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REFRIGERATION AND CRYOGENIC SYSTEMS A. M. Papadopoulos and C. J. Koroneos Aristotle University of Thessaloniki, Greece Keywords : Refrigeration cycles, refrigerants, heat pumps, absorption cycles, gas refrigeration cycle, cryogenic heat exchangers Contents 1. General Characteristics of Refrigeration and Cryogenic Installations 2. Actual Vapor-Compression Refrigeration Cycles 3. The Absorption Cycle 4. Gas Refrigeration Cycles 5. Cryogenic Installations and Gas Liquefaction 6. Heat Exchangers for Producing Low Temperatures Related Chapters Glossary Bibliography Biographical Sketches Summary The aim of this article is to provide the the basic information needed to understand the technology of producing low and very low temperatures, i.e., refrigeration and cryogenics. The article addresses the refrigeration, absorption, and liquefaction processes from the engineering point of view. It avoids a purely theoretical approach and it is not presenting technologies of limited, state of the art applications. In order to enable the comprehension of these applications, a brief presentation of the thermodynamics of the respective processes is also given. Emphasis is placed on the main technologies applied in contemporary refrigeration practice, like heat pump systems, single and multistage compression refrigerating systems, Lithium Bromide and Aqueous Ammonia absorption systems, and gas refrigeration systems. The Linde and the Claude gas liquefaction systems are presented as the most interesting cryogenic processes, as well as the heat exchangers used in those processes. For the reader who may be interested in a more detailed and in-depth knowledge of refrigeration and cryogenics technologies, the references mentioned in the bibliography will be helpful. 1. General Characteristics of Refrigeration and Cryogenic Installations It is a natural phenomenon that heat flows in the direction of decreasing temperature, i.e., from high temperature to low temperature regions. The reverse process, however, can take place, not in violation of the second law of thermodynamic but by the addition of work. The heat transfer from a low-temperature region to a high-temperature one requires therefore, intermediate devices and systems called refrigerators. The difference between a refrigeration and a cryogenic system lies in the achievable temperatures, with the dividing line being set at −100°F or −74°C. 1 of 21

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The methods used and the physical principles applied to achieve low temperatures are shown in Table 1. Table 1. Methods and principles for low temperature achievement Refrigerators are systems that operate on a cyclic principle involving a working fluid called a refrigerant. The working principle of a refrigerator is shown schematically in Figure 1. In this scheme QL is the magnitude of the heat removed from the refrigerated space at a given temperature TL, QH is the magnitude of the heat rejected to the warm environment at a temperature TL, and Wnet,in is the net work input needed to the refrigerator. QL and QH represent magnitudes and thus are positive quantities. In the same scheme, a heat pump is shown, in order to demonstrate the relation between the two applications, which is the inversion of the same function. The objective of the refrigerator is to remove heat from the cold medium, whilst the objective of the heat pump is to supply heat to a warm one. Both of these systems transfer heat from low to high temperature media with the input of work.

Figure 1. Refrigerator and heat pump working principle The objective of the refrigerator is to maintain the refrigerated space at a low temperature by removing heat from it. Discharging this heat to a higher temperature medium is only a necessary part of the operation, not an aim on its own. The objective of a heat pump, however, is to maintain a heated or cooled space at a given temperature. This is accomplished by absorbing heat from an energy source, such as the ground, water, or ambient air, and supplying this heat to a warmer or colder medium, such as a building, in winter or summer respectively. The performance of refrigerators and heat pumps is expressed in terms of the coefficient of performance (COP), which is defined as (1) (2) These relations can also be expressed in the rate form by replacing the quantities QL, QH and Wnet,in by respectively. Notice that both the COPR and COPHP can be greater than 1. A comparison of the equations (1) and (2) reveals that for fixed values of QL and QH: (3) This relation implies that COPHP >1 since COPR is a positive quantity. That is, a heat pump will function in the case of heating a space, at worst, as a resistance heater, supplying as much energy to the house as it consumes. In reality, however, part of QH is lost to the outside air through piping and other devices, and COPHP may drop below unity when the outside air temperature is too low. When this happens the system normally switches to a resistance heating mode. The cooling capacity of a refrigeration system 2 of 21

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from the refrigerated space is often expressed in terms of tons of refrigeration or in British Thermal Units (BTUs). The capacity of a refrigeration system that can freeze 1 ton of liquid water at 0°C into ice at 0°C in 24 hours is said to be 1 ton. One ton of refrigeration is equivalent to 211 kJ/min or 200 Btu/mon. 1.1 The Ideal Vapor-Compression Refrigeration Cycle Many of the impracticalities associated with the reversed Carnot cycle can be eliminated by vaporizing the refrigerant completely before it is compressed and by replacing the turbine with a throttling device, such as an expansion valve or capillary tube. The cycle that results is called the ideal vapor-compression refrigeration cycle, and it is shown schematically in Figure 2 and on a T-S diagram in Figure 3. The vapor-compression refrigeration cycle is the most widely used cycle for refrigerators, air conditioning systems, and heat pumps.

Figure 2. Schematic representation of the ideal vapor compression refrigeration cycle This cycle consists of four processes: 1 2 3 4

2: Isentropic compression 3: Heat rejection in a condenser at constant pressure 4: Throttling in an expansion device 1: Heat absorption in an evaporator at constant pressure

Figure 3. Temperature versus entropy diagram of the ideal vapor compression refrigeration cycle The cycle can be well described in five steps. Step 1: The refrigerant leaves the evaporator at state 1 as a low pressure, low temperature, saturated vapor. In an ideal cycle, the refrigerant enters the compressor without heat gain/loss from the environment. Step 2: The compression process compresses the refrigerant reversibly (without losses), leaving the refrigerant in a high temperature, high pressure, superheated vapor state. Step 3: Condensing brings the refrigerant out of superheated state and condenses it at a constant pressure. The refrigerant becomes a high pressure, medium

