Piping. Figure 1 shows a typical piping arrangement for two compressors operating in parallel off the same suction main. Suction mains should be laid out with the objective of returning only clean, dry gas to the compressor. This usually requires a suction trap sized adequately for gravity gas and liquid separation based on permissible gas velocities for specific temperatures. A dead-end trap can usually trap only scale and lubricant. As an alternative, a shell-and-coil accumulator with a warm liquid coil may be considered. Suction mains running to and from the suction trap or accumulator should be pitched toward the trap at 10 mm per metre for liquid drainage. It is also good practice to connect compressor suction branch piping above the centerline of the suction main.

In sizing suction mains and takeoffs from mains to compressors, consider how pressure drop in the selected piping affects the compressor size required. First costs and operating costs for compressor and piping selections should be optimized.

Figure 1. Schematic of Reciprocating Compressors Operating in Parallel

Good suction line systems have a total friction drop of 0.5 to 1.5 K pressure drop equivalent. Practical suction line friction losses should not exceed 0.01 K equivalent per metre equivalent length.

A well-designed discharge main has a total friction loss of 7 to 14 kPa. Generally, a slightly oversized discharge line is desirable to hold down discharge pressure and, consequently, discharge temperature and energy costs. Where possible, discharge mains should be pitched (10 mm/m) toward the condenser without creating a liquid trap; otherwise, pitch should be toward the discharge line separator.

High- and low-pressure cutouts and gages and lubricant pressure failure cutout are installed on the compressor side of the stop valves to protect the compressor.

Lubricant Separators. Lubricant separators are located in the discharge line of each compressor (Figure 1A). A high-pressure float valve drains lubricant back into the compressor crankcase or lubricant receiver. The separator can be placed away from the compressor, so any extra pipe length can be used to cool the discharge gas before it enters the separator. This reduces the temperature of the ammonia vapor and makes the separator more effective. A discharge gas heat exchanger may also be used to cool the gas before it enters the oil separator.

Liquid ammonia must not reach the crankcase. Often, a valve (preferably automatic) is installed in the drain from the lubricant separator, open only when the temperature at the bottom of the separator is higher than the condensing temperature. Some manufacturers install a small electric heater at the bottom of a vertical lubricant trap instead. The heater is actuated when the compressor is not operating. Separators exposed to cold must be insulated to prevent ammonia condensation. Venting the high-pressure gas in the oil separator to the crankcase or suction line also helps prevent ammonia condensation.

A filter is recommended in the drain line on the downstream side of the high-pressure float valve.

Lubricant Receivers. Figure 1B shows two compressors on the same suction line with one discharge-line lubricant separator. The separator float drains into a lubricant receiver, which maintains a reserve supply of lubricant for the compressors. Compressors should be equipped with crankcase floats to regulate lubricant flow to the crankcase.

Discharge Check Valves and Discharge Lines. Discharge check valves on the downstream side of each lubricant separator prevent high-pressure gas from flowing into an inactive compressor and causing condensation (Figure 1A).

The discharge line from each compressor should enter the discharge main at a 45° maximum angle in the horizontal plane so the gas flows smoothly.

Unloaded Starting. Unloaded starting is frequently needed to stay within the torque or current limitations of the motor. Most compressors are unloaded either by holding each cylinder’s suction valve open or by external bypassing. Control can be manual or automatic.

Suction Gas Conditioning. Suction main piping should be insulated, complete with vapor retarder to minimize thermal losses, to prevent sweating and/or ice build-up on the piping, and to limit superheat at the compressor. Additional superheat increases discharge temperatures and reduces compressor capacity. Low discharge temperatures in ammonia plants are important to reduce lubricant carryover and because compressor lubricant can carbonize at higher temperatures, which can cause cylinder wall scoring and lubricant sludge throughout the system. Discharge temperatures above 120°C should be avoided at all times. Lubricants should have flash-point temperatures above the maximum expected compressor discharge temperature.

Cooling. Generally, ammonia compressors are constructed with internally cast cooling passages along the cylinders and/or in the top heads. These passages provide space for circulating a heat transfer medium, which minimizes heat conduction from the hot discharge gas to the incoming suction gas and lubricant in the compressor’s crankcase. An external lubricant cooler is supplied on most reciprocating ammonia compressors. Water is usually the medium circulated through these passages (water jackets) and the lubricant cooler at a rate of about 2 mL/s per kilowatt of refrigeration. Lubricant in the crankcase (depending on type of construction) is about 50°C. Temperatures above this level reduce the lubricant’s lubricating properties.

For compressors operating in ambients above 0°C, water flow controlled by a solenoid valve in the inlet line is desirable to automate the system and prevent any refrigerant condensing above the pistons. When the compressor stops, water flow must be stopped to keep residual gas from condensing and to conserve water. A water-regulating valve, installed in the water supply line with the sensing bulb in the water return line, is also recommended. This type of cooling is shown in Figure 2.

The thermostat in the water line leaving the jacket serves as a safety cutout to stop the compressor if the temperature becomes too high.

