LOW-TEMPERATURE REFRIGERATION SYSTEM

 

George Brown, Lewis Refrigeration Co. Washington,
George Briley, Lewis Refrigeration Co.Texas.

This article generally covers low-temperature systems, including single-stage, compound or direct staging, and cascade staging. It reviews the fundamentals of refrigeration to give a more in-depth understanding of why one type of system may be selected for one installation and another.

In cryogenic engineering, temperatures closely approaching absolute zero aren't uncommon. However, such applications usually start at about -300 degrees F. and go downward. Referred to as ultra-low temperatures. Low-temperature refrigeration is defined as those applications requiring evaporator temperatures ranging from -20 to -250 degrees F. The vast majority of installations would be at temperatures above -125 degrees F. However, systems down to -250 are covered.

Mechanical refrigeration is the controlled heat removal (British Thermal Units), with temperatures maintained below atmospheric conditions. Removing heat by mechanical refrigeration requires the expenditure of mechanical or heat energy. Vapor-compression systems are one form of mechanical refrigeration. Such systems all have the characteristic of recalculating the refrigerant in a closed circuit. Such systems must reject the extracted heat to some form of sink, usually atmospheric air or surface water.

The refrigeration unit is a ton of refrigeration, usually abbreviated T.R., a rate unit. A ton of refrigeration is the heat removal necessary to freeze 2000# of 32-degree-F water in 24 hours. Since the latent heat of ice fusion is 144 BTU/lb, this means 288,000 BTU/day, 12,000 BTU/day, or 200 BTU/min.

Reviewing the basic principles of a single-stage vapor-compression refrigeration plant, Figure #1 shows the refrigerant vapor compressed by a piston compressor. Such a plant consists of four major equipment pieces: compressor, condenser, evaporator, and expansion valve. The refrigerant absorbs heat from a heat transfer medium that is cooled in the evaporator and then vaporizes. A compressor removes the vapors from the evaporator. It increases their pressure and temperature level and discharges the vapors into a condenser, along with the heat of compression. The heat of condensation is removed in the condenser when the vapor is liquefied by transferring it to a colder fluid utilizing a heat exchange surface. The liquefied refrigerant then flows through an expansion device, where its pressure and temperature are reduced to who's in the evaporator.

Using an "ideal "cycle, heat and flow balances can be analyzed using the Mollier Chart for the refrigerant. Figure # 2 shows a single-stage vapor compression cycle's Mollier or pressure-enthalpy diagram. The illustrations presented here deal with ammonia; however, the principles apply equally well to other refrigerants. This study of low-temperature refrigeration describes many refrigerants. In each case, the properties of the refrigerant must be considered.

Pressure-enthalpy diagrams are commonly called "P—H" diagrams. The development of Mollier Charts results from thermodynamic studies of the general energy equation as applied to steady-flow processes. Enthalpy measures the total heat or heat content at the various pressure points of the refrigerant's liquid, vapor, or mixture.

When work is done, or heat is transferred, the refrigerant changes its enthalpy. The enthalpy is read on the abscissa, in BTU per pound of refrigerant, and absolute pressure, in PSI, is read on the ordinate. The heavy curve to the left represents the saturated liquid line, and the heavy curve to the right represents the saturated vapor line. The refrigerant will be in both liquid and vapor form between the curves, whereas all points to the right of the saturated vapor line represent superheated gas. All points to the left of the saturated liquid curve represent a subcooled liquid condition.

Referring to this "ideal" cycle on the P-H diagram, the expansion of the high-pressure liquid to evaporator pressure is a constant–enthalpy process from D to A. As the liquid refrigerant expands to evaporator pressure, some flashes and cools the remainder to evaporator temperature. Assuming no friction, the process in the liquid-vapor mixture region is shown at constant pressure and temperature, 39 PSI and + 10 degrees F in this instance. At point B, the suction gas enters the compressor under saturated vapor conditions. Compression occurs from points B to C along a constant entropy line, assuming no friction or heat transfer. An "ideal" cycle; however, in actual compression, friction is involved, and so is heat transfer.

