Technical Paper #2. Refrigeration Piping: A Simplified Guide to a Modern Approach

Transcription

1 Technical Paper #2 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz Paul Orlando Toromont Process Systems Houston, Texas Abstract The paper examines issues faced by engineers, contractors and owners when designing, installing and operating refrigeration piping. It reviews piping mechanical design by discussing applicable piping codes and the suitability of piping, castings, bolting and gasket materials as well as welding methods for low-, normal- and high-temperature service. It then analyzes pipe-sizing methods for a variety of applications, emphasizing the effects of undersizing or oversizing liquid piping, vapor lines and control valves. This evaluation is followed by a discussion on piping layout issues, in particular, layout of compressor suction mains, wet suction returns, elevated equipment and condensers. The paper concludes with a presentation of special cases such as thermal expansion in liquid CO2 and subcooled refrigerant lines, vapor condensation in hot gas and instrument tubing, liquid hammer, liquid carryover, internal and external corrosion, and a comparison of pumped and gravity-fed overfeed systems IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico IIAR

2 ACKNOWLEDGEMENT The success of the technical program of the 27th Annual Meeting of the International Institute of Ammonia Refrigeration is due to the quality of the technical papers in this volume. IIAR expresses its deep appreciation to the authors, reviewers, and editors for their contributions to the ammonia refrigeration industry. Board of Directors, International Institute of Ammonia Refrigeration ABOUT THIS VOLUME IIAR Technical Papers are subjected to rigorous technical peer review. The views expressed in the papers in this volume are those of the authors, not the International Institute of Ammonia Refrigeration. They are not official positions of the Institute and are not officially endorsed. EDITORS M. Kent Anderson, President Chris Combs, Project Coordinator Gene Troy, P.E., Technical Director International Institute of Ammonia Refrigeration 1110 North Glebe Road Suite 250 Arlington, VA (voice) (fax) Ammonia Refrigeration Conference & Exhibition Fairmont Acapulco Princess Acapulco, Mexico

3 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando Introduction Numerous articles, handbooks and manuals in our industry discuss good piping practices for ammonia refrigeration systems. There is also a good understanding of the major issues related to the subject among the designers, contractors, and endusers of the refrigeration plants. The goal of this paper is not to repeat what has already been written on the topic of ammonia piping. Instead, the paper shall discuss the design issues that, in the authors opinion, are much ignored by the industry but contribute significantly to the problems encountered in refrigeration plants. The paper also presents modern piping design methods. Mechanical Design Issues This section covers the following mechanical design topics: Piping design for low temperatures Cast, malleable and nodular iron in refrigeration systems ASME piping classes ASHRAE 15 safety standard for refrigeration systems Piping Design for Low Temperatures Carbon steel is the material of choice for pipe, valves and other components used in ammonia refrigeration systems and screw compressor lubricant service. At low temperatures, however, carbon steel becomes brittle and loses flexibility and toughness. For safety reasons, piping codes establish guidelines to ensure that the materials used in piping systems provide adequate strength and flexibility at all service temperatures. The code most commonly used in industrial refrigeration piping is ASME B31.5, Refrigeration Piping Code, although some end users, especially outside the food Technical Paper #2 IIAR

4 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico processing industry, require designers and installers to conform to the more stringent ASME B31.3, Process Piping Code. (ASME B31.3, 2004; ASME, 2001) Prior to the 2001 edition, the ASME B31.5 code stated that impact tests were not required for ferrous material used in fabricating a piping system operating between -20 F and -150 F [-29 C and -101 C], provided that the maximum circumferential or longitudinal tensile stress, resulting pressure, thermal stresses, and bending stress between the supports did not exceed 40% of the allowable stress of the material at 100 F [38 C]. The latest edition of ASME B31.5 brought the requirements more in line with the methodology used in B31.3 and Section VIII of the ASME Boiler and Pressure Vessel (B&PV) Code. (ASME B&PV, 2004) The new rules establish the minimum temperature above which the material is exempt from impact test requirements based on the grade of carbon steel and the thickness of the material. For operating temperatures between -20 F and -50 F [-29 C and -46 C], if the system is to operate below the nominal impact-test exemption temperature, the code allows the designer to determine the additional reduction in minimum metal temperature for which impact testing is not required. This reduction in minimum temperature is based on the ratio of the design tensile stresses to the normal (room temperature) allowable stress for a given material. (Figure 1) For metal temperatures below -50 F [-46 C], the old 40% rule applies (i.e., impact testing is not required if this ratio is less than 40%.) Note that if the design temperature is below -55 F [-48 C] and the ratio is less than 30%, ASME B31.3 requires impact testing. Typically, for higher operating temperatures, designers specify SA-106 grade B seamless or SA-53 grade B ERW piping. For the pipe wall thicknesses typically encountered in refrigeration piping, these materials have a maximum allowable stress rated at temperatures above -20 F [-29 C]. For lower operating temperatures, the specifications quite often call for SA-333 grade 6 material, which is impact-tested by the producing mill at a minimum temperature of -55 F [-48 C]. 36 IIAR 2005 Technical Paper #2

