Refrigeration system with compressor load transfer means

A closed cycle, multiple compressor, multiple evaporator, refrigeration system of the type particularly adapted to supermarket applications having a common condenser, a multiplicity of refrigerated fixtures with associated evaporators for cooling, all operating at the same temperature and pressure, all discharging into a common compressor suction header and a series of parallel compressors pumping from the common suction header to the condenser, together with a second set of refrigerated fixtures having evaporators operating at a lower pressure than that of the first multiplicity of evaporators, discharging into a second, lower pressure, suction header, from where its effluent refrigerant is pumped by a second compressor system back to the common condenser. The two compressor suction headers are connected by a conduit containing a pressure regulating valve which senses suction pressure in the lower pressure system and, as necessary, transfers refrigerant from the higher pressure suction header to the lower pressure suction header so as to fully utilize the pumping capacity of the compressor system pumping from the low pressure suction header.

FIELD OF INVENTION 
This invention relates to a means for load sharing between various 
refrigeration compressors of a multiple compressor, parallel, 
refrigeration system wherein the total evaporator, refrigerant load is 
comprised of a plurality of evaporators which are operating at different 
refrigerant temperatures and pressures. Multiple compressor, parallel, 
system refrigeration plants are typically utilized in situations where the 
refrigeration load or demand varies from time to time, such as in 
supermarkets, manufacturing and food processing plants. 
DESCRIPTION OF PRIOR ART 
There is an increasing demand for more energy efficient means of providing 
refrigeration for today's food processing, storage and sales industries. 
These industries have for many years been confronted with an increasing 
demand for refrigeration coupled with increasing costs of energy to run 
the refrigeration systems. 
In the earliest stages of development of refrigeration technology each 
refrigeration fixture, cooled by its own evaporator, was supplied by a 
separate system, with separate compressors, condensers and control 
systems. A development which quickly followed was use of a single 
compressor and condenser to supply multiple evaporators, functioning 
independently of one another. 
The basic principle underlying the multiple evaporator refrigeration system 
was that one compressor could pump the low pressure, gaseous, refrigerant 
effluent from the evaporators into a common condenser. The evaporators 
would be flooded through their respective thermal expansion valves from 
the common supply of high pressure liquid refrigeration from the 
condenser. 
The typical multiple evaporator system in common use for many years and in 
production today, utilizes a thermal expansion valve and evaporator 
pressure regulator valve for each separate evaporator. The evaporator 
pressure regulator valve, located at the discharge of the evaporator, is 
set to control the pressure within the evaporator, and is thereby used to 
fix the saturation temperature of refrigerant in the evaporator. The 
thermal expansion valve senses the temperature of the evaporator near the 
discharge, and in conjunction with an input for evaporator pressure, 
operates to admit liquid refrigerant into the evaporator as necessary to 
keep the evaporator in a nearly flooded condition. 
Single compressor, multiple evaporator, refrigeration systems such as just 
described work satisfactorily in applications where the total 
refrigeration load is relatively small and constant. 
With single compressor systems, the compressor must necessarily be sized to 
handle the maximum load. Engineering and economics dictate practical 
limitations on the capacity and horsepower of compressors for single 
compressor systems. And in general, compressors of over 25 horsepower are 
not found in common usage in supermarkets, commercial food warehouses, 
storage facilities and processing plants because of these considerations. 
But even for systems requiring less than 25 compressor horsepower, single 
compressor systems are not necessarily the most desireable for two 
additional reasons. The first of which concerns variations in load demand. 
The second reason is that extremely large compressors would be needed to 
meet the demands of numerous evaporators connected to it. 
A basic requirement of the single compressor system is that the compressor 
be sized to meet the pumping requirements of all the evaporators operating 
at maximum capacity at one time. 
In practice, this is not the situation usually found. For example, the 
refrigeration load in a supermarket on a cold winter night will be 
substantially below the amount of refrigeration required on a hot summer 
afternoon when the supermarket is full of customers opening and closing 
cooler doors and reaching into open refrigerated display cases. 
During these periods of low refrigeration demand the evaporator pressure 
regulator valves for the various evaporators will be discharging less 
gaseous refrigerant from the evaporators to the suction of the compressor. 
This can result in serious inefficiencies if one large compressor is used. 
During periods of low demand the compressor will draw down the pressure in 
its suction header thereby increasing the pumping head and possibly 
damaging the compressor itself. 
