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LOW TEMPERATURE REFRIGERATION SYSTEMS
George Brown,Lewis Refrigeration Co. Washington,
George Briley,Lewis Refrigeration Co.Texas.

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

In cryogenic engineering, temperatures closely approaching absolute zero aren’t uncommon. However, such applications usually start at about -300oF., and go downward. This will be referred to as ultra low temperatures. Low temperature refrigeration is defined as those applications requiring evaporator temperatures in the range of -20 oF down to -250 oF. The vast majority of installations would be at temperature above -125 oF., however systems down to -250 oF are covered.

Mechanical refrigeration is the controlled removal of heat (British Thermal Units), with maintenance of temperatures below atmospheric conditions. To remove 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 unit of refrigeration is a ton of refrigeration usually abbreviated TR, a rate unit. A ton of refrigeration is the heat removal necessary to freeze 2000# of 32 oF water in 24 hours. Since the latent heat of fusion of ice 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, we find that figure #1 shows the refrigerant vapor compressed by a piston compressor. Such a plant consists of four major pieces of equipment, namely compressor, condenser, evaporator, and expansion valve. The refrigerant absorbs heat from a heat transfer medium being cooled in the evaporator, and 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 by means of hear exchange surface. The liquefied refrigerant then flows through an expansion device of some type where its pressure and temperature are reduced to whose in the evaporator.
Using an “ideal “cycle, hear and flow balances can be analyzed by means of the Mollier Chart for the refrigerant being used. Figure # 2 shows the Mollier, or pressure-enthalpy diagram, for a single- stage vapor compression cycle. The illustrations presented here deal with ammonia; however the principles apply equally as well to other refrigerants. Many different refrigerants are described in this study of low temperature refrigeration. In each case, the properties of the refrigerant must be considered.

The pressure-enthalpy diagrams are commonly called “P – H” diagrams. The development of Mollier Charts is a result of thermodynamic studies of the general energy equation as applied to steady flow processes. Enthalpy is a measure of the total heat, or heat content at the various pressure points of the liquid, or vapor, or mixture, of the refrigerant.

When work is done or heat is transferred, the refrigerant undergoes a change in enthalpy. The enthalpy is read on the abscissa, in BTU per pound of refrigerant, and absolute pressure, in psia, is read on the ordinate. The heavy curve to the left represents the saturated liquid line, and the heavy curve on the right represents the saturated vapor line. In between the two curves, the refrigerant will be in both liquid and vapor form, 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 of it flashes and cools the remainder of the liquid to evaporator temperature. Assuming no friction, the process in the liquid – vapor mixture region is shown at constant pressure and temperature, 39 psia and + 10 oF in this instance. At point B, the suction gas enters the compressor under saturated vapor conditions. Compression takes place from points B to C along a constant entropy line, which assumes no friction and no transfer of heat. This is an “ideal” cycle, however in actual compression, friction is involved and so is heat transfer.

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

1. 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. Pounds of ammonia circulated per minute per ton is equal to 200 (200 BTU/lb = TR) divided by the refrigerating effect, or 200+ 466 = 0.43 lbs. of ammonia per ton.

3. The ideal work of compression is equal to the difference in enthalpies at points C and B. In this case, 718-615-103 BTU/lb of ammonia. This ideal work of compression is 103 x 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 by the use of 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. Compression ratio is equal to the absolute discharge pressure divided by the absolute suction pressure. For the conditions shown, the CR would be 196+39= 5.0.

The P–H diagram discloses other important considerations in the study of basic refrigeration. The heat rejected to the condenser is equal to 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 greater than shown on the ideal cycle because of compressor inefficiencies and design. Required compressor BHP per TR increases greatly as evaporator temperature drops. Figure 3 gives a comparison of the condenser hear rejection, condenser tube surface and condenser water requirements at various evaporator temperatures.

