US3300996A

US3300996A - Variable capacity refrigeration system

Info

Publication number: US3300996A
Application number: US408272A
Authority: US
Inventors: Atwood Theodore
Original assignee: Allied Chemical Corp
Current assignee: Allied Corp
Priority date: 1964-11-02
Filing date: 1964-11-02
Publication date: 1967-01-31
Anticipated expiration: 1984-01-31
Also published as: DE1476691A1; GB1105190A

Description

Jan. 31, 1967 woo I 3,300,996

VARIABLE CAPACITY REFRIGERATION SYSTEM Filed Nov. 2, 1964 2 Sheets-Sheet 1 TEMPERATURE OF LIQUID LEAVING HEAT EXCHANGER F) I20 llO I00 90 80 70 I I l I l IIOO REFRIGERATING CAPACITY IN BTU /HOUR 00 I I I l 20 -IO 0 IO 20 3O 4O 5O 6O 70 I Set.)

TEMPERATURE OF SUCTION VAPOR LEAVING HEAT EXCHANGER (F) FIGURE I.

INVENTOR THEODORE ATWOOD LW WT ATTORNEY Jan. 31, 1967 T. ATwooD 3,300,996

VARIABLE CAPACITY REFRIGERATION SYSTEM Filed Nov. 2, 1964 2 Sheets-Sheet 2 VARIABLE CAPACITY REFRIGERATION SYSTEM.

FIGURE 2 INVENTOR THEODORE ATWOOD BY ALWWW AT ORNEY United States Patent C) 3,300 996 VARIABLE CAPACETY REFRIGERATION SYSTEM Theodore Atwood, Sparta, N.J., assignor to Allied Chemical Corporation, New York, N.Y., a corporation of New York Filed Nov. 2, 1964, Ser. No. 408,272

16 Claims. (Cl. 62-417) This invention relates to a variable capacity refrigeration system and more particularly, to a variable capacity refrigeration system employing a novel method of varying the capacity over a relatively wide range and causing the refrigerating capacity to parallel the load requirement.

In many refrigeration systems, such as air conditioning units, deep freezers and household refrigerators, variable capacity is commonly obtained by means of a simple onolf switch controlled by temperature conditions in some part of the system, ordinarily the evaporator, or in the media being controlled. This occasions frequent starting and stopping of the system which results in considerable wear of motors, switches and controls and consequent shorter equipment life. This arrangement is particularly impractical for large commercial units used for low temperature applications and in which there are constant variations in load.

As a result other means for varying the capacity of a refrigeration system have been developed which include various methods based upon providing compositions containing a plurality of circulating refrigerants having different capacities, with means to separate or vary the composition of the same, which particular mixtures or concentration of refrigerants define a range of specific capacity variation. Such systems employ a variety of conventional separating devices including distillation and sorption apparatus for the purpose of withdrawing one or more refrigerant components, in order to vary the capacity of such systems. As can readily be surmised, this results in an increase in the size and complexity of the refrigerating' unit requiring considerably elaborate and sophisticated equipment, particularly in coordinating such equipment with the cycle of the refrigerating system with which it is to be used. This creates structural problems and increases substantially the cost of design, installation and maintenance of the equipment.

Accordingly, it is an object of this invention to provide a simpler and more economical variable capacity refrigeration system than has heretofore been realized.

It is another object of the invention to provide a novel variable capacity refrigeration system in which the refrigerating capacity parallels the load requirement.

It is a further object of this invention to provide an improved variable capacity refrigeration system utilizing a refrigerant having a wide capacity variation responsive to change in demand on cooling requirements.

It is still yet another object of this invention to provide an improved variable capacity refrigeration system utilizing monochloropentafluoroethane as the refrigerant therein.

Further objects and advantages will become apparent from the following description of the invention in the specification, claims and drawings.

According to the practice of the invention, during operation of the normal refrigeration'cycle, that is, compression of evaporated refrigerant, return of compressed refrigerant through a condensing zone to an expansion valve, followed by evaporation; the evaporated refrigerant, or suction vapor, on its return to the compressor, is divided into two parallel streams. At least one of these streams is conducted through a heat exchange zone in indirect heat exchange relationship with compressed re- Patented Jan.

frigerant, as vapor or as condensate, responsive to a ternperature sensing means indicating increased or decreased demand on load in the evaporating zone. The flow of evaporated refrigerant may be proportioned between the parallel branches to effect the extent of pre-cooling of condensed refrigerant. Thus, as the temperature in the evaporating zone increases, indicating an increase in demand, part or all of total evaporated gas is flowed in indirect heat exchange relationship with condensed refrigerant to compensate for the refrigeration demand change. Conversely, upon temperature decrease in that zone or medium being cooled, the proportion of total flow of evaporated refrigerant subjected to such heat exchange may be reduced in conformity with the reduced demand.

