Control System For Low Temperature Refrigeration System

Abstract

In a compression refrigeration cycle employing a mixture of refrigerants and a single compressor wherein low temperatures are reached by one or more vapor-liquid separations in which the expanded and cooled liquid from each separation stage is evaporated and used to condense the vapors from that stage, high starting discharge pressures are avoided and rapid cool down to low temperature is achieved by incorporating a by-pass system for the throttling device feeding the evaporator which is responsive to predetermined operating characteristics of the cycle.

Description

This invention relates to compression refrigeration systems. More particularly, the present invention is concerned with a novel control system for rapidly achieving ultra-low temperatures in compression refrigeration systems.

U.S. Pat. No. 3,203,194 issued Aug. 31, 1965 to A. Fuderer describes a compression refrigeration system involving the use of a mixture of refrigerants circulating in a single refrigeration circuit employing a single compressor. In this system, the compressed mixture of refrigerant gasses undergoes a partial condensation in a first condensation stage and a liquid fraction, rich in the higher boiling refrigerant, is formed. Thereafter, the liquid fraction is separated from the remaining vapors and the vapors are transferred to a second condensation stage where they are condensed using the cold produced by the expansion and evaporation of the liquid component from the first condensation stage and by cold vapors returning to the compressor from the final evaporator. The final refrigeration temperature is achieved by throttling and at least partially evaporating the condensate from the second stage of condensation. The refrigeration circuit is closed by mixing the vapors, and any residual liquid leaving the final evaporator with the vapor formed by throttling the liquid rich in the higher boiling refrigerant and returning the vaporous mixture to the compressor.

Co-pending commonly assigned U.S. Pat. application, Ser. No. 43,108 filed June 3, 1970 and now abandoned, in the name of Dale Missimer and Daniel Lieberman and entitled "Low Temperature Refrigeration System", also describes a compression refrigeration system utilizing a single refrigeration circuit which circulates a mixture of refrigerants. In that application, multiple intermediate cooling stages and multiple vapor liquid separation steps are employed to achieve ultimate low temperatures in the cryogenic range with conventional compression equipment operating at compression ratios of less than 10:1.

While the above-described systems are effective in achieving temperatures approaching the cryogenic range at low pressures and compression ratios when operating normally after having achieved temperatures in the range for which they were designed, a problem has existed with regard to the commercial packaging and design of sealed systems incorporating the above refrigeration concepts. Specifically, it has been found that when these systems operate at higher temperatures, for example during start-up from room temperature or during power off and standby conditions, the refrigerants employed exert substantially higher vapor pressures and there is an increased likelihood of reaching compressor discharge pressures and compression ratios which would damage the compressor. In experimental systems, such a result is avoided by staging the refrigerant charge to the system, thereby limiting the volume of refrigerant flow until such time as lower temperatures are achieved and the system develops the capacity for accepting additional refrigerant. The use of staged charging techniques is obviously impractical in commercial systems which require that the full charge of refrigerant be permanently sealed into the system.

An alternative method for controlling vapor discharge pressures is described in the Fuderer patent and involves the use of a pressure regulated vapor tank on the discharge side of the compressor. This vapor tank serves to take low boiling refrigerant vapors out of circulation when the pressure on the discharge side of the compressor becomes too high and to return some of those vapors to the refrigeration cycle as the pressure is reduced. The use of vapor tanks is unacceptable in many commercial operations since it severely reduces refrigerant flow through the system and thereby greatly increases the time required to reach the optimum temperature range for which the system is designed.

It is an object of the present invention to provide a novel control system for maintaining the compressor discharge pressures and compression ratios of compression refrigeration systems within the safe limits of the compressor being employed.

It is another object of this invention to provide a by-pass system for rapid start-up of compression refrigeration systems employing multiple refrigerants and intermediate cooling stages.

Yet another object of the present invention is to provide a sealed compression refrigeration system containing a full refrigerant charge which is capable of rapidly achieving low temperatures without damage to compression equipment or other system components.

