refrigeration petrochemical

SECTION 14 Refrigeration Selection of a refrigerant is generally based upon temperature requirements, availability, economics, and previous experience For instance, in a natural gas processing plant, ethane and propane may be at hand; whereas in an olefins plant, ethylene and propylene are readily available Propane or propylene may not be suitable in an ammonia plant because of the Refrigeration systems are common in the natural gas processing industry and processes related to the petroleum refining, petrochemical, and chemical industries Several applications for refrigeration include NGL recovery, LPG recovery, hydrocarbon dew point control, reflux condensation for light hydrocarbon fractionators, and LNG plants FIG 14-1 Nomenclature BP GHP h h′VD k m n P Q r S T W X η σ ρ ∆h = = = = = = = = = = = = = = = = = = Brake Power, kW gas power defined by Eq 14-7b enthalpy, kJ/kg isentropic enthalpy, kJ/kg specific heat ratio refrigerant flow, kW number of stages pressure, kPa (abs) heat duty, kW compression ratio entropy, kJ/(kg • K) temperature, °C work of compression, kW weight fraction isentropic efficiency surface tension, N/m density, kg/m3 enthalpy change, kJ/kg Subscripts L = liquid state V = vapor state A, B, C, D = denote unique points of operation on P-H diagrams b = bypass i = isentropic cd = condenser ref = refrigeration s = suction d = discharge n = stage number T = Total 1, = stage number DEFINITIONS OF WORDS AND PHRASES USED IN REFRIGERATION Effect, refrigerating: the rate of heat removal by a refrigerant in a refrigeration system It is equal to the difference in specific enthalpies of the refrigerant at two designated thermodynamic states Expansion valve: a valve for controlling the flow of refrigerant to an evaporator or chiller Flash gas: the gas resulting from the instantaneous evaporation of refrigerant by a pressure reducing device, such as a control valve Frost Plug: small diameter closed nozzle protruding from the side of an insulated vessel which indicates liquid level in the vessel by accumulation of frost Halocarbons: a family of refrigerants consisting of fluorinated and/or chlorinated hydrocarbons Hot gas bypass: warm discharge gas recycled to chiller for maintaining system’s operating integrity at minimum load conditions Liquid refrigerant receiver: a vessel in a refrigeration system designed to ensure the availability of adequate liquid refrigerant for proper functioning of the system and to store the liquid refrigerant when the system is pumped down Refrigerant: the fluid used for heat transfer in a refrigeration system, which absorbs heat at a low temperature and low pressure and rejects heat at a higher temperature and a higher pressure Accumulator: a storage vessel for liquid refrigerant; also known as surge drum Bubble point: the temperature at which the vapor pressure of the liquid refrigerant equals the absolute external pressure of the liquid-vapor interface Capacity, refrigerating system: the cooling effect produced by the total enthalpy change between the refrigerant entering the evaporator and the refrigerant leaving the evaporator Chiller, evaporator: a heat exchanger in which the liquid refrigerant is vaporized by a process stream which is in turn cooled Compression ratio: ratio of outlet to inlet absolute pressures for a compressor Condenser: a heat exchanger in which the refrigerant, compressed to a suitable pressure, is condensed by rejection of heat to a cooling medium Cooling medium: any substance whose temperature is such that it is used, with or without change of state, to lower the temperature of refrigerant either during condensing or subcooling 14-1 risk of contamination, while ammonia may very well serve the purpose Hydroflourocarbons have been used extensively because of their nonflammable characteristics Fig. 14-1 provides a nomenclature for this section followed by a glossary MECHANICAL REFRIGERATION can be defined as X (hVB – hLB) and substituting from Eq. 14-2, the effect becomes: Effect = hVB – hLA The refrigeration duty (or refrigeration capacity) refers to the total amount of heat absorbed in the chiller by the process, generally expressed as kW The refrigerant flow rate is given by: Refrigeration Cycle The refrigeration effect can be achieved by using one of these cycles: • Vapor compression-expansion • Absorption • Steam jet (water-vapor compression) By utilizing the Pressure-Enthalpy (P-H) diagram, the refrigeration cycle can be broken down into four distinct steps: • Expansion • Evaporation • Compression • Condensation The vapor-compression refrigeration cycle can be represented by the process flow and P-H diagram shown in Fig 14-2 Qref m = (hVB – hLA) Since point B is inside the envelope, vapor and liquid coexist In order to determine the amount of vapor formed in the expansion process, let X be the fraction of liquid at pressure PB with an enthalpy hLA The fraction of vapor formed during the expansion process with an enthalpy hVB is (1-X) Equations for the heat balance and the fraction of liquid formed are: Eq 14-5 Compression Step — The refrigerant vapors leave the chiller at the saturation pressure PC The corresponding temperature equals TC at an enthalpy of hVB The entropy at this point is SC These vapors are compressed isentropically to pressure PA along line C – D´ (Fig. 14-2) The isentropic (ideal) work, Wi, for compressing the refrigerant from PB to PA is given by: Wi = m (h´VD – hVB) Eq 14-6 The quantity h´VD is determined from refrigerant properties at PA and an entropy of SC Since the refrigerant is not an ideal fluid and since the compressors for such services not operate ideally, isentropic efficiency, ηi, has been defined to compensate for the inefficiencies of the compression process The actual work of compression, W, can be calculated from: Wi m (hVD – hVB) W = η = = m (hVD – hVB) ηi i Expansion Step — The starting point in a refrigeration cycle is the availability of liquid refrigerant Point A in Fig 14-2 represents a bubble point liquid at its saturation pressure, PA, and enthalpy, hLB In the expansion step, the pressure and temperature are reduced by flashing the liquid through a control valve to pressure PB The lower pressure, PB, is determined by the desired refrigerant temperature, TB (point B) At point B the enthalpy of the saturated liquid is hLB, while the corresponding saturated vapor enthalpy is hVB Since the expansion step (A – B) occurs across an expansion valve and no energy has been exchanged, the process is considered to be isenthalpic Thus the total stream enthalpy at the outlet of the valve is the same as the inlet, hLA Eq 14-4 Eq 14-7 The enthalpy at discharge is given by: (h´VD – hVB) h + hVB VD = ηi Eq 14-7a The work of compression can also be expressed as: W BP = 3600 Eq 14-7b where 3600 kJ/h = kW (See Section 13 for a discussion on compressors.) Condensation Step — The superheated refrigerant leaving the compressor at PA and TD (Point D in Fig. 14-2) is cooled at nearly constant pressure to the dew point temperature, TA, and refrigerant vapors begin to condense at constant temperature (X) hLB + (1 – X) hVB = hLA Eq 14-1 (hVB – hLA) X = (hVB – hLB) Eq 14-2 During the desuperheating and condensation process, all heat and work added to the refrigerant during the evaporation and compression processes must be removed so that