5 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando Suppliers charge a substantial premium for impact-tested carbon steel pipe and fittings. A recent informal survey showed that SA-333 grade 6 piping had a 10% to 20% higher cost than SA-106 grade B seamless piping, and a 40% to 50% higher cost than SA-53 grade B ERW. The premium is even higher for large forgings used to manufacture flanges, unions and other butt-welded fittings. Figure 2 details a price comparison among the different types of material. For packaged refrigeration systems, the cost of low-temperature carbon piping as a percentage of total job costs is not as large as for field-installed systems. The piping runs are typically short on packaged units and use of low-temperature specialty piping material is usually limited to compressor suction lines. The cost becomes significant for packaged units if the customer is concerned that the piping is suitable for the normal boiling temperature (saturation temperature) of the refrigerant at atmospheric pressure and specifies that all piping shall be rated for this temperature. In the case of ammonia, the equivalent saturation temperature at sea-level atmospheric pressure is F [-33.3 C]. For R-22, the equivalent saturation temperature is F [-40.7 C] and for R-507, F [-46.7 C]. In the packaged refrigeration systems supplied to refineries, hydrocarbons such as propane and propylene are quite often the refrigerants of choice; their saturation temperatures are F and F [-39.1 C and C] respectively. Some customers require that the piping system be rated for these low temperatures at full design working pressures. The situation is dramatically different for ammonia refrigeration systems in typical food processing installations, which have hundreds of feet of large diameter pipe, valves and fittings operating at metal temperatures below the minimum -20 F [-29 C]. At times, large overfeed systems have substantial wet suction lines or liquid lines that operate at temperatures of -40 F [-40 C] or lower. Valve manufacturers use either cast/ductile iron or impact-tested cast steel materials to ensure that their products can operate at suitably low temperatures. This same Technical Paper #2 IIAR

6 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico practice applies to manufacturers of pressure regulators, level switches, etc. However, to meet piping code requirements, system suppliers must either install expensive impact-tested piping or prove that the piping system will not be overstressed at low temperatures considering the reduced maximum allowable stresses of the piping material. In refrigeration systems, operating pressures at low temperatures are typically very low because the vapor pressure of the refrigerant falls as the temperature drops. For example, ammonia vapor pressure at 100 F is psia [38 C/14.41 bar]. At -35 F [-37 C], it is only 12.1 psia [0.823 bar]. It is apparent that the stresses imposed on the piping due to direct refrigerant pressure will be generally low. The majority of the stress that the designer will have to consider is a thermal stress due to expansion or contraction of the piping when its operating temperature fluctuates between ambient and operating temperatures. Currently, system designers may choose from several software packages to verify the stresses imposed on piping. By using this software, the system designer can prove the suitability of non-impact-tested, less-expensive piping materials for operation at low temperatures. In most cases, a software analysis for large systems proves to be beneficial because non-impact-tested piping costs far less than impact-tested materials. Many companies use three-dimensional modeling as their standard method for laying out the piping in the plant, which makes the analysis quite simple. The models are used to estimate the labor required for the installation, to prepare bills of materials, and to create isometric drawings for pipe spools. Once the model has been created, the pipe runs can be evaluated using the stress analysis software. If the initial analysis shows that the allowable stresses would be exceeded, the designer can reduce them by modifying the proposed layout and location of the anchor points. Installations designed in such a manner are safe, less expensive and meet the piping code requirements. This type of design process does not just apply 38 IIAR 2005 Technical Paper #2

7 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando to low-temperature applications; designers may check stresses in other types of piping systems as well, including hot gas lines or compressor discharge lines. Figures 3 and 4 illustrate a sample analysis of a -50 F [-46 C] wet suction line and 120 F [49 C] hot gas line. In the example, the suction main and hot gas runs were first drawn using computer-aided design (CAD) software based on a typical roof routing of the piping mains. Once the mains had been designed, the pipe supports were added to the drawing, and the system was then evaluated using piping-stressanalysis software. The analysis showed that both SA106 grade B and SA 333 grade 6 piping materials were suitable for the installation, based on the pipe size and thickness, ambient temperature range, operating pressures and temperatures, and fluid density. Process requirements define operating temperatures. While the designer cannot change this constraint, he or she can lay out the plant pipe in such a manner to limit the imposed thermal stresses. When laying out system pipe runs, no single rule assures that the proposed layout will not exceed the maximum allowable stresses for the material. Instead, a common sense approach that considers the behavior of pipe subjected to temperature fluctuations should guide the design process. The system design must allow the pipe to move without breaking the welds and prevent it from becoming overly stressed while it experiences temperature reductions or increases. One good method to mitigate such stresses in a pipe run is to add a few elbows to create a loop that can flex when the pipe run expands or contracts. Also, vessels and attached pipe can be installed such that thermal expansion and contraction will not cause forces acting on the nozzles to exceed the maximum allowable limits. Like pipe, all other components must be suitable for operation at the design operating temperature. The ASME B31.5 piping code does not require the use of impact-tested carbon steel bolting for temperatures above -50 F [-46 C] as long as the bolting used is high strength A193 grade B7. The designer also needs to confirm that the elastomers used for gaskets and seals are not only compatible with the Technical Paper #2 IIAR