It has now been discovered that excessive compressor discharge pressures and compression ratios can be avoided during periods of start-up and temperature pull-down by permitting uncondensed refrigerant vapors to return to the low pressure side of the refrigeration system without passing through the primary throttling device feeding the low temperature evaporator of the refrigeration cycle. It has also been discovered that the utilization of such by-pass systems results in the more rapid attainment of the normal operating conditions of the system.

While not wishing to be limited to any particular theory, it is presently believed that the by-pass system of the invention avoids excessive discharge pressures and results in more rapid cooling for several reasons. First, the volume on the low pressure or suction side of the refrigeration cycle is greater than the volume on the high pressure or discharge side of the compressor. Thus, eliminating the final throttling step, the vapors are permitted to more easily flow to the low pressure or suction side of the cycle thereby increasing the pressure on the side of the system which has more volume. As a result, a larger amount of the non-condensed low boiling point vapor fractions can be retained in the system at a lower discharge pressure. Second, the by-pass system allows the partially condensed low boiling point vapors on the high pressure side of the cycle to flow at a higher rate than the flow rates which can be achieved when the final throttling device is not by-passed, thereby resulting in quicker cooling of the final evaporator. Moreover, since the larger flow of suction vapors to the compressor can more readily absorb the low boiling point rich vapors from the high pressure side of the system, the system appears to lower the partial vapor pressure of the evaporating condensate employed in the high pressure side of the cascade condenser. The lower partial pressure results in lower evaporating temperatures for the liquid and thus, condensation on the high pressure side of the cascade condenser commences sooner and at lower temperatures.

The invention will be further understood by reference to the accompanying drawings where:

FIG. 1 is a schematic representation of a refrigeration system employing one form of the by-pass system of the invention; and

FIG. 2 is a schematic representation of a modified form of the by-pass system shown in FIG. 1.

Referring specifically to FIG. 1, a mixture of two or more refrigerants having different boiling points is charged into a single closed refrigeration circuit generally identified as 10 through a service valve 12 or other conventional charging means such as a tube, pipe, or the like which will be sealed after the charging step. The amount of each refrigerant charged to the system may be predetermined by weight or, in the case of lower boiling point refrigerants, by allowing each refrigerant gas to circulate through the system until a predetermined partial pressure and a predetermined total pressure for the system are reached.

Subsequent to the charging step the vapors are aspirated by a compressor 14 and pass through line 16 to condenser 18 where partial condensation occurs. Condensation occurs by heat exchange with ambient air forced over condenser pipes 20 by a fan 22, or alternatively, condensation may be carried out using a readily available source of water. The partially condensed refrigerant mixture flows through line 24 to an optional auxiliary condenser 26 where, after the system is in operation, further condensation may occur by heat exchange with the cooler vapors returning to compressor 14 from the final evaporator 28 through line 30. The utilization of an auxiliary condenser is not critical to the refrigeration system but such additional heat exchange at this point and at other similar points throughout the system serve to improve thermodynamic efficiency. The partially condensed refrigerant mixture leaves auxiliary condenser 26 through line 32 and passes to a vapor-liquid separator 34. The liquid at this point is rich in the higher boiling refrigerant or refrigerants of the mixture. Nevertheless, each fraction will contain at least minor amounts of each of the refrigerants in the mixture.

The liquid separated in separator 34 passes through an optional dryer-strainer 36 where particulate matter is filtered from the stream and residual moisture is removed and then through a capillary tube 38 which throttles the liquid, i.e., the pressure on the liquid drops from the discharge pressure to the suction pressure. The use of a capillary tube is not critical and any expansion or throttling device such as a float valve, thermal expansion valve, or other device well known in the refrigeration art may be employed in lieu of capillary tube 38. A portion of capillary tube 38 may be disposed in a heating exchange relationship at point 40 with line 30 through which cold vapors are returning to the compressor from the final evaporator 28. As noted previously, such an arrangement is not critical but serves to further cool the condensate and improve thermodynamic efficiency. In lieu of the heat exchange between capillary tube 38 and line 30 a heat exchanger could be employed at point 40. Such an exchanger would be particularly useful where expansion or throttling was achieved by means of a valve or other throttling device located below strainer 36, rather than the capillary tube as shown.