Additionally, in preferred aspects of the invention, the How of compressed refrigerant is regulated in response to changes in demand on refrigeration capacity, so as to divert at least a portion thereof through the heat exchange zones referred to above, either before or subsequent to condensation; diversion being effected of part or all of the compressed refrigerant, depending upon the magnitude and direction of the demand change and upon the proportion of evaporated gas being utilized for the pre-cooling.

In practice of the invention, wide flexibility of refrigeration capacity, available in the evaporating zone, can be achieved with constant uninterrupted delivery from the compressor. The minimum refrigeration is delivered when none of the evaporated gas passes in heat exchange relationship with the condensed refrigerant and all of the v compressed refrigerant vapor passes in indirect heat exchange with the evaporated gas. To meet maximum refrigeration demand, all of the evaporated gas is permitted to pass in indirect heat exchange with all of the condensed refrigerant and none of the compressed refrigerant vapor is permitted to pass in heat exchange relationship with the evaporated gas. Intermediate refrigeration demands, of course, are met by varying the relative extent of heat exchange between evaporated refrigerant and compressed and condensed refrigerant. A satisfactory arrangement, includes delivery of compressed gas to one of the heat exchange zones and compressed condensate to the other. it will be evident to persons skilled in the art that control of the flows of compressed gas or condensate to the one or both heat exchange zones, whichever may be used, permits a nicety of control of refrigeration capacity, precisely satisfying increases or decreases in demand for that capacity.

Although the invention is applicable to systems containing commercial refrigerants, the choice of which, of course, depends upon the refrigeration temperature required; it is especially well adapted and functions most advantageously with refrigerants which are capable of operating satisfactorily over a relatively wide range of evaporating temperatures; for example, a range of F. to 50 F. An excellent refrigerant, particularly well adapted for use in the system of the invention, is monochloropentafluoroethane which is capable of operation with satisfactory efiiciency at evaporating final temperatures within the range of 50 F. to 50 F. and, of course, at corresponding pressures.

FIG. 1 is a curve showing the relationship between refrigerating capacity, and temperature of liquid and vapor leaving the heat exchanger.

FIG. 2 is a schematic showing of the refrigeration system including controls.

Making reference to FIG. 1, there is illustrated a curve indicating the range of capacity of GENETRON-IIS. The data is based on a condensing temperature of F. and an evaporating temperature of 20 F. The refrigerating capacity in B.t.u./hr. (vertical scale) is plotted against the temperature of the suction vapor leaving the heat exchanger in degrees F. (lower horizontal scale). The corresponding temperature of the liquid leaving the heat exchanger, which liquid is in indirect heat exchange relationship with the suction vapor, may be read from the upper horizontal scale in F. It can readily be seen from the curve in FIG. 1, that the curve" is very nearly a straight line forming an angle of about 45 with the horizontal, indicating a wide, fairly constant capacity variation responsive to the change in demand on GENE- TRON-llS. To illustrate, it can be seen that an increase or decrease of 10 F. of the suction vapor will result in a corresponding increase or decrease of about 50-70 B.t.u./hr./lb. in refrigerating capacity.

Referring to FIG. 2, there is illustrated diagrammatically a typical system according to the practice of the invention showing refrigerant compressor 1 and thermal expansion valve 5, interconnected with compressed refrigerant line 7, which line contains condenser 2 and liquid receiver 3. Expanded refrigerant line 9, containing evaporator coil 4, interconnects expansion valve 5 and the suction side of compressor 1. Between evaporator coil 4 and the compressor, expanded refrigerant line 9 is divided into branched

parallel segments

19 and 20.

Branches

19 and 20 each contain a heat exchange zone indicated generally at 6 and 21, respectively, comprising in the latter case,

as shown in the drawing, a coil disposed in by-p-ass 7A of line 7, disposed in heat exchange relationship with evaporated refrigerant branch 20. Valves 10, disposed in line 7 and 7A, preferably connect with temperature sensing device 13, hereinafter described in more detail by means of actuating line 23 and regulate the quantity of compressed refrigerant gas flowing through heat exchange zone 21 and of compressed refrigerant gas flowing directly to condenser 2.

