
from the compressor may be directed to the storage coil 124 to operate the 
storage coil as a condenser and to store heat emanating from another part 
of the system such as the inside air coil 112 or outside air coil 110 
which would be acting as evaporator coils. 
As in FIG. 1, the hermetic compressor unit 12 is further provided with a 
line 100 leading from the common high side 42 for cylinders 3 and 4 to 
permit, in a selective manner under the control of solenoid operated 
control valve V11 and by way of a check valve 101, the high pressure 
compressed refrigerant vapor to be directed to a heat exchange coil 
functioning to reject heat or to store heat in addition to storage coil 
124. The storage coil 124 is immersed within the mass of heat storage 
liquid or the like media M which readily receives and gives up heat to 
that coil 124 depending upon system demands. 
Reference will now be made to typical system operating conditions 
illustrating the utility of the present invention as applied to a 
representative heat pump system. Reference to FIG. 2a shows the basic 
components of the system under heat pump heating mode with relatively mild 
outdoor ambient. 
Referring next to FIG. 2a, the heat pump system of the present invention is 
considered as having the outside air coil 110, the inside air coil 112, 
and the inside hydronic coil 122, as the only existing coils within the 
system with the system lacking storage coil 124 and its controls and 
attendant equipment. 
In fact, while the system is shown as including an inside hydronic coil 122 
for independently conditioning space 148, such coil could be eliminated, 
and the invention would have equal application to a two-coil heat pump 
system consisting of only outside air coil 110 and inside air coil 112. In 
such case, the receiver 116 could also be eliminated, with the inside air 
coil 112 feeding directly to subcooler 118 by connection of line 162 
directly to line 138 at connection point 156. Under the realization that 
the system could be simplified to that degree, a typical system under 
operation during three step heating starting with mild ambient conditions 
will now be discussed. 
As may be further appreciated, one aspect of the present invention resides 
in the start up of the two speed motor, of course at low speed and under 
conditions in which the compressor is totally unloaded. This is achieved 
by energization of solenoid operated control valve V1c as shown in 
dash-dot line fashion, which opens line 102 between the suction return 
line 60 and the compressor discharge line 96, such that the suction and 
discharge sides of all four cylinders 1, 2, 3 and 4 are connected together 
with resultant driving of the cylinders under no load, non-gas compression 
conditions. 
After the unload-start sequence is completed and solenoid operated control 
valve V1c de-energizes, and with the four way valve 114 in the heating 
mode as shown, discharge line 96 is connected to line 128 leading to the 
indoor air coil 112 and the indoor hydronic coil 122 and the line 126 from 
the outside air coil 110, is feeding and is connected to the suction 
return line 60 including accumulator 120, compressed refrigerant 
discharging from the compressor, via discharge or outlet manifold 48, is 
directed to the inside air coil 112 and the inside hydronic coil 122. Only 
solenoid operated control valves V2 and V9 are energized, while solenoid 
operated control valves V1c, V2c, V3, V4, V5, V6, V7, V8, V10, V11, V12 
are not energized. De-energization of the solenoid operated control valve 
V1c terminates the mechanical short between the suction and discharge 
sides of the compressor by shutting off the connection between discharge 
line 96 and suction line 60. Energization of solenoid operated control 
valve V2 permits liquid refrigerant flow from the receiver 116 by way of 
accumulator coil 164 and subcooler coil 166 to the outdoor air coil which 
functions as an evaporator coil under heating mode. 
The two speed motor is maintained energized by motor control 107 at low 
speed operation. The de-energization of solenoid operated control valve 
V2c takes cylinders 1 and 2 off the line by shutting off line 60a leading 
from the suction return line 60 to the low side or inlet 30 for cylinders 
1 and 2 of cylinder head 24. Condensed refrigerant flows from the inside 
air coil 112 and the inside hydronic coil 122 to the receiver 116 and from 
the receiver through liquid refrigerant supply line 162 to the outside air 
coil 110 acting as the system evaporator where it is expanded through the 
use of a suitable expander and returned, after absorbing heat, to the 
suction return line 60 through line 126 and the four way valve 114. 