For the same single – stage vapor compression system at an evaporator temperature to -20 oF the P – H diagram would look something similar to figure #4. In this instance, the important considerations as analyzed on the previous diagram are shown in the tabulation. The differences in ideal cycles for +10 oF evaporator conditions versus -20 oF conditions are obvious. As the evaporator temperature decreases, the compression ratios get higher and the theoretical H.P. /TR increases. These figures are determined by comparing ideal cycles on a Mollier Chart. Unfortunately, in an actual plant, the compressor inefficiencies mentioned earlier cause the cycle figures to be slightly misleading. 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 compressor has one or more pistons to compress the gas. As the piston makes its stroke toward top dead center, some clearance will always exist between the end of the stroke and the top head. The exact amount is determined by many factors such as manufacturing tolerances, valve design, and others. Clearance in the vicinity of 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 of the gas from being ejected from the cylinder. The gas remaining is under discharge pressure conditions, which is 1496 psia for the figures shown. Actually, pressure within the cylinder would be a few pounds higher because of pressure drop thru the valves. As the piston reverses direction and starts the suction stroke, no gas can enter from the evaporator side until the clearance gas has its pressure reduced to that corresponding 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 greater pressure spread between discharge and suction conditions and this result in a greater travel of the piston before useful work commences. If the piston travels one- half of its stroke before staring to admit new gas, because of this re-expansion process, then only 50% of the total displacement can perform useful work. Thus, the reciprocating compressor does not pump the same amount of gas when operating under varying compression ratios. The ratio of actual gas handling capacity of a piston 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 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 bean set, the volumetric efficiency is primarily a function of compression ratio. Figure #6 shows volumetric efficiency versus compression ratio for a typical ammonia compressor.

The volumetric efficiency, or the ability of the compressor to do useful work, decreases as the compression ratio increases. Since volumetric efficiency is a major factor in determining the actual capacity of a compressor, designers have staged the compressors to keep compression ratios at an optimum point.
Figure #7 shows graphically why consideration should be given to a staged system, starting at about -10 oF., to increase volumetric efficiency and system performance. A direct staged system is basically a combination of two or more cycles, using the same refrigerant in the circuit. Normally the expressions “compound system” and “booster system” are used interchangeably even though there is a slight difference in the two systems.

Figure #8 shows how the basic components could be arranged for a two-stage compound compression system. Two – stage plants are practical from approximately -10 oF down to about -70 oF. Three – stage compound compression would start around -70 oF and go down to about -120 oF. 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 is shown. It accomplished two purposes, namely it cools the superheated discharge gas from the low stage compressor and it cools the high pressure liquid between condenser and evaporator. This type intercooler will maintain a high liquid pressure for better operation of the refrigerant controls at the evaporator. A flash type intercooler is sometimes used, in which all of the liquid is flashed to the intermediate pressure. This has 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. In a booster system, the intermediate pressure is set by a side load or loads. The compound system results in the best horsepower per ton of refrigeration. The side load, that differentiates 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 can 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 has been shown to be 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 psia. In actual practice, it is customary to add about 5 psig to the ideal pressure. Since full load operation, and maximum condensing pressure, usually exists for a short period of time only, this deliberate imbalance of work will give better overalls plant efficiency cover a years operation. Also, the selection of the intermediate pressure may depend upon the availability of machine displacement increments. In booster operation, the high stage machine will usually be working on a higher compression ratio as the high temperature evaporator pressure is likely to be 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 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 of the high temperature, high pressure liquid is flashed at the suction pressure of the high stage machine to cool the remaining liquid to 50. oF as the temperature of the subcooled liquid is purely arbitrary. A 20 oF approach to the intermediate temperature is normally reasonable for this type intercooler which is normally referred to as a Shell and Coil intercooler. The flash gas developed in the intercooler must be pumped by the high stage compressor. 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 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 oF, one pound of ammonia gas occupies about 4.8 cubic feet, whereas one pound of ammonia gas at 20 oF occupies about 14.7 cubic feet.

Figure #10 tabulates the results of single –stage versus compound compression. These figures reveal theoretical values as calculated by using an ideal cycle on the Mollier chart, and actual requirements as 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 have been 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 is reasonably close. The high stage V.E. may be a little higher than that of the low stage compressor. A considerable amount of heat is absorbed by the suction gas before it passes thru the suction valves and into the cylinder. The low stage compressors cylinders handle a gas that is less dense than that handled by the high stage machine. Thus, this gas absorbs a greater amount of 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 on the same compression ratios. The possible error in this assumption would be very small.

Aside from the mater of high operating costs 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 9 absolute compression ratios for this type compressor. As compression ratio increases, so does compressor discharge temperature. Excessive heat is of primary concern as it is the forerunner of high maintenance costs due to lubrication breakdown, and valve carbonization. Suction superheat must be controlled very closely especially on ammonia systems operating at high compression ratios. Operation at low compression ratios will reduce wear on bearings and other moving parts of the compressor. As a rough approximation, applications having a compression ratio of about 7:1 and higher should be staged or compounded. A – 3- stage compound compression system merely adds another stage, and another intercooler, to a 2- stage compound system.