Figure 2. Jacket Water Cooling for Ambient Temperatures Above Freezing

For compressors where ambient temperatures may be below 0°C, a way to drain the jacket on shutdown to prevent freeze-up must be provided. One method is shown in Figure 3. Water flow is through the inlet line normally closed solenoid valve, which is energized when the compressor starts. Water then circulates through the lubricant cooler and the jacket, and out through the water return line. When the compressor stops, the solenoid valve in the water inlet line is deenergized and stops water flow to the compressor. At the same time, the drain line normally open solenoid valve deenergizes and opens to drain the water out of the low point to wastewater treatment. The check valves in the air vent lines open when pressure is relieved and allow the jacket and cooler to be drained. Each flapper check valve is installed so that water pressure closes it, but absence of water pressure allows it to swing open.

For compressors in spaces below 0°C or where water quality is very poor, cooling is best handled by using an inhibited glycol solution or other suitable fluid in the jackets and lubricant cooler and cooling with a secondary heat exchanger. This method for cooling reciprocating ammonia compressors eliminates fouling of the lubricant cooler and jacket normally associated with city water or cooling tower water.

Figure 3. Jacket Water Cooling for Ambient Temperatures Below Freezing

Piping. Rotary vane compressors have been used extensively as low-stage compressors in ammonia refrigeration systems. Now, however, the screw compressor has largely replaced the rotary vane compressor for ammonia low-stage compressor applications. Piping requirements for rotary vane compressors are the same as for reciprocating compressors. Most rotary vane compressors are lubricated by injectors because they have no crankcase. In some designs, a lubricant separator, lubricant receiver, and cooler are required on the discharge of these compressors; a pump recirculates lubricant to the compressor for both cooling and lubrication. In other rotary vane compressor designs, a discharge lubricant separator is not used, and lubricant collects in the high-stage suction accumulator or intercooler, from which it may be drained. Lubricant for the injectors must periodically be added to a reservoir.

Cooling. The compressor jacket is cooled by circulating a cooling fluid, such as a water/glycol solution or lubricant. Lubricant is recommended, because it will not freeze and can serve both purposes (Figure 4).

Helical screw compressors are the choice for most industrial refrigeration systems. All helical screw compressors have a constant-volume (displacement) design. The volume index V_i refers to the internal volume ratio of the compressor (i.e., the reduction in volume of the compressed gas from suction to discharge of the compressor). Capacity control is accomplished by use of a slide valve, bypass ports, or by controlling the speed [variable-frequency drive (VFD)]. The slide valve and bypass ports control capacity by only using a portion of the screw(s) for the compression process.

Figure 4. Rotary Vane Booster Compressor Cooling with Lubricant

Some compressors are designed with a fixed V_i. When V_i is fixed, the compressor functions most efficiently at a certain compression ratio (CR). In selecting a fixed V_i compressor, consider the average CR rather than the maximum CR. A guide to proper compressor selection is based on the equation V_i^k = CR, where k = 1.4 for ammonia. For example, a screw compressor operating at 265 kPa suction and 1350 kPa discharge has a CR = 5.09. Therefore, V_i = 3.2 (V_i = CR¹/k). Thus, a compressor with the V_i at or close to 3.2 is the best selection. Because ambient conditions vary throughout the year, the average condensing temperature may be 24°C (969 kPa). With the lower discharge pressure, the average compressor CR is 3.65 and the ideal V_i is 2.52. Therefore, a compressor with the V_i at or close to 2.5 is the proper selection to optimize efficiency.

Some compressors are equipped with a variable V_i control. This makes compressor selection simpler, because the volume index can vary for different operating conditions. Therefore, the internal compression ratio can automatically match the external pressure ratio. Typically, screw compressors with variable V_i can control between 2.2 and 5.0 V_i. Variable-V_i compressors are beneficial over a wide range of system pressure ratios to improve efficiency as condensing pressures vary.

Piping. Oil-flooded screw compressors are the most common type of screw compressor used in refrigeration. Introduced in the late 1950s as an alternative to dry compressors with a symmetric rotor profile, oil-flooded compressors rapidly gained acceptance in many conventional reciprocating and small centrifugal applications. These compressors typically have oil supplied to the compression area at a volume rate of about 0.5% of the displacement volume. Some of this oil is used for lubricating the bearings and shaft seal. Typically, paraffinic or naphthenic mineral oils are used, though synthetics are being used more frequently on some applications.

The oil fulfills three primary purposes: sealing, cooling, and lubrication. The oil tends to fill any leakage paths between and around the screws. This provides a good volumetric efficiency even at high compression ratios. Normal volumetric efficiency exceeds 85% even with a compression ratio of 25. The oil sealing also helps maintain good volumetric efficiency with decreased operating speeds. The cooling function of the oil transfers much of the heat of compression from the gas to the oil, keeping typical discharge temperatures below 90°C. This allows high compression ratios without the danger of oil breakdown. The oil’s lubrication function protects the bearings, seals, and screw contact areas.