Refrigerant vapor entering cylinder 3 through line 62, passes over the 
hermetic motor stator 68 and rotor 70 for cooling the same and passing to 
the low side or inlet 38 of cylinder head 26 via port 80 for compression 
along with a second portion of the return gas by way of line 60b and check 
valve 82 to the low side or inlet 40 for cylinder 4 of the same cylinder 
head 26. Thus, only two cylinders 3 and 4 are operating to compress 
refrigerant under low speed motor operation. The de-energization of valve 
V3 prevents the subcooler from being fed liquid. The inside hydronic coil 
and the inside air coil seek a common condenser pressure level, since they 
are both in parallel. Under some conditions of operation, either coil 
could back up with a degree of refrigerant liquid and the receiver 116 is 
necessary for charge balance. Liquid coming out of the receiver 116 passes 
through the suction accumulator 120 where a limited degree of subcooling 
takes place, then passes into the subcooler and through subcooler coil 166 
and directly to the outside air coil as shown but is not further subcooled 
at this point since solenoid operated control valve V3 is de-energized, 
cutting off flow through subcooler return line 86. 
If the outdoor thermostat OT defines the second step of heating, 
appropriately solenoid operated control valve V2c is energized, opening 
the line 60a between the suction return line 60 and the common low side or 
inlet 30 to compressor cylinders 1 and 2 for cylinder head 24, and thus 
placing cylinders 1 and 2 in compression along with cylinders 3 and 4. At 
the same time, appropriately, the solenoid operated control valve V3 is 
energized and is shown in dotted line fashion. Under the second step 
heating mode, not only is refrigerant vapor flowing to cylinders 1 and 2 
for compression, but some liquid refrigerant, bled from line 162, passes 
to the subcooler, where it is expanded to subcool the liquid refrigerant 
within subcooler coil 166 of the liquid refrigerant line 162. The bled 
portion of liquid refrigerant, as vapor, passes at an intermediate 
pressure above suction but below full compression to point 64, where it 
enters the hermetic housing through the end bell 16, discharging over the 
stator 68 and rotor 70 of the two speed motor 20, cooling the same, 
pressurizing the compressor crank case and finally entering cylinder 3 for 
compression through port 80 which opens to the inlet or low side 38 of 
cylinder head 26 leading only to cylinder 3 of that portion of the 
compressor. Since the subcooler return vapor is at pressure higher than 
that within the suction return line 60, only the subcooler refrigerant 
vapor can pass to cylinder 3, the suction return line refrigerant in vapor 
form being directed to cylinders 1 and 2 through line 60a and cylinder 4 
through line 60b. Check valve 66 within line 60 and check valve 90 within 
line 84 limits relatively high pressure subcooler return vapor flow to 
cylinder 3. Under these conditions of operation, the compressor operates 
at maximum capacity at low speed. High speed operation would not be 
initiated until the outside air temperature drops further under the 
setting determined by the outside thermostat OT. When the outside air 
temperature as sensed by thermostat OT drops to the point whereby home 
heat loss is starting to approach the capacity of the machine when 
operating with all four cylinders at low speed, then the outdoor air 
thermostat OT will initiate a control sequence which would change control 
operation to one in which the indoor thermostat IT would not allow the 
compressor to go off. 
The indoor thermostat IT, when in its neutral position or off position, 
would permit operation of the compressor at low speed with only two 
cylinders operational, that is, cylinders 3 and 4, with solenoid operated 
control valve V2c closed and solenoid operated control valve V3 closed. 
This constitutes the first stage system operation under low outdoor 
temperature conditions. 
As the indoor temperature continues to drop, solenoid operated control 
valves V2c and V3 are energized simultaneously, adding cylinders 1 and 2 
to the compression process with the motor still running at low speed and 
effecting subcooler operation with liquid refrigerant flow to the 
subcooler 118 via line 192. This constitutes a second step heating under 
cold ambient conditions. 
Again, if the indoor temperature drops further, appropriately the motor 
control 107 is energized to change the motor connection to the stator from 
four pole to two pole and double the speed of the compressor motor 20. 
Once in high speed mode, it may be further desirable to allow an 
additional control sequence permitting the solenoid operated control 
valves V2c and V3 to be de-energized once the heating load is overbalanced 
with four cylinder operation at high speed, eliminating the subcooling 
process and removing cylinders 1 and 2 from the compression process. Thus, 
the compressor in high speed mode would cycle between two and four 
cylinders in operation, and during four cylinder operation the subcooler 
is cut into the refrigeration circuit. If this mode is selected, then the 
indoor thermostat IT will have the following control sequence during low 
ambient operation. The normally off position will allow low speed 
operation with only cylinders 3 and 4. The first step of heating will 
cause a shift to high speed, but still unloading. The second step of 
heating causes all four cylinders in the subcooler to be activated when 
under high speed mode. 