the cycle can be completed by reaching Point A (the starting point) on the P-H diagram, as shown in Fig. 14-2 (hLA – hLB) (1 – X) = (hVB – hLB) Eq 14-3 By adding the refrigeration duty to the heat of compression, we calculate the condensing duty, Qcd, from: Evaporation Step — The vapor formed in the expansion process (A-B) does not provide any refrigeration to the process Heat is absorbed from the process by the evaporation of the liquid portion of the refrigerant As shown in Fig. 14-2, this is a constant temperature, constant pressure step (B-C) The enthalpy of the vapor at point C is hVB Physically, the evaporation takes place in a heat exchanger referred to as an evaporator or a chiller The process refrigeration is provided by the cold liquid, X, and its refrigerant effect Qcd = m [(hVB – hLA) + (hVD – hVB)] = m (hVD – hLA) Eq 14-8 The condensing pressure of the refrigerant is a function of the cooling medium available — air, cooling water, or another refrigerant The cooling medium is the heat sink for the refrigeration cycle Because the compressor discharge vapor is superheated, the refrigerant condensing curve is not a straight line It is a combi- 14-2 nation of desuperheating and constant temperature condensing This fact must be considered for proper design of the condenser Two-Stage System — Savings in the 20% range can often be achieved with a two-stage refrigeration system and interstage flash economizer Additional savings can be realized by removing process heat at the interstage level rather than at the low stage level A typical two-stage system with an intermediate load is shown in Fig 14-5 with data for pure propane System Pressure Drop — Some typical values for pressure drops that must be considered are: Condenser pressure drop Line hydraulic losses Evaporator to Compressor* Compressor to Condenser Condenser to Receiver 20 to 50 kPa Three-Stage System — Additional power savings can be achieved by using a three-stage compression system As with a two-stage system, flash economization and/or an intermediate heat load can be used The savings, while not as dramatic as the two stage versus one-stage, can still be significant enough to justify the additional equipment A typical three stage propane system is shown in Fig. 14-6 0.7 to 10 kPa 0.7 to 14 kPa 3.5 to   kPa *  This is an important consideration in refrigeration services with low suction pressure to compressor System Configuration — Energy consumption is frequently reduced as the number of stages is increased For a propane refrigeration system, Fig. 14-7 illustrates the effect of Refrigeration Stages Refrigeration systems utilizing one, two, three, or four stages of compression have been successfully operated in various services The number of levels of refrigeration generally depends upon the number of compression stages required, interstage heat loads, economics, and the type of compression FIG 14-3 One-Stage Refrigeration System One-Stage System — A typical one-stage refrigeration system is shown in Fig. 14-3 where the data are for pure propane refrigerant Fig. 14-4 illustrates a process application of a single level chiller and the associated cooling curve FIG 14-2 Process Flow Diagram and Pressure-Enthalpy Diagram FIG 14-4 Single-Stage Cooling, Chilling and Heating Curves 14-3 FIG 14-5 Two-Stage Refrigeration System FIG 14-6 Three-Stage Refrigeration System 14-4 interstages without using refrigeration at intermediate levels However, the installation cost of such refrigeration systems increases as the number of stages increases The optimum overall cost will be a function of the specific system and has to be determined for a set of economic criteria (26.4) (106) m 1 = = 77 650 kg/h (820 – 480) (10.6) (106) 2 = m = 44 170 kg/h (870 – 630) The compression power for refrigeration can be reduced further by shifting refrigerant load from cooler levels to warmer levels Fig. 14-8 shows a refrigeration system using two levels of chilling The gas is initially chilled to –1°C with –4°C propane and then to –37°C with –40°C propane The selection of the – 4°C level results from equal compression ratios for each stage The interstage pressure and corresponding refrigerant temperature may be fixed by either equipment or process conditions Equal compression ratios per stage are chosen whenever possible to minimize horsepower where m1 is the flowrate through the first stage chiller, and m2 is the flowrate through the second stage chiller Liquid flow to the first-stage chiller (77 650 kg/h) is provided by flashing the liquid refrigerant from the refrigerant receiver at 49°C and bypassing the second-stage chiller In order to determine the flow of liquid refrigerant from the receiver, consider the heat and material balances shown in Fig. 14-9 Here, let mb (kg/h) denote the refrigerant bypassing the second-stage chiller The chiller produces 44 170 kg/h of refrigerant vapor at –4°C These vapors flow through the second stage suction drum, and leave overhead The liquid required from the second stage flash drum for the first stage chiller comes from the quantity mb Example 14-1 — Calculate the power and condenser duty required for the process shown in Fig. 14-8 using propane refrigeration Design condensing temperature is 49°C The pressure drop from the chillers to the compressor suction is 10 kPa The pressure drop from compressor discharge to the receiver is 70 kPa By material balance, we find the vapors leaving the second stage suction drum as mb + 44 170 – 77 650 or mb – 33 480 kg/h By heat balance around the suction drum, we can determine the amount of refrigerant, mb: Solution Steps: In order to determine the interstage refrigeration level for a two-stage system, determine the ratio per stage: Pd ⁄n = r  Ps  (mb – 33 480) (870) + (77 650) (480) = mb (630) + (44 170) (870) mb = 126 180 kg/h Eq 14-9 From the propane vapor pressure curve: Pd = 1670 kPa (abs) + 70 kPa = 1740 kPa (abs) FIG 14-8 Ps = 108 kPa – 10 kPa = 98 kPa (abs) Two-Level Chilling, Two-Stage Cooling System 1740 ⁄2 r =  98  = 4.21 Thus the second stage suction pressure is: Ps2 = (98) (4.21) = 412 kPa The first stage discharge pressure is: Pdl = 412 + = 420 kPa From the vapor pressure curve for propane, the refrigeration temperature at 420 kPa (abs) is –4°C Converting kW duties to kJ/h and substituting enthalpy values from Section 24, into Eq 14-5, we find the refrigerant flowrate through each chiller: FIG 14-7 Effect of Staging on a Propane Refrigeration System Stages, n Refrigeration Duty, kW 293 293 293 Refrigeration Temperature, °C –40 –40 –40 Refrigerant Condensing   Temperature, °C 38 38 38 Compression Requirements, kW 218 176 167 Reduction in BP, % Base 19.2 23.3 Condenser Duty, kW 511 469 462 Change in condenser duty, % Base –8.2 –9.6 14-5 In order to calculate isentropic work for the first stage, it is necessary to determine the isentropic enthalpy at 412 kPa (abs) Fig. 24-20, the first stage inlet entropy equals 3.85 kJ/kg • K, and the corresponding isentropic enthalpy at 412 kPa (abs) is 890 kJ/kg Hence, the compression required for the two-stage propane refrigeration system becomes: BPT = 2013 + 4732 = 6745 kW Using Eq 14-7a, the second stage discharge enthalpy is: The ideal change in enthalpy = 890 – 820 = 70 kJ/kg For propane refrigerant k = 1.13, compression ratio, r, of 4.21 and the isentropic efficiency, ηi of 0.75, the