Similarly, by-pass 7B for line 7 contains a coil 24, disposed in heat exchange relationship with evaporated refrigerant branch 19 and in the same manner as described above, the How of condensed compressed refrigerant is proportioned between by-pass line 713 and line 7, by operation of valves 8 located in line 7 and by-pass line 7B, preferably controlled by temperature sensing device 13, by means of actuating lines 25. Alternatively, valves 8, or valves 10 for that matter, may be controlled manually. Hence, valves 8 regulate the quantity of compressed refriger ant liquid flowing through heat exchange zone 6 and of compressed refrigerant liquid flowing directly to evaporating coil 4 through thermal expansion valve 5.

Valves

17 and 18 serve to regulate flow of evaporated refrigerant as between the divided parallel branches, 19 and 20, of the evaporated (expanded) refrigerant line 9 and can conveniently be valves which shut off flow to one branch, while permitting total flow through the other branch, such as solenoid valves. Alternatively and preferably, a proportioning valve 16 is provided which permits flow of evaporated refrigerant to take place to a lesser or greater extent in either

branch

19 or 20, or both, thus providing a modulating type flow a s,between

segments

19 and 20. I r

Valve pairs

8 and 10 may be manipulated together with valve 16 or

valves

17 and 18 so that evaporated refrigerant fiow can take place to the compressor, through either of

segments

19 and 20 of expanded refrigerant line, or through both in any proportion, in (or out of) heat exchange relationship with compressed refrigerant passing through heat exchange zones 6 and 21. Thus, when sensing unit 13 responds to increased temperature in cooling space 14 surrounding evaporator 4, indicating an increased load; it causes control unit 12 to operate

valve pairs

8 and 10 and/or valve 16 or, in place of valve 16,

valves

17 and 18, if such are provided, by means of the

appropriate actuating lines

15, 23 and 25, which connect the various valves with control unit 12, to direct a greater proportional flow of expanded refrigerant gas through heat exchange zone 6. When the cool expanded refrigerant vapor in segment 19 passes in indirect heat exchange relationship with the warmer compressed condensate, conducted via by-pass 7B through heat exchange zone 6, the compressed condensate will be cooled before expansion and its ultimate capacity for cooling during expansion will be increased in compensation of the increased demand. If the cool expanded refrigerant vapor in segment 20 is directed to pass in indirect heat exchange relationship with the very warm refrigerant discharge vapor from the outlet of compressor I, conducted via by-pass 7A through heat exchange zone 21, the warm discharge vapor will be pre-cooled somewhat before condensation and subsequent evaporation, however, the relatively large offsetting effect of the increased volume of further expanded refrigerant vapor in the compressor, will cause less refrigerant to be pumped through the system and thus result in a net loss in refrigerating capacity. Although the temperature and hence volume of the evaporated gas will also be increased after indirect heat exchange with the compressed condensate; the effect of change in evaporated gas volume is more than offset by the increase in evaporator capacity resulting from lower condensate temperature. It can be seen that the greatest extremes of system capacity are obtainable when one heat exchange zone is employed, to the exclusion of the other. The use of heatexchange zone 6, alone, affords the maximum system capacity because it provides the maximum amount of regenerative heat exchange between the evaporated suction vapor and the compressed refrigrant. A falling load, signaled by a lower cooling space temperature, would result in switching or modulating the evaporated refrigerant flow away from heatexchange zone 6, thus avoiding precooling of compressed condensate and thus minimizing the refrigerants capacity for cooling, thereby decreasing the overall capacity of the unit for cooling. The greater the amount of evaporated refrigerant, thus diverted away from heat exchange zone 6, if routed through heat exchange 21; the lower will be the ultimate capacity of the system for cooling due to the increase of expansion of evaporated gas in the compressor, as explained heretofore.

Heat exchange zone 21, when used alone, or in conjunction with heat exchange zone 6, serves a double purpose. In addition to its function for lowering capacity, described above, heat supplied from the warm compressed refrigerant vapor contained in compressed refrigerant vapor by-pass conduit 7A, to the refrigerant flowing in the compressor via expanded refrigerant segment 20, tends to keep the expanded refrigerant vapor in vaporous state and thus serves to protect the compressor against liquid slugging. Accordingly, when both heat exchange zones are employed, and increased proportion of evaporated (or expanded) refrigerant, responsive to increased load, is passed through heat exchange zone 6; heat exchange zone 21 serves as a slugging preventative means. However, when neither heat exchange zone 6 nor heat exchange zone 21 is employed; it is desirable to provide superheat to the refrigerant flowing to the compressor, to prevent liquid slugging in the compressor. Any available, auxiliary heat source may be used for this purpose. Examples of convenient heat sources include: ambient air, condensing water, electric heater, etc. Use of slugging preventative means will, of course, result in som but insignificant sacrifice of ultimate refrigerating capacity.