8 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico refrigerant, but that they will maintain sealing capabilities at low temperatures. Neoprene, which is the most common elastomer used in ammonia systems, is usually rated for temperatures above -40 F [-40 C]. For temperatures below this threshold, the designer may have to consider either PTFE gaskets or spiral wound stainless steel gaskets that are suitable for very low operating temperatures. Also, the welding procedures used in fabricating piping systems should be rated for the minimum metal design temperature. A suitable Welding Procedure Specification (WPS) and Procedure Qualification Record (PQR) must be prepared in accordance with Section IX of the B&PV Code, and a set of pipe coupons must be impact-tested at the minimum design metal temperature to be used for the procedure. The main reason for the impact-testing requirement is to establish the heat input required to achieve the desired weld toughness at the design temperature. Although the B&PV Code does not specifically require that the welding consumables be impact-tested for the minimum metal temperature, it may be beneficial to utilize impact-tested or low-temperature-rated consumables to assure that the coupon passes the impact test. Also, the welders must be currently qualified to perform the approved WPS. It is worth pointing out that common stainless steel materials such as type 304, 304L, 316 or 316L do not require impact testing, nor does aluminum. Stainless steel and aluminum are commonly used in the construction of evaporator coils in air units. Cast, Malleable and Nodular Iron in Refrigeration Systems Valve manufacturers commonly use cast, malleable and nodular types of iron. Most of the relief valves and regulators used in ammonia refrigeration are made of gray cast iron or ductile iron materials. The ASME B31.5 code allows cast and malleable types of iron as long as they are not used for hydrocarbon or flammable fluid service at temperatures above 300 F [149 C] or operating pressures above 300 psig [20.4 barg]. Restrictions against nodular iron are somewhat different; the code prohibits the use of nodular iron above 1000 psig [20.4 barg], which obviously 40 IIAR 2005 Technical Paper #2

9 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando would not be a common case with ammonia refrigeration. Notably, the code allows the usage of cast, malleable and nodular iron for temperatures above -150 F [-101 C] without impact testing. ANSI/ASME Piping Classes The ASME B16 Committee developed several standards to ensure consistency in the design and application of valves, flanges, fittings and gaskets. The committee defined several pressure-temperature classes for a variety of cast and forged materials. The two main steel flange classes defined by the ASME B16.5 standard that are commonly used in ammonia refrigeration piping are Class 150 and 300. (ASME, 2003) The similar classes for cast and nodular irons are defined as Class 125 and 250. The class defines the maximum allowable working pressure for which a given component can be used depending on the system operating temperature. Figure 5 details the basic pressure ratings for both classes and common carbon and stainless steel materials. Class 150 and 300 represent nominal 150-psig [10.2-barg] and 300-psig [20.4-barg] designs, respectively. For carbon and stainless steel fittings, the maximum allowable working pressure decreases as temperature increases. For example, for Class 150 carbon steel flanges, the maximum allowable working pressure at 100 F [38 C] is 284 psig [19.3 barg] and falls to 200 psig [13.6 barg] at 392 F [200 C]. The examination of the pressure-temperature charts for each class shows that if the refrigeration system is designed entirely for 300-psig [20.4 barg] maximum allowable working pressure, Class 150 valves, flanges or other components should not be used. As indicated above, for Class 150, the maximum pressure rating at any temperature is less than 300 psig [20.4 barg]. This restriction is also applicable to compressor suction flanges or flanges supplied with pressure vessels, heat exchangers or pumps. Some compressor and pump manufactures utilize Class 150 flanges at compressor or pump suction connections; thus, they should not be used if the low side of the system is protected with a relief valve set for 300 psig [20.4 barg]. Technical Paper #2 IIAR

10 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico ANSI/ASHRAE and Mechanical Design ASHRAE , a model refrigeration safety code, provides recommendations for the design pressures for refrigeration systems. The code states that all equipment should be designed for a 29 Hg vacuum [23 Torr]. For the low side of ammonia systems, the code requires 138 psig [9.39 barg] minimum design pressure, which is equivalent to 80 F [27 C] saturation temperature. For the high side, with an evaporative or water-cooled condenser, the code recommends a 210-psig [14.3-barg] design, which corresponds to 104 F [40 C] saturation temperature. If air-cooled condensers are used, then the minimum high side design pressure should be 280 psig [19.0 barg], equivalent to a 122 F [50 C] saturation temperature. If a designer expects ambient or inside temperatures hotter than 80 F [27 C], he/she should then design the system for the pressure equivalent to the maximum expected temperature, with at least a 15% margin of safety to account for tolerance in the relief valve set pressure. For example, for a 95 F [35 C] ambient, which is a common temperature experienced in the summer in the southern U.S., the equivalent saturation gauge pressure is 181 psig [12.3 barg]. At 100 F [38 C], the ammonia vapor pressure is 197 psig [13.4 barg]. For these temperatures, even a 200-psig [13.6-barg] low side design appears to be insufficient. Designing the low side for 250 psig [17.0 barg] can prevent the relief valves from opening and releasing ammonia during elevated ambient temperatures. For systems designed for 250 psig [17.0 barg] and operating temperatures below 212 F [100 C], Class 150 flanges and fittings can still be used on the compressors, pumps, vessels and heat exchangers. At times, designers specify the high side of the system for 300 psig [20.4 barg]. Many ammonia systems use evaporative condensers operating at 181 psig [12.3 barg], with occasional excursions to 210 from 240 psig [14.3 to 16.3 barg]. The differential between these pressures and the 300-psig [20.4-barg] relief valve set 42 IIAR 2005 Technical Paper #2