Oil injection to the screw compressor is normally achieved by one of two methods:

Some screw compressors may use a combination of both methods to achieve proper oil injection. The pump may only operate for a period of time when the compression ratio is below a set value. This option is shown schematically in Figure 5.

Oil injection requires an oil separator to remove the oil from the high-pressure refrigerant. Oil separators are designed to satisfy requirements of system type, refrigerant, and heat transfer equipment being used. Modern separation equipment routinely limits oil carryover to the refrigeration system to less than 5 mg/kg of oil in proportion to the circulated refrigerant.

Figure 5. Screw Compressor Flow Diagram with Optional Oil Pump

Because the oil absorbs a significant amount of the heat of compression in oil-flooded operation, oil cooling must be used to maintain low discharge temperatures. Common oil-cooling methods include the following:

It is important to remove the condensed liquid from the condenser continuously to clear the condensing heat transfer surface for fresh vapor. Properly designed piping around the condensers and receivers keeps the condensing surface at its highest efficiency by draining liquid ammonia out of the condenser as soon as it condenses and keeping air and other noncondensables purged.

Horizontal Shell-and-Tube Condenser and Through-Type Receiver. Figure 9 shows a horizontal water-cooled condenser draining into a through (top inlet) receiver. Ammonia plants do not always require controlled water flow to maintain pressure. Usually, pressure is adequate to force the ammonia to the various evaporators without water regulation. Each situation should be evaluated by comparing water costs with input power cost savings at lower condenser pressures.

Evaporative condensers are typically selected based on the desired maximum condensing pressure and the design wet-bulb temperature at the installed location. The 1% design wet bulb is the temperature that will be equaled or exceeded 1% of the year, or 87.6 hours. The resultant condensing pressure only equals or exceeds the design condition 1.0% of the time if the design wet-bulb temperature and peak design refrigeration load occur coincidentally. This peak condition depends less on condenser size and is more a function of how the load is calculated, what load diversity factor exists or is used in the calculation, and what safety factor is used in the calculations.

Figure 9. Horizontal Condenser and Top Inlet Receiver Piping

Location. If an evaporative condenser is located with insufficient space for air movement, the effect is the essentially same as that imposed by an inlet damper, and the fan may not deliver enough air. Such a location may also increase the chances of recirculating evaporative condenser discharge air, which reduces condenser capacity because the recirculated air has an increased wet-bulb temperature. For forced-draft condensers, the high inlet velocity causes a low-pressure region to develop around the fan inlet, inducing flow of discharge air into that region. If the obstruction is from a second condenser, the problem can be even more severe because discharge air from the second condenser flows into the air intake of the first. Induced-draft condensers are theoretically less prone to recirculation; however, inadequate space between multiple induced-draft condensers can still result in recirculation.

Prevailing winds can also contribute to recirculation. In many areas, winds shift with the seasons; wind direction during the peak high-humidity season is the most important consideration.

The tops of condensers should always be higher than any adjacent structure to eliminate downdrafts that might induce recirculation. Where this is impractical, discharge hoods can be used to discharge air far enough away from the fan intakes to avoid recirculation. However, the additional static pressure imposed by a discharge hood must be added to the fan system. Fan speed can be increased slightly to obtain proper air volume.

In areas with ambient temperatures below 0°C, water in the evaporative condenser drain pan and water circuit must be kept from freezing. When the temperature is at or below freezing, the evaporative condenser may be able to operate as a dry-coil unit and still maintain acceptable condensing pressures. In some cases (depending on weather, loads, and amount of condenser capacity), the water pump(s) and piping could be drained and secured for the season.

Another way to keep water from freezing is to drain the water in the condenser sump to an indoor water tank, as shown in Figure 10. When outdoor temperature drops, the condensing pressure drops, and a pressure switch with its sensing element in the discharge pressure line stops the water pump; the water is then drained into the tank. An alternative is to use a thermostat that senses water or outdoor ambient temperature and stops the pump at low temperatures. Exposed piping and any trapped water headers in the evaporative condenser should be drained into the indoor water tank.

Air volume capacity control methods include fan cycling in response to pressure controls, and two-speed or VFD fan motors.

Figure 10. Evaporative Condenser with Inside Water Tank

Installation. For a relatively small refrigeration system, a single evaporative condenser used with a through-type (top inlet) receiver can be connected as shown in Figure 11. The receiver must always be at the same pressure as the condensing pressure in the condenser to allow gravity to drain condensed liquid ammonia out of the condenser tube bundle. This is accomplished by locating the receiver below the elevation of the evaporative condenser outlets, by equalizing the receiver to the evaporative condenser inlet piping, and by ensuring that condenser drain line piping velocity is sized for sewer drainage (i.e., only partially full of liquid). For a single condenser circuit, this is accomplished by equalizing through the condenser drain line and limiting velocity to 0.5 m/s.