The present invention advantageously offers the acceptable alternative 
under solid state control of either effecting a speed change for the 
compressor drive motor or unloading the compressor by taking one or more 
cylinders out of the compression process. 
By further viewing FIG. 2a, it may be appreciated that the basic air source 
heat pump system involves two condensers in parallel under heating mode, 
that is, the inside air coil 112 and the inside hydronic coil 122. One is 
a refrigerant to air condenser and the other is a refrigerant to water 
condenser (operable only during the heating mode). The purpose of the 
inside hydronic condenser as shown is to allow a degree of hot water heat 
in typical retrofit applications or even new applications where the 
hydronic system is maintained at approximately 100.degree. or 110.degree. 
F. condensing temperature, thus allowing a degree of comfort during the 
colder weather that would be unobtainable with direct air systems. The 
perimeter of the residence may be provided with soft heat, while obviously 
the air flow will take care of the balance of the requirements. This would 
prevent cold spots near walls, etc., in the residence or other space being 
conditioned. 
To achieve cooling of the space within enclosure 144 to be conditioned 
under the system componentry discussed in FIG. 2a as being operable within 
that system, it is necessary only to energize solenoid operated control 
valve V1c to effect unload start up, and thence upon de-energization of 
solenoid operated control valve V1c, energization of solenoid operated 
control valves V5 and V9 in a basic system, with the four way valve 114 
being shifted to cooling mode operation, wherein discharge line 96 
connects to line 126 and suction return line 60 connects to line 128 which 
extends to the inside air coil 112. Under this type operation, with the 
compressor being driven at low speed or high speed, loaded or unloaded, 
that is, with all four cylinders 1, 2, 3 and 4 or only cylinders 3 and 4 
in the compression process, by energization of solenoid operated control 
valve V5, refrigerant flows to the outside air coil 110 which acts as a 
condenser, FIG. 2f. Solenoid operated control valve V2 is de-energized in 
this case and condensed liquid refrigerant passes to receiver 116 through 
line 182 and check valve 184. Further, the liquid refrigerant R from the 
receiver passes by way of liquid refrigerant line 162 through the 
accumulator and subcooler coils and by way of tap point 198 through line 
194, check valve 200, solenoid operated control valve V1 (which is 
energized) and line 138 to the inside air coil which is now functioning as 
the system evaporator for cooling of the space 146 to be conditioned. 
The presence of the check valve 142 prevents refrigerant flow to the inside 
hydronic coil 122, thus coil 122 does not function as an evaporator coil 
during this cooling mode. Cooling mode operation may be appropriately 
controlled in terms of multiple steps including either second to third 
stage operation by capacity control or cylinder removal from the 
compression process or speed change for the two speed motor 20. However, 
it is preferred that in cooling mode there is no speed change for northern 
latitude use. 
The present invention advantageously incorporates a heat storage coil as at 
124 for purposes of storing excess available heat under certain ambient 
FIG. 2g, and indoor, FIG. 2d, temperature conditions, for removing heat 
from storage when such heat is needed, FIGS. 2b and 2h, and for permitting 
temporary storage of heat which otherwise could be discharged to the 
atmosphere by way of the outside air coil 110 for instance, under low 
efficiency heat exchange conditions while permitting at a later time the 
discharge of the waste, stored heat under ambient temperature conditions 
more favorable to efficient surface heat transfer. For instance, in 
cooling the space to be conditioned at 146 within enclosure 144, because 
of high temperature ambient daytime conditions, system efficiency may be 
improved by rejecting heat from the refrigeration loop to storage media M 
by way of the storage coil 124 during the day, and subsequently removing 
heat from storage at night for discharge to the outside air under ambient 
temperature reduction on the order of 20.degree. or so (the difference 
between daytime and night time ambient temperature). 