11 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando point permits the valve to remain securely closed, preventing weeping that can occur when the differential is less than 25 psi [1.7 barg], as would be the case had the design pressure been chosen at 250 psig [17.0 barg]. Sizing of Piping Runs New Tools for Selecting Refrigerant Pipe Sizes Design engineers may choose among a wide variety of selection charts for sizing refrigerant piping. These charts are generally based on the Darcy equation, which is suitable for single-phase flow of both liquids and gases as long as the average flow velocity does not exceed approximately 0.3 Mach. For vapor flow, the equation can be used only when the pressure losses in the pipe run are relatively low, which is the case for a majority of pipe runs. This limitation is due to the fact that for vapors, as the pressure decreases, the specific volume increases. Thus, for the same mass flow rate the average flow velocity increases. However, because the Darcy equation assumes a constant average flow velocity for calculation of the pressure drop, it becomes less accurate as vapor pressure drops increase in size. This Darcy-based methodology does not address two-phase flow situations, common for all overfeed systems. The ammonia refrigeration industry has circumvented this problem by the practice of selecting wet suction lines one size larger than the dry suction line required for the given duty. Also, some refrigeration system designers and pipe sizing software developers have adopted Beattie s simplistic method of calculating pressure drop in two-phase flow. (Beattie, 1982) For some time, the process industry has used simulation software to aid engineers in designing process equipment, such as absorbers, strippers, mixers, etc. Industries that compress gases also use this software, especially for compression of wet mixed hydrocarbon gases. Further, the software aids in the design of carbon dioxide Technical Paper #2 IIAR

12 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico purification and liquefaction plants. The software output is essential engineering data that is then used for selecting all heat exchangers, pressure vessels and piping for single- and two-phase flows in the plant. A typical Process Flow Diagram, a software output report, and an example of pipe sizing are included in Figures 6 through 8. There are many advantages to using simulation software for plant design. The software allows for a detailed analysis of the plant performance, including predicting system performance under different operating situations. It eliminates guessing in the design process by rigorously calculating all relevant flow data. It assures that all pipe sizes are correct and designed for the actual plant requirements. The software output can also be used to confirm that compressors, pumps and other equipment selected have enough capacity. The software allows for optimizing plant performance by allowing the designer to evaluate different what-if scenarios. Obviously, the software can also be a handy tool for analyzing existing installations. The Effects of Undersizing or Oversizing Piping Systems Liquid Lines: It is essentially impossible to oversize liquid lines from a flow and pressure drop point of view. The designer has to be aware that the larger-thannecessary liquid lines will contribute to overall plant liquid inventory, which in turn will require a larger receiver volume, adding cost to the installation. Undersized liquid piping causes liquid pressure to drop, which in turn will cause flash gas to form during flow. Flash gas has an adverse impact on the performance of expansion valves. This problem is especially acute for liquid lines containing saturated liquid. For pumped liquid overfeed or pressure-fed liquid systems, excessive pressure drop in liquid piping can result in insufficient liquid pressure at the expansion valve to overcome the pressure drop at the required flow through the evaporator. When analyzing liquid lines, the designer has to be aware of the consequences of installing larger-than-necessary liquid control valves. In the case of thermostatic 44 IIAR 2005 Technical Paper #2

13 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando expansion valves, if the valve is too big, it can become unstable, which in turn can result in poor control and eventual flooding of the evaporator. The valve seats of automated or hand-operated expansion valves for flooded evaporators will wear out faster if forced to operate barely open for long periods. On the other hand, if the control or expansion valve is too small, the evaporator will starve. Also, some solenoid valves require a minimum pressure differential to operate properly. If liquid line solenoid valves are oversized, there may not be enough flow to create a pressure drop adequate to hold the valve open. Designers should also take advantage of the fact that the plant can operate more efficiently at lower-than-design ambient dry- or wet-bulb temperatures, which occur the majority of the time. At low ambient conditions, the differential pressure between the high and low side of the system can be substantially less than the design differential pressure, which reduces the required compressor horsepower. Therefore, both liquid and hot gas lines should be sized for flow at the reduced differential pressure to take advantage of these conditions. Otherwise, the plant will pay an expensive penalty due to the excessive power consumption. Suction Lines: Special care should be taken when sizing suction lines. The lower the design suction temperature, the greater the effect of undersizing. Undersized low temperature suction piping and components will result in increased compressor horsepower requirements for the same refrigeration duty. As the pressure decreases, the gas specific volume and the horsepower-per-ton increase (assuming constant condensing pressure). The lower suction pressure requires a greater compression ratio to compress the refrigerant vapor to condensing pressure. As the suction pressure decreases, the required compressor volumetric flow increases. For the same duty, this may result in having to provide larger compressor capacity, thus increasing equipment costs. A good approach to avoid excessive capital costs is to establish the optimum economical velocity based on a comparison of initial and operating costs. (Richards, 1984) Based on the historical relationship Technical Paper #2 IIAR