required compression power for the first stage is obtained from Eq 14-7b: BP1 = (70) (77 650) (0.75) (3 600) = 2013 kW Using Eq 14-7a we determine the first stage discharge enthalpy is: 70 hvld = + 820 = 913 kJ/kg 0.75 A material balance around the second compression stage yields the total refrigerant flow: mT = m1 + (mb – 33 480) = 77 650 + (126 180 – 33 480) = 170 350 kg/h A heat balance at the second compression stage entrance yields the second stage inlet enthalpy: (913) (77 650) + (870) (126 180 – 33 480) hv2s = (170 350) = 890 kJ/kg From Section 24, the inlet entropy at 412 kPa (abs) and 890 kJ/kg is 3.85 kJ/(kg • K), and the isentropic enthalpy at 1740 kPa (abs) is 965 kJ/kg Substituting into Eq 14-6, the ideal enthalpy change across the second stage as: ∆h = 965 – 890 = 75 kJ/kg The required compression power for the second stage is determined from Eq 14-7b: (75) (170 350) 2= BP = 4732 kW (0.75) (3600) 75 V2d = H + 890 = 990 kJ/kg 0.75 Substituting into Eq 14-8 yields the condenser duty for the two-stage propane refrigeration system: Qcd = (990 – 630) (170 350) = (6.133) (107) kJ/h = 17 036 kW From Section 24 the second stage discharge temperature at 1740 kPa and enthalpy of 990 kJ/kg is 80°C Condensing Temperature Condensing temperature has a significant effect on the compression power and condensing duty requirements Mehra3 illustrated the effect of the condensing temperature on refrigeration requirements for one, two, and three stage systems Results for a one-stage propylene refrigeration system are summarized in Fig. 14-10 Fig. 14-10 illustrates that the colder the condensing temperature, the lower the power requirements for a given refrigeration duty Traditionally, the heat sinks for most refrig- eration systems have been either cooling water or ambient air If cooling water or evaporative condensing is utilized, a 27 to 38°C temperature can be achieved Section 11 provides wet and dry bulb temperature data Fig. 14-10 also indicates, to a certain extent, the effect on operations between summer and winter conditions as well as between day and night operations Refrigerant Subcooling Subcooling liquid refrigerants is common in refrigeration systems Subcooling the refrigerant reduces the energy requirements It is carried out when an auxiliary source of cooling is readily available, and the source stream needs to be heated Subcooling can be accomplished by simply installing a heat exchanger on the appropriate refrigerant and process streams FIG 14-9 FIG 14-10 Data for Heat and Material Balances Effect of condensing Temperature Condensing   Temperature, °C 16 27 38 49 60 Refrigeration Duty,   kW 293 293 293 293 293 Refrigeration   Temperature, °C –46 –46 –46 –46 –46 Compression   Requirement, kW 157 199 248 320 413 Change in BP, % –36.6 –19.8 Base 28.8 66.4 Condenser Duty,   kW 451 492 539 613 709 –16.3 –8.7 Base 13.6 31.5 Change in Condenser   Duty, % 14-6 frigeration system has to be designed to provide a total of 15 740 kW at –40°C in addition to 2930 kW at –20°C and 2050 kW at 7°C Example 14-2 — Consider installing an 880 kW subcooler on the liquid propane refrigerant from the receiver at 49°C in Example 14-1 for the two-stage propane refrigeration system The second stage of this system is shown in Fig. 14-11 Freon (CFC) Refrigerant Phase Out Solution Steps: Clorinated fluorocarbons (commonly called Freon) have been used for many years as effective refrigerants in many applications However, the stability of these compounds, coupled with their chlorine content, has linked them to the depletion of the earth’s protective ozone layer As a result, these compounds have been phased out of production and usage globally Hydrofluorocarbons (HFC) have been developed as an alternative By performing the heat balance around the subcooler and the second stage suction drum, the liquid refrigerant flowrate to the subcooler is determined to be 139 515 kg/h When comparing this to the earlier flowrate of 170 350 kg/h, the refrigerant flow is reduced by 30 835 kg/h By heat balance around the subcooler, we determine the enthalpy of liquid propane refrigerant leaving the subcooler is 615 kJ/kg which corresponds to a temperature of 43°C Refrigerant HFC-410a has been developed to replace chlorodifluoromethane (R-22 ) This compound is reasonably close to R-22 in performance Fig 14-12 shows a comparison of HFC410a and R-22 The refrigerant power requirement is quite similar but the operating pressures are higher for HFC-410a The flowrate of refrigerant through the second stage chiller becomes (10.6) (106) m2 = = 41 569 kg/h (870 – 615) As a result of subcooling, the flow of refrigerant through the second stage chiller has been reduced from 44 170 kg/h to 41 569 kg/h The lower flowrates result in reduced compression power, condenser duty, and reduced size of piping and equipment These benefits must be balanced against the installed cost of the subcooler exchanger HFC-134a has been developed to replace dichlorodifluoromethane (R-12) This compound is reasonably close to R-12 in performance but differences in equipment design and operation must be taken into account in the replacement Fig 14-13 shows a comparison of HFC-134a and R-12 for an example application One of the important differences is the higher compression ratio necessary for this refrigerant Refrigerant Properties Refrigerant For Reboiling Physical properties of pure component refrigerants in common use are given in Fig. 14-15 The vapor pressure curves for ethane, ethylene, propane, prpylene, and Refrigerant 22 (R-22) are available in Sections 23 and 24 or references 2, 5, 9, and 10 Figs 14-35 through 14-37 contain properties for HFC-410a Properties for HFC-134a are given in Figs 14-38 through 14-40 References 12 and 13 contain additional data for these refrigerants Refrigerants have been successfully used for reboiling services wherever applicable conditions exist Reboiling is similar in concept to subcooling — heat is taken out of the refrigeration cycle In reboiling service, the heat removed from the refrigerant condenses the refrigerant vapor at essentially constant temperature and pressure The liquid refrigerant produced in a reboiler service is flashed to the next lower pressure stage to produce useful refrigeration The refrigerant condensing pressure is a function of the reboiling temperature Enthalpy data are necessary in designing any refrigeration system Pressure-enthalpy diagrams for pure ethane, ethylene, propane, propylene, and R-22 are available in Section 24 of this data book or references 2, 5, 9, and 10 References 12 and 13 contain additional information for these refrigerants Refrigerant Cascading In the cascading of refrigerants, warmer refrigerants condense cooler ones Based on the low temperature requirements of a process, a refrigerant that is capable of providing the desired cold temperature is selected For example, the lowest attainable temperature from ethane refrigerant is –85°C (for a positive compressor suction pressure), whereas the lowest temperature level for propane is –40°C (for a similar positive pressure) FIG 14-11 Refrigerant Subcooling In a refrigeration cycle, energy is transferred from lower to higher temperature levels economically by using water or ambient air as the ultimate heat sink If ethane is used as a refrigerant, the warmest temperature level to condense ethane