The throttled, low pressure liquid emanating from capillary 38 is intermixed at point 42 with the cold returning vapors in line 30. This liquid is evaporated in cascade condenser 44 and the absorption of heat causes the condensation of the low-boiling point vapor fraction which leaves separator 34 and passes through line 43 to cascade condenser 44.

When the system is operating normally, the by-pass systems of the invention are closed and the final liquid fraction from cascade condenser 44 passes through a second optional dryer strainer 46, is throttled in capillary tube or other throttling device 48 and, if desired, further cooled by heat exchange with returning vapors in line 30 at exchange point 50. The liquid at evaporator inlet 52 is at the coldest system temperature and essentially at the suction pressure and is partially or completely evaporated in evaporator 28 to achieve the final refrigeration temperature of the system. The refrigeration circuit is closed by returning the vapors and any residual liquid from evaporator 28 through line 30 back to compressor 14, the vapor being mixed with additional refrigerant fractions, e.g., the expanded and evaporated higher boiling liquid fraction emanating from capillary 38, along its path of travel as previously described and the cycle is repeated.

The refrigeration system as described above is substantially identical to the system described in U.S. Pat. No. 3,203,194. It is capable of producing low refrigeration temperatures, when operating, utilizing mixtures of refrigerants having different boiling points with conventional compressors operating at compression ratios of less than 10:1.

FIG. 1 also illustrates one form of the by-pass system of the invention. This system includes valve 54 positioned near the outlet end of optional dryer-strainer 46 and line 56 which runs from valve 54 to the inlet 52 of evaporator 28. During normal operation, valve 54 is closed and the refrigeration cycle operates as previously described. Thus, the condensate leaving cascade condenser 44 passes through expansion device 48 and then through final evaporator 28. However, during start-up of the refrigeration cycle, valve 54 will be open and the vapors and condensate from cascade condenser 44 will pass through valve 54 and line 56 to the inlet end 52 of evaporator 28 thereby by-passing expansion device 48.

The selection of a particular flow control device for the by-pass system is not critical so long as it meets certain functional characteristics. Thus valve 54 may be either a temperature or pressure actuated valve of a construction generally well known to persons skilled in the art.

A typical pressure actuated valve which is useful in the system of the invention is designed to sense the upstream or compressor discharge pressure and open as this pressure increases above a pre-set level. In a typical pressure actuated valve the discharge pressure of the system which acts to open the valve will be counterbalanced by a spring and atmospheric pressure acting in a downward or closing direction. Moreover, the port size of the valve will be selected so as to handle the vapor flow without permitting the outlet pressure from the valve to influence its operating characteristics while being small enough to avoid chattering. In preferred valve arrangements, the valve seat and plug shape may be designed to permit a modulated flow rather than merely operating only at the full open or full closed position although the pressure range between the full open and pre-set shut-off positions will be relatively small, e.g., 10 to 50 psi. While the operating range for the valve is not critical it may function to by-pass the final throttling step, as least in part, until the discharge pressure reaches a level of within 10 to 150 psi of the final discharge pressure of the system.

A typical temperature actuated valve which is useful in the system of the invention is designed to sense the refrigerant temperature leaving the final evaporator and shut-off when this temperature reaches within a pre-determined level, e.g., within 5.degree. to 100.degree. F., preferably 10.degree. to 20.degree. F. of the final evaporator temperature for which the system is designed. The thermal expansion valve will normally be installed in parallel with the capillary tube or other expansion device feeding the final evaporator and will have an oversized port or orifice, e.g., 3 to 5 times the size required for a standard duty thermal expansion valve.

Irrespective of which type of by-pass valve is selected the valve will be set to function at a predetermined level based upon the operating characteristics of the refrigeration system in which it is employed. These characteristics will vary depending upon the nature of the refrigerants employed and the maximum safe operating pressure and compressor ratios for the compressor employed in the system. Typically a pressure actuated type would be wide open at discharge pressures above 300 psig and will close tight below 250 psig with a compressor suitable for air-conditioning with R-22 refrigerant (about 2.5 cfm displacement per h.p.). For an air-conditioning range R-12 compressor having about 4 cfm displacement per h.p., the actuation range would be full open at 200 psig and closed at 160 psig discharge pressure.