14 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico between these costs, several design recommendations for the compressor suction lines have been developed. One contractor proposed the following recommendations in a form of overall pressure drop per 100 feet of equivalent length of pipe: -40ºF 0.15 to 0.20 psi -20ºF 0.25 psi +15ºF 0.5 psi +35ºF 1 psi Another economics-based method is described in IIAR s Ammonia Refrigeration Piping Handbook. (IIAR, 2004) The Handbook includes several tables for both suction and discharge lines. Again, the overall pressure loss must be considered when selecting isolation and control valves. For isolation valves, angle or butterfly designs are recommended because they require less pressure drop than globe valves. Control valves should be selected and sized to allow proper operation of the valves at the lowest pressure drop that is economically feasible. However, if the system design requires a greater pressure drop for a certain piece of equipment (e.g., multiple evaporators operating at different evaporating temperatures connected to a common suction header), then the designer should disregard this guideline. Check valves should also be specified taking into account the overall pressure loss. When selecting check valves, both the operating and cracking pressure drop required to lift or open the valve should be considered. Cracking pressure is the pressure drop required for the valve to open and overcome valve spring pressure and/or disk or piston weight. Oversized piston-type check valves will cause the piston to operate too close to the seat, which, in the event of a minor upset, will cause the piston to bounce off the seat. Also, when excessive forces are required to open check valves, the resulting pressure drop fluctuation across the valve may lead to bouncing or chattering and eventually damage the valve. The use of split-wafer check valves allows minimum cracking force and low operating pressure drop if the valve is 46 IIAR 2005 Technical Paper #2

15 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando properly sized and installed. These valves are available with low-torque springs that have very low valve cracking pressures. In summary, when rating and selecting compressors, the designer must consider all combined pressure drops between the evaporator and the compressor flange. Hot Gas Lines: During hot gas defrost, the required hot gas duty is about twice the refrigeration capacity of the evaporator that operates at a 10ºF [5.6K] temperature difference. Hot gas piping should be sized to accommodate the volume of hot gas required to condense this amount of vapor. Hot gas headers should be sized to accommodate all of the evaporators scheduled to defrost at the same time. The volume of hot gas flow is determined by the difference between the compressor discharge pressure and the highest defrost pressure regulator setting on any of the evaporators in the system. Piping should be designed to permit adequate flow even when the condensing pressure is relatively low, such as during winter months. At lower-than-design ambient conditions, undersized hot gas lines waste energy unnecessarily and prolong defrost cycles. Extended defrost cycles give the evaporators less time to recover between cycles, potentially compromising room temperatures or causing product quality issues. Piping Layout and Special Issues Thermal Expansion of Subcooled Liquid Subcooling: Subcooled liquid is frequently used in ammonia refrigeration plants and is produced by cooling liquid below its saturation temperature at the prevailing pressure. Subcooling may increase the efficiency of the refrigeration cycle and prevent the formation of flash gas in downstream piping. In a two-stage system, for example, by exchanging heat between liquid supply A (which feeds low-temperature loads) and liquid supply B (which flashes off and is directed to the high-stage Technical Paper #2 IIAR

16 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico compressor), liquid supply A would become subcooled and its refrigerating effect would increase. While the flash gas would increase the volume normally handled by the high-stage compressor, the subcooling effect would reduce the amount of gas handled by the low-stage compressor. However, because the high stage compressor operates more efficiently in terms of Btuh/BHP than the low stage, the process yields a net gain in efficiency. Subcooling is required in piping that suffers excessive frictional pressure losses or that must be delivered to elevated evaporators. In the case of elevated evaporators fed with saturated ammonia at 95 F [35 C], for every 3.92 ft [1.19 m] of elevation increase, the liquid vapor pressure falls by 1 psi [0.068 bar], causing subsequent flashing of liquid. Preventing flash gas is also important for sizing liquid expansion valves. Most of the selection charts and software assume that pure liquid exists upstream of the expansion device. If this is not the case, the device will either not operate properly, or will be too small for the duty, and may eventually fail. Thermal expansion of cold liquid: Consider a case in which we have trapped subcooled liquid ammonia in a 200 length of 2 schedule 40 pipe. If the subcooled liquid were allowed to warm up from 10 F to 95 F [-12 C to 35 C], for example, during system shutdown, it would expand, but how much? The density of subcooled liquid ammonia at 10 F [-12 C] is approximately 40.9 lb/ft 3 [656 kg/m 3 ]. Therefore, the specific volume of the liquid is ft 3 /lb [ m 3 /kg]. At 95 F [35 C], its density would be 36.7 lb/ft 3 [589 kg/m 3 ]; specific volume would be ft 3 /lb [ m 3 /kg]. Thus, each pound of trapped liquid would increase in volume by cubic feet [0.18 liters]. 48 IIAR 2005 Technical Paper #2