From point C, the superheated discharge gas is rejected by the condenser. After the gas is cooled to saturated conditions, liquid condensation occurs through the liquid-vapor region over to the liquid line at point D. The P-H diagram fully describes the refrigeration process. In this figure, the evaporator temperature is +10 degrees F, and the condensing temperature is 96 degrees F. The diagram yields this information:

1. The refrigerating effect per pound of ammonia circulated through the evaporator equals the enthalpy difference between points B and A. In this case, 615-149-466 BTU/lb. of ammonia.
2. The pounds of ammonia circulated per minute per ton equals 200 (200 BTU/lb = T.R.) divided by the refrigerating effect, or 200 + 466 = 0.43 lbs. of ammonia per ton.

3. The ideal work of compression equals the difference in enthalpies at points C and B. In this case, 718-615-103 BTU/lb of ammonia. This perfect compression work is 103 × 0.43 lb/NH3/ TR=44 BTU/TR.

4. The theoretical horsepower per ton is found by converting work of compression per ton into horsepower utilizing the conversion factor 42.42 BTU/min = 1 H.P. In this case, our theoretical H. P. would be 44+42.42 or 1.04 H. P. per ton.

5. The compression ratio equals the absolute discharge pressure divided by the absolute suction pressure. For the conditions shown, the C.R. would be 196+39= 5.0.

The P–H diagram discloses other essential considerations in studying basic refrigeration. The heat rejected by the condenser equals the enthalpy at C minus the enthalpy at D, or 569 BTU/lb. of ammonia. This amounts to about 245 BTU/TR. The horsepower and heat rejection in an actual plant are more significant than shown on the ideal cycle because of compressor inefficiencies and design. Required compressor BHP per T.R. increases significantly as evaporator temperature drops. Figure 3 compares the condenser hear rejection, tube surface, and water requirements at various evaporator temperatures.

For the same single-stage vapor compression system at an evaporator temperature of -20 degrees F, the P – H diagram would look similar to Figure #4. In this instance, the critical considerations, as analyzed in the previous diagram, are shown in the tabulation. The differences between ideal cycles for +10 degree F evaporator and -20 degree F conditions are apparent. The compression ratios increase as the evaporator temperature decreases, and the theoretical H.P. /TR increases. These figures are determined by comparing ideal cycles on a Mollier Chart. Unfortunately, the compressor inefficiencies mentioned earlier cause the cycle figures to be slightly misleading in an actual plant. Higher compression ratios mean higher discharge temperatures and lower volumetric efficiencies.

Figure # 5 shows a typical piston and cylinder in a reciprocating compressor. This type of compressor has one or more pistons to compress the gas. As the piston makes its stroke toward the top dead center, some clearance will always exist between the end of the stroke and the top head. Factors such as manufacturing tolerances, valve design, and others determine the exact amount. A clearance of around 5% is reasonable in most modern compressors. A theoretical indicator card for the ideal cycle shows pressure versus volume within the cylinder. As the piston reaches the end of its compression stroke, the clearance prevents all the gas from being ejected from the cylinder. The figures show that the remaining gas is under discharge pressure conditions, which is 1496 PSI. The pressure within the cylinder would be a few pounds higher because of the pressure drop through the valves. As the piston reverses direction and starts the suction stroke, no gas can enter from the evaporator side until the clearance gas's pressure reduces to suction. At some point on its suction stroke, the clearance volume is such that this "leftover" gas reaches a pressure condition that enables new gas to enter the cylinder through the suction valve. Higher compression ratios mean a more significant pressure spread between discharge and suction conditions, which results in more substantial piston travel before practical work commences. If the piston travels one-half of its stroke before starting to admit new gas, because of this re-expansion process, then only 50% of the total displacement can perform valuable work. Thus, the reciprocating compressor does not pump the same amount of gas when operating under varying compression ratios. The ratio of a piston's actual gas handling capacity to its theoretical displacement is called volumetric efficiency. Designing valves and valve plates to maintain low clearance volume is of prime importance to a manufacturer. Other factors, such as the rate of heat transfer from a cylinder wall, valve and piston leakage, and wire drawing in the valves, affect volumetric efficiency. However, once the compressor design has been set, the volumetric efficiency is primarily a function of the compression ratio. Figure #6 shows a typical ammonia compressor's volumetric efficiency versus compression ratio.