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temperature, saturated liquid. Step 4: Once condensed, the refrigerant expands adiabatically (without heat transfer) and reversibly (i.e., at constant enthalpy) in the expansion valve. The refrigerant leaves the expansion valve as a low pressure, low temperature, low enthalpy vapor. Step 5: At constant pressure and in an adiabatic fashion, the refrigerant enters the evaporator and evaporates. It is during the evaporation that heat transfer occurs from the refrigerated space to the refrigerant. 2. Actual Vapor-Compression Refrigeration Cycles An actual vapor-compression refrigeration cycle differs from the ideal one in several ways, owing mostly to the irreversibilities that occur in various components. Two common sources of the irreversibilities are fluid friction (which causes pressure drops) and heat transfer to or from the surroundings. The T-S diagram of an actual vapor compression refrigeration cycle is shown in Figure 4. In the ideal cycle, the refrigerant leaves the evaporator and enters the compressor as saturated vapor. This cannot be accomplished in practice, however, since it is not possible to control the state of the refrigerant so precisely. Instead the system is designed so that the refrigerant is slightly superheated at the compressor inlet. This slight overdimensioning ensures that the refrigerant is completely vaporized when it enters the compressor. Also, the line connecting the evaporator to the compressor is usually very long, thus the pressure drop caused by fluid friction and heat transfer from the surroundings to the refrigerant can be very significant. The result of superheating, heat gain in the connecting line, and the pressure drops in the evaporator and the connecting line, is an increase in the specific volume. This leads to an increase in the power-input requirements to the compressor, since steady-flow work is proportional to the specific volume. The compression process, however, will involve frictional effects that increase the entropy, and heat transfer which may increase or decrease the entropy, depending on the direction. Therefore the entropy of the refrigerant may increase (process 1-2), or decrease (process 1-2') during the actual compression process, depending on which effect dominates. The compression process 1-2 may be even more desirable than the isentropic compression process since the specific volume of the refrigerant and thus the work input requirements are smaller in this case. Therefore, the refrigerant should be cooled during the compression process whenever it is practical and economical to do so. In the ideal case, the refrigerant is assumed to leave the condenser as saturated liquid at the compressor exit pressure. In actual situations, however, it is unavoidable to have some pressure drop in the condenser as well as in the lines connecting the condenser to the compressor and to the throttling valve. Also, it is not easy to execute the condensation processes with such precision that the refrigerant is saturated liquid at the end and it is undesirable to route the refrigerant to the throttling valve before the refrigerant is completely condensed. Therefore the refrigerant is subcooled somewhat before it enters the throttling valve. The refrigerant enters the evaporator with a lower enthalpy and thus can absorb more heat from the refrigerated space. The throttling valve and the evaporator are usually located very close to each other, so that the pressure drop in the connecting line is small.

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Figure 4. Temperature versus entropy diagram of the actual vapor compression refrigeration cycle 2.1 Selecting the Right Refrigerant When designing the refrigeration system, there are several refrigerants to choose from, such as chloroflurocarbons (CFCs), ammonia, hydrocarbons (propane, ethane, ethylene, etc.), carbon dioxide, air (in the air conditioning of aircraft), and even water (in applications above the freezing point). The right choice of refrigerant depends on the situation at hand. CFCs such as R-11, R-12, R-22 and R-502 account for over 90% of the market in the United States today. Ethyl ether was the first commercially used refrigerant in vapor compression systems in 1850, followed by ammonia, carbon dioxide, methyl chloride, sulfur dioxide, butane, ethane, propane, isobutane, gasoline, and chloroflurocarbons, among others. The industrial and heavy commercial sectors were very satisfied with ammonia and still are, although ammonia is toxic. The advantages of ammonia over other refrigerants are focused in its low cost, higher COPs (and thus lower energy costs), more favorable thermodynamic and transport properties and thus higher heat transfer coefficients (therefore it leads to smaller and lower-cost heat exchangers), greater detectability in the event of a leak, and no effect on the ozone layer. The major drawback of ammonia is its toxicity, which makes it unsuitable for domestic use. Ammonia is predominantly used in food refrigeration facilities such as the cooling of fresh fruits, vegetables, meat and fish; refrigeration of beverages and diary products such as beer, wine, milk, and cheese; freezing of ice cream and other foods; and ice production and low-temperature refrigeration in the pharmaceutical and other process industries. It is remarkable that the refrigerants used in the light commercial and household sectors such as sulfur dioxide, ethyl chloride, and methyl chloride are highly toxic. The widespread publicity of a few instances of serious leaks which resulted in illnesses and deaths in the 1920s caused public pressure to ban or limit the use these refrigerants, resulting in a need for the development of safe refrigerants for household use. At the request of the Frigidaire Corporation, the General Motors research laboratory developed R 21, the first member of the CFC family of refrigerants in 1928. Of several CFCs developed, the research teams concluded that R 12 was the most suitable refrigerant for commercial use, and gave the CFC family the trade name "Freon". Commercial production of R 11 and R 12 began in 1931 by a company jointly formed by General Motors and E.I. du Pont de Nemours and Co., Inc. The versatility and low-cost of CFCs made them the refrigerants of choice. CFCs were also widely used in aerosols, foam insulation, polystyrol insulating materials, and in the electronic industry as solvents to clean computer chips. R 11 is used primarily in large-capacity water chillers serving air conditioning systems in buildings. R 12 is used in domestic refrigerators and freezers, as well as in automotive air conditioners. R 22 is used in window air conditioners of commercial buildings, in large industrial refrigeration systems, and is competitive with ammonia. A mixture of R 115 and R 22, called R 502, is the dominant refrigerant used in commercial refrigeration system such as those in supermarkets,