Control unit 12 may operate electrically or by pressure and accordingly, actuating

lines

15, 23 and 25 can be electric or pressure lines depending on choice of control. In a given case, depending on design requirements, any of the actuating lines may be deactivated and the valve pairs 8 and 10, and valve 16 or

valves

17 and 18 preset to permit a predetermined refrigerant fiow with respect to heat exchange zones 6 and 21. Sensing unit 13 may be any one of a number of conventional thermostatic control dcvices responsive to temperature change,

such as a sensing bulb similar to thermal valve sensing bulb ll, a bimetallic strip, or the like.

In an illustrative cycle, according to the invention, compressor 1 compresses and discharges used, relatively warm, gaseous refrigerant via compress-ed refrigerant conduit '7, directly to. condenser 2, by-passing heat exchange zone 21. Condenser 2 serves to cool the gas and cause it to give up its latent heat of evaporation under high pressure. Under such conditions, the gas is converted to a liquid, which is collected in receiver 3, which serves to store said condensate until needed. The compressed condensate then flows to evaporating unit 4, which is at reduced pressure, through valve pair 8, either directly through conduit 7 or indirectly through conduit 7B, which latter route passes through heat exchange zone 6. In the evaporator the condensate evaporates, absorbing heat from surrounding medium 14 when converting to gaseous form, thus cooling said medium. The flow of compressed condensate to evaporator 4 is regulated by thermal expansion valve 5. This valve is connected to a temperature sensitive bulb 11 located at the outlet of the cooling unit. When this bulb is warmed, expansion of the fluid in it causes a diaphragm or bellows to which it is connected (not shown in the drawing) to expand, which in turn tends to open an expansion valve needle situated in valve 5 by means of electric or pressure actuating line 26, thus enabling the cooling coils of evaporator t to fill more completely and therefore cool more etficiently. Upon cooling of the cooling coils, temperature in expanded refrigerant line to which the thermal valve sensing bulb 11 is attached, drops and the pressure in said bulb decreases due to the refrigerant relieving pressure in the bellows, thus causing the expansion valve to be shut off. This, of course, reduces flow of refrigerant cooling fluid to the cooling coils when cooling fluid is not needed in such high quantities. The above-described control, which regulates flow of cooling fluid to the cooling coils of evaporator 4, operates independently of sensing device 13 and control unit 12 according to the practice of the invention. These mechanisms serve to control the flow of evaporated refrigerant, after it leaves the cooling coils in evaporator 4, through expanded refrigerant line or conduit 9 which is split or branched at the point indicated by 16, so that evaporated refrigerant can flow to the compressor via alternate routes. One route carries the evaporated refrigerant through segment 19 and in heat exchange relationship with coil 24 of heat exchange zone 6. The other route carries the evaporated refrigerant directly to the compressor, by-passing heat exchange zone 6. The latter route may be optionally provided with an available auxiliary heat source, such as described heretofore, to assist in preventing liquid slugging in the compressor. Segments l9 and of expanded refrigerant line 9 may, of course, connect with the compressor by means of separate suction inlets, however, as shown in the drawing, it is more expedient to reunite said segments at a point beyond either of heat exchange zones 6 or 21.

According to this embodiment, when temperature rises in cooling space 14, indicating a heavier load, sensing device 13 is triggered which signals control unit 12, which in turn operates proportioning valve 16 to permit evaporated refrigerant to flow through segment 19 and heat exchange zone 6 and also valve pair 8 to permit compressed condensate to flow through condiut 7B and heat exchange zone 6. At the same time control unit 12 opcrates valve pair Hi to direct compressed refrigerant away from by-pass conduit 7A and heat exchange zone 21. The relatively cool evaporated refrigerant in segment 19 absorbs heat from the relatively warm compressed condensate in coil 24-. The compressed condensate, thus pre cooled, before flowing through cooling coils in evaporator 4, has increased capacity for absorbing heat from cooling space lid. When temperature in cooling space 14 falls, indicating a lighter load, the above-described sequence of events is triggered except that valve 16, is operated so as to cause evaporated refrigerant to flow through segment 20, by-passing heat exchange zone 6 and valve pair 8 is operated so as to direct compressed condensate directly to evaporator 4, by-passing conduit 78 and heat exchange zone 6. Thus, the pre-cooling effect of compressed condensate, feeding to the evaporator is diminished, resulting in a relative lowering of the refrigeratingcapacity of the system. The overall result of the above-described arrangement is that the refrigeration capacity of the system at a given time parallel the load requirement.