Figure 11. Single Evaporative Condenser with Top Inlet Receiver

Liquid Traps. For a through-type (top inlet) receiver, liquid traps are needed at the outlets when two or more condensers or condenser coils are installed so that variations in pressure drop through each condenser circuit do not interrupt continuous liquid draining from the condenser (Figure 12). An equalizer line is also necessary to ensure free drainage from condensers by keeping receiver and condenser pressures equal. For example, assume a 10 kPa pressure drop in the operating condenser in Figure 12, which produces a lower pressure (1290 kPa) at its outlet compared to the idle condenser (1300 kPa) and the receiver (1300 kPa). The trap creates a liquid seal so that a liquid height h of 1700 mm (equivalent to 10 kPa) builds up in the vertical individual condenser drain piping and not in the condenser coil.

Figure 12. Two Evaporative Condensers with Trapped Piping to Receiver

The trap must have enough height above the vertical liquid leg to accommodate a liquid height equal to the maximum pressure drop encountered in the condenser. The example shows the extreme case of one unit on and one off; however, the same phenomenon occurs to a lesser degree with two condensers of differing pressure drops when both are in full operation. Substantial differences in pressure drop can also occur between two different brands of the same size condenser or even different models produced by the same manufacturer.

The minimum recommended height of the vertical leg is 1500 mm for ammonia. This vertical dimension h is shown in all evaporative condenser piping diagrams. This height is satisfactory for operation within reasonable ranges around normal design conditions, and is based on the coil’s maximum condensing pressure drop. If service valves are installed at the coil inlets and/or outlets, the pressure drops imposed by these valves must be accounted for by increasing the minimum 1500 mm drop-leg height by an amount equal to the valve pressure drop in height of liquid refrigerant (Figure 13).

Figure 13. Method of Reducing Condenser Outlet Sizes

Figure 14. Piping for Shell-and-Tube and Evaporative Condensers with Top Inlet Receiver

Noncondensable gases can accumulate in an ammonia refrigeration system from incomplete evacuation, fluid chemical breakdown, and air entering any portion of the system operating at less than atmospheric pressure. These gases travel with the refrigerant and into the condenser, but do not condense and cannot pass through the liquid refrigerant filling the condenser outlet traps. Unless purged from the condenser circuits, they blanket the heat transfer surface and degrade condenser capability. Purging is accomplished by dedicated equipment connected through solenoid valves to taps located at the top of each condenser outlet. It is important to ensure that each solenoid valve is energized individually and not in combination with another purge solenoid valve on another condenser tube bundle, because this would connect the condenser outlets and bypass the traps.

Figures 14, 15, and 16 show various piping arrangements for evaporative condensers.

Figure 15. Piping for Parallel Condensers with Surge-Type Receiver

Horizontal Shell-and-Tube Condenser and Through-Type Receiver. Figure 9 shows a horizontal water-cooled condenser draining into a through (top inlet) receiver. Ammonia plants do not always require controlled water flow to maintain pressure. Usually, pressure is adequate to force the ammonia to the various evaporators without water regulation. Each situation should be evaluated by comparing water costs with input power cost savings at lower condenser pressures.

Figure 16. Piping for Parallel Condensers with Top Inlet Receiver

Water piping should be arranged so that condenser tubes are always filled with water. Air vents should be provided on condenser heads and should have hand valves for manual purging.

Receivers must be below the condenser so that the condensing surface is not flooded with ammonia. The piping should provide (1) free drainage from the condenser and (2) static height of ammonia above the first valve out of the condenser greater than the pressure drop through the valve.

The drain line from condenser to receiver is designed on the basis of 0.5 m/s maximum velocity to allow gas equalization between condenser and receiver. See Table 2 for sizing criteria.

Figure 17. Parallel Condensers with Top Inlet Receiver

Parallel Horizontal Shell-and-Tube Condensers. Figure 17 shows two condensers operating in parallel with one through-type (top inlet) receiver. The length of horizontal liquid drain lines to the receiver should be minimized, with no traps allowed. The shells are equalized by keeping liquid velocity in the drain line less than 0.5 m/s. The drain line can be sized from Table 2.

Air-Cooled Condensers. This condensing mode is less common than evaporative, but is receiving more attention as water, chemical treatment and sewer costs rise. Because air-cooled equipment works by sensible heat transfer and the approach between dry-bulb temperature and condensing temperature is unlikely to be less than 11 K, the increase in condensing temperature and compressor power must be factored into any selection process. Piping arrangements and airflow routing are similar to those shown for evaporative condensers, but it is crucial that air-cooled condensing tubes are installed level to drain liquid effectively. A secondary heat transfer surface of plate or spiral fins is common, and fin spacing is usually 1.7 to 3.2 mm. Airflow in blow-through or draw-through configurations deposits any atmospheric contaminants between the fins, and monthly cleaning is usually required. See Figure 7 in Chapter 39 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment for an example of this setup.

Figure 19 shows a flooded evaporator with a close-coupled low-pressure vessel for feeding ammonia into the coil and automatic water defrost.

The lower float switch on the float column at the vessel controls opening and closing of the liquid-line solenoid valve, regulating ammonia feed into the unit to maintain a liquid level. The hand expansion valve downstream of the solenoid valve should be adjusted so that it does not feed ammonia into the vessel more quickly than the vessel can accommodate while raising the suction pressure of gas from the vessel no more than 7 to 14 kPa.