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because it allows low temperatures at evaporators while operating at single-stage compression. The ozone crisis has caused a major stir in the refrigeration and air conditioning industry, leading to skepticism over these widespread refrigerants. During the 1970s it became evident that CFCs deplete the ozone layer that protects the Earth from harmful ultraviolet radiation, while they enhance the greenhouse effect, hence preventing infrared radiation from escaping the Earth. Their significant role both in the greenhouse effect and in the ozone layer depletion became a major scientific and public issue. Since then, and as a result of various national and international legislative actions, the use of some CFSs has been banned and they are being phased-out of use in many countries. In particular fully halogenated CFCs such as R 11, R 12, and R 115, are held responsible for the depletion of the ozone layer. The non-fully halogenated refrigerants, such as R 22, have only about 5% of the ozone-depleting capability of R 12. As a result of these developments, research has focused on the development of refrigerants which are friendly to the ozone layer and at the same time do not contribute to the greenhouse effect. The recently developed chlorine-free R 134a is expected to replace R 12. Two important parameters that need to be considered in the selection of a refrigerant are: the temperatures of the two media, i.e., the refrigerated space and the environment, with which the refrigerant is to exchange heat. To have heat transfer at a reasonable rate, a temperature difference of 5 to 10°C should be maintained between the refrigerant and the medium with which it is exchanging heat. If a refrigerant space is to be maintained at 10°C, for example, the temperature of the refrigerant should remain at about 20°C while it absorbs heat in the evaporator. The lowest pressure in a refrigeration cycle occurs in the evaporator, and the pressure should be above atmospheric pressure to prevent any air leakage into the refrigeration system. Therefore, a refrigerant should have a saturation pressure of 1 atm (0.101325 MPa) or higher, at 20°C in this particular case. Ammonia, R 12 and R 134a are substances with such properties. The temperature, and thus the pressure, of the refrigerant on the condenser side depend on the medium to which heat is rejected. Higher COPs, achievable through lower temperatures in the condenser, can be maintained if the refrigerant is cooled by liquid water instead of air. However, the use of water as a cooling medium can not be justified by its economic feasibility, except in large industrial refrigeration systems. The temperature of the refrigerant in the condenser can not fall below the temperature of the cooling medium (about 20°C for a household refrigerator system) and the saturation pressure of the refrigerant at this temperature should be well below its critical pressure, if the heat rejection process is to be approximately isothermic. If no single refrigerant can meet the temperature requirements, then two or more refrigeration cycles with different refrigerants can be used in series. Such a refrigeration system is called a cascade system. Other desirable characteristics of a refrigerant include being non-toxic, non-corrosive, nonflammable, and chemically stable, having a high enthalpy of vaporization in order to minimize the needed mass flow rate, and of course an affordable cost. In the case of heat pumps, the minimum temperature and pressure for the refrigerant may be considerably higher, since heat is usually extracted from media that are well above the temperatures encountered in refrigeration systems. 2.2 Heat Pump Systems

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Heat pumps are generally more expensive to purchase and install than other heating systems, but considering their life cycle they lead to lower total costs, as they count for lower energy consumption. This explains why, in spite of their initial cost, the popularity of heat pumps is increasing. About one third of all single-family homes built in the United States in 1984 were using heat pumps. A study for southern European countries predicts that by 2005 every third household will also feature one. The most common energy source for heat pumps is atmospheric air (air-to-air systems), although water and soil are also used as energy basins. The major problem with air source systems is frosting, which occurs in humid climates when the temperature falls below 2 5°C. The frost accumulation on the evaporator coils is highly undesirable since it disables the heat transfer process. The coils can be defrosted, by reversing the heat pump cycle, that is, by running it as an air conditioner (in winter), but this results in a decline of the system’s efficiency. Watersource systems usually use well water from depths of up to 80 m in the temperature range of 5 18°C, and they do not have a frosting problem. They typically have larger COPs but are more complex, more expensive, and require easy access to a large reservoir of water, such as underground water. Soil-source systems are also rather complex, as they require long tubing (buried pipes) placed deep in the ground where the soil temperature is relatively constant. The COP of heat pumps usually ranges between 1.5 and 4, depending on the particular system used and the temperature of the source. Recently developed heat pumps, that imply variable-speed electric driven motors, are at least twice as energy efficient as its predecessors, in particular under partial load conditions. Both the capacity and the efficiency of a heat pump fall significantly at low temperatures. Therefore, most air source heat pumps require a supplementary heating system such as an electrical resistance heater or an oil or gas fired boiler. Since water and soil temperatures remain fairly constant, supplementary heating may not be required for water-source or soil-source systems. In this case the heat pump system must be fairly with respect to its capacity to cope with the maximum heating load, which therefore may result in questionable economics. Heat pumps and air conditioners have the same mechanical components. Therefore, it is not economical to have two separate systems to meet the heating and cooling requirements of a house or a building. The same system can be used as a heat pump in the winter and air conditioner in the summer, as shown in Figure 5. This is accomplished by adding a reversing valve to the cycle. As a result of this modification, the condenser of the heat pump (located indoors) functions as the evaporator of the air conditioner in summer. Similarly, the evaporator of the heat pump (located outdoors) serves as the condenser of the air conditioner. This feature increases the competitiveness of the heat pump. Such dual-purpose units, also known as window or split-units, have been commonly used in motels in the USA. Since the late eighties they have become particularly popular in retrofitting at domestic and public building in southern European countries.

Figure 5. Design principle of a heat pump

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Heat pumps are most competitive in areas that have large cooling loads during the cooling season and a relatively small heating load during the heating season. In these areas, the heat pump can meet the entire cooling and heating needs of residential or commercial buildings. The heat pump is least competitive in areas where the heating load is significant and the cooling load is small, such as in the northern parts of the Unites States or in the northern European countries. 2.3 Multistage Compression Refrigeration Systems When the fluid used throughout cascade refrigeration systems is the same, the heat exchanger between the stages can be replaced by a mixing chamber (called flash chamber), since it has better heat transfer characteristics. Such systems are called multistage compression refrigeration systems. A two-stage compression refrigeration system is shown in Figure 6.