is its critical temperature of about 32°C This temperature requires unusually high compression ratios — making an ethane compressor for such service complicated and uneconomical Also in order to condense ethane at 32°C, a heat sink at 29°C or lower is necessary This condensing temperature is a difficult cooling water requirement in many locations Thus a refrigerant such as propane is cascaded with ethane to transfer the energy from the ethane system to cooling water or air An example of a cascaded system is shown in Fig. 14-14, where an ethane system cascades into a propane system The condenser duty for the ethane system is 9000 kW This duty becomes a refrigeration load for the propane system along with its 6740 kW refrigeration at –40°C Therefore, the propane re14-7 Fig 14-12 Comparison Example of R-22 and HFC-410* R-22 HFC-410a Psuction, kPa (abs) 214.4 353 Pdisch 1965 3068 Compr Ratio 9.19 8.7 kW/MW 78.6 78.9 Condenser Load kJ/kJ 1.52 1.51 *–23°C Chiller, 49°C Condensing, 69 kPa (ga) Condenser DP FIG 14-13 Theoretical Cycle Comparison of R-12 and HFC-134a* Capacity (as % of R-12) Compressor Exit temperature °C Exit pressure, kPa (abs) Compression ratio R-12 HFC-134a 100 99.7 86.8 1348.6 83.1 1473.4 4.1 4.7 *Conditions: Condenser, 54.4°C; Evaporator, 1.7°C; Comp Suction, 26.7°C; Expansion Device, 51.7°C After defining the lowest refrigerant level and the condensing temperature, the pressure at the evaporator and condenser can be established from the vapor-pressure curve for a specific refrigerant All examples and data in this section are based upon pure component properties In practice, pure hydrocarbon refrigerants are not always available Impurities may cause significant deviations in design and performance One-Stage Systems — Figs. 14-16 through 14-20 provide data for estimating gas power and condenser duty requirements for one-stage refrigeration systems using ethylene, propane, propylene, R-22, and HFC 410a refrigerants Two-Stage Systems — The data for estimating gas power and refrigerant condenser duty requirements for two-stage refrigeration systems utilizing ethylene, propane, propylene, R-22, and HFC 410a are shown in Figs. 14-22 through 14-26 Three-Stage Systems — The data for estimating gas power and condenser duty requirements for three-stage refrigeration systems utilizing ethylene, propane, propylene, and R-22 are presented in Figs. 14-27 through 14-31 Example 14-3 — Estimate the power and condenser duty requirements for a single stage propylene refrigeration system that will provide 26.4 (106) kJ/h of process chilling at a refrigerant level of –29°C Solution Steps The unit BP for this example from Fig. 14-19 is 565 kW per MW of refrigeration duty at an evaporator temperature of –29°C and a condenser temperature of 38°C And, from Fig. 14-19, the condenser duty factor equals 1.565 MW per MW of refrigeration duty for the same evaporator and condenser temperatures Hence, the total power and condenser duty are: Power and Condenser Duty Estimation Since many gas processing plants require mechanical refrigeration, generalized charts5 were developed to aid in a modular approach for designing refrigeration systems Because of the complexity of generalizing refrigeration systems, the charts have been developed for four of the most common refrigerants: ethylene, propylene, propane, and Refrigerant 22 In order to apply these curves to most of the commercially available compressors, a polytropic efficiency of 0.77 was assumed The polytropic efficiency was converted into an isentropic efficiency1 to include the effects of compression ratio and specific heat ratio (k = Cp/Cv) for a given refrigerant For well balanced and efficient operation of the compressor, an equal compression ratio between stages was employed The refrigeration level is defined as the temperature of the dew point vapor leaving the evaporator The pressures at the compressor suction and side load inlet nozzles were adjusted by 10 kPa to allow for pressure drop These charts also include a 35 kPa pressure drop across the refrigerant condenser for ethylene, and a 70 kPa drop for propane, propylene, and Refrigerant 22 Before developing any system, one must define refrigerant temperature and condensing temperature of the refrigerant based on the medium used for condensing To achieve maximum energy conservation and minimum energy cost, it is necessary to match the process conditions and refrigeration compressor design to obtain the best efficiency BP = (565) (7.325) = 139 kW Qcd = (7325) (1.565) = 11 464 kW Heat Exchanger Economizing — An alternative to flash economizing of the refrigeration cycle is to use a heat exchanger to accomplish an economizing step Fig 14-21 shows an example economizer using a heat exchanger The heat exchanger is a chiller which uses some of the condensed refrigerant to subcool the balance of the condensed refrigerant stream The refrigerant used for the chilling is then fed to the interstage (or second stage) of the refrigeration compressor The subcooled refrigerant is then used for process chillers The subcooled refrigerant produces less unusable vapor when flashed to suction drum conditions than a refrigerant stream that is not subcooled Thus the use of the heat exchanger effectively shifts vapor from the low stage of compression to the high stage, thus saving power The resultant process impact is very similar to the flash economization previously discussed Design and Operating Considerations The following are some of the important parameters that should be considered while designing any refrigeration system to provide a safe, reliable, and economical operation Oil Removal — Oil removal requirements from evaporators are related to the type of the refrigerant, lubricant, evaporator, and compressor used in the refrigeration cycle Fig. 14-32 illustrates the application of an oil reclaimer in a propane refrigerant cycle In order to remove oil from the refrigerant, a 14-8 Liquid Surge and Storage — All refrigeration systems should have a liquid surge and storage vessel, commonly called a receiver A surge vessel is required on all systems where the operating charge in the evaporator(s) and the condenser(s) varies due to variable load conditions In addition to accommodating a varying refrigerant charge, the receiver drains the condenser(s) of liquid so that the effective condensing surface is not reduced by liquid backing up The refrigerant charge in a receiver may vary over a wide range, from a minimum at full load to a maximum at no load slip stream of refrigerant from the bottom of the chiller is drained into the reclaimer where hot propane refrigerant from the compressor discharge is used to evaporate the refrigerant into the compressor suction The oil is removed from the bottom of the reclaimer Similar arrangements can be utilized for other hydrocarbon and ammonia refrigerants Operation may be designed for either manual or automatic Where halocarbon refrigerants and/or synthetic lubricants are employed, it is imperative that the oil reclaimer system be approved by the compressor manufacturer FIG 14-14 Cascade Refrigeration System 14-9 Systems with inadequate surge vessels often cause problems as they lose the liquid seal due to load variations that always occur Surge vessels or receivers are relatively inexpensive and when sizing them, consideration should be