Volumetric efficiency, or the compressor's ability to do practical work, decreases as the compression ratio increases. Since volumetric efficiency is a significant factor in determining a compressor's capacity, designers have staged compressors to keep compression ratios at an optimum point.
Figure 7 graphically shows why a staged system, starting at about -10 Degrees F., should be considered to increase volumetric efficiency and system performance. A direct staged system combines two or more cycles using the same refrigerant in the circuit. Typically, the expressions "compound system" and "booster system" are used interchangeably, even though there is a slight difference between the two systems.

Figure #8 shows how the essential components could be arranged for a two-stage compound compression system. Two-stage plants are practical from approximately -10 Degrees F down to about -70 Degrees F. Three–stage compound compression would start around -70 Degrees F and go down to about -120 Degrees F. These figures pertain to the use of rotary and reciprocating compressors. Compound centrifugal systems are slightly different, as several impellers can be used within a single – casing. In the case described in Figure 8, a single refrigerant is used, and it flows in series through the two stages. In the discharge of the low-stage compressor, a closed gas and liquid inter-stage cooler are shown. It accomplished two purposes. Namely, it cools the superheated discharge gas from the low-stage compressor and the high-pressure liquid between the condenser and evaporator. This type of intercooler will maintain a high liquid pressure to operate the refrigerant controls at the evaporator better. A flash-type intercooler is sometimes used, in which all the liquid is flashed to the intermediate pressure. This provides the advantage of maximum cooling of the liquid. Earlier, it was mentioned that there is a slight difference between compound systems and booster systems. In a compound system, the work done in each compression stage is nearly equal, and the intermediate pressure is selected to provide this condition. A booster system sets the intermediate pressure by a side load or loads. The compound system results in the best horsepower per ton of refrigeration. The side load differentiating a booster system from a compound system is usually directed into the inter-stage cooler. If a separate suction scrubber had been used, the side load could be brought directly into the suction line of the high-stage machines.

Figure #9 shows the direct staged system on a Mollier diagram. In this instance, the work of compression is equally divided between the first and second stages. The ideal intermediate pressure is determined by linear measurement on the Mollier Chart to find the midpoint between the pressure extremes. Or, the same answer will result by taking the square root of the product of those pressure extremes. As shown in Figure #9, the ideal intermediate pressure should be 60 PSI. Adding about 5 PSI to the perfect pressure is customary in actual practice. Since entire load operation and maximum condensing pressure usually exist for a short period only, this deliberate imbalance of work will improve overall plant efficiency to cover a year's operation.

Furthermore, the selection of the intermediate pressure may depend upon the availability of machine displacement increments. In booster operation, the high-stage machine will usually work on a higher compression ratio, as the high-temperature evaporator pressure is likely lower than the ideal intermediate pressure of compound compression. If the booster load or side load is small compared to the high-temperature system and the booster compression ratio is low, the intermediate cooling system intercooler can sometimes be omitted. In Figure #9, the dotted line marked "intermediate flash" indicates that some high-temperature, high-pressure liquid is flashed at the suction pressure of the high-stage machine to cool the remaining liquid to 50. Degree F as the temperature of the subcooled liquid is purely arbitrary. A 20-degree F approach to the intermediate temperature usually is reasonable for this type of intercooler, normally referred to as a Shell and Coil intercooler. The high-stage compressor must pump the flash gas developed in the intercooler. Therefore, the high-stage compressor capacity must be adequate to handle the evaporator load plus the flash gas. As a general rule, the intercooling duty will require that the high-stage compressor have a capacity of about 10% to 25% higher than that of the low-stage compressor. This does not indicate a relationship in compressor displacement between high and low-stage machines. The gas density is much greater at intermediate (intercooling) conditions than at evaporator conditions. For example, at 30 Degrees F, one pound of ammonia gas occupies about 4.8 cubic feet, whereas one pound at 20 Degrees F occupies about 14.7 cubic feet.