Temperature actuated valves, or more suitably temperature-suction pressure thermal expansion valves which sense and control from superheat have a charge such that they remain open until the suction line leaving the evaporator is reduced to within 10.degree. to 20.degree. F. of the final low temperature. Such a valve, when installed in parallel with a thermal expansion valve used for final control of feed to the evaporator, will have a charge in the power assembly having a boiling point about 10.degree. to 30.degree. F. higher than that of the refrigerants exiting from the evaporator, when compared at the same suction pressures. This may involve the use of a mixed or "cross" charge in the by-pass valve power assembly.

FIG. 2 illustrates a slightly modified form of the by-pass system. In this arrangement a pair of complimentary temperature and/or pressure controlled solenoid actuated or modulating type valves 60 and 62 function as a by-pass control system. During normal operation valve 62 will be closed and valve 60 will be open so that the condensate leaving cascade condenser 44 travels through expansion device 48 and final evaporator 28 in the normal fashion. However, during start-up, valve 60 will be closed and valve 62 will remain open until such time as the solenoid operating valves 60 and 62 are energized, depending upon the operating characteristics of the refrigeration cycle as previously discussed at which time valve 62 will close and valve 60 will open. With valve 62 in the open position and valve 60 in the closed position, the condensate leaving cascade condenser 44 will by-pass both expansion device 48 and final evaporator 28 and pass immediately to return line 30 at junction 64. It will be evident to those persons skilled in the art that the solenoid valves may be actuated by the action of a temperature or pressure controller or by a suitable timing circuit. It is also evident that in lieu of two solenoid actuated valves a single three-way valve may be employed.

While the by-pass system of the invention has been schematically represented with respect to the refrigeration cycle illustrated in Fuderer U.S. Pat. No. 3,203,194 it will be obvious to those persons skilled in the art that the system is equally applicable to refrigeration systems utilizing multiple vapor-liquid separators, multiple cascade condensers, and any desired number of refrigerants provided only that the boiling points of the individual refrigerants in the mixture are sufficiently far apart to permit a reasonable separation into distinct liquid and vapor phases at each stage of the system. Such systems and refrigerants are described in co-pending commonly assigned U.S. application, Ser. No. 43,108, filed June 30, 1970 and include refrigerant mixtures wherein each refrigerant in the mixture will differ in boiling point from the next closest boiling refrigerant by 40.degree. to 250.degree. F., preferably 90.degree. F. Any of the well-known refrigerants set forth in standard refrigerant tables and charts may be employed.

The invention will be further understood by reference to the following illustrative example:

A storage freezer having a 4.5 cubic foot workspace volume and a capability of reaching -100.degree. F. was operated for the purpose of determining the effect of the by-pass system of the invention. The freezer was first tested with a refrigeration cycle similar to that illustrated in FIG. 1 but not including the by-pass system. The refrigerants employed were R-503, an azeotrope of trifluoromethane (R-23) and chlorotrifluoromethane (R-13) and R-114, dichlorotetrafluoroethane. The start-up pressure of the system was 350 psig and it took 90 minutes to reach a final evaporator temperature of -100.degree. F. In addition, in repeated tests the system had difficulty in starting up after each "off" cycle and on one occasion the cascade condenser went into self-refrigeration.

The system described above was subsequently operated with a by-pass control system including a thermal expansion valve having a suitably selected port size and a power assembly set up for 55 psig miximum operating pressure, and a "cross-charged" R-13 charge, in parallel with the capillary tube feeding the final evaporator. The maximum starting discharge pressure was normally only 280 psig. The chamber reached -100.degree. F. in 45 minutes.

Similar tests on the same system with a by-pass system comprising a pressure actuated valve resulted in pull-down to -100.degree. F. in 60 minutes at starting discharge pressures up to approximately 300 psig.