given to: (1) a volume equal to 100% of the system inventory at 80% full level, and (2) the availability and quantity of refrigerant makeup rocating, centrifugal, and screw compressors These systems will suffer from the same corrosion problems as defined above, but to a lesser extent However, since they are generally used at lower temperatures, water in the system can freeze the control valve and in the evaporator Refrigerant dryers are required in these systems A good purge system is also required Vacuum Systems — Refrigeration systems can operate with a suction pressure below atmospheric pressure These vacuum systems require special considerations: • A  mmonia has been employed with reciprocating, centrifugal, and screw compressors in vacuum service for many years Since water will not freeze in the presence of ammonia and the aqua-ammonia formed is only slightly corrosive, this type system has few problems during operation A good purge system is recommended  here hydrocarbons are used with reciprocating com• W pressors (which employ rod “packing”), air can enter the compressor and possibly form a hazardous mixture Extreme care should be taken where such systems are used These systems must have a manual or automatic purge system Double acting packing should be employed Considerations for Vacuum Refrigeration Systems: • W  here hydroflourocarbons (such as HFC-134a, HFC410a, and other low pressure, high volume refrigerants are employed with centrifugal compressors, the deep vacuums may “draw” air and moisture through flanges, seals, etc This water-oxygen combination in the presence of halocarbons forms acid and causes “crevice corrosion” of the tubes along with some other problems A positive purge system must be employed and frequent monitoring of the moisture content in the refrigerant is suggested Eliminate all flanges where possible Weld all piping Use weld in-line valves Use steel “refrigeration type” stop valves with “back seating” feature and seal caps in lieu of hand wheels All suction line valves should be angle valves to reduce pressure drop • H  igh pressure hydroflourocarbons HFC-134a, HFC-410a, and others are employed in vacuum systems with recip- FIG 14-15 Physical Properties of Common Refrigerants1, 4, 9, 10, 11 ASHRAE Refrigerant Number Chemical Name Chemical Formula Molecular Mass Normal Freezing Boiling Critical Critical Point °C Point °C Temperature Pressure @ 101.325 @ 101.325 °C kPa (abs) kPa (abs) kPa (abs) Liquid Viscosity mPa • s Liquid Thermal Conductivity W/(m • °C) Specific Toxicity Heat UL Group Ratio Classification k = Cp/Cv 11 Trichlorofluoromethane CC13F 137.4 23.8 198.0 4413 –111 0.421 @ NBT 0.395 @ 30°C 0.0876 @ NBT 0.0862 @ 30°C 1.13 114 Dichlorotetrafluroethane CC1F2OC1F2 170.0 3.6 145.7 3268 –94 0.44 @ NBT 0.32 @ 30°C 0.0701 @ NBT 0.0633 @ 30°C 1.09 –158 0.358 @ NBT 0.206 @ 30°C 0.0897 @ NBT 0.0678 @ 30°C 1.14 –160 0.33 @ NBT 0.192 @ 30°C 0.1203 @ NBT 0.0857 @ 30°C 1.18 5a 0.1147 @ NBT 0.1056 @ 30°C 1.09 5b 12 Dichlorodifluoro methane 22 Chlorodifluoro methane CC12F2 CHC1F2 120.9 86.5 –29.8 –40.8 112.0 96.0 4116 4937 600 N-Butane C4H10 58.1 –0.5 152.0 3797 –138 0.213 @ NBT 0.159 @ 30°C 290 Propane C3H8 44.1 –42.1 96.7 4249 –187 0.21@ NBT 0.101 @ 30°C 0.1315 @ NBT 0.0969 @ 30°C 1.14 5b Propylene C3H6 42.1 –47.7 91.7 4600 –185 0.15 @ NBT 0.089 @ 30°C 0.1419 @ NBT 0.0987 @ 30°C 1.15 5b –183 0.168 @ NBT 0.039 @ 30°C 0.1419 @ NBT 0.0831 @ 30°C 1.19 5b 0.1921 @ NBT 0.0537 @ 30°C 1270 170 Ethane C2H6 30.1 –88.6 –12.8 4880 1150 Ethylene C2H4 28.1 –103.8 9.2 5041 –169 0.17 @ NBT 0.07 @ 30°C 1.24 5b 50 Methane CH4 16.0 –161.5 –82.6 4604 –182 0.118 @ NBT 0.1904 @ NBT 1.305 5b –78 0.25 @ –15°C 0.207 @ 30°C 0.5019 @ 0°C 0.5019 @ 0°C 1.29 717 Ammonia NH3 17.0 –33.3 132.4 11 280 410a HFC-410a CH2F2/ CHF2CF3 72.58 –51.6 72.13 4926.1 0.336 @ NBT 0.112 @ 30°C 0.127 @ NBT 0.087 @ 30°C 134a HFC-134a CHF2CF3 102.03 –26.1 101.1 4065.2 0.39 @ NBT 0.205 @ 30°C 0.109 @ NBT 0.087 @ 30°C 14-10 FIG 14-23 Two-Stage Ethylene Refrigeration System 800 750 700 650 600 500 450 1.80 400 350 1.70 -45 -40 -35 -30 1.60 300 Refrigerant Condensing Temperature, °C 250 1.50 200 1.40 -45 -40 -35 -30 150 1.30 100 1.20 50 1.10 1.00 -100 -95 -90 -85 -80 -75 -70 -65 -60 -55 Evaporator Temperature, °C 14-16 -50 -45 -40 -35 -30 MW of Condenser Duty per MW Refrigeration Duty Gas Power per MW Refrigeration Duty, kW 550 FIG 14-24 Two-Stage Propane Refrigeration System 800 600 Q1 60 500 55 50 400 45 40 300 Refrigerant-condensing temperature,°C 35 Gas Power per MW Refrigeration Duty, kW 700 30 25 200 20 15 100 -40 -30 -10 -20 10 20 30 2.0 1.9 MW of Condenser Duty per MW of Refrigration Duty 1.8 1.7 1.6 1.5 60 Refrigerant-condensing temperature °C 55 1.4 1.3 1.2 1.1 1.0 50 45 40 35 30 25 20 15 Evaporator Temperature, °C 14-17 40 50 60 FIG 14-25 Two-Stage Propylene Refrigeration System 1000 900 700 Q1 60 55 600 50 500 45 40 400 35 30 Refrigerant-condensing temperature,°C 300 25 20 Gas Power per MW Refrigeration Duty, kW 800 15 200 100 -40 -30 -20 -10 10 20 2.0 1.9 1.8 MW of Condenser Duty per MW of Refrigration Duty 60 1.7 55 1.6 1.5 50 Refrigerant-condensing temperature °C 45 40 1.4 1.3 35 30 25 1.2 1.1 20 15 1.0 Evaporator Temperature, °C 14-18 30 40 50 60 Centrifugal Compressors — At the normal process temperatures encountered in gas processing, a three or four wheel centrifugal compressor is normally required for refrigeration service This offers the opportunity of utilizing multiple interstage flash economizers and permits multiple chiller temperature levels for further reductions in power Centrifugal compressor capacity is controlled by speed variation or suction or discharge pressure throttling Discharge throttling can cause surge It is also possible to recirculate refrigerant discharge vapors to the compressor suction during operation at lower loading in order to avoid surge problems Such recirculation results in wasted power and is one of the primary drawbacks to utilizing centrifugal units For more details on Centrifugal Compressors, refer to Section 13 Reciprocating Compressors — Process temperatures gen­ erally dictate two stage compression in a reciprocating machine This affords the opportunity for one interstage economizer, and also one additional level of chilling In a conventional refrigeration system, the first stage cylinder is normally quite large as a result of the low suction pressure The economizer also reduces first stage volume, cylinder diameter, and consequently rod load Capacity adjustment is accomplished by speed variation, FIG 14-26 variable clearance on the cylinders, valve lifters, and recirculation of refrigerant vapor to the suction As with centrifugal compressors, recirculation does result in wasted power It is also possible to throttle the refrigerant suction pressure between the chiller and compressor in order to reduce cylinder capacity However, suction pressure control can result in wasted power and the possibility of below atmospheric suction pressure, which should be avoided For more details on Reciprocating Compressors, refer to Section 13 Screw Compressors — Screw compressors have been used in refrigeration systems for many years They can be employed with all refrigerants The limitation for suction pressure is about 21 kPa with standard discharge pressures at 2400 kPa Discharge pressures of over 5000 kPa are also available Screw