17 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando The volume of the pipe in our example, 200 of 2 schedule 40, is calculated: 200 x ft 3 /linear foot = 4.66 ft 3 [0.132 m 3 ]. The mass of this volume of ammonia at 10 F [86.6 kg at -12 C] is: 4.66 ft 3 x 40.9 lbs/ft 3 = 190 lbs. After warming to 95 F [35 C], the ammonia will expand to a volume of: 190 x = 5.17 ft 3 [0.147 m 3 ] The net increase in liquid volume, then, would be: = 0.51 ft 3 [0.015 m 3 ]. However, liquid ammonia cannot be compressed. Therefore, as trapped liquid warms, the pipe will at first expand and stretch to accommodate the larger volume and rising pressure. As the pressure exceeds the mechanical limits of the pipe, it will fail. Fortunately, the designer of the system can address this problem by providing the liquid a place in which to expand, such as an upstream or downstream vessel or heat exchanger. Isolation valves at both ends of a liquid line should also be tagged or labeled, advising the operator not to close both valves simultaneously. In addition, liquid relief valves can be installed and piped either to a relief header or a low side pressure vessel. In carbon dioxide installations, it is a common practice to protect subcooled and saturated liquid lines with pop-up hydrostatic relief valves. These relatively small and inexpensive devices prevent the rising of hydrostatic pressure in the piping system by relieving expanding liquid to the atmosphere. A similar problem is encountered in low temperature cascade systems. These systems utilize various refrigerants, including carbon dioxide. Because the critical temperature of carbon dioxide is below typical ambient temperatures, it will develop a high pressure at relatively moderate temperatures. For example, saturated carbon dioxide s vapor pressure at 50 F [10 C] is approximately 652 psia [44.4 bar]. In the Technical Paper #2 IIAR

18 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico case of R-23, another commonly used low temperature refrigerant, the vapor pressure at the same temperature is 472 psia [32.1 bar]. In those installations, any trapped saturated liquid must be vented to prevent the system pressure from rising above its maximum allowable design pressure. Hydrostatic relief valves can be used for this purpose. Vapor Condensation in High Pressure Lines One problem commonly overlooked is the effect of low ambient temperatures on high-pressure refrigerant vapor lines, encountered when routing compressor discharge lines or high-pressure vent lines through air-conditioned spaces or when high-pressure pipes are installed outside the building. At 95 F [35 C] condensing temperature, the ammonia vapor pressure is about psia [13.31 bar]. If a system is shut down under this condition and the pipe pressure is maintained at the pressure equivalent to the high ambient temperature, then once the pipe wall temperature starts falling, the vapor refrigerant in the pipe will begin to condense. Condensing liquid may flow back down the compressor discharge lines and accumulate on top of the compressor discharge check valve. If the check valve is not fully closed, liquid will drain into the compressor s oil separator. If the discharge line is large enough and the oil separator does not have a heat source sufficiently great to boil off all of the condensed liquid, the refrigerant will increasingly dilute the compressor oil. At high levels of oil dilution, the oil viscosity can fall so far that the oil will not provide sufficient bearing lubrication, causing compressor failure once the compressor is restarted. An easy and relatively inexpensive solution to this problem is to use inverted traps for individual compressor discharge risers at the discharge header, or trap discharge lines in which any condensed liquid can collect. The trap can be either drained with 50 IIAR 2005 Technical Paper #2

19 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando a liquid drainer to a high or low side vessel, or liquid can be allowed to boil off once the system is restarted. Additionally, heat tracing and insulation may be applied to maintain a warm pipe surface temperature to prevent liquid from condensing. Differential pressure transmitters are not commonly used in ammonia refrigeration. However, where they are used to measure liquid level in the high-pressure vessel, it is important to heat-trace and insulate the vapor leg of the device. Otherwise, on a cold day, liquid will condense in the tubing and stack up on the low side of the transmitter. This condition may cause the transmitter to falsely indicate low liquid levels in the vessels. Likewise, the tubing runs to the pressure transmitters and pressure transducers on the high side of the system should not be trapped. They should be installed to allow ammonia condensate to drain freely into the pipe. Layout of Suction Piping and Liquid Carry-over It is very important to prevent overfed liquid from reaching refrigerant compressors. In a liquid-vapor separator vessel, the liquid-vapor flow velocity is reduced below a certain critical velocity, and the vapor can no longer fully entrain liquid droplets. The liquid droplets fall out of the vapor stream and settle into the liquid pool. The separated vapor is saturated (but not wet), and is collected to flow into the compressor. The key distinction between saturated and wet vapor is that saturated vapor does not carry any free liquid droplets, while wet vapor includes liquid. All liquid overfeed systems have wet return lines. The compressor suction lines should be dry. The maximum allowable flow velocity is usually determined from a well-known industry equation: V max = k ρ l ρ v ρ v Technical Paper #2 IIAR