Referring to FIG. 2b, the improved heat pump system of the present 
invention is shown under a heating mode condition, operating at high speed 
and with all four cylinders under the compression process. The operating 
conditions are not dissimilar from the third step heating as discussed 
with respect to FIG. 2a. However, in this case, solenoid operated control 
valve V4 is energized along with solenoid operated control valve V7 to 
permit thermal energy to be extracted from storage and delivered via the 
compressor to the inside air coil 112 and the inside hydronic coil 122 for 
heating respectively, enclosure spaces 146 and 148, along with heat 
extracted from the outside air by way of outside air coil 110. The system 
is shown after start up, so that solenoid operated control valve V1c is 
de-energized, while capacity control solenoid operated control valve V2c 
is energized and refrigerant within the suction return line 60 is 
available to all four cylinders, although under system operation as shown 
in FIG. 2b, the lower pressure suction gas returning by way of the four 
way valve 114 which is conditioned for heating mode will enter the low 
side 30 of cylinder head 24 for cylinders 1 and 2, but will be effectively 
blocked by way of check valves 66 and 82 from flowing to cylinders 3 and 4 
respectively. Additionally, solenoid operated control valves V2 and V4 are 
energized, opening the outside air coil 110 and the storage coil 124 to 
the liquid refrigerant within the liquid refrigerant line 162 downstream 
of the subcooler 118. Refrigerant vapor returns from the outside air coil 
110 through line 126 to the four way valve 114 while refrigerant returns 
from the storage coil (both storage coil 124 and the outside air coil 110 
are acting as evaporators) by way of line 84 to the low side 40 of 
cylinder 4 of the compressor via line 92 upon energization of solenoid 
operated control valve V7 and de-energization of solenoid operated control 
valve V6, within lines 172 and 84, respectively. 
Further, by energization of solenoid operated control valve V3, liquid 
refrigerant bled from the refrigerant line 162 and expanding within the 
subcooler and about subcooler coil 166, causes the intermediate pressure 
vapor to be returned via line 86 to connection point 64 where it enters 
the interior of the hermetic compressor unit 12 end bell 16 through line 
62 and acts to cool the drive motor components and pressurize the crank 
case, entering the low side 38 of cylinder head 26 for recompression by 
cylinder 3 via port 80. The vapor pressure of the return from subcooler is 
higher than that of the refrigerant vapor within line 84 from the storage 
coil 124 or the refrigerant vapor within the suction return line 60 and 
thus check valves 90 and 66, respectively prevent refrigerant vapor from 
the storage coil and from the outside air coil from mixing with the 
subcooler return and passing through cylinder 3 for recompression. 
It is noted that under this type of operation solenoid operated control 
valves V1c, V1, V5, V6, V8, V10, V11, V12 are de-energized while as stated 
previously, solenoid operated control valves V2c, V2, V3, V4, V7 and V9 
are energized. With change in inside load requirements and outside ambient 
temperature conditions, it may not be necessary to remove heat from 
storage, and in that case, solenoid operated control valves V4 and V7 may 
be de-energized forcing refrigerant to circulate only through the outside 
air coil 110 for picking up heat which is then directed to the inside air 
coil 112 and the inside hydronic coil 122 as discussed previously. 
Obviously, under the heating mode, the motor speed may be changed from 
high speed to low speed and vice versa and the compressor loaded or 
unloaded by removal of cylinders 1 and 2 from the compression process 
under automatic control provisions in response to temperature sensed by 
the outside thermostat OT and the inside thermostat IT adjacent coil 110 
and within enclosure space 146. 
Alternatively, it may be desired that the storage coil 124 carry the load 
totally due to unfavorable ambient air conditions, in which case solenoid 
operated control valves 2 and 7 would be de-energized, while solenoid 
operated control valves 4 and 6 would be energized, terminating 
refrigerant flow from the liquid refrigerant line 162 to the outside air 
coil 110 and causing all of the liquid refrigerant to be directed to the 
storage coil 124, whereupon by expansion thermal energy is picked up from 
storage and directed to all four cylinders (if desired by energization of 
the capacity control solenoid operated control valve V2c) through four way 
valve 114 and by way of the suction return line 60. 
Further, it is apparent that if the pressure level in the return line from 
the storage coil 124 were sufficiently high, in comparison to the pressure 
of the refrigerant vapor within line 86 returning from the subcooler 118, 
some of the refrigerant vapor would flow through check valve 90 to mix 
with the subcooler return vapor and enter through the hermetic compressor 
unit 12 and low side 38 leading to cylinder 3, while the remaining 
refrigerant vapor returning to the compressor from storage would pass 
through line 92 to the low side 40 of the same cylinder head 26 for 
recompression by cylinder 4. 