Numerous and varied control systems have been developed over the years to 
avoid compressor damage in such situations. Two methods well known in the 
art involve cycling the compressor on and off during periods of low 
demand, and load by-pass systems. While these systems are effective to 
protect the compressor they are inefficient in other respects. Cycling a 
large compressor on and off wastes power and usually shortens compressor 
lifetime. Load by-passing, of course, represents a total waste of energy 
and pumping power. 
A practical solution to these problems inherent to the single, large 
compressor system has been well known in the art for many years. A series 
of smaller compressors, connected in parallel are used in lieu of one 
large compressor. If, for example, the maximum design refrigeration load 
was 80 tons, four compressors capable of pumping the refrigerant necessary 
for 20 tons each of cooling would be installed, instead of one compressor 
with four times the capacity. 
The control system normally installed in parallel compressor systems senses 
the common suction header pressure for the four compressors and turns the 
compressors on and off in sequence as necessary to meet pumping demands. 
In this way the amount of pumping capacity being utilized at any given 
time will be roughly equivalent to the demand. This can represent 
significant power savings and operating cost reductions, particularly in 
supermarket applications. 
There is, however, one significant problem remaining which is common to 
both single and parallel compressor systems which supply multiple 
evaporators. That is, the compressor suction pressure must be lower than 
the pressure in the lowest pressure, coldest, evaporator connected to the 
system. If an evaporator designed to provide cooling, but not freezing 
were connected in parallel with an evaporator designed to maintan subzero 
temperatures the pressure in the compressors' suction header would have to 
be lower than the design pressure of the subzero evaporator. 
For example, the design pressure for an evaporator contained in a 
supermarket beverage cooling fixture using R-502 Refrigerant normally 
ranges from 50 to 55 psi and the design pressure for an evaporator 
designed for use in a supermarket frozen food fixture using R-502 normally 
ranges from 13 to 15 psi. If both evaporators were connected to a common 
discharge, compressor suction, header the pressure in that suction header 
would have to be maintained below 13 psi. If it were not, then there would 
be a back pressure against the evaporator pressure regulator valve for the 
colder evaporator, hence no discharge of gaseous refrigerant from that 
evaporator. 
The net result is that the refrigerant discharge from the higher pressure, 
warmer, evaporator operating at 50 to 55 psi would have to be 
substantially reduced as it entered the suction header before it was again 
pumped into the high pressure side of the refrigeration system. To 
accomplish this pressure reduction, the pumping capacity of the system 
would have to be necessarily increased in order to draw down the pressure 
in the common discharge header. Two inefficiencies become readily 
apparent; the first is the cost of pumping refrigerant across a greater 
pressure differential than is necessary; the second is the increased 
equipment costs incurred through the requirement for more pumping 
capacity. 
Since pumping across the minimum possible pressure differential always 
results in savings in pump operating costs and smaller pump capacity 
equipment requirements, parallel and single compressor systems are usually 
designed to serve evaporators operating at relatively uniform refrigerant 
temperatures and pressures. In the case of the modern supermarket, two 
parallel compressor systems are commonly installed. The first parallel 
compressor system pumps from the evaporators of the fixtures used for 
cooling and a second parallel compressor system from the evaporators of 
the fixtures used for freezing. For the past several years this type of 
general classification of evaporators, those for freezing and those for 
cooling has been the accepted catagorization for determining to which 
compressor the evaporators are connected. 
In addition to the advantages of the parallel compressor system previously 
described, there remains one additional advantage and that is the ability 
to handle the defrosting cycles of the various evaporators. It is 
economically essential in today's modern supermarket, that the 
refrigeration fixtures have automatic defrost capabilities. There are 
numerous designs for automatic defrosting, which are well known in the 
art. 
One of the most common methods is to isolate the evaporator of the fixture 
to be defrosted from the parallel refrigeration system, and then to employ 
electric heaters to defrost the fixture and evaporator contained within 
it. Automatic defrosting is done periodically, usually by means of a 
shutoff valve, downstream of the evaporator pressure regulator valve; said 
shutoff valve being controlled by a solenoid and timer-clock. 
When the shutoff valve closes, there is no flow of refrigerant through the 
evaporator. While the heaters are on and the fixture and evaporator coils 
are being defrosted, the liquid refrigerant trapped in the evaporator will 
evaporate and collect superheat, thereby increasing the pressure in the 
evaporator. When the defrost cycle is completed and the heaters are off, 
the shutoff valve opens and the high pressure superheated refrigerant 
quickly discharges to the compressor suction header. 