In general, in order for a substance to be useful as a refrigerant, it must have a low boiling point and in passing from a liquid to a gas, must absorb a high quantity of heat per pound. It is also desirable that the specific volume of the gas be small in order to minimize equipment size. Ideally, refrigerants should be .nonflammable, stable non-toxic, non-corrosive, non-explosive and non-injurious to lubricants used in the system. In the system according to the invention, it is desirable that the refrigerant employed have an additional thermodynamic property: viz. that the ratio of its specific heat of the vapor to its latent heat of evaporation be high. The refrigerant should also be capable of operation over a wide range of evaporator temperatures. It has been found that GENETRON-IIS (chloropentafluoroethane) combines all of the above-described properties. It has a boiling point of 37.7 F., is highly stable, inert, nonflam mable, has no evident toxicity and is non-corrosive. Moreover, GENETRON-llS exhibits a wide capacity variation responsive to change in demand on cooling requirements, which capacity can readily be varied by use of heat exchange means such as employed herein.

Other refrigerants may be employed which have the above-described combination of properties. A particular class of compounds, within which such refrigerants may be found, are the halogenated hydrocarbons containing one or more fluorine and/or chlorine atoms. Representative compounds from this class, which would be useful in the practice of theinvention, are octafluorocyclobutane, perfluoropropane and chloroperfiuoropropane. Useful refrigerants from without this class of chemical compounds, sulfur hexafluoride being exemplary, may be readily surmised by practitioners of the invention by study of thermodynamic data without departing from the spirit of the invention. In a like manner, it is obvious that variations and embodiments may be developed which do not depart from the spirit of the invention and accordingly applicant intends to be limited only by the reasonable scope of the appended claims.

The superior performance of GENETRON-llS, or other refrigerants with similar properties, in the subject system, as compared with ordinary refrigerants, can read ily be appreciated by comparing Examples 1 and 2 with Examples 3-6, as follows:

EXAMPLE 1 The system is as above-described in the illustrative cycle, excepting that all evaporated refrigerant (suction gas) is directed through segment 19, so as to pass in indirect heat exchange relationship with refrigerant condensate flowing through coil 24 of heat exchange zone 6. (See FIG. 2.) The refrigerant employed is GENTRON- (chloropentafluoroethane) on a F. condensing -20 F. evaporating cycle. Heat exchange zone 6 is so designed as to be capable of an approach of 10 F. between leaving liquid and vapor temperature. This system equalizes with a suction gas temperature of 62 F. and a refrigerant condensate temperature of 72 F. Under these conditions, the specific volume of the suction vapor is 1.56 cu. ft./lb. The net refrigerant effect per pound of refrigerant is 29.1 B.t.u. For a compressor pumping at 1 cu. ft/minute, this results in a mass flow rate of 38.4 lbs/hr. and a cooling capacity of 1120 B.t.u./hr. at the load. This figure represents the maximum capacity attainable under the recited conditions.

EXAMPLE 2 The system is identical to that described in Example 1 excepting that all evaporated refrigerant (suction gas) is directed through segment 20, so as to pass in indirect heat exchange relationship with compressed refrigerant flowing through coil 22 of heat exchange zone 21. (See FIG. 2.) The relatively warm compressed refrigerant gas gives up heat to the relatively cool suction vapor. (This may either be sensible heat or latent heat depending on the amount of super-heat present.) Heat exchange zone 21 is sized to maintain the same suction gas temperature as before (+62 F.) and the specific volume of the suction vapor accordingly remains as before, viz. 1.56 cu. ft./1b. Although the compressor pumps, as before, at a mass flow rate of 38.4 lbs./hr.; the net refrigerating effect is now only 15.4 B.t.u./hr. and so the cooling capacity of the unit is only 592 B.t.u./hr. This is a capacity reduction of 47% from the capacity obtained in Example 1 and represents the minimum capacity attainable under the recited conditions.

Under intermediate loads, sensing element 13 acts to proportion the flow between heat exchange zones 6 and 21 (or between heat exchange zone 6 or no heat exchange zone), thereby either increasing or decreasing the effective cooling, until a balance point is reached. An infinite number of intermediate steps are possible between the maximum and minimum load conditions.