Having thus described the general nature as well as preferred embodiments of the invention and true scope will now be pointed out in the appended claims:

Claims

What is claimed is:

1. In a compression process for refrigeration using a mixture of refrigerants having different boiling points which process comprises the steps of compressing a vaporous mixture of refrigerants, partially condensing the compressed refrigerant vapors in a first condensation stage, separating the condensate into first liquid and first vapor phases, throttling said first liquid phase to a lower pressure, evaporating said throttled first liquid phase to absorb heat from and at least partially condense said separated first vapor phase, throttling the at least partially condensed vapors to a lower pressure, at least partially evaporating said throttled condensed vapors in an evaporator to produce the final refrigerating temperatures, mixing the refrigerant leaving said evaporator with the said throttled first liquid phase and closing the circuit by recycling the mixture of refrigerants to the compressing step, the improvement comprising by-passing the final throttling step feeding the evaporator when compression is commenced and continuing by by-pass said throttling step until the said refrigerants reach a predetermined operating condition.

2. The process of claim 1 wherein said final throttling step is by-passed until the discharge pressure is 10 to 150 psi more than the final discharge pressure of said process.

3. The process of claim 1 wherein said final throttling step is by-passed until the temperature of the refrigerant in said evaporator is within 5.degree. to 100.degree. F. of the lowest evaporator temperature for said process.

4. The process of claim 1 including at least one additional intermediate cooling stage subsequent to said first condensation stage and prior to separating the condensate into first liquid and first vapor phases, each said intermediate cooling step comprising the steps of separating the condensate from said first condensation stage into a vapor phase and a liquid phase, throttling the liquid phase to a lower pressure; evaporating the throttled liquid phase to absorb heat from and at least partially condense the separated vapor phase and mixing the evaporated liquid phase with said mixture of refrigerants being recycled from said final evaporator to said compressing step.

5. The process of claim 1 wherein said refrigerant mixture comprises two to six individual refrigerants.

6. The process of claim 1 in which any separated liquid phase is cooled by an exchange of heat with said recycling refrigerant returning to the compressing step prior to said evaporation of said separated and throttled liquid.

7. A compression refrigeration apparatus comprising:

a. compression means for compressing a vaporous mixture of refrigerants having different boiling points;

b. condenser means for at least partially condensing said compressed refrigerant vapors;

c. at least one intermediate cooling stage, said stage comprising vapor-liquid separating means, throttling means for reducing the pressure on the separated liquid phase, heat exchanger means for at least partially condensing said separated vapor and evaporating said throttled liquid phase, circulation means for passing said at least partially condensed vapor to the next succeeding cooling stage, and circulation means for returning said evaporated liquid phase to said compressor means;

d. final throttling means for reducing the pressure of the at least partially condensed vapors from said intermediate cooling stage;

e. evaporator means for at least partially evaporating said throttled condensed vapors to produce the final refrigerating temperature;

f. by-pass means interposed between said intermediate cooling stage and said evaporator said by-pass means having an open position and a closed position and adapted to move between said positions in response to predetermined operating characteristics of the refrigerants; and

g. recycling means for returning the at least partially evaporated refrigerant mixture to said compressor means.

8. The apparatus of claim 7 wherein said by-pass means is a pressure actuated valve.

9. The apparatus of claim 8 wherein said valve is responsive to the discharge pressure of said apparatus.

10. The apparatus of claim 7 wherein said by-pass means is a temperature responsive valve.

11. The apparatus of claim 10 wherein said valve is a thermal expansion valve which is actuated by the evaporator temperature and pressure.

12. The apparatus of claim 7 which includes two to five successive intermediate cooling stages.

13. The apparatus of claim 7 including a series of mixing means for joining the recycling refrigerant mixture emanating from said evaporator means with the throttled separated liquid from each of said successive intermediate cooling stages prior to reaching said compression means.

14. The apparatus of claim 7 further including at least one auxiliary condenser means for permitting heat exchange between said recycling refrigerant mixture and said at least partially condensed vapors.

15. The apparatus of claim 7 further including at least one heat exchange means for permitting an exchange of heat between said separated liquid and said recycling refrigerant mixture.