Figure #10 tabulates the results of single–stage versus compound compression. These figures reveal theoretical values calculated using an ideal cycle on the Mollier chart and actual requirements computed from compressor efficiencies at varying compression ratios. The last two columns show interesting comparisons between total compressor displacements required per ton and actual brake horse powers per evaporator ton. Identical volumetric efficiencies is shown for the low-stage compressor and the high-stage compressor based on their having equal compression ratios. This may not be completely accurate, but it is reasonably close. The high-stage V.E. may be slightly higher than the low-stage compressor. The suction gas absorbs considerable heat before it passes through the suction valves and into the cylinder. The low-stage compressor cylinders handle a less dense gas than that managed by the high-stage machine. Thus, this gas absorbs more heat per pound of ammonia circulated. Initial heating does affect volumetric efficiency. Therefore, in actual practice, there may be some variation in volumetric efficiency, even though identical compressors may be used at the same compression ratios. The possible error in this assumption would be minimal.

Aside from the high operating expenses of single-stage compression in the low-temperature range, the compressor maintenance cost and manufacturer's limitations must be considered. Most manufacturers' have a limitation of about nine absolute compression ratios for this type of compressor. As the compression ratio increases, so does the compressor discharge temperature. Excessive heat is of primary concern as it is the forerunner of high maintenance expenses due to lubrication breakdown and valve carbonization. Suction superheat must be controlled closely, especially on ammonia systems operating at high compression ratios. Operation at low compression ratios will reduce wear on bearings and other compressor moving parts. As a rough approximation, applications having a compression ratio of about 7:1 and higher should be staged or compounded. A – 3- 3-stage compound compression system adds another stage and another intercooler to a 2- 2-stage compound system.

Both of these systems, single stage, and compound, would generally use any of the low-pressure refrigerants, such as R-12, R-22, R-502, ammonia propylene, or propane when capacity is such that reciprocating or rotary compressors would be involved. Lubricated reciprocating compressors pump some lubricating oil along with the discharge gas; however, oil-free refrigeration compressors are available in ringless or Teflon ring designs. Unless effectively removed from the system, this oil causes problems in low-temperature work below about -40 Degrees F. The oil is heavier than ammonia and is not miscible; thus, it sinks to the bottom and is relatively easy to remove from a vessel. Using discharge line oil separators and other oil removal procedures is good practice. It is good practice to remove the oil before it gets to a low-temperature evaporator, where it will become dense and more difficult to drain from the system. Oil in most halo carbon, all the Hydrocarbon refrigerants, is completely miscible or soluble at higher temperatures. ** With R-22, some of the oil begins to separate at about 0 Degrees F. At around -40 Degrees F, most of the oil has formed into a sticky, stiff layer floating on top of the refrigerant.