Figure 6. Two-stage compression refrigeration system with flash chamber In this system, the liquid refrigerant expands in the first expansion valve to the flash chamber pressure. Part of the liquid vaporizes during the process. This saturated vapor (state 7) is mixed with the superheated vapor from the low-pressure compressor (state 2), and the mixture enters the high-pressure compressor at state 3. This is, in essence, a refrigeration process. The saturated liquid (state 7) expands through the second expansion valve into the evaporator, where it picks up heat from the refrigerated space. The compression process in this system resembles a two-stage compression with interceding cycles and hence, and the compressor work decreases. However, attention should be paid to the interpretations of the areas on the temperature entropy diagram in Figure 7, since the mass flow rates are not the same in all parts of the cycle.

Figure 7. Temperature versus entropy diagram of a two- stage refrigeration cycle 2.3.1 Example Consider a two-stage compression refrigeration system operating between the pressure limits of 0.8 and 0.14 MPa. The working fluid is refrigerant R 12. The refrigerant leaves the condenser as a saturated liquid and is throttled to a flash chamber operation at 0.32 MPa. Part of the refrigerant evaporates during this flashing process, and this vapor is mixed with the refrigerant leaving the low-pressure

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compressor. The mixture is then compressed to the condenser pressure by the high-pressure compressor. The liquid in the flash chamber is throttled to the evaporator pressure and cools the refrigerated space as it vaporizes in the evaporator. Assuming the refrigerant leaves the evaporator as a saturated vapor and both compressors are isentropic, it is possible to determine (a) the fraction of refrigerant which evaporates in the chamber, (b) the amount of heat removed from the refrigerated space and the compressor work per unit mass of refrigerant flowing through the condenser, and (c) the coefficient of performance COP. 2.3.2 Solution The two-stage compression refrigeration cycle is shown on a T-S diagram in Figure 8. The enthalpies of the refrigerant at various stages are determined from the R 12 tables and are indicated on this T-S diagram. The fraction of refrigerant which evaporates as it is throttled to the flash chamber is simply the equality at state 6, which is: (4) The amount of heat removed from the refrigerated space and the compressor work input per unit mass of refrigerant flowing through the condenser is: (5)

Figure 8. Temperature versus enthalpy diagram of a two- stage refrigeration cycle. The enthalpy at the state 9 is determined from an energy balance in the flash chamber:

(6)

Also s9 = 0.7074 kJ/(kg.K), thus the enthalpy at state 4 (0.8 MPa, s4 = s9) is h4 = 207.76 kJ/kg. Substituting Win = (1 - 0.2)[(191.97 - 177.87)kJ/kg] + (207.76 - 191.18)kJ/kg = 27.86kJ/kg the coefficient of performance is determined from:

(7)

(8) 2.4 Multipurpose Refrigeration System with a Single Compressor

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Some applications require refrigeration at more than one temperature. This can be accomplished by using a separate throttling valve and a separate compressor for each evaporator operating at different temperatures However, such systems will be bulky and probably too expensive. A more practical and economical approach will be to route all the exit streams from the evaporators to a single compressor that will handle the compression process for the entire system. The example of an ordinary refrigerator freezer-unit is presented schematically in Figure 9 and the temperature entropy diagram of its cycle is presented in Figures 10. Most refrigerated goods have a high water content, and the refrigerated space must be maintained above the ice point (at about 5°C) to prevent freezing. The freezer compartment however, is maintained at about −15°C. Therefore, the refrigerant should enter the freezer at about −25°C to have heat transfer at a reasonable rate to the freezer. If a single expansion valve and evaporator were used, the refrigerant would have to circulate in both compartments at about −25°C, which would cause ice formation in the neighborhood of the evaporator coils and dehydration of the produce. This would not be acceptable to a household. This problem can de eliminated by throttling the refrigerant to a higher pressure (hence temperature) for use in the refrigerated space and then throttling it to the minimum pressure for use in the freezer. The entire refrigerant leaving the freezer compartment is subsequently compressed by a single compressor to the condenser pressure.

Figure 9. Simplified schematic diagram of a refrigerator freezer unit

Figure 10. Temperature versus enthalpy diagram for a refrigeration-freezer unit with one compressor 3. The Absorption Cycle Absorption systems are characterized by the fact that the refrigerant is absorbed by the absorbent on the low pressure side of the system and is given up on the pressure side. Its major difference to the vapor compression cycle, presented in Figure 5 above, lies in the process by which the low-pressure refrigerant is transformed to a high-pressure vapor. In the vapor compression cycle the low-pressure vapor is being mechanically compressed. In an absorption cycle, however, the low-pressure vapor is absorbed into a solution at low-pressure condition, subsequently pumped to a higher pressure, and then heated to produce a high-pressure vapor. The advantage of the absorption cycle lies in the fact that as a liquid has been pumped to a higher pressure, less mechanical input is required than for the pumping of the gaseous medium in the vapor compression cycle. One can therefore consider the mechanical compression 10 of 21

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being replaced by a thermal or thermochemical compression. Figure 11 shows the schematic function of an absorption cycle. The refrigerant cycle (path 2 3 4 1) is identical in both the vapor-compression cycle and the absorption cycle. However, the compression function path (1 2) is replaced with a solution circuit and a pump. Hence, instead of having mechanical work Win as an input, thermal input (i.e., heat) Qin is being used.

Figure 11. Schematic representation of the absorption refrigeration cycle The main principle of operation can be summarized in the following elements: Evaporator. System water flows through the tubes. The refrigerant is sprayed over the surface of the tubes, it vaporizes and hence removes heat from the system water. Absorber. The refrigerant vaporized in the evaporator is absorbed by a spray of the absorbent. The heat given up as the vapor is condensed and absorbed is removed by the cooling water that flows through the tubes in the absorber. Generator. A heat source, like steam or solar heated water, flows through the tubes and provides the heat required to separate the refrigerant from the dilute solution coming in from the absorber. The refrigerant vapors are led to the condenser and the concentrated solution returns to the absorber. Condenser. Cooling water flowing through the tubes condenses the refrigerant coming from the generator. The condensed refrigerant flashes through an expansion valve (an orifice) back to the evaporator. The absorption refrigeration cycle is a process that can be feasible provided the absorbent features two properties: · ·

Low vapor pressure in the condenser, so that only the refrigerator will evaporate, and Low specific heat in order to minimize thermal losses in the process.