Both of these systems single stage and compound would normally use any of the low pressure refrigerants, such as R-12, R-22, R-502, ammonia or propylene or propane when capacity is such that reciprocating or rotary compressors would be involved. Lubricated reciprocating compressors pump some amount of lubricating oil along with the discharge gas, however oil free refrigeration compressors are available in ring less or Teflon ring design. Unless effectively removed from the system, this oil causes problems in low temperature work below about -40 oF. Oil is heavier than ammonia, and no miscible, thus it sinks to the bottom and is relatively easy to remove from a vessel. It is good practice to use discharge line oil separators, and other oil removing procedures. It is good practice to remove the oil before it gets to a low temperature evaporator and where it will become viscous and more difficult to drain from the system. Oil in most of the halocarbon, all of the Hydrocarbon refrigerants, is completely miscible or completely soluble at the higher temperature. ** With R-22, some of the oil begins to separate about 0 oF. At around -40 oF 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. For thermal expansion fed evaporators, it is also doubtful whether this oil can be returned. Needless to say, R-22 is not the preferred refrigerant for temperatures below about -40 oF. 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 direst expansion return risers. However for evaporators those 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 weight more at 40 oF, 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 evaporator temperature is decreased. Oil is usually detrimental in the evaporator in either dissolved or free condition, 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 oil return systems must be employed. For systems of higher capacity, the use of 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 oF 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 conditions down to evaporator conditions. The flash gas downstream of each float valve is either brought into the compressor between impellers, or carried along to the evaporator. In this instance, there is no desuperheating 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 series flow of a single refrigerant; however a desuperheater or flash type intercooler is used between compressors as shown in Figure #13. This desuperheats 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 desuperheater/intercooler. It is also possible to use a combination of centrifugal and reciprocating compressors in compound arrangement as shown by Figure #15. Direct expansion type intercoolers are shown here. The flash gas only partially desuperheats 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 compressor. 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. Where high capacity systems require the use of large diameter evaporators, a refrigerant recirculating pump and a sprayed tube bundle becomes a practical application. To submerge tubes in a liquid refrigerant, as in conventional flooded shell and tube coolers, results in a refrigerant pressure increase 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 oF, as Fig. #17 shows. At about this level, the overall temperature spread from evaporator to condenser exceeds the practical range of one refrigerant. Using the conventional refrigerants such as R-12 or R-22, the gas volumes per ton of refrigeration become too great at the lower temperatures. For example, using R-22 in an evaporator at -150 oF, the low stage compressor must pump about 250 cubic feet of gas per TR, whereas ethylene would only require about 8 CFM/TR. At -130 oF, R-as has a pressure of 0.41 psia, 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 oF and -250 oF, 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 own closed refrigerant circuit. Between the stages is a heat exchanger, called the cascade condenser, that serves both as condenser for lower stage and evaporator for the ext higher stage. With reference to a 2 – stage system, the high stage is a standard low temperature system of the type previously described. It uses the conventional refrigerants, that can be condensed at normal pressures by wither air or water. The low stage also functions like any other low temperature system, with the exception being the pressure characteristics of the refrigerant. The cascade system enables the designer to choose refrigerants at their most economical and practical operating points, namely; low compression ratios, low discharge temperatures and pressures, low gas cfm/tr, low BHP/TR, and economical piping requirements.

Figure #19 compares 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 at temperatures below that obtainable in a water-cooled condenser, hence requiring condensation at a low temperature. Critical pressure is the saturation pressure at the critical temperature. These low temperature refrigerants can’t be condensed by air or water and used in a conventional high stage refrigerant cycle. Commercially available refrigeration equipment is suitable for a maximum operating pressure of about 300 psig to 350 psig.

Figure # 20 shows the basic 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 halocarbon high pressure refrigerants has come about in the past 15 years. They are normally very expensive on a per pound basis. The hydrocarbon refrigerants, such as ethylene or propylene, are inexpensive; however, they are highly flammable. When using hydrocarbon, it is normally recommended that the use be limited to evaporator pressures above atmospheric. If operating in a vacuum and air enters the system, an explosion hazard could exist. 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 at gaskets and seals. It is much better from an operating standpoint, to have refrigerant leak out than air to 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 in the low stage of non-hydrocarbon refrigeration system as small as possible. Note that expansion tanks have been shown in the low stage piping. Liquid R-13 at 80 oF would be at a pressure of 521 psig, which is well above 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 into which the refrigerant would be transferred before being allowed to warm up to ambient conditions. This would either require gravity drainage or a pumping system not affected by electrical power failures.