Attempts to bleed off this sticky mass of oil to oil still are almost impossible. It is also doubtful whether this oil can be returned for thermal expansion-fed evaporators. Needless to say, R-22 is not the preferred refrigerant for temperatures below about -40 Degrees F. Prolonged operation at these temperatures will spread a blanket of oil across the evaporator or pump receiver that literally chokes ebullition and gas return to the compressors. R-12 and oil remain more miscible at these lower temperatures, and the oil can be partially removed by use of oil stills or by holding the proper mass velocity in the direst expansion return risers. However, for evaporators that are fed by thermal expansion valves, vertical risers between the evaporator and compressor should be avoided in low-temperature work. The higher the temperature, the higher the density for a particular gas, i.e., one cubic foot of any gas will weigh more at 40 Degrees F than one cubic foot at any lower temperature. With less dense suction gas in a low-temperature system, 1000 FPM would correspond to a very low mass velocity and would not return the oil. Thus, velocity must be increased as the evaporator temperature is decreased. Oil is usually detrimental to the evaporator in either dissolved or free conditions, as it reduces the heat transfer coefficient. There are several oil recovery methods which will be described later. Consideration must be given to the refrigerant involved and the temperature level. This determines whether the oil is floating on top, on the bottom, or distributed throughout the evaporator and dictates which type of oil return systems must be employed. For higher-capacity systems, using centrifugal compressors will effectively eliminate the oil problem. By using four or more impellers and the proper refrigerant, it is possible to produce four or more impellers, and the proper refrigerant, it is possible to product evaporator temperatures down to about -50 degrees F and lower with this type of system (See Fig. II). A flash-type intercooler (See Fig. 12) of one or more stages is used to cool the liquid from condensing to evaporator conditions. The flash gas downstream of each float valve is either brought into the compressor between impellers or carried to the evaporator. In this instance, there is no superheating of discharge gas, but there are improvements in cycle efficiency. Figure # 12 shows a one-stage flash intercooler in detail.

A compound centrifugal system still uses the series flow of a single refrigerant; however, a superheater or flash-type intercooler is used between compressors, as shown in Figure #13. This superheats the discharge gas of the low-stage compressor before it enters the suction of the high-stage compressor to prolong the compressor life and prevent carbonization of the lubricating oil. Figure #14 describes a typical superheater/intercooler. It is also possible to use a combination of centrifugal and reciprocating compressors in a compound arrangement, as shown in Figure #15. Direct expansion-type intercoolers are shown here. The flash gas only partially superheats the discharge gas of the two compressors in the lower stages. A water-cooled gas cooler assists in this function between the middle and upper-stage compressors. The low-temperature evaporator employs a refrigerant recirculating pump and a spray-type cooler. Most low-temperature systems of low capacity use direct expansion-type evaporators. In contrast, high-capacity systems require the use of large-diameter evaporators, a refrigerant recirculating pump, and a sprayed tube bundle to become practical applications. To submerge tubes in a liquid refrigerant, as in conventional flooded shell and tube coolers, increases refrigerant pressure due to hydrostatic pressure. This gives a corresponding rise in effective evaporator temperature and can result in a temperature penalty depending on the refrigerant, the refrigerant depth, and the operating temperature level. The static head penalty in the shell and tube evaporator can be practically eliminated if the liquid level is kept below or near the bottom of the heat transfer surface and a pump is used to spray the liquid over the surface. Figure #16 describes the spray-type cooler in more detail, with tubes omitted for simplicity.

A compound centrifugal system can operate at evaporator temperatures of -150 Degrees F, as Fig. #17 shows. At about this level, the overall temperature spread from evaporator to condenser exceeds the practical range of one refrigerant. Using conventional refrigerants such as R-12 or R-22, the gas volumes per ton of refrigeration become too great at lower temperatures. For example, using R-22 in an evaporator at -150, the low-stage compressor must pump about 250 cubic feet of gas per T.R., whereas ethylene would only require about 8 CFM/TR. At -130 Degrees F, R-as has a pressure of 0.41 psi and requires about 200 CFM of gas per T.R. This is the application where Cascade staging is recommended. The usual range for cascading is between -130 degree F and -250 Degree F, however there are many circumstances where a cascade system is a wise selection at much higher temperatures.