For these reasons two absorbent-refrigerant mixtures are used widely commercially: 1.

Lithium bromide (LiBr) water mixture, which is also known as aqueous LiBr mixture, and 2. Water ammonia mixture or aqueous ammonia mixture. There are absorption cycles that use other solution combinations and refrigerants, like methylene chloride and dimethol, but they are not commercially marketed. 3.1 The Lithium Bromide - Water Absorption Cycle One of the two main absorption cycles used for refrigeration is the basic lithium bromide-water (LiBr) absorption cycle, which has been in use since the 1940s in refrigeration applications. It is schematically presented in Figure 12. Its key

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characteristics are determined by the pressure-temperature-concentration (PTX) equilibrium chart like the one presented in Figure 13, the solution cycles, the individual components, and the chemical additives. Its functional performance is mainly characterized by the part-load operation and the maintenance of the system. The main points to be kept in mind are the following: · · ·

· · ·

In this cycle LiBr is the absorbent and water is the refrigerant. LiBr is a salt that can cause corrosion. Chemical additives are used to reduce or eliminate the corrosion potential in the systems components. Crystallization of the absorbent used to be a problem in the past, but has been reduced by using microprocessors and direct digital controls (DDC) for the systems operation. The system operates at high vacuum conditions. The output of an LiBr system can be as low as 4°C. Purge systems maintain the vacuum integrity of the machine and are also used to remove non condensables from the machine.

Figure 12. Schematic representation of the LiBr absorption refrigeration cycle LiBr absorption systems are categorized by the number of times the solution is heated to produce refrigerant vapors. This is referred to as the number of effects. Most LiBr absorption systems are single effect or double effect systems. A single effect system uses the input heat once, while a double-effect system uses the heat input for one deabsorption effect and subsequently uses the warm refrigerant vapors as the heat source for the second effect.

Figure 13. Pressure-Temperature-Concentration (PTX) diagram 3.2 Single Effect Systems This section describes the compressor function of a LiBr absorption machine in the absorption cycle. The condenser, expansion device, and evaporator components of the LiBr absorption machine were discussed in the previous section and perform the same function as those used in vapor-compression systems. The metering device is much simpler since the pressure differences are small. The metering can be achieved with an orifice and the height of refrigerant liquid. Condensers and evaporators in absorption systems typically require pumping and spray headers to ensure wetted surfaces and heat transfer. 3.3 Double Effect Systems A double-effect absorption machine has two stages of generation to separate the refrigerant from the absorbent. The temperature of the heat source required to drive

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the high stage generator must be higher than that used for a single-effect machine. Generally, a double-effect system is preferred to a single effect system when there is a heat source at a sufficient temperature to power a double-effect machine. The initial cost of a double-effect machine is higher than that of a single-effect machine, however, when recovered heat is utilized, the COP and the resulting cooling output are maximized with a double-effect machine. 3.4 The Aqueous Ammonia Absorption Cycle The second most widely used refrigeration system is based on the aqueous ammonia absorption cycle (AAR). This system is mainly met in industrial applications, in chemical processes, in the food industry, and in drying processes. It is also suitable for solar energy applications. The main points to be kept in mind concerning the AAR cycle are the following: · ·

·

· ·

Water is the absorbent and ammonia is the refrigerant (i.e., the opposite of the LiBr cycle). The cooling output is in most cases liquid ammonia, separated from the ammonia in the AAR machine, as opposed to the chilled water used in LiBr absorption systems. Due to the fact that ammonia is the refrigerant, the output of an AAR system can be as low as −60°C. The ammonia and water vapor are separated by using a fractional distillation process, because a small amount of water is converted to the vapor state within the temperatures and pressures prevailing in the absorption cycle. It is not possible to use copper in the construction of AAR systems, as copper and ammonia react adversely with each other. If the system is installed in a closed space, ventilation is required according to the standards for ammonia refrigeration.

AAR systems are categorized according to the number of evaporator stages. A single-stage system has one stage of evaporation/absorption, and a two-stage system has two stages of evaporation/absorption. The number of stages of evaporation can be increased in order to reach the cooling and the refrigeration temperatures required by the design parameters. AAR systems are tailor-made systems for each application. The operational principles of an AAR system are similar to those of LiBr absorption machines, as presented in Figure 14.

Figure 14. Schematic representation of the AAR cycle In an AAR absorption cycle, the low-pressure refrigerant vapor from the evaporator is converted to high-pressure refrigerant by using a thermal compressor rather than a mechanical compressor. As the properties of the refrigerant/absorbent solution cannot be illustrated directly in a P-H diagram, other charts and tables have been developed to aid in the analysis of the thermal compression cycle. There are several charts or tables available to determine the properties of the aqueous ammonia solution at the different locations in the AAR machine. There are three common methods used to obtain ammonia-water property data. 13 of 21