In many cases with refrigerants, such as R-13, an expansion tank is the most practical solution. The tank, and low side piping and vessels, must have enough expansion volume to contain the full charge as vapor, 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 side rather than the high side. Even with a relatively small charge in the low stage, the required expansion tank can be huge. 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, valving or relief arrangement must to follow to relieve the high side pressure.

Just as with the high stage system, consideration must be given to oil removal in lubricated systems, subcooling, intercooling, and desuperheating. The liquid refrigerant returning from the cascade condenser will be saturated, or at best only slightly subcooled. This is similar to the compound system using a flash type intercooler. The liquid line must be properly insulated and perhaps provided with a separate subcooler if the evaporators are at a distance and a static lift is required to the evaporator. It is sometimes desirable to provide a water-cooled desuperheater 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 must be handled by the compressor, but this undesirable effect is offset by the increase in refrigerating 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 a ability of the oil in the compressor or to meet material standards for low temperature. Liquid to suction heat exchangers may be used for this purpose. Compressor manufacturers have varying requirements for minimum suction temperatures however, -100 oF is a practical limit.

Evaporator temperature or pressure control would be similar to any other compressor capacity control for the equipment involved. The high stage compressor – 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. The use of a low boiling point refrigerant for the low stage will tend to offset this by requiring less compressor displacement. Also, such a refrigerant may allow the evaporator circuit to take a greater pressure drop, reducing temperature penalty. For example, an evaporator loads resulting in a pressure drop of 0.5 psi at -100 oF will cause temperature rises of approximately 8 o and 0.7 o for R-12 and R-13 respectively. The ability to take a higher pressure drop with a thermal expansion fed evaporator would also improve heat transfer and oil return.

Compressor ratings can be closely approximated by the system analysis just outlined. The net displacement required 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. Required brake horsepower can also be approximated by conventional methods.

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

Thus, at maximum expected ambient temperature, the system pressure will not exceed 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. To avoid motor overloads on star-up due to high back pressure, the low stage compressor should have an adjustable, and automatic, unloading capability. 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 restrictor valve. As the compressor suction and discharge pressures approach design conditions, no more gas would be relieved to the expansion tank.

The low stage compressors, or compressors, should be carefully checked for suction and discharge superheats, for compressor protection. The difference between the suction line temperature and the suction temperature corresponding to suction pressure in the suction superheat. At these temperatures, the suction line could be heavily frosted and still be operating at excessive superheat. Thermocouples should be used to determine suction line temperatures.
Figure #21, shows the basic components and piping of an ethylene-propylene cascade system. This system is designed to condense high pressure ethylene gas, and to store it as a liquid at atmospheric pressure. Ethylene is the 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 circuit condenses the flash gas. The storage tanks are vacuum-jacketed vessels, designed for atmospheric pressure only. Fifty percent compressor units are provided. Figure 22 shows the Mollier diagram for the ethylene circuit.

As previously discussed, in all low temperature refrigeration systems that use lubricated compressors some type of oil return system must be employed. The design of such systems is more of an art than science. Figure 23 described a R22 Refrigeration System with a -20 oF to -40 oF evaporator temperature with a two stage lubricated compressor, an oil still and return system. The evaporator is a vertical shell and coil type which is not conductive to good oil return and thus several drain points for 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 oF, it is possible 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 easily 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 oF. Below this point, oil drainage becomes increasingly difficult.

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 for this type evaporator and any other types 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 eliminating the oil, the evaporator efficiency which normally drops with oil fouling is always at the maximum to the refrigerant side. Thus evaporators normally designed for lubricated refrigeration compressor systems would be oversized for an oil free compression system and would operate at 100% efficiency at all times.

The screw compressor is also adaptable to refrigeration systems. This compressor has characteristics that make it highly adaptable to refrigeration duty for low temperature as well as 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 a reciprocating compressor. 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 and usually 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 up to 20 without damaging the machine. Oil injection into the rotors decreases noise and also decreases 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 are also available that are miscible with R-22 at temperatures to -50 dF.