Turning next to FIG. 2c, the heat pump system of the present invention is 
illustrated again under a heating mode. However, in this case there are no 
heating or cooling requirements for the enclosure 144 and specifically the 
spaces 146 and 148 to be conditioned. However, thermal energy is available 
from the outside air under favorable system efficiency conditions to 
permit that thermal energy to be stored by way of storage coil 124. In 
this mode of operation, solenoid operated control valves V1c, V1 (and for 
purposes of illustration), V4, V5, V6, V7, V11, V12 are de-energized, 
while solenoid operated control valves V2c, V2, V3, V8 and V10 are 
energized. Since the outdoor air coil 110 is acting as an evaporator and 
an outdoor air heat source, the vaporized refrigerant returning to the 
compressor by way of the four way valve 114 is directed through the 
suction return line 60 to all four cylinders since solenoid control valve 
V2c is energized and refrigerant is available to cylinders 1 and 2 through 
line 68. With the solenoid operated control valve V3 energized, there is 
refrigerant for subcooling, and check valve 66 prevents refrigerant vapor 
within the suction return line 60 entering end bell 16 of the hermetic 
compressor unit 12 for cooling of the motor windings, pressurization of 
the compressor crank case and movement to the low side 38 of the cylinder 
head 26 by way of port 84 compression by cylinder 3, this being achieved 
by subcooler return vapor at relatively high pressure. Cylinder 4 is fed 
through line 60b. 
The compressed refrigerant vapor at high pressure being discharged from the 
discharge manifold 48 by way of line 96 cannot pass to the inside air coil 
112 and the inside hydronic coil 122 since the solenoid operated control 
valve V9 is de-energized. However, since solenoid operated control valve 
V8 is energized, this opens line 206, leading to one side of the storage 
coil 124, permitting compressed refrigerant vapor to enter the storage 
coil for condensation therein and transfer of heat to the storage media M. 
With the solenoid operated control valve V4 de-energized, refrigerant 
return from the storage coil is permitted through line 186 and check valve 
188 to the receiver, the liquid refrigerant R within the receiver flows 
through the accumulator coil 164 and subcooler coil 166 with excellent 
subcooling and to the outside air coil 110 which is acting as the 
evaporator for the system, since solenoid operated control valve V2 is 
energized. Heat is absorbed from the atmosphere and the refrigerant vapor 
returns through check valves 129 and line 126 through the four way valve 
114, where after passing through accumulator 120 it enters cylinders 1, 2 
and 4 of the compressor for recompression. Thus, the storage tank 
temperature can be built up under mild temperature conditions where there 
is no system requirement to either cool or heat the enclosure spaces 146 
and 148 and flow to the inside air cool 112 and the inside hydronic coil 
122 can be terminated. Operation can be effected at low or high speed and 
with two or more cylinders. 
As discussed previously, the improved heat pump system of the present 
invention may be advantageously employed to effect daytime cooling of the 
enclosure 144 and particularly space 146 by operating the system under a 
cooling mode in which the inside air coil 118 functions as the system 
evaporator. However, it may be under given ambient condition, that it will 
be required during the night to in fact add heat by reversing the system 
and employing the inside air coil 112 and the inside hydronic coil 122 as 
condensers. Under such conditions, it is preferred that the heat removed 
from the space, such as 146 to be conditioned during daytime, be stored by 
storage coil 124 during the day and then removed from storage under 
relatively high thermal efficiency conditions and supplied to the 
enclosure 144 at night rather than extract heat from the atmosphere by way 
of the outside air coil 110. During daytime, therefore, referring to FIG. 
2d, under the certain ambient temperature conditions, heat may be stored 
within the storage media in comparison to the ambient heat rejection 
conditions, solenoid operated control valves V1c, V2, V4, V5, V7, V6, 
V11,V12 are de-energized, while solenoid operated control valves V2c, V1, 
V3, V8, V9 and V10 are energized. The four way control valve is shifted to 
cooling mode operation such that discharge line 96 is connected to line 
126 leading to the storage coil 124, while the line 128 including the 
solenoid operated control valve V9 is connected to the suction return line 
60. With solenoid operated control valve 5 de-energized, outside air coil 
110 cannot act as a waste heat discharge and the heat is stored by the 
storage media M with all refrigerant flow from the compressor passing 
through the storage coil 124 via line 206 and control valve V8, with 
solenoid operated control valve V6 closed. Energization of solenoid 
operated control valve V10 permits the condensed refrigerant to pass 
through the receiver 116 where the liquid refrigerant after passing 
through the accumulator and subcooler coils 164 and 166 and being 
subcooled is directed by way of line 194, with the solenoid operated 
control valve V1 energized, to the inside air coil 112 where the liquid 
refrigerant is expanded; coil 112 functioning as an evaporator for cooling 
the enclosure space 146. 