Figure 19. Arrangement for Automatic Defrost of Air Blower with Flooded Coil

The static height of liquid in the vessel should be sufficient to flood the coil with liquid under normal loads. The higher float switch is to signal a high level of liquid in the vessel. It should be wired into an alarm circuit or possibly a compressor shutdown circuit if there is no other compressor protection. The float switches and/or columns should be insulated. With flooded coils having horizontal headers, distribution between the multiple circuits is accomplished without distributing orifices.

A combination evaporator pressure regulator and stop valve is used in the suction line from the vessel. During operation, the regulator maintains a nearly constant back pressure in the vessel. A solenoid coil in the regulator mechanism closes it during the defrost cycle. The liquid solenoid valve should also be closed at this time. One of the best means of controlling room temperature is a room thermostat that controls the effective setting of the evaporator pressure regulator.

A spring-loaded relief valve is used around the suction pressure regulator and is set so that the vessel is kept below 860 kPa (gage). Other suction line pressure control arrangements, such as a dual pressure regulator, can be used to eliminate the extra piping of the relief valve.

A solenoid valve unaffected by downstream pressure is used in the water line to the defrost header. The defrost header is constructed so that it drains at the end of the defrost cycle and the downstream side of the solenoid valve drains through a fixed orifice.

Unless the room is kept above 0°C, the drain line from the unit should be wrapped with a heater cable or provided with another heat source and then insulated to prevent defrost water from refreezing in the line.

Water line length in the space leading up to the header and the length of the drain line in the cooled space should be minimized. A flapper or pipe trap on the end of the drain line prevents warm air from flowing up the drain pipe and into the unit.

An air outlet damper may be closed during defrosting to prevent thermal circulation of air through the unit, which affects the temperature of the cooled space. The fan is stopped during defrost. This type of defrosting requires a drain pan float switch for safety control. If the drain pan fills with water, the switch overrides the time clock to stop flow into the unit by closing the water solenoid valve.

There should be a 5 min delay at the end of the water spray part of the defrosting cycle so water can drain from the coil and pan. This limits ice build-up in the drain pan and on the coils after the cycle is completed.

When the cycle finishes, the low-pressure vessel may be at about 517 kPa (gage). When the unit is opened to the much-lower-pressure suction main, some liquid surges out into the main; therefore, it may be necessary to gradually bleed off this pressure before fully opening the suction valve, to prevent thermal shock. Generally, a suction trap in the engine room removes this liquid before the gas stream enters the compressors.

The type of refrigerant control shown in Figure 20 can be used on brine spray units where brine is sprayed over the coil at all times to pick up the condensed water vapor from the airstream. The brine is reconcentrated continually to remove water absorbed from the airstream.

When a system has only one evaporator, a high-pressure float control can be used to keep the condenser drained and to provide a liquid seal between the high and low sides. Figure 20 shows a brine or water cooler with this type of control. The high-side float should be located near the evaporator to avoid insulating the liquid line.

The amount of ammonia in this type of system is critical: the charge must be limited so that liquid will not surge into the suction line under the highest loading in the evaporator. Some type of suction trap should be used. One method is to place a horizontal shell above the cooler, with suction gas piped into the bottom and out the top. The reduction of gas velocity in this shell causes liquid to separate from the gas and drop back into the chiller.

Figure 20. Arrangement for Horizontal Liquid Cooler and High-Side Float

Coolers should include a liquid indicator. A reflex glass lens with a large liquid chamber and vapor connections for boiling liquids and a plastic frost shield to determine the actual level should be used. A refrigeration thermostat measuring chilled-fluid temperature as it exits the cooler should be wired into the compressor starting circuit to prevent freezing.

A flow switch or differential pressure switch should prove flow before the compressor starts. The fluid to be cooled should be piped into the lower portion of the tube bundle and out of the top portion.

For multiple evaporator systems, low-side float valves are used to control the refrigerant level in flooded evaporators. The low-pressure float in Figure 21 has an equalizer line from the top of the float chamber to the space above the tube bundle and an equalizer line out of the lower side of the float chamber to the lower side of the tube bundle.

For positive shutoff of liquid feed when the system stops, a solenoid valve in the liquid line is wired so that it is only energized when the brine or water pump motor is operating and the compressor is running.

Figure 21. Piping for Evaporator and Low-Side Float with Horizontal Liquid Cooler

A reflex glass lens with large liquid chamber and vapor connections for boiling liquids should be used with a plastic frost shield to determine the actual level, and with front extensions as required. These chambers or columns should be insulated to prevent false levels caused by heat transfer from the surrounding environment.

Usually a high-level float switch is installed above the operating level of the float to shut the liquid solenoid valve if the float should overfeed.

Because copper and copper-bearing materials are attacked by ammonia, they are not used in ammonia piping systems. Steel or stainless steel piping, fittings, and valves of the proper pressure rating are suitable for ammonia gas and liquid.