Figure #18 describes the basis cycle for cascade staging. Such a cascade system consists of one circuit or stage, with its refrigerant, and as many succeeding stages as are necessary, each with its closed refrigerant circuit. Between the stages is a heat exchanger, called the cascade condenser, that serves as a condenser for the lower stage and an evaporator for the higher stage. About a 2 – 2-stage system, the high stage is a standard low-temperature system of the type previously described. It uses conventional refrigerants that can be condensed at normal pressures by either air or water. The low stage also functions like any other low-temperature system, except for the pressure characteristics of the refrigerant. The cascade system enables the designer to choose refrigerants at their most economical and practical operating points: low compression ratios, low discharge temperatures and pressures, low gas CFM/tr, low BHP/TR, and economical piping requirements.

Figure #19 compares the pressure and temperature characteristics of several refrigerants. The refrigerants suitable for the low stage of a cascade system have desirable pressures at the evaporator level. However, they have critical pressures below that obtainable in a water-cooled condenser, requiring condensation at a low temperature. Critical pressure is the saturation pressure at the necessary temperature. These low-temperature refrigerants can't be condensed by air or water and are used in a conventional high-stage refrigerant cycle. Commercially available refrigeration equipment is suitable for a maximum operating pressure of about 300 PSI to 350 PSI

Figure # 20 shows the essential components of a cascade system that would employ R-13 in the low stage, condensing the R-13 with a compound system using R-12. Cascade refrigeration systems are not new. They have been used by the Petro Chemical industry for many years, using hydrocarbon refrigerants in both low and high stages. The availability of non-flammable and non-toxic halo carbon high-pressure refrigerants has come about in the past 15 years. They are generally expensive on a per-pound basis. Hydrocarbon refrigerants, such as ethylene or propylene, are inexpensive; however, they are highly flammable. When using hydrocarbon, it is usually recommended that the use be limited to evaporator pressures above atmospheric. An explosion hazard could exist if operating in a vacuum and air enters the system. Operation above atmospheric pressure is desirable, though not mandatory, with any refrigerant to minimize purging problems. Because of the temperature extremes, expansion and contraction sometimes cause leaks in gaskets and seals. From an operating standpoint, it is much better to have refrigerant leak out than air leak in, especially when using halocarbon refrigerants. Because of expansion and contraction, the cascade condenser and the low-stage evaporator sometimes use a U-tube arrangement rather than a fixed tube bundle. The low-stage refrigerant condenses inside the tubes, and the high-stage refrigerant evaporates on the outside. The larger systems usually use a spray recirculation system in the cascade condenser and evaporator to keep the static head penalty down. It is imperative to keep the refrigerant charge as small as possible in the low stage of a non-hydrocarbon refrigeration system. Note that expansion tanks are shown in the low-stage piping. Liquid R-13 at 80 degrees F would be at a pressure of 521 PSI, which is well above the maximum working pressure for most standard equipment. Thus, if standard equipment is to be used, it is necessary to make provisions to maintain reasonable pressures during inoperative periods. It is also possible to provide a high-pressure receiver to transfer the refrigerant before being allowed to warm up to ambient conditions. This would require gravity drainage or a pumping system not affected by electrical power failures.

In numerous instances, with refrigerants such as R-13, an expansion tank is the most practical solution. The tank, low-side piping, and vessels must have enough expansion volume to contain the full vapor charge at a reasonable pressure at the highest expected ambient conditions. The usual practice is to pipe the expansion tank, or tanks, into the low side of the low-stage system. Less expansion volume is required if installed on the low rather than the high side. The tank of expansion needed can be huge, even with a relatively small charge in the low stage. When using a hydrocarbon refrigerant, it is sometimes more practical to flare the gas rather than provide expansion volume. If located on the low side, a valving or relief arrangement has to be followed to relieve the high-side pressure.