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Temperature-concentration charts contain information on solution concentration, solution temperature, saturation pressure, enthalpy of saturated vapor, vapor composition, and enthalpy of saturated liquid. A temperature-concentration chart is typically referred to as a Ponchon diagram. Enthalpy-concentration charts are typically used to describe ammonia-water property data at a single pressure. However, if the high and low pressures are known, the two pressures may be juxtaposed in the same diagram. · Property tables can be developed from the temperature-concentration and enthalpy-concentration charts. All these data can be found in the literature. 4. Gas Refrigeration Cycles The standard of comparison for all power cycles, i.e. the Carnot cycle, and its reverse, the reversed Carnot cycle, are identical except that the reversed Carnot cycle operates “backward.” This leads to the conclusion that the power cycles can be used as refrigeration cycles by simply reversing them. In fact, the vapor-compression refrigeration cycle is a essentially a modified Rankine cycle operating in reverse. Another example is the reversed Stirling cycle, which is the cycle on which Stirling refrigerators operate. Possibly the most interesting gas refrigeration cycle is the reversed Brayton cycle, presented in Figure 15. If the ambient is at a given Temperature T0 and the refrigerated space is to be maintained at a temperature TL, the gas is compressed during process 1 2, as shown in Figure 16. The high pressure and high temperature gas at state 2 is cooled at a constant pressure to T0 by rejecting heat to the ambient air. This is followed by an expansion process in a turbine, during which the gas temperature drops to T4. Finally the cool gas absorbs heat from the refrigerated space until its temperature rises to T1. All the processes described above are internally reversible, and the cycle executed is the ideal gas refrigeration cycle. In actual gas refrigeration cycles, the compression and expansion process will deviate from the isentropic ones, and T3 will be higher than T0 unless the heat exchanger is infinitely large. The COP for this cycle may be expressed as (9)

Figure 15. Schematic diagram of a gas refrigeration cycle

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Figure 16. Temperature versus entropy diagram of a gas refrigeration cycle Despite the low COP of the gas refrigeration cycles, these systems have two advantages: They consist of simple components with limited weight and are therefore particularly suitable for aircraft cooling. They can also incorporate regeneration, which makes them suitable for liquefaction of gases and cryogenic applications of extremely low temperatures. 5. Cryogenic Installations and Gas Liquefaction Cryogenic installations are used to produce “very cold” temperatures, a term arbitrarily used for temperature lower than -74°C (-100°F). Until the 1940s liquefied gases were mainly produced in laboratories. The liquefaction of gases has always been an important area of refrigeration because many important scientific and engineering processes depend on liquefied gases. Some examples of such processes are the separation of oxygen and nitrogen from air, preparation of liquid propellants for rockets, study of material properties at low temperatures, and study of phenomena such as superconductivity. At temperatures above the critical-point value, a substance exists in the gas phase only. Hence, a fluid’s temperature must be reduced below its critical temperature so that it will become liquid. The critical temperatures of helium, hydrogen, and nitrogen, three of the most commonly used liquefied gases, are −268, −240 and −147°C respectively. Such low temperatures cannot be produced with the refrigeration cycles and systems described in the above paragraphs. A solution to this problem is the liquefaction of gases by utilizing the “Joule Thompson” effect. According to this one can achieve the temperatures needed for gas liquefaction by throttling. In order to achieve a temperature drop by throttling, the initial temperature should be lower than the inversion temperature and the Joule Thompson coefficient positive. The maximum temperature drop occurs in this case if the throttling starts at a state on the inversion curve. Several cycles are used successfully for the liquefaction of gases. The two most commonly used are the Linde and the Claude liquefiers. 5.1 The Linde Liquefier The Linde liquefier is the simplest way for liquefying gases, as can be seen from its schematic operation principle in Figure 17.

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Figure 17. Schematic diagram of a Linde system The make up gas, which has to be cleaned of contaminants that may solidify during the process, is mixed with the uncondensed portion of the gas from the previous cycle, and the mixture at state 2 of the temperature-entropy diagram in Figure 18 is compressed by means of an isothermal process due to intercooling. The high-pressure gas is cooled, in an after-cooler by a cooling medium or by a separate external refrigerator system, to state 4. The gas is further cooled in a regenerative counterflow heat exchanger by the uncondensed portion of gas from the previous cycle to state 5, and it is throttled at an expansion valve to state 6, which is a saturated liquid-vapor mixture state. The liquid produced is collected, in state 7, as the final product and the vapor (state 8) is driven back, through the refrigerator, to cool the high pressure gas approaching the throttling valve. Finally, the gas is mixed with fresh makeup gas and the cycle is repeated. This cycle is also used for the solidification of gases, like CO2, whose critical state is at 31°C and 73.85 MPa.

Figure 18. Temperature versus enthalpy diagram of a Linde system for liquefying gases 5.2 The Claude Liquefier The Claude liquefier is derived from the Linde system with the following modification: the expansion valve of the Linde process is replaced by an expander, resulting in an increase of the produced liquid gas, as is shown in Figure 19. This change can also lead to the utilization of the work obtained by the expander, which can be a turbine or a reciprocating one, to drive the compressor, hence enabling higher overall performance of the system. In order to prevent damage to the expander due to the presence of two phases of the working medium, the expander receives the working medium at an intermediate state.

Figure 19. Schematic diagram of a Claude system 6. Heat Exchangers for Producing Low Temperatures In all the applicable cycles, heat exchangers are used to enable the extraction or addition of heat from and to the working medium. The same applies in cryogenic applications like the liquefaction of gases, as described in the previous section. The major difference in cryogenic heat exchangers is their operating temperature,