compressors are gaining popularity in the gas processing industry Screws can operate over a wide range of suction and discharge pressures without system modifications There are essentially no compression ratio limitations with ratios up to 10 being used They operate more efficiently in the to ratio and are comparable in efficiency to reciprocating compressors within this range Automatic capacity control can provide capacity adjustments from 100% down to 10% with comparable reduction in power requirements Condenser Duty and Gas Power for Two Stage HFC-410a Refrigerant FIG 14-27 800 Condenser Duty and Gas Power for Three Stage R-22 Refrigerant 600 1.70 400 1.65 Condensing Temperature, °C 500 35°C40°C 30°C 45°C MW of Condenser Duty per MW of Refrigeration Duty Gas Power per MW Refrigerant Duty, kW 700 50°C 300 200 100 -40 -30 -20 -10 10 20 30 40 1.60 1.55 1.50 1.45 1.40 1.35 50 1.30 CONDENSING TEMPERATURE,°C 45 40 1.25 1.20 35 1.15 30 1.10 25 1.05 Evaporator Temperature, °C 1.00 -40 -30 -20 -10 10 20 30 40 50 Evaporator Temperature,°C 1.80 1.60 Condensing Temperature, °C 1.50 1.40 Gas Power per MW Refrigeration Duty, kW MW of Condenser duty per MW Refrigerant Duty 1.70 35°C40°C 30°C 45°C 50°C 1.30 1.20 1.10 1.00 -40 -30 -20 -10 10 20 30 700 650 600 550 500 450 400 350 300 350 300 250 200 150 100 50 -40 50 CONDENSING TEMPERATURE,°C 45 40 35 30 25 -30 -20 -10 10 20 40 Evaporator Temperature,°C Evaporator Temperature, °C 14-19 30 40 50 FIG 14-28 Three-Stage Ethylene Refrigeration System Q1 14-20 FIG 14-29 Three-Stage Propane Refrigeration System 800 600 Q1 500 °C 60 °C 55 400 °C 50 Gas Power per MW Refrigeration Duty, kW 700 45 Refrigerant condensing temperature °C 300 40 °C 35 °C 200 30 °C 25 °C 100 20 °C 15 °C -30 -40 -10 -20 10 20 30 40 50 1.9 1.8 MW of Condenser Duty per MW of Refrigration Duty 1.7 60 °C 55 °C 1.6 1.5 50 °C 45 °C 40 °C 1.4 35 °C Refrigerant condensing temperature 30 1.3 °C 25 °C 1.2 20° C 15° C 1.1 1.0 Evaporator Temperature, °C 14-21 60 FIG 14-30 Three-Stage Propylene Refrigeration System 1000 800 Q1 700 °C 60 600 °C 55 500 °C 50 Gas Power per MW Refrigeration Duty, kW 900 Refrigerant condensing temperature 45 400 °C 40 °C 300 35 °C 30 °C 200 25 °C 100 20 °C 15 °C -40 -30 -20 -10 10 20 30 40 2.2 2.1 MW of Condenser Duty per MW of Refrigration Duty 2.0 1.9 1.8 1.7 1.6 60 °C Refrigerant condensing temperature 55 1.5 °C 50° 1.4 1.3 C 45° C 40° C 35° 1.2 1.1 1.0 C 30° C 25° C 20° C 15° C Evaporator Temperature, °C 14-22 50 60 Screw compressors normally operate at 3600 rpm direct coupled to motor drives However, they can operate over a range of speeds from 1500 to 4500 rpm Engine drives, gas turbines, and expanders can also be used as drivers Rotary Compressors — There is a limited application for large rotary compressors This is the low-temperature field in which the rotary serves the purpose of a high volume low-stage or booster compressor These booster compressors are applied at saturated suction conditions ranging from –87°C to –21°C with R-12, R-22, ammonia, and propane refrigerants Available units range in power from to 450 kW and in displacement from to 102 m3/min in a single unit The most common conventional refrigerants, HFC-410a and propane, exhibit atmospheric boiling temperatures of –51.6°C and –42°C, respectively Lower temperatures can be obtained utilizing propylene, ethane, and ethylene, which have atmospheric boiling temperatures of –48°C, –89°C, and –104°C, respectively However, these refrigerants require the use of a cascade system because condensation at ambient temperatures is FIG 14-32 Oil Reclaimer To compresso r suction Mixed Refrigerants Cryogenic processes which remove heat below ambient temperature generally use pure compounds as refrigerants in a closed mechanical refrigeration system However, when it is not necessary to remove the heat at a practically constant temperature, it may be advantageous to use a mixture of refrigerants In a proper composition, a mixed refrigerant can minimize temperature differences between the process stream and the refrigerant during heat exchange This match provides an efficient chilling system FIG 14-31 Refrigerant chiller PI Hot propane vapor from compressor discharge Oil reclaimer To refrigeratio n interstage Waste oil Condenser Duty and Gas Power for Three Stage HFC-410a Refrigerant FIG 14-33 700 Process Chilling Curves Gas Power per MW Refrigerant Duty, kW 600 500 Condensing Temperature, °C 400 45°C 30°C 35°C 40°C 50°C 300 200 100 -40 -30 -20 -10 10 20 30 40 20 30 40 Evaporator Temperature, °C 1.70 MW of Condenser duty per MW Refrigerant Duty 1.60 1.50 Condensing Temperature, °C 1.40 45°C 30°C 35°C 40°C 50°C 1.30 1.20 1.10 1.00 -40 -30 -20 -10 10 Evaporator Temperature, °C 14-23 not feasible One alternative is the use of a mixed refrigerant; for example, ethane-propane The ethane lowers the evaporation temperature while still permitting condensation at ambient temperatures, albeit at considerably higher pressures The shape of the refrigerant vaporization curve is a function of the composition of the mixed refrigerant In Fig. 14-33 the composition of the mixed refrigerant is methane 8 mol %, ethylene 37 mol %, and propane 55 mol % Fig. 14-33 compares the shape of process chilling curves for an ethylene refrigerant cycle with a mixed refrigerant cycle.6 Some of the design parameters7 to be considered while evaluating the application of a mixed refrigerant cycle include: • Compressor suction pressure • Shape of vaporization curve FIG 14-34 • Compressor discharge pressure and compression ratio Refrigeration System Checklist Indication High Compressor Discharge Pressure High Process Temperature Inadequate Compressor Capacity Inadequate Refrigerant Flow to Economizer or Chiller • Type of controls • Type of compressor Causes Check accumulator temperature If the accumulator temperature is high, check:   Condenser operation for fouling   High air or water temperature   Low fan speed or pitch   Low water circulation If condensing temperature is normal, check for:   Non-condensables in refrigerant   Restriction in system which is creating pressure drop Check refrigerant temperature from chiller If refrigerant temperature is high and approach temperature on chiller is normal, check:   Chiller pressure   Refrigerant composition for heavy ends contamination   Refrigerant circulation or kettle level (possible inadequate flow resulting in superheating of refrigerant)   Process overload of refrigerant system If refrigerant temperature is normal,, and approach to process temperature is high,, check:   Fouling on refrigerant side (lube oil or moisture)   Fouling on process side (wax or hydrates)   Process overload of chiller capacity Check:   Process overload of refrigerant system   Premature opening of hot gas bypass   Compressor valve failure   Compressor suction pressure restriction   Low compressor speed Check:   Low accumulator level   Expansion valve capacity   Chiller or economizer level control malfunction   Restriction in refrigerant flow (hydrates or ice) Mixed refrigerants present the problem of component segregation with the lighter components concentrating in the receiver, and the heavier components concentrating in the chiller unless the refrigerant is totally vaporized Because of the high condensing pressure, mixed refrigerants significantly increase the power per ton of