As with the high-stage system, consideration must be given to oil removal in lubricated systems, subcooling, intercooling, and de-superheating. The liquid refrigerant returning from the cascade condenser will be saturated or slightly subcooled at best. This is similar to the compound system using a flash-type intercooler. The fluid line must be adequately insulated and perhaps provided with a separate subcooled if the evaporators are at a distance and a static lift is required for the evaporator. It is sometimes desirable to provide a water-cooled de superheater on the discharge gas of the low-stage lubricated compressor to aid oil separation and to reduce cascade condenser load. A liquid-to-suction heat exchanger has the same advantages on the low-stage circuit as on the high, namely, liquid subcooling and helping to provide dry vapor to the compressor suction. The heat exchanger will increase the volume of suction vapor that the compressor must handle, but the increase in refrigerating effect offsets this undesirable effect. Thus, there is no appreciable change in required compressor displacement. For some types of compressors, it is sometimes necessary to raise the temperature of the low-stage suction gas to above some predetermined point to maintain the lubricating ability of the oil in the compressor or to meet material standards for low temperatures. Liquid-to-suction heat exchangers may be used for this purpose. Compressor manufacturers have varying requirements for minimum suction temperatures, however, -100 Degree F is a practical limit.

Evaporator temperature or pressure control would resemble any other compressor capacity control for the equipment involved. The high-stage compressor's capacity is usually controlled by the discharge pressure (condensing pressure) of the low stage to condense within the desired limits. The low stage requires a high-pressure cutout to prevent compressor operation at excess pressures during pull-down or when the high-stage capacity is unable to hold desired condensing pressures in the low stage. Conventional means should be used to prevent motor overloads due to high suction pressures on the low-stage compressor.

1. The cascade system has two disadvantages compared to the compound system.
To achieve heat transfer between the condenser of the low stage and the succeeding stage evaporator, there must be a temperature gradient across the cascade condenser. This is a penalty in the form of increased size and power for the same refrigerating effect. However, practical considerations overshadow this penalty on lower temperature applications.

2. The cascade system will probably be higher in first cost because of the requirement for a cascade condenser and expansion volume. Using a low boiling point refrigerant for the low stage will tend to offset this by requiring less compressor displacement. Furthermore, such a refrigerant may allow the evaporator circuit to take a greater pressure drop, reducing the temperature penalty. For example, evaporator loads resulting in a pressure drop of 0.5 psi at -100 Degrees F will cause temperature rises of approximately 8 degrees and 0.7 degrees for R-12 and R-13, respectively. Taking a higher pressure drop with a thermal expansion-fed evaporator would also improve heat transfer and oil return.

The system analysis just outlined can closely approximate compressor ratings. The required net displacement may be calculated using the heat load, refrigerating effect, and the specific volume of refrigerant at compressor suction conditions. Manufacturer's data can be used to determine approximate overall volumetric efficiency from the known compression ratio, and required compressor displacement can be calculated. Conventional methods can also approximate the required brake horsepower.

For a cascade system such as the one shown in Figure 20, an expansion tank or tanks are usually used, and the vapor-charging technique is usually used on the low stage. The entire system is allowed to warm up to ambient temperature conditions. Then, the calculated refrigerant charge is introduced into the system at a rate slow enough to allow the pressure build-up to correspond to ambient room conditions. For any ambient condition, the pressure must not exceed the calculated pressure-temperature relationship for the refrigerant.

Thus, at the maximum expected ambient temperature, the system pressure will not exceed the design working pressures of the expansion tank. When the high-stage system is started and brought down to its design evaporator temperature, the low-stage gas condenses to liquid in the cascade condenser. The liquid level in all shells should be noted before the low-stage compressor is started. The low-stage compressor should have an adjustable and automatic unloading capability to avoid motor overloads on star-up due to high back pressure. Another method sometimes used to relieve high discharge pressure on start-up and quickly reduce suction pressure on thermal expansion-fed evaporators is to install a load limit valve or automatic relief valve between the high side and the expansion tank. The refrigerant, thus stored in the expansion tank, would be allowed to return slowly to the suction side through a restriction or valve. As the compressor suction and discharge pressures approach design conditions, no more gas would be released into the expansion tank.