The compressor is operating at full capacity with solenoid operated control 
valve V2c energized and line 60a open. Solenoid operated control valve V3 
is also energized so that the subcooler is operating to subcool the liquid 
refrigerant being fed in this case to the inside air coil 112, and wherein 
the refrigerant vapor returning by way of subcooler return line 86 to the 
compressor feeds to cylinder 3 through the hermetic casing. Being at a 
higher pressure than that of the suction return line 60, it thus 
pressurizes the crank case and prevents, because of check valve 66, 
refrigerant vapor within the return line 60 from mixing with the subcooler 
return vapor and passing to cylinder 3 via cylinder head low side 38. 
Again, temperature signals emanating from the outside thermostat OT and the 
inside thermostat IT control the energization of the solenoid operated 
control valves and the shifting of the four way valve between heating and 
cooling modes. Obviously, with the suitable control unit 107, compressor 
22 and the system components, the system can operate unloaded with two 
cylinders, loaded with four cylinders, as the case may be, and at high or 
low speed with heat being added to storage and removed from the space 146 
being conditioned with the inside air coil functioning as the evaporator 
for the closed loop system. 
FIG. 2d represents operation under four cylinders with the two speed motor 
20 operating at high speed. The system controls may be such that under 
cooling mode, the motor would always operate at low speed and there would 
be either two cylinders operating or four cylinders depending upon 
energization of the solenoid operated control valve 2c. Further, while the 
illustrated system preferably employs a control program via unit 107, 
where solenoid operated control valves V2c and V3 are energized 
simultaneously and de-energized simultaneously, this could be modified to 
permit full capacity operation for the compressor but without subcooler 
operation in which case solenoid operated control valve V2c would be 
energized but solenoid operated control valve V3 would remain 
de-energized. 
Further with respect to the system operating in accordance with FIG. 2d, 
while during the day the system operates under a cooling mode to provide 
light cooling to the space 146 of the enclosure 144 under mild ambient 
conditions, it may be necessary at night to in fact heat the same space 
which was cooled during the daytime. This is achieved, FIG. 2f, by simple 
reversal of the four way valve from cooling mode to heating mode, wherein 
the discharge line 96 is connected to line 128 while the line 126 is 
connected to the suction return line 60. The thermostat ST or other 
temperature sensitive device sensing the temperature of the storage media 
M within the storage device 125 causes, under programmed control, at unit 
107 the selection of the storage coil 124 as the source of that heat 
rather than the outside air coil 110. The storage coil then acts as the 
evaporator for the system with the inside air coil 112 and the inside 
hydronic coil 122 acting as condensers. Vaporized refrigerant is returned 
to the compressor through line 172 with the solenoid operated control 
valves V4 and V6 energized, solenoid operated control valve V10 
de-energized and directing the refrigerant to the compressor through the 
suction return line 60. Assuming again that the compressor is operating 
with all four cylinders 1, 2, 3 and 4 involved in the compression process, 
and with the subcooler operating by energization of solenoid operated 
control valve V3 refrigerant vapor is returned to compressor cylinders 1, 
2 and 4 by way of the suction return line 60 and lies 60a and 60b. The 
subcooler return vapor being at a higher pressure than that of the suction 
return line 60 causes the check valve 66 to operate to prevent refrigerant 
vapor flow from the suction return line 60 to cylinder 3 but permits that 
cylinder 3 to receive all of the vapor returned from the subcooler. 
Reference to FIG. 2e shows the system operating in a heat rejection mode 
wherein the heat previously stored within the storage media is removed 
from storage by operation of the storage coil 124 as an evaporator and 
with the outside air coil 110 being employed as a heat reject condenser. 
No heating or cooling requirements exist for the enclosure 144 although 
the system permits the inside air coil 112 and the inside hydronic coil 
122 to function as condensers for heating spaces 146 and 148, if desired, 
while at the same time dissipating heat from storage to the outdoor air, 
or alternatively, refrigerant may be directed reversely through the indoor 
air coil 112 to effect an evaporation action and removal of heat from the 
space 146 while still dissipating heat from storage to the outside air 
through outdoor air coil 110. 