EXAMPLE 3 The system is identical to that of Example 1, excepting that monochlorodifiuoromethane is used as refrigerant. With the 120 F. condensing/20 F. evaporating cycle, the 10 F. approach in the heat exchange zone results in suction gas at a temperature of 72 F. In order to obtain a cooling capacity of 1120 B.t.u./hr. with heat exchange zone 6 in use, a compressor of .71 cu. ft./minute pumping capacity is required.

EXAMPLE 4 This system is identical to that of Example 3, excepting that all evaporated refrigerant (suction gas) is directed through heat exchange zone 21. System capacity is reduced to 913 B.t.u./hr., representing only an 18% reduction from maximum capacity.

EXAMPLE 5 This system is identical to that of Example 3, excepting that dichlorodilluoromethane is used as refrigerant. With the 120 F. condensing/20 F. evaporating cycle, the F. approach in the heat exchange zone results in suction gas at a temperature of 62 F. In order to obtain a cooling capacity of 1120 B.t.u./hr. with heat exchange zone 6 in use, a compressor of 1.09 cu. ft./minute pumping capacity is required.

EXAMPLE 6 This system is identical to that of Example 5, excepting that all evaporated refrigerant (suction gas) is directed through heat exchange zone 21. System capacity is reduced to 865 B.t.u./hr., representing only a 23% reduction from maximum capacity.

What is claimed is:

1. A variable capacity refrigeration system comprising in combination:

(a) a compressor,

(b) an evaporator,

(c) a compressed refrigerant conduit, including a condenser, a compressed refrigerant vapor portion and a compressed refrigerant condensate portion, connecting the compressor with the evaporator,

(d) an evaporated refrigerant conduit connecting the evaporator with the compressor,

(e) said evaporated refrigerant conduit being divided into two parallel branches,

(f) at least one of the compressed refrigerant vapor and compressed refrigerant condensate portions of the compressed refrigerant conduit being disposed in heat exchange relationship with a parallel branch of the evaporated refrigerant conduit, provided that if both the compressed refrigerant vapor portion of the compressed refrigerant conduit and the compressed refrigerant condensate portion of the compressed refrigerant conduit are so disposed, such portions of the compressed refrigerant conduit are disposed in heat exchange relationship with separate parallel branches of the evaporated refrigerant conduit,

(g) valve means disposed and arranged to regulate the proportion of evaporated refrigerant fiow as between said parallel branches,

(h) the valve means being responsive to temperature changes in a cooling space surrounding the evaporator,

the compressor, condenser and evaporator being connected in circuitous flow relationship.

2. A variable capacity refrigeration system according to claim 1 in which one portion of the compressed refrigerant conduit is disposed in heat exchange relationship with one of the parallel branches of the evaporated refrigerant conduit, said portion of the compressed refrigerant conduit being the compressed refrigerant vapor portion.

3. A variable capacity refrigeration system according to claim 1 in which one portion of the compressed refrigerant conduit is disposed in heat exchange relationship with one of the parallel branches of the evaporated refrigerant conduit, said portion of the compressed refrigerant conduit being the compressed refrigerant condensate portion.

4. A variable capacity refrigeration system comprising in combination:

(a) a compressor,

(b) an evaporator,

(c) said evaporated refrigerant conduit being divided into two parallel branches,

(f) the compressed refrigerant condensate portion of the compressed refrigerant conduit being disposed in heat exchange relationship with one of the parallel branches of the evaporated refrigerant conduit,

(g) valve means disposed and arranged to regulate the proportion of evaporated refrigerant flow as between said parallel branches,

(h) the valve means being responsive to temperature changes in a cooling space surrounding the evaporator, a higher temperature therein signaling a heavier load and causing a greater proportional flow of refrigerant through the branch of the evaporated refrigerant conduit disposed in heat exchange relationship with the compressed refrigerant condensate portion of the compressed refrigerant conduit, 9. lower temperature therein signaling a lighter load and causing a greater proportional fiow of refrigerant through the other branch of the evaporated refrigen ant conduit,

5. A variable capacity refrigeration system according to claim 4 in which solenoid valves are employed to direct refrigerant flow as between the parallel branches of the evaporated refrigerant conduit.

6. A variable capacity refrigeration system according to claim 4 in which a proportioning valve is employed to direct the refrigerant flow as between the parallel branches 9 of the evaporated refrigerant conduit, thus providing a modulating type control.