The low-stage compressors should be carefully checked for suction and discharge superheats for compressor protection. The difference between the suction line and the suction temperature corresponds to the suction pressure in the superheat. The suction line could be heavily frosted at these temperatures and still operate at excessive superheat. Thermocouples should be used to determine suction line temperatures.
Figure #21 shows an ethylene-propylene cascade system's essential components and piping. This system is designed to condense high-pressure ethylene gas and to store it as a liquid at atmospheric pressure. Ethylene is a low-stage refrigerant. The low-stage compressor is a two-stage, non-lubricated ring-less type reciprocating compressor. The high-stage system uses a five–stage centrifugal compressor with propylene as the refrigerant. Note that the cascade condenser uses a double tube bundle. One circuit cools the incoming ethylene, and the other condenses the flash gas. The storage tanks are vacuum-jacketed vessels designed for atmospheric pressure only. Fifty percent of compressor units are provided. Figure 22 shows the Mollier diagram for the ethylene circuit.

As previously discussed, some oil return system must be employed in all low-temperature refrigeration systems that use lubricated compressors. The design of such systems is more of an art than science. Figure 23 describes an R22 Refrigeration System with a -20 Degree F to -40 Degree F evaporator temperature with a two-stage lubricated compressor, an oil spill, and a return system. The evaporator is a vertical shell and coil type, which is not conducive to good oil return, and thus, several drain points for the oil/refrigerant mixture are shown with sight glasses. These drains are manually adjusted to obtain the maximum oil return if oil return is possible. As the temperature may vary, some 20 degrees F, it is likely that the oil accumulation in the R-22 will move, and thus, several drain points are necessary. If this was an R-12 system, the oil drain could be from one point ……..preferably the bottom of the evaporator, as oil remains miscible in the R-12 at lower temperatures. In ammonia systems, oil may be quickly drained from the lowest point in the evaporator, receiver, or suction trap. Manual oil drains from ammonia systems will remove the oil relatively easily down to a temperature of about -35 Degree F. Below this point, oil drainage becomes increasingly tricky.

Figure #24 describes a system similar to the one shown in Figure 23; however, it employs a labyrinth piston-type oil-free compressor. This excellent compressor is for this type of evaporator and any other type of flooded or spray-type evaporator operating at low temperatures. The lower the temperature, the more advantageous the oil-free system becomes. In overall operation, the oil-free system eliminates the oil return problem. By removing the oil, the evaporator efficiency, which usually drops with oil fouling, is always at the maximum on the refrigerant side. Thus, evaporators typically designed for lubricated refrigeration compressor systems would be oversized for an oil-free compression system and operate at 100% efficiency.

The screw compressor is also adaptable to refrigeration systems. This compressor has characteristics to make it highly adaptable to refrigeration duty for low and high temperatures. The compressors are highly reliable, with running times of over 25,000 hours between maintenance periods. The compressors are positive displacement rotary machines with the same relative operating characteristics as reciprocating. They have only four bearings and a thrust bearing as wearing parts. Capacity reduction is available down to 10%, either automatic or manual. These compressors can pump liquid without causing damage and are stable over the entire operating range.

Both oil-free and oil-injected screw compressors are available. The oil-free units require four shaft seals, operate at relatively high speeds of 4000 to 10,000 rpm, and have a limited pressure differential. The oil-injected screw machines operate at 3600 to 4500 rpm.

A typical flow diagram, figure #25, shows an oil-free ammonia refrigeration system for an ammonia plant using rotary screw compressors. Figure 26 shows an oil-injected screw compressor, lubrication, and control system. The oil-injected screw compressor is unique in that it can operate continuously at absolute compression ratios of up to 20 without damaging the machine. Oil injection into the rotors decreases noise and reduces the required shaft horsepower. These compressors operate at efficiencies similar to standard reciprocating refrigeration compressors.

** The addition of 3% pentane of R-12 to a low-temperature R-22 system (below -10dF) will provide a good vehicle for oil return. New synthetic oils miscible with R-22 at temperatures of -50 Degrees F are also available.

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