refrigeration Chillers Kettle Type Chiller — The most common type of chiller employed in the gas processing industry is the kettle type The refrigerant is expanded into the shell of the kettle where a level is maintained to completely submerge the process tube bundle A level control maintains the proper amount of refrigerant in the kettle When using a kettle type chiller, care should be taken to provide adequate vapor disengaging space above the operating level of liquid refrigerant This type chiller improperly designed and operated is probably the largest single cause of compressor failure due to liquid carryover The following equation allows approximation of allowable refrigerant load: Allowable refrigerant load in kg/h per m vapor space (S.F.) (ρV) (503 700) σ = ρ – ρ (0.869)  L V Eq 14-10 where S.F = Safety Factor = 1⁄2 HTRI has a detailed method for determining the chiller sizing taking into account the vapor space requirement Plate-Fin Chillers — Modern cryogenic plants frequently employ plate-fin exchangers for gas cooling and chilling When the design calls for a hot gas-gas exchanger, a gas chiller, and a cold gas-gas exchanger in sequence, then it may be convenient to put these services in single plate-fin exchanger Also, platefin exchangers offer significant savings for low temperature application where stainless steel is needed for shell and tube units Significant pressure drop savings can be realized by using single or multiple units for chilling services 10 For other types of heat exchangers, refer to Sections and System Controls Level Controls — External cage (displacer-type) level controls are the most commonly used in refrigeration services and are probably the most reliable and easy to maintain instruments However, because the float chamber is external to the refrigeration system, it is imperative that the float chamber and connecting lines to the chiller be adequately sized and well insulated Vaporization of refrigerant (due to heat leak) in the 14-24 float chamber can result in difficulty in maintaining proper level Internal float level controls eliminate this problem, but present some problems in instrument maintenance Where evaporative condensers are used, several methods can be employed to control condensing pressure depending upon the ambient temperature and type of installation A differential pressure device is also frequently used for chiller level control; it affords good control when properly installed The high pressure side connection from the liquid phase should be large, well insulated, and installed in such a way that lubricating oil cannot accumulate and cause erroneous readings The low pressure side connection to the vapor phase should be uninsulated and possibly even liquid sealed or heated to prevent liquid accumulation The same problem exists in level indication External gauge glasses should have large connecting lines to the chillers and good insulation Bull’s-eye sight glasses are much better for direct indication of chiller level and normally not present any maintenance problem other than cleaning the glass Frost plugs are sometimes used and give an approximate level indication while requiring no maintenance Pressure Controls — Refrigerant compressor high suction pressure control may be desirable when there are multiple refrigerant compressors in the system Without high suction pressure control, loss of one refrigerant compressor can result in overloading of the other compressors and loss of all units in an unattended operation However, suction pressure control can also result in power waste if the compressor suction is throttled unnecessarily The refrigerant compressor hot gas bypass is used to prevent compressor suction pressure from getting too low If the process load decreases, the hot gas bypass will open to maintain a satisfactory compressor suction pressure in an unattended plant If hot gas bypass remains open, the compressor capacity should be adjusted to reduce bypassing in order to conserve energy Screw compressors need no such arrangements as they can be automatically unloaded to satisfy the suction pressure settings Evaporator Temperature — The evaporator (or chiller) temperature is normally controlled by controlling the refrigerant pressure on the chiller This may be accomplished by using back pressure valves, refrigerant compressor speed, or hot gas bypass around the compressor Low Ambient Controls — All refrigeration systems should have low ambient controls where ambient temperature is below 4°C These controls, which maintain a preset pressure differential between the condenser and the evaporator pressures, are necessary for continuous operation and for start-up at low ambients There are several approaches to these controls: For air coolers used as condensers, louvers, air recirculation systems, and fan cycling are employed For both shell and tube condensers and air coolers, condensing pressure can be controlled by installing a pressure regulating valve actuated by condensing pressure set at a minimum predetermined pressure in the line between the condenser and the liquid receiver In addition, a small pressure regulator set at a predetermined pressure is installed in a line between the discharge line and the liquid receiver This regulator will direct enough hot gas to the receiver to keep the pressure high enough to operate the evaporator liquid control valve Where a shell and tube condenser is used, a water flow control valve operated by condenser pressure can be utilized This type control may cause sediment and scaling in the condenser • The condenser can be selected to operate as an air cooler at temperatures below 0°C (water system shut down and drained) while employing a fan cycling controller • A system as described above in item 2 can be employed • Where the system is indoors with the condenser outdoors, an indoor water sump can be employed with a fan cycling controller Refrigerant System Troubleshooting Figure 14-34 contains a check list for troubleshooting refrigeration systems This is not an exhaustive list but rather a handy guide to prompt inspection of the system depending on the operating problems experienced ABSORPTION REFRIGERATION Even though absorption refrigeration has seen little use in the gas processing industry, it does have application In areas where there is low cost natural gas, where a low level heat source is available, or where electrical rates have risen dramatically, absorption refrigeration may be an economical way to attain modest temperature level refrigeration In circumstances where unused boiler capacity is available in summer months, absorption units can be utilized to produce refrigeration Lithium Bromide-Water Systems The lithium bromide absorption refrigeration cycle8 operates on the simple principle that, under low absolute pressure, water will boil at a low temperature Fig. 14-41 shows a schematic arrangement of lithium bromide-water system The system uses heat to efficiently produce refrigeration The lower shell is divided into absorber and evaporator sections while the upper shell consists of the generator and condenser sections The evaporator section contains the refrigerant, water A coil, through which the cooling system water circulates, is inserted into the evaporator to establish a heat exchange The refrigerant gains heat from the cooling system water, and