7.' -A variable capacity refrigeration system comprising in combination:

(a) a compressor,

(b) an evaporator,

(d) an evaporated refrigerant conduit'connecting the evaporator with the compressor,

(f) at least one of the compressed refrigerant vapor and compressed refrigerant condensate portions .of the compressed refrigerant conduit being disposed in heat exchange relationship with a parallel branch of the evaporated refrigerant conduit, provided that if both the compressed refrigerant vapor portion of the compressed refrigerant conduit and the compressed refrigerant condensate portion of the compressed refrigerant conduit are so disposed, such portions of the compressed refrigerant conduit are disposed in heat exchange relationship with separate parallel branches of the evaporated refrigerant conduit,

(h) the valve means being responsive to temperature changes in a cooling space surrounding the evaporator and (i) means to circulate a refrigerant in closed circuitous flow relationship from compressor to condenser to evaporator to compressor.

8. The variable capacity refrigeration system of claim 7 wherein the circulating refrigerant is a member selected form the group consisting of monochloropentafiuoroethane, octafluorocyclobutane, perfluoropropane, chloroperfiuoropropane and sulfur hexafiuoride.

9. The variable capacity refrigeration system of claim 7 wherein the refrigerant is a halogenated hydrocarbon containing at least one halogen atom selected from the group consisting of chlorine and fluorine.

10. The variable capacity refrigeration system of claim 9 wherein the refrigerant is monochloropentafluoroethane.

11. A variable capacity refrigeration system comprising in combination:

(a) a compressor,

(b) an evaporator,

(c) a compressed refrigerant conduit, including a condenser, a compressed refrigerant vapor portion and a compressed refrigerant condensate portion, connecting the compressor with the vaporator,

(f) a portion of the compressed refrigerant conduit being disposed in heat exchange relationship with one of the parallel branches of the evaporated refrigerant conduit, said portion of the compressed refrigerant conduit being the compressed refrigerant condensate portion,

(g) proporti-oning valve means disposed and arranged to regulate the proportionate flow of refrigerant as between the parallel branches,

(h) the proportioning valve mean being responsive to temperature changes in a cooling space surrounding the evaporator, a higher temperature therein signaling a heavier load and causing a greater proportional flow of refrigerant through the branch of the evaporated refrigerant conduit disposed in heat exchange relationship with the compressed refrigerant condensate portion of the compressed refrigerant conduit, a lower temperature therein signaling a lighter load and causing a greater proportional flow of refrigerant through the other branch of the evaporated refrigerant,

(i) means to circulate monochloro-pentafiuoroethane in closed circuitous flow relationship from compressor to condenser to evaporator to compressor.

12. A variable capacity refrigeration system according to claim 11 in which solenoid valves are substituted for the proportioning valve which serves to distribute the refrigerant flow as between the parallel branches of the evaporated refrigerant conduit.

13. A variable capacity refrigeration system according to claim 11 in which heat is supplied to the other of the parallel branches of the evaporated refrigerant conduit.

14. The method of obtaining variable capacity in a refrigerant system which comprises:

(a) circulating a refrigerant possessing a wide capacity variation responsive to change in demand on cooling requirements in closed circuitous flow through a refrigeration system comprising a compressor, an evaporator, a compressed refrigerant conduit including a condenser, a compressed refrigerant vapor portion and a compressed refrigerant condensate portion, connecting the compressor with the evaporator, and an evaporated refrigerant conduit connecting the evaporator with the compressor in split parallel flow relationship,

(b) causing the evaporated refrigerant to flow to a greater or lesser extent through the split segments of the evaporated refrigerant conduit, one of which segments is disposed in heat exchange relationship with the condensate portion of the compressed refrigerant conduit,

(0) regulating the amount of How as between the split segments responsive to temperature changes in the cooling space surrounding the evaporator,

(d) permitting a greater proportional flow through that segment disposed in heat exchange relationship with the compressed refrigerant condensate portion of the compressed refrigerant conduit when temperatures in the cooling space surrounding the evaporator register higher, thus signaling a heavier load and (e) causing a lesser proportional flow through that segment when lower temperatures are registered in the cooling space surrounding the evaporator, signaling a decreased load and thus enabling the refrigeration capacity to parallel the load requirement.

15. The method of obtaining variable capacity in a refrigeration system according to. claim 14, in which the temperature of evaporated refrigerant in that segment not in heat exchange relationship with the compressed refrigerant condensate portion of the compressed refrigerant conduit is raised by heat exchange means before passing into the compressor.