With a parallel compressor system, this rapid, short term, increased load 
is easily handled by the standby pumping capacity. With a single 
compressor system this extra load must be handled by the single compressor 
and during periods of high refrigeration demand can overload the 
compressor and slow the rate of cooldown of the defrosted evaporator from 
defrost temperature back to normal operating temperature. 
But even with the advantages of the parallel compressor system, there 
remain limitations created by operating cost considerations. In many 
modern applications, especially supermarkets, the use of two separate 
parallel compressor systems may not meet all the refrigeration needs. 
There are oftentimes refrigeration loads which fall halfway between the 
cooler and freezer classification and other loads which require super cold 
evaporator temperatures substantially below the temperatures and pressures 
of normal freezer fixture evaporators. Examples of these intermediate 
temperature supermarket fixtures are meat and delicatessen coolers. The 
ice cream display case normally requires a super cold evaporator. 
These intermediate refrigeration fixtures, falling between the cooler and 
freezer parallel compressor systems and the super cold fixtures have 
traditionally been handled by separate refrigeration systems. Of course, 
by the use of separate compressor systems, all of the problems heretofore 
described as inherent to the single compressor system once again appear. 
Recent developments in the art aimed at solving this problem, have 
utilized what is commonly referred to as a satellite compressor to handle 
these intermediate and super cold refrigeration fixtures. 
In the typical satellite compressor installation the intermediate or super 
cold evaporator receives its supply of high pressure liquid refrigerant 
from the condenser receiver of the cooler or freezer parallel compressor 
systems respectively. Control of temperature and refrigerant flow through 
the evaporator is achieved in the same manner as in the parallel system 
evaporators. However, the discharges from these evaporators are not 
returned to the parallel compressor suction header, but rather, to 
separate compressors which complete the refrigeration cycle by pumping the 
refrigerant back to the parallel compressor system's common condenser. 
By utilizing the satellite compressor two significant savings are realized; 
the first is the elimination of the need for additional condensers, 
receivers, and the associated hardware for an entirely separate system; 
the second is that by not connecting the discharge of the intermediate or 
super cold evaporators to the cooler or freezer parallel compressor 
suction headers, we eliminate the need to reduce the suction pressures of 
the parallel compressor systems. This allows each parallel compressor 
system to operate at the highest possible suction pressure and still draw 
its suction from the majority of the evaporators connected to the system. 
The result is a substantial operating cost savings because the parallel 
compressors, serving the majority of the evaporators, operating at the 
same temperatures and pressures, are pumping across a minimum pressure 
differential, and the intermediate temperature and super cold evaporators 
are served by separate compressors pumping across greater pressure 
differentials than their respective parallel compressor systems. 
The drawback is, as previously stated, that the satellite compressors must 
be sized for the maximum loads, thereby creating unused and unneeded 
pumping capacity during low load periods of time. In addition, to the 
wasted pump capacity, problems of cycling the pump and overtaxing the pump 
during discharge after defrost of its evaporator are present in the 
satellite system. 
In practice, it is common to have, during low load periods, the first stage 
parallel compressor running, the second stage parallel compressor cycling 
on and off, and the satellite compressor cycling on and off. 
OBJECTS OF THE INVENTION 
Accordingly, it is the basic object of this invention to provide a more 
efficient and less costly parallel compressor, multiple evaporator 
refrigeration system by means of a novel method for optimizing and 
maximizing the use of the pumping capacity of each compressor within the 
parallel compressor system. 
Another object of this invention is to provide a means whereby the pumping 
capacity of a satellite compressor, attached to a parallel compressor 
system, is fully utilized at all times. 
A third object of this invention is to eliminate the need for cycling, or 
by-passing the load of, a satellite compressor and to level the load on 
the satellite compressor during periods of defrosting. 
A final object of this invention is to provide a means whereby a series of 
compressors, each serving an evaporator or evaporators, operating at 
different temperatures and pressures, can be connected together to create 
a parallel system, in which the compressor serving the coldest evaporator, 
or evaporators, becomes the first stage compressor and the remaining 
compressors are staged, in sequence, from the compressor serving the next 
coldest evaporator to the one serving the warmest evaporator or 
evaporators. 
SUMMARY OF THE INVENTION 
These objects are achieved in a system having a common condenser, and two 
separate compressors each independently pumping from an evaporator 
operating at a different temperature and pressure, by connecting the 
compressor suction header of the compressor pumping from the higher 
pressure evaporator to the suction header of the compressor pumping from 
the lower pressure evaporator by means of a connecting header containing a 
pressure regulating valve. This pressure regulating valve senses pressure 
in the low pressure compressor suction header and is preset to maintain a 
certain minimum pressure in the low pressure compressor suction header. 