However, in the illustrated embodiment of the present invention, as per 
FIG. 2e, only solenoid operated control valves V4, V5, V7, V12 are 
energized, while solenoid operated control valves V1c, V2c, V1, V2, V3, 
V6, V8, V9, V10 and V11 are de-energized. The machine is operating at low 
speed (or perhaps under high speed conditions) but with solenoid operated 
control valve V2c de-energized refrigerant vapor returning from the 
storage coil 124 after vaporization and pick up of heat returns through 
line 84, check valve 90 and line 62 to end bell 16, passing over the motor 
components for cooling the same, entering the compressor crank case and 
also passing to cylinder 3 via port 80 for recompression, while a further 
portion of that refrigerant vapor passes through line 92 to the low side 
40 of the compressor head 26 for recompression by way of cylinder 4 and 
for common discharge from the high side 42 of that cylinder to discharge 
line 96 via discharge manifold 48. 
The compressed refrigerant vapor passes through check valve 98 and through 
the four way valve 114 to line 126 since lines 96 and 126 are connected by 
the four way valve 114 with the four way valve in cooling mode position. 
The energization of the solenoid operated control valve V5 permits the 
refrigerant vapor to flow to the outside air coil 110, now acting as the 
system condenser, and discharging the heat at relatively low temperature 
night time ambient conditions with the temperature in the range of 
70.degree. to 75.degree. F. The condensed refrigerant flows through the 
check valve 184 to the receiver 116 via line 182. Liquid refrigerant R 
returns to the storage coil 124 for vaporization, through liquid 
refrigerant line 162 and passage through accumulator coil 164 and 
subcooler coil 166. With the solenoid operated control valves V2c and V3 
de-energized, obviously there is no liquid refrigerant for subcooling 
purposes and no subcooling return vapor within return line 86. Heat is 
then dissipated from the storage coil under low motor load and high system 
efficiency conditions. Solenoid operated control valve V12 within line 182 
is open. 
With respect to the illustrated embodiment of the invention and the various 
modes of operation, it may be appreciated that additional changes both in 
the control format and in the structural aspects of the heat pump system 
and the compressor may be made without departing from the spirit of this 
invention. For instance, the invention is broadly directed to a heat pump 
system incorporating a multi-cylinder reciprocating compressor and the 
compressor in simplified form may comprise three cylinders or four 
cylinders. However, multiple cylinders performing the function of a given 
first, second, third and fourth cylinder may be accomplished in commercial 
practice, particularly for large heat pump systems. For instance, the 
compressor may constitute more than two cylinder heads or may be three, 
six, nine or twelve cylinders, or a four cylinder machine may be enlarged 
to incorporate eight or twelve cylinders acting in banks of two and three 
respectively and functioning for the individual cylinders of the four 
cylinder reciprocating compressor illustrated in the embodiment of the 
invention found within the drawings. Further, indoor and outdoor 
thermostats and a media thermostat supply control signals to the control 
unit 107, and solenoid operated control valves control the flow of 
refrigerant within the circuit. The control valves could be other than 
solenoid operated, and the control scheme can be modified to accomplish 
the same purposes without departing from the invention. The two speed 
motor preferably comprises a single phase alternating current motor but 
obviously it could be a two phase, a three phase, or a direct current 
motor. 
In a typical control format for operating control device 107, the 
connections to the various solenoid operated control valves and based on 
inputs received from the storage thermostat ST, outdoor thermostat OT and 
indoor thermostat ID, are provided in the chart below. 
______________________________________ 
TYPICAL SYSTEM CONTROL FORMAT 
Outside Air Thermostat 
______________________________________ 
Indoor Thermostat 
Moderate Ambient 
Low Ambient 
Neutral Position 
Unit Off 3 & 4 Low Speed 
First Step Heat 
3 & 4 Low Speed 
1, 2, 3 & 4 Low 
Speed Subcooler if 
Air Source 
Second Step Heat 
1, 2, 3 & 4 Low 
1, 2, 3 & 4 High 
Speed (No Subcool- 
Speed Subcooler 
er) ON 
______________________________________ 
While the invention has been particularly shown and described with 
reference to a preferred embodiment thereof, it will be understood by 
those skilled in the art that various changes in form and details may be 
made therein without departing from the spirit and scope of the invention.