16. The method of obtaining variable capacity in a refrigerant system which comprises:

(b) causing the evaporated refrigerant to flow to a greater or lesser extent through the split segments of the evaporated refrigerant conduit, one of which segments is disposed in heat exchange relationship with the compressed refrigerant condensate portion of the compressed refrigerant conduit,

(c) raising the temperature of evaporated refrigerant 1 1 in the other segment of the evaporated refrigerant conduit, before the evaporated refrigerant contained therein passes into the compressor,

(d) regulating the amount of flow of evaporated refrigerant as between the split segments responsive to temperature changes in the cooling space surrounding the evaporator,

(e) permitting a greater proportional flow through that segment in heat exchange relationship with the compressed refrigerant condensate portion of the compressed refrigerant conduit, when temperatures in the cooling space surrounding the evaporator register higher, thus signaling a heavier load and 1 2 (f) causing a lesser proportional flow through that segment when lower temperatures are registered in the cooling space surrounding the evaporator, signaling a decreased load, thus enabling the refrigeration capacity to parallel the load requirement.

References Cited by the Examiner UNITED STATES PATENTS 2,022,774 12/1935 Kucher 62-496 2,223,900 12/1940 Pownall 62-513 2,691,273 10/1954 Kramer 62-513 X MEYER PERLIN, Primary Examiner.

Disclaimer 3,300,996.-The0cl0re Atwood, Sparta, NJ. VARIABLE CAPACITY RE- FRIGERATION SYSTEM. Patent dated Jan. 31, 1967. Disclaimer filed June 20, 1968, by the assignee, Allied Chemical Corporation. Hereby enters this disclaimer to

claims

1, 2, 7 8, 9 and 10 of said patent.

[Oyfim'al Gazette November 19, 1.968.]

Claims

14. THE METHOD OF OBTAINING VARIABLE CAPACITY IN A REFRIGERANT SYSTEM WHICH COMPRISES: (A) CIRCULATING A REFRIGERANT POSSESSING A WIDE CAPACITY VARIATION RESPONSIVE TO CHANGE IN DEMANT ON COOLING REQUIREMENTS IN CLOSED CIRCUITOUS FLOW THROUGH A REFRIGERATION SYSTEM COMPRISING A COMPRESSOR, AN EVAPORATOR, A COMPRESSED REFRIGERANT CONDUIT INCLUDING A CONDENSER, A COMPRESSED REFRIGERANT VAPOR PORTION AND A COMPRESSED REFRIGERANT CONDENSATE PORTION, CONNECTING THE COMPRESSOR WITH THE EVAPORATOR, AND AN EVAPORATED REFRIGERANT CONDUIT CONNECTING THE EVAPORATOR WITH THE COMPRESSOR IN SPLIT PARALLEL FLOW RELATIONSHIP, (B) CAUSING THE EVAPORATED REFRIGERANT TO FLOW TO A GREATER OR LESSEAR EXTEND THROUGH THE SPLIT SEGMENTS OF THE EVAPORATED REFRIGERANT CONDUIT, ONE OF WHICH SEGMENTS IS DISPOSED IN HEAT EXCHANGE RELATIONSHIP WITH THE CONDENSATE PORTION OF THE COMPRESSED REFRIGERANT CONDUIT, (C) REGULATING THE AMOUNT OF FLOW AS BETWEEN THE SPLIT SEGMENTS RESPONSIVE TO TEMPERATURE CHANGES IN THE COOLING SPACE SURROUNDING THE EVAPORATOR, (D) PERMITTING A GREATER PROPORTIONAL FLOW THROUGH THAT SEGMENT DISPOSED IN HEAT EXCHANGE RELATIONSHIP WITH THE COMPRESSED REFRIGERANT CONDENSATE PORTION OF THE COMPRESSED REFRIGERANT CONDUIT WHEN TEMPERATURES IN THE COOLING SPACE SURROUNDING THE EVAPORATOR REGISTER HIGHER, THUS SIGNALING A HEAVIER LOAD AND (E) CAUSING A LESSER PROPORTIONAL FLOW THROUGH THAT SEGMENT WHEN LOWER TEMPERATURES ARE REGISTERED IN THE COOLING SPACE SURROUNDING THE EVAPORATOR, SIGNALING A DECREASED LOAD AND THUS ENABLING THE REFRIGERATION CAPACITY TO PARALLEL THE LOAD REQUIREMENT.