The set point, or minimum desired low pressure compressor suction header 
pressure, is empirically determined and is that suction pressure at which 
the compressor pumping from the low pressure suction header will be fully 
and optimally utilized. In practice it has been determined that the 
optimal set point for minimum pressure in the low pressure suction header 
should be set at that point which represents full utilization of the 
pumping capacity of the low pressure compressor. 
During periods of low demand on the total refrigeration system the 
pressures in both the higher and lower pressure suction headers will be 
drawn down by their respective compressors. As the pressure in the low 
pressure compressor suction header falls below the predetermined set point 
of the pressure regulating valve, the valve opens, allowing gaseous 
refrigerant from the high pressure suction header to flow through the 
connecting header to the low pressure suction header as necessary to 
maintain the minimum preset pressure. 
If the demand on the total refrigeration system falls below the preset, 
optimal, pumping capacity of the low pressure compressor as determined by 
the minimum low pressure suction header pressure, a control system 
operates to turn off the unneeded high pressure compressor and all of the 
pumping requirements of both the higher and lower pressure evaporators 
will be handled by the low pressure compressor. If total refrigeration 
demand continues to fall, the low pressure compressor may, as determined 
by the control system installed, be cycled on or off, or, have its pumping 
load by-passed. 
In this manner the compressor serving the lowest pressure evaporator will 
always be optimally utilized first, and during periods of low 
refrigeration demand, a minimum number of compressors will be running or 
cycling on and off. 
It should be readily apparent to those skilled in the art that this 
invention will work as well with parallel evaporator and compressor 
systems. The connecting header and pressure regulating valve can be used 
to connect the suction header of a compressor, or series of parallel 
compressors, serving a plurality of evaporators, all operating at 
approximately the same temperature and pressure, to the suction header of 
a compressor, or series of parallel compressors, serving a set of lower 
pressure evaporators. Likewise, it should be apparent that this invention 
can be used in applications where there are several compressors serving 
evaporators at widely varied temperatures and pressures.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS 
FIG. 1 shows to advantage the features of this invention when used in a 
refrigeration system which has a common condenser 3 and three evaporators 
18, 17 and 16 each being independently served by compressors 22, 23 and 24 
respectively. In FIG. 1, each of the three evaporators is designed to 
control at a different temperature with evaporator 18 being the coldest, 
evaporator 17 set for an intermediate temperature and evaporator 16 for 
the warmest temperature. 
Refrigerant is admitted to each of the three evaporators through thermal 
expansion valves 5 from a common refrigerant supply header 4 which, in 
turn, is supplied from the system condenser 3. 
The evaporators 16, 17 and 18 remove heat from the fixtures to be cooled by 
means of passing subcooled refrigerant through the evaporators. Control 
over the saturation temperature of the refrigerant within each of the 
evaporators and the flow of refrigerant through them is accomplished by 
two valves, thermal expansion valves 5 located at the inlets to each of 
the evaporators, and evaporator pressure regulator valves 21, 20 and 19 at 
the discharges of each of the evaporators. 
The actual pressure in each of the evaporators is controlled by evaporator 
pressure regulator valves. And since the saturation temperature of the 
refrigerant is a direct function of pressure within the evaporator. The 
higher the pressure in the evaporator, the higher the saturation 
temperature of the refrigerant entering into the evaporator through the 
thermal expansion valve, hence the higher the temperature at which the 
evaporator operates. In effect, the pressure at which the evaporator 
pressure regulator valves 21, 20 and 19 are set to control at determines 
the temperature at which the evaporators will operate at. 
The thermal expansion valves 5 are designed to keep the evaporators 16, 17 
and 18 in, as nearly as possible, a flooded condition. This is 
accomplished by the thermal expansion valves 5 through the use of two 
sensory inputs, one is evaporator pressure and the second is the 
temperature of the refrigerant in the vicinity of evaporator discharge. By 
combining these two inputs, thermal expansion valves 5 sense the presence 
of superheat at the discharge of each evaporator. 
In FIG. 1 evaporator 18 is designed to control at the coldest temperature, 
hence its evaporator pressure regulator valve 19 is set to control at a 
pressure lower than that of evaporator pressure regulating valve 20 for 
the intermediate temperature evaporator 17. Likewise, the evaporator 
pressure regulating valve 21 for the warmest evaporator 16 will be set at 
a higher pressure than evaporator pressure regulator valve 20. 