Ammonia piping should conform to ASME Standard B31.5, and to IIAR Standard 2, which states the following:

Couplings, elbows, and tees for threaded pipe are for a minimum of 21 MPa design pressure and constructed of forged steel. Fittings for welded pipe should match the type of pipe used (i.e., standard fittings for standard pipe and extra-heavy fittings for extra-heavy pipe).

Tongue-and-groove or ANSI flanges should be used in ammonia piping. Welded flanges for low-side piping can have a minimum 1 MPa design pressure rating. On systems located in high ambients, low-side piping and vessels should be designed for 1.4 to 1.6 MPa. The high side should be 1.7 MPa if the system uses water-cooled or evaporative cooled condensing. Use 2.1 MPa minimum for air-cooled designs.

Joints between lengths of pipe or between pipe and fittings can be threaded if the pipe size is 32 mm or smaller. Pipe 40 mm or larger should be welded. An all-welded piping system is superior.

Threaded Joints. Many sealants and compounds are available for sealing threaded joints. The manufacturer’s instructions cover compatibility and application method. Do not use excessive amounts or apply on female threads because any excess can contaminate the system.

Welded Joints. Pipe should be cut and beveled before welding. Use pipe alignment guides and provide a proper gap between pipe ends so that a full-penetration weld is obtained. The weld should be made by a qualified welder, using proper procedures such as the Welding Procedure Specifications, prepared by the National Certified Pipe Welding Bureau (NCPWB).

Gasketed Joints. A compatible fiber gasket should be used with flanges. Before tightening flange bolts to valves, controls, or flange unions, properly align pipe and bolt holes. When flanges are used to straighten pipe, they put stress on adjacent valves, compressors, and controls, causing the operating mechanism to bind. To prevent leaks, flange bolts are drawn up evenly when connecting the flanges. Flanges at compressors and other system components must not move or indicate stress when all bolts are loosened.

Union Joints. Steel (21 MPa) ground joint unions are used for gage and pressure control lines with screwed valves and for joints up to 20 mm. When tightening this type of joint, the two pipes must be axially aligned. To be effective, the two parts of the union must match perfectly. Ground joint unions should be avoided if at all possible.

Piping should be at least 2.3 m above the floor. Locate pipes carefully in relation to other piping and structural members, especially when lines are to be insulated. The distance between insulated lines should be at least three times the thickness of the insulation for screwed fittings, and four times for flange fittings. The space between the pipe and adjacent surfaces should be three-fourths of these amounts.

Hangers located close to the vertical risers to and from compressors keep the piping weight off the compressor. Pipe hangers should be placed no more than 2.4 to 3 m apart, depending on pipe size, and within 0.6 m of a change in direction of the piping. Hangers should be designed to bear on the outside of insulated lines. Sheet metal sleeves on the lower half of the insulation are usually sufficient. Where piping penetrates a wall, a sleeve should be installed, and where the pipe penetrating the wall is insulated, it must be adequately sealed.

Piping to and from compressors and to other components must provide for expansion and contraction. Sufficient flange or union joints should be located in the piping so components can be assembled easily during installation and also disassembled for servicing.

Table 1 presents practical suction line sizing data based on 0.005 K and 0.01 K differential pressure drop equivalent per metre total equivalent length of pipe, assuming no liquid in the suction line. For data on equivalent lengths of valves and fittings, refer to Tables 10, 11, and 12 in Chapter 1. Table 2 lists data for sizing suction and discharge lines at 0.02 K differential pressure drop equivalent per metre equivalent length of pipe, and for sizing liquid lines at 0.5 m/s. Charts prepared by Wile (1977) present pressure drops in saturation temperature equivalents. For a complete discussion of the basis of these line sizing charts, see Timm (1991). Table 3 presents line sizing information for pumped liquid lines, high-pressure liquid lines, hot-gas defrost lines, equalizing lines, and thermosiphon lubricant cooling ammonia lines.

Many controls available for single-stage, high-temperature systems may be used with some discretion on low-temperature systems. If the liquid level is controlled by a low-side float valve (with the float in the chamber where the level is controlled), low pressure and temperature have no appreciable effect on operation. External float chambers, however, must be thoroughly insulated to prevent heat influx that might cause boiling and an unstable level, affecting the float response. Equalizing lines to external float chambers, particularly the upper line, must be sized generously so that liquid can reach the float chamber, and gas resulting from any evaporation may be returned to the vessel without appreciable pressure loss.

The superheat-controlled (thermostatic) expansion valve is generally used in direct-expansion evaporators. This valve operates on the difference between bulb pressure, which is responsive to suction temperature, and pressure below the diaphragm, which is the actual suction pressure.

The thermostatic expansion valve is designed to maintain a preset superheat in suction gas. Although the pressure-sensing part of the system responds almost immediately to a change in conditions, the temperature-sensing bulb must overcome thermal inertia before its effect is felt on the power element of the valve. Thus, when compressor capacity decreases suddenly, the expansion valve may overfeed before the bulb senses the presence of liquid in the suction line and reduces the feed. Therefore, a suction accumulator should be installed on direct-expansion low-temperature systems with multiple expansion valves.