because of low pressure maintained in the evaporator, quickly reaches saturation temperature and vaporizes, cooling the system water The remainder of the cycle deals with reclaiming this refrigerant The affinity of lithium bromide for water causes the refrigerant vapor to be absorbed by the strong solution in the absorber section The diluted (weak) solution is pumped into the generator, where steam or hot water is used to drive the water out of the solution as a vapor The vapor passes into the condenser and changes back to liquid which returns to the evaporator to be reused Meanwhile, the strong solution left in the generator flows back to the absorber to complete the cycle The lowest chilled water temperature achieved by this system is 6°C and typically the unit operates between 6°C and 10°C with varying refrigeration capacity loads Aqueous Ammonia System Refrigeration can be provided by using waste heat with the water-ammonia absorption cycle This cycle was originally em- 14-25 ployed in the 1800’s and has been refined over the years It lost its economic value in the 1930’s as the more efficient centrifugal and reciprocating compressor systems became inexpensive Due to its basic inefficiency, the ammonia absorption system cannot be justified unless low level waste heat is available, such as low pressure steam or hot process streams FIG 14-36 Vapor Thermal Conductivity of HFC-410a at Atmospheric Pressure13 19 Ammonia absorption capacities have been designed in sizes from a minimum of 703 kW at –46°C and 1055 kW at –7°C to a maximum of 8792 kW at –46°C and 17 584 kW at –7°C Most systems would employ shell and tube condensers and absorbers; 18 Suva® 410a 17 FIG 14-35 16 Thermal Conductivity, mW/m, °C Pressure vs Temperature of HFC-410a13 10000 100 Vapor Pressure, kPa 1000 15 R22 14 13 12 11 10 10 -100 -80 -60 -40 -20 20 40 –20 Temperature, °C Courtesy of Dupont HFC-410a Refrigerants 20 40 60 Temperature, °C 80 100 120 Courtesy of Dupont HFC-410a Refrigerants FIG 14-37 Pressure Enthalpy Diagram for HFC-410a13 Courtesy of Dupont HFC-410a Refrigerants 14-26 FIG 14-39 FIG 14-38 12 Vapor Thermal Conductivity of HFC-134a at Atmospheric Pressure12 Pressure vs Temperature of HFC-134a FIG 14-40 Pressure-Enthalpy Diagram for HFC-134a12 14-27 FIG 14-41 Lithium Bromide-Water Refrigeration System Cooling Water Out Condenser Generator Excess Process Heat (Waste heat) Capacity Control Valve Out Chilled Water In Cooling Water In Evaporator Absorber Cycle-Guard Valve Solution Heat Exchanger Hermetic Refrigerant Pump Hermetic Solution Pump Diluted LiBr Concentrated LiBr Refrigerant Water FIG 14-42 Flow Sheet of an Ammonia Absorption System Cooling water NH3 vapor Condenser NH3 vapor Fluid to be chilled Evaporator Rectifier Aqua NH3 NH3 liquid NH3 liquid NH3 liquid Waste heat Aqua film absorber NH3 vapor Ammonia receiver Generator NH3 liquid Bottoms Weak aqua Reflux pump (if required) Strong aqua Strong aqua Aqua pump Strong aqua Weak aqua 14-28 Strong aqua tank Cooling water however, evaporative cooled absorbers have been used Air cooled condensers and absorbers could also be used Various schemes can be used for supplying the water requirements of the condensers and absorbers other than series flow Parallel flow can be used to reduce absorber size and the heat input to the system The heat source will govern the generator design The generator may be finned-surface heat exchangers with aqueous solution pumped through the tubes for vapor heating mediums or double-pipe heat exchangers for liquid heating mediums Fig. 14-42 shows a flow sheet of an ammonia absorption system Reliability — Ammonia absorption systems are normally installed with spare aqua pumps and spare reflux pumps offering a comparison to centrifugal, reciprocating, and screw compressor systems that have a spare compressor-motor train Downtime from failure of mechanical items is negligible due to the 100% spare pumps Design Flexibility — Ammonia absorption systems are usually custom designed for each specific application Evaporator temperatures down to –51°C are possible Systems can be designed one stage or two stage for several different evaporator temperatures Systems can be increased in size and evaporator temperatures raised or lowered by the addition of heat exchange surface Evaporator temperatures are related to heat input temperature Raising the temperature of the heat source lowers the possible evaporator temperature The evaporators remain 100% efficient at all times as the refrigeration is oil-free There is no need to add oil fouling factors to the evaporator design, thus saving to 10% in the evaporator cost The choice of evaporator design has no limitations Applications — The ammonia absorption system has many applications It can produce refrigeration from waste heat for almost any kind of application in the chemical and petroleum industry Waste steam has been used as the heat source in many installations in the chemical and petroleum industry providing temperatures from 10°C to –46°C Process vapor streams and hot oil have also been used as heat sources Exhaust gases from gas turbines would be an excellent source of heat and this heat would normally be capable of providing low temperature refrigeration due to its high temperatures Supplemental firing can also be added for peak loads Stack gases of many kinds could also be used as a heat source REFERENCES Elliott Multistage Compressors, Bulletin P-25A, Elliott Co., Jeanette, PA, 1975 Starling, K E., “Fluid Thermodynamic Properties for Light Petroleum Systems,” Gulf Publishing, Houston, TX, 1973 Mehra, Y R., “Refrigeration Systems for Low-Temperature Processes,” Chem Eng., July 12, 1982, p. 94 Sibley, H W., “Selecting Refrigerants for Process Systems,” Chem Eng., May 16, 1983, p. 71 Mehra, Y R., “Refrigerant Properties of Ethylene, Propylene, Ethane and Propane,” Chem Eng., Dec. 18, 1978, p. 97; Jan. 15, 1979, p. 131; Feb. 12, 1979, p. 95; Mar. 26, 1979, p. 165 Kaiser, V., Becdelievre, C and Gilbourne, D. M., “Mixed Refrigerant for Ethylene,” Hydro Processing, Oct. 1976, p. 129 Kaiser, V., Salhi, O and Pocini, C., “Analyze Mixed Refrigerant Cycles,” Hydro Processing, July 1978, p. 163 8 Carrier Hermetic Absorption Liquid Chillers, Form 16JB-3P, Carrier Corporation, Syracuse, NJ, 1975 E I DuPont de Nemours & Co., Bulletins G-1, C-30, S-16, T-11, T-12, T-22, and T-114D, Wilmington, DE 19898 10 ASHRAE “Thermodynamic Properties of Refrigerants,” 1791 Tullie Circle N.E., Atlanta, GA 30329 11 Underwriters’ Laboratories Reports MH-2375, MH-3134, MH2630, and MH-3072 12 E.I Dupont de Nemours & Co., “Dupont™ HFC-134a Properties, Uses, Storage and Handling.” 13 E  I Dupont de Nemours & Co., “DupontTM HFC-410a Properties, Uses, Storage and Handling.” 14-29 NOTES: 14-30 ... Two-Stage Refrigeration System FIG 14-6 Three-Stage Refrigeration System 14-4 interstages without using refrigeration at intermediate levels However, the installation cost of such refrigeration. .. effect of Refrigeration Stages Refrigeration systems utilizing one, two, three, or four stages of compression have been successfully operated in various services The number of levels of refrigeration. .. by a glossary MECHANICAL REFRIGERATION can be defined as X (hVB – hLB) and substituting from Eq. 14-2, the effect becomes: Effect = hVB – hLA The refrigeration duty (or refrigeration capacity)