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which often lies below −100°C. At such temperatures the design, the construction, and the material properties of the exchangers have to be able to cope with effects like the contraction occurring due to the increased stresses caused by the low temperatures. The materials used must therefore be able to resist high stress changes and shock loads. For this, tough materials should be used. As most materials become brittle at temperatures below −30°C, they are liable to fracture at full stress. Hence materials should be chosen which do not become brittle at low temperatures, such as stainless steels, aluminum, copper, and copper-nickel alloys. Even if these materials fail, the risk of a sudden brittle fracture is limited, providing time for measures to avoid a complete failure. Another problem that has to be dealt with is the necessity of removing the moisture entirely, in order to prevent its freezing and the resulting destruction of the exchangers tubes. To avoid such risks, the units have to be dried before entering operation. This is usually achieved by circulating warm air through the exchanger. Thermal insulation is another feature that differentiates cryogenic heat exchangers from conventional ones. As the temperature difference between the heat exchanger and the environment is large, a very effective thermal insulation is necessary to reduce heat transfer from the environment to the system. In the case of an air separation plant, where the temperatures are around −200°C, a single layer insulation of low thermal conductivity can be used. In the case of very low temperatures, however, such as in the liquefaction of hydrogen or helium, techniques involving multiple insulation layers have to be used. These include vacuum layers around the heat exchanger, layers of higher boiling liquid above the vacuum, layers of low-conductivity packing powder over the liquid layer, and radiant insulating layers, like aluminum foil, around the packing powder layer. In such cases the insulation cost becomes a significant part of the heat exchanger’s total cost. Finally, the controls used in cryogenic heat exchangers, like valves and thermostats, have to be designed and constructed to cope with the low temperature conditions. In refrigeration units where the produced refrigeration must be transferred to various places, magnetic and constant pressure valves are mainly used. Temperature control is achieved by thermostats. Their performance depends not only on their features but also on the positions in which they are placed. 6.1 Types of Heat Exchangers used for Cryogenic Purposes The Linde and the Hamspon heat exchangers were the first types used for cryogenics, however their efficiency is no longer considered to be satisfactory. The most frequently used systems are instead: (a) The coil-tube exchanger, (b) The switching regenerator, (c) The reversing exchanger, and (d) Compact units. In certain applications one can also find heat pipes systems. A brief description of the four most commonly met exchanger types is given in the following sections. 6.2 Coil-Tube Heat Exchangers The coil-tube heat exchanger is an exchanger that consists of different groups of tube

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bundles that carry different liquids. These tube bundles are coiled around a core cylinder in a number of successive layers and end up at collectors at each end of the exchanger’s shell, as can be seen in Figure 20. The tubes form layers coiled in opposite directions and separated in order to allow the shell-side fluid to contact all the tubes. High-pressure gasses flow inside the tubes, while low-pressure cold gas flows on the shell side. Hence, if the exchanger is used as an evaporator, the refrigeration medium evaporates outside the tubes and the obtained gas or liquid, which is going to be refrigerated, is driven through the tubes. This kind of evaporator is used for liquid and gas refrigeration in the process industry.

Figure 20. Coil-tube heat exchanger (Courtesy of Linde A. G., 1977) The coil-tube heat exchanger has become widely used, as it features a series of advantages. Heat exchange is possible with more than two fluids and the exchanger can cope with high pressure and large temperature ranges. Large heat transfer areas of up to 25 000 m2 are possible. Despite this, the design is fairly compact. The shape and geometry of the exchanger can be adapted to requirements. There is no limitation to the size of the system, as far as its manufacturing is concerned. Contraction due to low temperatures can occur without inducing undue stresses, and vibration problems are not encountered. 6.3 Switching Regenerators Switching regenerators used for cryogenics operate in a way similar to those used for high temperature regeneration. The warm and cold streams of gasses are led periodically through towers, which contain high heat capacity and high surface area material, such as thin, corrugated strips of aluminum. The main reason for this exchanger type becoming popular in the process industry is that the manufacturing of very large systems is possible, enabling the large-scale production of liquefied gases. A further advantage lies in the fact that moisture and carbon dioxide can be removed from the feed stream by condensation and subsequently be evaporated into the outgoing waste stream. In order to achieve this the operating conditions have to be precisely controlled. 6.4 Reversing Heat Exchangers Reversing heat exchangers present an alternative to switching regenerators. They consist of a number of coaxial finned tubes, through which the gasses are led. The flow direction of the gases is reversed periodically in order to remove the impurities. This feature is the main advantage of reversing heat exchangers compared to switching regenerators, as one can produce extremely pure gases, like oxygen for medical purposes. A further advantage is that the size of reversing heat exchangers is much smaller than that of switching regenerators. Finally, the energy and the

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maintenance costs of reversing heat exchangers are lower than the respective costs of switching regenerators. However, their initial cost is higher. 6.5 Compact Heat Exchangers Compact heat exchangers, like the one shown in Figure 21, are made of aluminum or aluminum alloys. They feature high surface to volume ratio, enabling the construction of very compact units, with limited weight. Considering the fact that their heat transfer efficiency is very high, exceeding 90%, and that the yield strength (physical property of the material) of aluminum increases with lowering temperatures, they are suitable for applications where high efficiencies are needed at a limited volume and weight. These advantages offset the initial cost, which is rather high.

Figure 21. Compact heat exchanger (Courtesy of The Trane Company, 1984) Related Chapters Click Here To View The Related Chapters Glossary Brayton cycle

Carnot cycle

Critical point Enthalpy

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:was first proposed by George Brayton around 1870. It is used for gas turbines, which operate on an open cycle, where both the compression and expansion processes take place in rotating machinery. The open gas-turbine cycle can be modeled as a closed cycle by utilizing the air-standard assumptions. The combustion process is replaced by a constant-pressure heat-addition process from an external source, and the exhaust process is replaced by a constantpressure heat-rejection process to the ambient air. The ideal Brayton cycle is made up of four internally reversible processes 1-2 Isentropic compression (in a compressor) 2-3 Constant pressure heat addition 3-4 Isentropic expansion (in a turbine) 4-1 Constant pressure heat rejection :was first proposed in 1824 by French engineer Sadi Carnot, is composed of four reversible processes-two isothermal and two adiabatic, and can be executed either in a closed or a steady-flow system. :is defined as the point at which the saturated liquid and saturated vapor states are identical. :H (from the Greek word enthalpien, which means to heat) is a property and is defined as the sum of the internal energy U and the PV product. 6/2/2012 14:34

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Entropy Isentropic

Rankine cycle

Resistance heater Stirling cycle

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:(from a classical thermodynamics point of view) is a property designated S and is defined as dS=(dQ/T)int rev. :is the stagnation state when the stagnation process is reversible as well as adiabatic (i.e.,isentropic). The entropy of a fluid remains constant during an isentropic stagnation process. :is the ideal cycle for vapor power plants. The ideal Rankine cycle does not involve any internal irreversibilities and consists of the following four processes 1-2 Isentropic compression in a pump 2-3 Constant pressure heat addition in a boiler 3-4 Isentropic expansion in a turbine 4-1 Constant pressure heat rejection in a condenser :is an electricity driven air heater. :is made up of four totally reversible processes 1-2 T constant expansion (heat addition from the external source) 2-3 v constant regeneration (internal heat transfer from the working fluid to the regenerator) 3-4 T constant compression (heat rejection to the external sink) 4-1 v constant regeneration (internal heat transfer from the regenerator back to the working fluid) T-S relate the Tds product to other thermodynamic properties. The first Gibbs relation is Tds = du + Pdv. The second Gibbs relation is Tds = dh - vdP.