System transients during pulldown can be managed by controlling compressor capacity. Proper load control reduces compressor capacity so that energy requirements stay within motor and condenser capacities. On larger systems using screw compressors, a current-sensing device reads motor amperage and adjusts the capacity control device appropriately. Cylinders on reciprocating compressors can be unloaded for similar control.

Alternatively, a downstream, outlet, or crankcase pressure regulator can be installed in the suction line to throttle suction flow if the pressure exceeds a preset limit. This regulator limits the compressor’s suction pressure during pulldown. The disadvantage of this device is the extra pressure drop it causes when the system is at the desired operating conditions. To overcome some of this, the designer can use external forces to drive the valve, causing it to be held fully open when the pressure is below the maximum allowable. Systems using downstream pressure regulators and compressor unloading must be carefully designed so that the two controls complement each other.

Compressor and evaporator capacity controls are similar for multi- and single-stage systems. Control methods include compressor capacity control, hot-gas bypass, or evaporator pressure regulators. Low pressure can affect control systems by significantly increasing the specific volume of the refrigerant gas and the pressure drop. A small pressure reduction can cause a large percentage capacity reduction.

System load usually cannot be reduced to near zero, because this results in little or no flow of gas through the compressor and consequent overheating. Additionally, high pressure ratios are detrimental to the compressor if it is required to run at very low loads. If the compressor cannot be allowed to cycle off during low load, an acceptable alternative is a hot-gas bypass. High-pressure gas is fed to the low-pressure side of the system through a downstream pressure regulator. The gas should be desuperheated by injecting it at a point in the system where it is in contact with expanding liquid, such as immediately downstream of the liquid feed to the evaporator. Otherwise, extremely high compressor discharge temperatures can result. The artificial load supplied by high-pressure gas can fill the gap between the actual load and the lowest stable compressor operating capacity. Figure 31 shows such an arrangement.

Microprocessor- and computer-based control systems are the norm for control systems on individual compressors as well as for whole-system control. Almost all screw compressors use microprocessor control systems to monitor all safety functions and operating conditions. These machines are frequently linked together with a programmable controller or computer for sequencing multiple compressors so that they load and unload in response to system fluctuations in the most economical manner. Programmable controllers are also used to replace multiple defrost time clocks on larger systems for more accurate and economical defrosting. Communications and data logging allow systems to operate at optimum conditions under transient load conditions even when operators are not in attendance.

Figure 31. Hot-Gas Injection Evaporator for Operations at Low Load

Most lubricants are immiscible in ammonia and separate out of the liquid easily when flow velocity is low or when temperatures are lowered. Normally, lubricants can be easily drained from the system. However, if the temperature is very low and the lubricant is not properly selected, it becomes a gummy mass that prevents refrigerant controls from functioning, blocks flow passages, and fouls heat transfer surfaces. Proper lubricant selection and management is often the key to a properly functioning system.

In two-stage systems, proper design usually calls for lubricant separators on both the high- and low-stage compressors. A properly designed coalescing separator can remove almost all the lubricant that is in droplet or aerosol form. Lubricant that reaches its saturation vapor pressure and becomes a vapor cannot be removed by a separator. Separators that can cool the discharge gas condense much of the vapor for consequent separation. Using lubricants that have very low vapor pressures below 80°C can minimize carryover to 2 or 3 mg/kg. Take care, however, to ensure that refrigerant is not condensed and fed back into the compressor or separator, where it can lower lubricity and cause compressor damage.

In general, direct-expansion and liquid overfeed system evaporators have fewer lubricant return problems than do flooded system evaporators because refrigerant flows continuously at good velocities to sweep lubricant from the evaporator. Low-temperature systems using hot-gas defrost can also be designed to sweep lubricant out of the circuit each time the system defrosts. This reduces the possibility of coating the evaporator surface and hindering heat transfer.

Flooded evaporators can promote lubricant build-up in the evaporator charge because they may only return refrigerant vapor back to the system. In ammonia systems, the lubricant is simply drained from the surge drum. At low temperatures, this procedure is difficult if the lubricant selected has a pour point above the evaporator temperature.

Lubricant Removal from Ammonia Systems. Most lubricants are miscible with liquid ammonia only in very small proportions. The proportion decreases with the temperature, causing lubricant to separate. Ammonia evaporation increases the lubricant ratio, causing more lubricant to separate. Increased density causes the lubricant (saturated with ammonia at the existing pressure) to form a separate layer below the ammonia liquid.

Unless lubricant is removed periodically or continuously from the point where it collects, it can cover the heat transfer surface in the evaporator, reducing performance. If gage lines or branches to level controls are taken from low points (or lubricant is allowed to accumulate), these lines will contain lubricant. The higher lubricant density is at a lower level than the ammonia liquid. Draining lubricant from a properly located collection point is not difficult unless the temperature is so low that the lubricant does not flow readily. In this case, keeping the receiver at a higher temperature may be beneficial. Alternatively, a lubricant with a lower pour point can be selected.

Lubricant in the system is saturated with ammonia at the existing pressure. When the pressure is reduced, ammonia vapor separates, causing foaming.