The gaseous refrigerant from evaporator 18 is discharged through evaporator 
pressure regulating valve 19 into the low pressure suction header 25, from 
where it is pumped by compressor 22 through check valve 15 into the 
parallel system compressor discharge header 2 for return to condenser 3. 
In a like manner evaporators 17 and 16, discharging through evaporator 
pressure regulator valves 20 and 21, pass effluent refrigerant to the 
intermediate and high pressure suction headers 26 and 27. 
When the entire system is operating at design capacity the pressure in low 
pressure suction header 25 will be lower than the pressure in intermediate 
pressure suction header 26, which in turn is lower than the pressure in 
the high pressure suction header 27. 
In this invention the medium pressure suction header 26 is connected to the 
low pressure suction header 25 by the first stage connecting header 33. 
The first stage pressure regulator valve 28 in the first stage connecting 
header 32 senses the pressure in the low pressure suction header 25. 
During periods of low demand on evaporator 18 low pressure compressor 22 
through operation will reduce the pressure in the low pressure suction 
header 25 below the designed set point of the first stage pressure 
regulator valve 28. 
The design set point is that pressure at which compressor 22 is running at 
optimal capacity. When the pressure in suction header 25 falls below the 
set point, the first stage pressure regulator valve 28 opens to discharge 
medium pressure gaseous refrigerant from medium pressure suction header 26 
into low pressure suction header 25. 
In a like manner, high pressure suction header 27 is connected to medium 
pressure suction header 26 by second stage connecting header 34. Second 
stage pressure regulator valve 29 senses the pressure in medium pressure 
suction header 26 and when the pressure in medium pressure suction header 
26 falls below its designed pressure, the second stage pressure regulator 
valve 29 opens allowing the high pressure gaseous refrigerant of high 
pressure suction header 27 to pass through the second stage connecting 
header 34 to the medium pressure suction header 26. 
In this manner, the gaseous refrigerant in suction headers 27 and 26 will 
pass through the second stage connecting header 34 and the first stage 
connecting header 33 to low pressure suction header 25 during periods of 
time of low demand on the total refrigeration system. 
By such a design, low pressure compressor 22 will be pumping all of the 
effluent gaseous refrigerant up to its maximum design capacity. When low 
pressure compressor 22 reaches its maximum design capacity the pressure in 
low pressure suction header 25 will begin to rise and first stage pressure 
regulator valve 28 will close at its predetermined set point. As the load 
further increases medium pressure compressor 23 will turn on to pump the 
effluent medium pressure gaseous refrigerant in medium pressure suction 
header 26, that is no longer being discharged to the low pressure suction 
header 25 through the first stage pressure regulator valve 28. If the load 
on the medium pressure compressor 23 increases to the point where the 
pressure in the medium pressure suction header 26 rises above the set 
point of the second stage pressure regulator valve 29, then second stage 
pressure regulator valve 29 will close and high pressure compressor 24 
will turn on. 
In this manner, the economies of a parallel compressor system are fully 
utilized in that only the minimum amount of compressor capacity is 
operating during periods of low demand and the optimal design efficiency 
of each compressor is fully utilized during periods of higher demand in 
that each compressor, during periods of higher demand, only pumps across 
the minimum pressure differential. 
FIG. 2 shows an embodiment of the invention that is typical of 
refrigeration system installations in supermarkets. Evaporators 8 with 
their respective thermal expansion valves 5 and evaporator pressure 
regulator valves 6 are all designed to operate at approximately the same 
temperature. Since each of the evaporators 8 are operating at the same 
temperature, the pressures of the gaseous refrigerant discharged through 
evaporator pressure regulator valves 6 are approximately equal. The 
parallel evaporators 8 all discharge to compressor suction header 7. 
Parallel system compressors 1 each draw from suction header 7 and are 
designed and controlled so that one compressor is running nearly 
continuously with the second and third compressors turning on in sequence, 
as necessary, as the pressure in suction header 7 increases. 
This sequential control is normally accomplished by use of a pressure 
sensing device which senses pressure in the compressor suction header 7. 
When the refrigeration system is operating a first compressor 1 will 
normally run continuously, only turning off if it reduces pressure in the 
compressor suction header 7 below a predetermined set point. As load on 
evaporator 8 increases, pressure in suction header 7 will rise. As this 
pressure increases above a predetermined set point, a second compressor 1 
is turned on. Likewise, if pressure in compressor suction header 7 
continues to rise, a third compressor 1 will be turned on. 