20 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico where: V max = maximum vapor flow velocity (ft/sec) k = flow coefficient ρ l = liquid density (lb/ft 3 ) ρ v = vapor density (lb/ft 3 ) Separation vessel manufacturers provide the maximum vessel capacity, expressed in most cases in terms of tons of refrigeration. The capacity depends on the vessel size, operating temperature, maximum allowable vessel flow velocity, vapor density, and mass flow rate. For example, consider a load of 80 tons of refrigeration (TR) [280 kw] at -40 F [-40 C] with a 90 F [32 C] liquid supply. We wish to calculate the mass flow. First, we determine the refrigerating effect of each pound of ammonia by subtracting the enthalpy of the liquid (143 Btu/lb) from the vapor enthalpy (597 Btu/lb). The result, 454 BTU/lb [293 W-hr/kg], is the refrigerating effect. We convert the TR units into Btu/hr and multiply by the refrigerating effect to obtain the mass flow. The estimated mass flow is: 12,000 Btu 1 lb 80TR x x =2,115 lb/hr [961 kg/hr] Ton hr 454 Btu Ammonia vapor density at -40 F [-40 C] is 0.04 lb/ft 3 [0.64 kg/m 3 ]. We divide the mass flow by the density to obtain the volumetric flow rate: 2,115 lb hr 1 ft x 3 1 hr x = 881 ft 3 /min [25.0 m 3 /min] 0.04 lb 60 min Therefore, for a given vessel internal diameter, the flow velocity inside the vessel can be easily established by dividing the volumetric flow rate by the vessel crosssectional area and comparing it with the maximum allowable velocity. 52 IIAR 2005 Technical Paper #2

21 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando The plant designer can depend on experienced refrigeration vessel manufacturers to provide a vessel with a suitable capacity for the design conditions or follow the maximum separating velocity guidelines published by ASHRAE. Several additional factors must be considered to assure that the objective of liquid separation is achieved. First, the designer must consider off-design conditions (e.g., system pull-down, heavy evaporator defrost duty) during which the actual flow may be well above the steady state flow conditions. Second, all the wet and dry suction vapor piping should be sloped back to the separation vessel to make sure that any liquid condensed or entrained in the stream is returned into the separation vessel. Third, all individual compressor suction lines should be piped to the top of the main header in lieu of the bottom or side to prevent liquid refrigerant from draining back down the suction riser during the off-cycle. Trapped lines, obviously, does not slope back to the vessel, so individual suction lines and mains should not be trapped. Traps will accumulate liquid until it is pushed back to the separator, which in turn can overload the separator and create overall system instability. If it is not possible to drain the liquid back to the vessel, then a means to return liquid must be provided (e.g., collection pots, a double riser return). Another issue that needs to be considered while designing liquid-vapor separation systems is how to return the collected liquid back into the system. If the separator operates at an intermediate pressure, then liquid can be drained back to the low side of the system utilizing the differential pressure between the intermediate and low sides. If the separator also acts as an intercooler, collected liquid can vaporize during desuperheating of the boosters discharge vapor. The low side separators, however, require a source of heat to boil off the refrigerant. In ammonia systems, this is typically achieved with a boil-off coil installed in the bottom of the vessel through which warm liquid refrigerant is circulated. Another suitable, but much less efficient way, is to use an electric heater or vessel heat tracing. It is also possible to collect liquid in a liquid trap and return it by pushing it with discharge gas, or a pump, into an intermediate pressure vessel. Technical Paper #2 IIAR

22 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico Risers for flooded shell-and-tube evaporators operate below their flooding velocities much of the time and must be large enough to allow separated liquid to drain back while allowing vapor to flow out of the exchanger. Pumped and Gravity-fed Evaporators Pumped systems are common for liquid overfeed applications. Gravity-fed (also called thermosyphon) systems are also frequently used for evaporators and for cooling compressor oil. Piping of pumped systems is quite simple. The refrigerant pump develops sufficient head, normally psi, to deliver liquid to even the most remote evaporators in the plant. The piping from the pump to the evaporators can be routed in any manner and the system will operate properly as long as the pump flow is large enough for the duty and pump head is greater than the pressure drop between the pump and the evaporator. In the case of gravity fed evaporators, the driving force to supply liquid to the evaporator is the elevation difference between the liquid level above the evaporator and the evaporator itself, plus the difference between the density of liquid in the supply line and the liquid-vapor mix in the return piping. This force must overcome liquid line pressure losses, pressure drop across the evaporator and the static head of the liquid-vapor mix in the return line. It should be noted that for gravity systems, the amount of liquid supplied to the evaporator would depend on the evaporator heat load. As the heat load increases, the amount of liquid vaporized increases, causing the density of vapor liquid-vapor mix leaving the evaporator to decrease. Thus, as the heat load increases, the amount of liquid flowing into the evaporator also increases, which increases the pressure losses (which are proportional to the square of the flow velocity). 54 IIAR 2005 Technical Paper #2