Bibliography Air Liquide. Catalogues of companies active in the refrigeration and cryogenics sector: BASF, LINDE, Trane. ASHRAE (1985). ASHRAE handbook Fundamentals. Atlanta: American Society of Heating Refrigerating and Air-Conditioning Engineers Inc. ASHRAE (1994). ASHRAE handbook Refrigeration. Atlanta: American Society of Heating Refrigerating and Air-Conditioning Engineers Inc. Cengel Y. A., and Boles M. A. (1994). Thermodynamics, An engineering approach, pp. 472 New York: Mc Graw Hill Inc. [This is a very comprehensive approach to gas power cycles.]

486.

Dorgan C. B., Leigth S. P., and Dorgan C. E. (1995). Application Guide for Absorption Cooling/Refrigeration using Recovered Heat, pp. 55 93. Atlanta: American Society of Heating Refrigerating and Air-Conditioning Engineers Inc. [This is an excellent handbook for applications in this field.] Faires V. M., and Simmang C. M. (1978). Thermodynamics, pp. 460 475. New York: MacMillan Publishing Co. Inc. [The basics of thermodynamic processes are presented in a synoptic and illustrating way.] Flynn T. (1996). Cryogenic Engineering, New York: Marcel Dekker Inc. [A comprehensive handbook on cryogenics, with emphasis being placed on the engineering aspects, that provides extensive data for most cryogenic problems.] Gupta J. P. (1989). Working with Heat Exchangers, pp. 262 267. Washington: Hemisphere Publishing Corporation. [A synoptic approach to the use of heat exchangers in the field of cryogenics.] Pittsburgh Corning Europe (1992). Foamglas Industrial Insulation Handbook. Waterloo: PC Europe. [A very comprehensive handbook on industrial insulation applications, with detailed case studies.]

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Schmidt F. W., Henderson R. E., and Wolgemuth C. H. (1993). Introduction to Thermal Sciences, pp. 268 285. John Wiley and Sons Inc. [The thermodynamics of heat exchangers are presented in this book, accompanied by an appendix that features tables, charts and data for most of the problems to be met in this field.] Shah R. K. (1990). Heat Exchanger Design. Washington: Hemisphere Publishing Corporation. Sotiropoulos V. A. (1986). Industrial Refrigeration, pp. 191 216 and 359 371. Giahoudis Giapoulis Press [In Greek. A textbook with well presented case studies of industrial refrigeration.] Stoecker W. F., and Jones J. W. (1982). Refrigeration and Air-Conditioning. New York: McGraw Hill.

Biographical Sketches Agis M. Papadopoulos. Born in Thessaloniki (GR) in 1966, he graduated as a Diploma Mechanical Engineer (Dipl.-Eng.) from the Aristotle University of Thessaloniki in 1989. After a Master of Science in Energy Conservation and the Environment at the School of Mechanical Engineering of Cranfield University (UK) in 1990, he earned his Doctorate on the feasibility of solar systems at the Department of Mechanical Engineering at the Aristotle University of Thessaloniki in 1994. Between 1994 and 1998 he was a Teaching Associate, teaching Energy and Business Economics at the Department of Industrial Engineering at the University of Thessaly in Volos. Between 1995 and 1998 he was a visiting Professor on Energy Resources Management at the Department of Business Administration at the University of Macedonia, Thessaloniki. In 1998 he was appointed as Assistant Professor at the Department of Mechanical Engineering at the School of Engineering, Aristotle University of Thessaloniki, in the area of Energy Systems, Economics and Policies. He has participated in a series of national and European research projects and is author or co-author of more than 50 papers and 3 textbooks on energy management and conservation issues. His research topics are energy conservation and the rational use of energy in buildings and also energy resources economics and management. He is an active member of the Hellenic Technical Chamber (TEE), the Hellenic Society of Operational Research (EEEE), the German Society for Rational Energy Use (GRE), and the International Association of Energy Economics (IAEE). Christopher J. Koroneos studied Chemical Engineering in Columbia University where he also earned his Doctorate. His research interests are in the areas of Environmental Engineering, Energy Engineering, Process Engineering, Environmental Process Synthesis, and Life Cycle Analysis. At the present time he is a Teaching Associate and Visiting Professor at the Laboratory of Heat Transfer and Environmental Engineering of the Mechanical Engineering Department of Aristotle University of Thessaloniki, in Greece. His professional experience includes being Associate Professor at the Department of Chemical Engineering at Columbia University and head of Program Development of Earth Engineering Center at Columbia University. He has many years experience in the industry working as Senior Research Engineer, Process Engineer, Process Development Engineer, and Consultant. He has multiple professional affiliations and many professional awards.

To cite this chapter A. M. Papadopoulos and C. J. Koroneos, (2004), REFRIGERATION AND CRYOGENIC SYSTEMS, in Air Conditioning - Energy Consumption and Environmental Quality, [Ed. Matheos Santamouris], in Encyclopedia of Life Support Systems (EOLSS), Developed under the Auspices of the UNESCO, Eolss Publishers, Oxford ,UK, [http://www.eolss.net] [Retrieved February 6, 2012]

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