Similarly, as the refrigeration load decreases, the operating compressors 1 
will draw the pressure in compressor suction header 7 down, and as 
pressure falls below predetermined set points the compressors 1, are 
sequentially turned off. 
Evaporator 9 is designed to operate at a lower saturation temperature, 
hence a lower pressure. Evaporator 9 receives its refrigerant from the 
common refrigerant supply header 4 through its thermal expansion valve 5 
and discharges its effluent refrigerant through its evaporator pressure 
regulator valve 10 to its own, satellited compressor 12 through its own 
suction header 11. 
Satellite compressor 12, is connected to the parallel system and pumps its 
discharge through check valve 15 directly into the common compressor 
discharge header 2. 
Because evaporator 9 is operating at a lower temperature and pressure, the 
pressure in suction header 11 is necessarily lower than would be required 
in suction header 7. To prevent the inefficiencies created by reducing the 
pressure in suction header 7 and discharging the effluent refrigerant from 
evaporator 9 into suction header 7 and thereby requiring the parallel 
system compressors 1 to pump across a greater pump head than is necessary, 
the effluent from evaporator 9 is discharged into a separate suction 
header 11 and satellite compressor 12 is installed. 
Connecting header 13 containing pressure regulating valve 14 is installed 
between suction headers 7 and 11. Pressure regulating valve 14 senses the 
pressure of suction header 11. When the refrigeration demand on evaporator 
9 is low and when the pressure in suction header 11 falls below a designed 
set point, pressure regulator valve 14 opens and the higher pressure 
effluent gaseous refrigerant in suction header 7 flows through connecting 
header 13 to suction header 11. 
In this manner, satellite compressor 12 is always operating at its maximum 
design capacity or, in the event of extremely low refrigeration demand on 
the entire system, carrying the entire pumping requirements of the 
refrigeration system. 
In FIG. 3 an additional embodiment of the invention is shown to advantage. 
In this embodiment provisions are made to transfer a load back and forth 
between suction headers 11 and 7 during time periods when evaporator 9 is 
in a defrost cycle. 
FIG. 3 illustrates one of the more common methods of controlling a defrost 
cycle, and that is to install the defrost suction shutoff valve 30 in 
suction header 11 upstream of compressor 12. During the defrost cycle 
shutoff valve 30 closes, thereby stopping the return flow of effluent 
gaseous refrigerant from evaporator 9. When shutoff valve 30 is closed, 
electrical heaters 35 in the vicinity of evaporator 9 are turned on to 
remove frost from the cooling coils of evaporator 9. 
In installations not using the connecting header 13 and pressure regulating 
valve 14, compressor 12 would necessarily have to be shut off or a load 
by-pass system installed during the time when shutoff valve 30 is closed. 
By using this invention, during periods of time when evaporator 9 is in 
the defrost cycle and shutoff valve 30 is closed, pressure regulating 
valve 14 allows gaseous effluent refrigerant from compressor suction 
header 7 to flow through connecting header 13 to compressor 12, thereby 
allowing compressor 12 to be run at its maximum designed load even while 
evaporator 9 is in defrost. 
At the end of evaporator's 9 defrost cycle, shutoff valve 30 opens and the 
now high pressure, hot gaseous effluent refrigerant in evaporator 9 is 
allowed to pass through evaporator pressure regulating valve 10 and the 
now open shutoff valve 30 into suction header 11. Check pressure relief 
valve 32 is installed in overload connecting header 31 which connects 
suction header 11 to suction header 7. As the shutoff valve 30 opens and 
the hot refrigerant enters into suction header 11 at the end of the 
defrost cycle, pressure in suction header 11 rapidly increases. If the 
pressure in suction header 11 increases above the pressure in suction 
header 7 then check pressure relief valve 32 opens, discharging the high 
pressure effluent refrigerant from suction header 11 to suction header 7. 
By the operation of check pressure relief valve 32, the fluctuations in 
load on compressor 12 are further minimized, thus minimizing the wear and 
tear on compressor 12. 
Having thus described in detail preferred designs which embody the concepts 
and principles of the invention and which accomplish the various objects, 
purposes and aims thereof, it is to be appreciated and will be apparent to 
those skilled in the art that many physical changes could be made in this 
invention without altering the inventive concepts and principles embodied 
therein. Hence, it is intended that the scope of this invention be limited 
only to the extent indicated in the appended claims.