23 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando If the system is designed for a 3 to 1 overfeed ratio at the given heat load, then the liquid supply line will carry 3 times as much liquid as necessary for vaporization. Thus, the return line will carry two times more liquid than vapor in terms of mass flow. If the piping is undersized or pressure losses through the evaporator are too high, the system will self-compensate by moving less than design overfeed ratio through the evaporator because the flow velocity will decrease and the density difference between supply and return will increase. In the case of gravity-fed systems, it is important that the evaporator s surge drum is sized properly and capable of storing fluctuating and transient liquid levels without tripping the high-liquid-level switch installed on the drum. Other important issues requiring attention in gravity-fed systems are oil management and prevention of oil and water accumulation in the evaporator coil. The most efficient way to prevent oil accumulation in the evaporator is to avoid drawing liquid from the bottom of the surge drum, which permits oil to accumulate on the bottom, and to provide for draining oil from the bottom of the surge drum and the coil. Both pumped and gravity systems can be utilized for a variety of evaporators and refrigerants. In the process refrigeration industry, thermosyphon arrangements are popular for heat exchangers that cool viscous liquids, such as ethylene glycol. In those cases, the viscous brine flows through the shell side of the evaporator while liquid refrigerant is supplied to the tube side of the exchanger. For the same reason, oil flows on the shell side of a typical shell-and-tube thermosyphon oil cooler. Not only ammonia, but also a variety of other refrigerants are used in pumped or gravity systems in process refrigeration. The installations utilizing the other refrigerants are designed in a similar fashion. However, for most of these, the overfeed ratio is usually lower, because the amount of mass required for vaporization is much greater. As a result, the velocity through the exchangers is high enough for a reasonable heat transfer coefficient even at low overfeed ratios. Technical Paper #2 IIAR

24 2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico Avoiding Liquid Hammer Special precautions must be taken in the design and layout of refrigeration piping to avoid the conditions that allow for the sudden deceleration of liquid flow, which can result in failure of components and/or piping. This phenomenon is commonly known as liquid hammer. It can occur when high-pressure vapor is suddenly introduced in a low-pressure line partially filled with liquid. The liquid becomes entrained, forming a slug that grows and accelerates until it seals the pipe off completely and is propelled with the full velocity of gas flowing in the pipe. Liquid hammer can occur during defrost either when pressure is released too quickly from an evaporator at the end of the cycle, or, when condensate that forms in a hot gas line is propelled with the hot gas when released into an evaporator. Because hot gas velocities can easily reach 100 feet per second, the pressure of the liquid slug upon impact can exceed 3,000 psig [200 bar] (IIAR, 1992) Soft gas defrost cycles may be used to prevent this situation and piping should be designed to prevent liquid slugs from hitting a dead end in the piping. Cleanliness With carbon steel piping, particulates and moisture potentially may remain in the system after assembly. These contaminants can wreak havoc with components such as solenoids, check valves and regulators, plug strainers and orifices, and worst of all, damage compressors. After assembly, the piping system should be cleaned and dried. Steps should be taken to reduce the amount of contaminants initially introduced into the system: To reduce the amount of rust in the system, piping should be kept capped at all times until it is welded to the system. Welding slag should be removed or procedures used that reduce the amount of slag introduced into the system. 56 IIAR 2005 Technical Paper #2

25 Refrigeration Piping: A Simplified Guide to a Modern Approach Paul Danilewicz and Paul Orlando Procedures should be used to keep foreign materials from being placed or forgotten in the system such as rags, welding rods, soap stones and other items and checks performed to insure such items are not left in the system before sealing. Prior to system start up, socks should be installed in strainers to collect the initial mass of particulate. Water Contamination The introduction of moisture in the system should be minimized. Purchasing ammonia from a reliable vendor who provides refrigerant grade ammonia with certification is one step to help reduce moisture content. The sometimes-overlooked situation during installation is when a pipe penetration, through the building wall, for example, is sealed on the cold side but not sealed on the warm side. When the cold side is subject to temperatures below the dew point of the air on the warm side, water will condense inside the pipe even in dry climates. If the water is not removed, it will introduce a large volume of moisture into the system. The refrigeration system should be initially dried out to an acceptably low moisture content and should be periodically checked. The effects of moisture contamination can be found in IIAR Bulletin No (IIAR, 1986) Purgers should be used to remove air in the systems that operate in a vacuum. A method of removing air-borne moisture should also be provided. An ammonia/water distiller is a good way to periodically remove moisture. Excessive oil is also considered to be a contaminant in the system. Equipment should be in place to reduce the amount of oil introduced into the system by having properly sized oil separators. The oil pots and other oil accumulation point drains to remove accumulated oil from the system should also be provided. Where applicable, systems should be in place to return or recycle oil back to the compressors. Technical Paper #2 IIAR