Draining lubricant from ammonia systems requires special care. Ammonia in lubricant foam normally starts to evaporate and produces a smell. Operators should be made aware of this. On systems where lubricant is drained from a still, a spring-loaded drain valve, which closes if the valve handle is released, should be installed.

Stop Valves. These valves, also commonly called shutoff or isolation valves, are generally manually operated, although motor-actuated units are available. ASHRAE Standard 15 requires these valves in the inlet and outlet lines to all condensers, compressors, and liquid receivers. Additional valves are installed on vessels, evaporators, and long lengths of pipe so they can be isolated in case of leaks and to facilitate pumping out for servicing and evacuation. Sections of liquid piping that can experience hydraulic lockup in normal operation must be protected with a relief device (preferably vented back into the system). Only qualified personnel should be allowed to operate stop valves.

Installing globe-type stop valves with the valve stems horizontal lessens the chance (1) for dirt or scale to lodge on the valve seat or disk and cause it to leak or (2) for liquid or lubricant to pocket in the area below the seat. Wet suction return lines (recirculation system) should use angle valves or globe valves (with their stems horizontal) to reduce the possibility of liquid pockets and reduce pressure drop.

Welded flanged or weld-in-line valves are desirable for all line sizes; however, screwed valves may be used for 32 mm and smaller lines. Ammonia globe and angle valves should have the following features:

Consider seal cap valves in refrigerated areas and for all ammonia piping. To keep pressure drop to a minimum, consider angle valves (as opposed to globe valves).

Control Valves. Pressure regulators, solenoid valves, check valves, gas-powered suction stop valves, and thermostatic expansion valves can be flanged for easy assembly and removal. Alternative weld-in line valves with nonwearing body parts are available. Valves 40 mm and larger should have socket- or butt-welded companion flanges. Smaller valves can have threaded companion flanges.

A strainer should be used in front of self-contained control valves to protect them from pipe construction material and dirt.

Solenoid Valves. Solenoid valve stems should be upright, with their coils protected from moisture. They should have flexible conduit connections, where allowed by codes, and an electric pilot light wired in parallel to indicate when the coil is energized.

Solenoid valves for high-pressure liquid feed to evaporators should have soft seats for positive shutoff. Solenoid valves for other applications, such as in suction, hot-gas, or gravity feed lines, should be selected for the pressure and temperature of the fluid flowing and for the pressure drop available.

Relief Valves. Safety valves must be provided in conformance with ASHRAE Standard 15 and Section VIII, Division 1, of the ASME Boiler and Pressure Vessel Code. For ammonia systems, IIAR Bulletin 109 also addresses the subject of safety valves.

Dual relief valve arrangements allow testing of the relief valves (Figure 32). The three-way stop valve is constructed so that it is always open to one of the relief valves if the other is removed to be checked or repaired.

Sections of piping that can be isolated between hand valves or check valves can be subjected to extreme hydraulic pressures if cold liquid refrigerant is trapped in them and subsequently warmed. Additional pressure-relieving valves for such piping must be provided.

Figure 32. Dual Relief Valve Fitting for Ammonia

Chapter 10 covers insulation and vapor retarders. Insulation and effective vapor retarders on low-temperature systems are very important. At low temperatures, the smallest leak in the vapor retarder can allow ice to form inside the insulation, which can totally destroy the integrity of the entire insulation system. The result can significantly increase load and power usage.

A typical two-stage, two-temperature screw compressor system provides refrigeration for high- and low-temperature loads (Figure 36). For example, the high-temperature stage supplies refrigerant to all process areas operating between −2 and 10°C. A –8°C intermediate suction temperature is selected. The low-temperature stage requires a −37°C suction temperature for blast freezers and continuous or spiral freezers.

Figure 36. Compound Ammonia System with Screw Compressor Thermosiphon Cooled

The system uses a flash intercooler that doubles as a recirculator for the −8°C load. It is the most efficient system available if the screw compressor uses indirect lubricant cooling. If refrigerant injection cooling is used, system efficiency decreases. This system is efficient for several reasons:

When plant refrigeration capacity must be increased and the system is operating below about 70 kPa (gage) suction pressure, it is usually more economical to increase capacity by adding a compressor to operate as the low-stage compressor of a two-stage system than to implement a general capacity increase. The existing single-stage compressor then becomes the high-stage compressor of the two-stage system. When converting, consider the following:

This section is based on a technical paper by Briley and Lyons (1992). Several methods are used for defrosting coils in areas below 4°C room temperature:

The evaporator (air unit) in a liquid recirculation system is circuited so that refrigerant flow provides maximum cooling efficiency. The evaporator can also work as a condenser if the necessary piping and flow modifications are made to receive hot-gas defrost (HGD) refrigerant vapor instead of liquid. When in this mode and with fans shut down, HGD vapor raises the primary and secondary surface temperature of the coil enough to melt any ice and/or frost so that it drains off. The HGD condenses to liquid in the process and is returned to the refrigeration system. Note that the substantial majority of heat transferred to the coil in this condensation process is latent, not sensible. Although this method is effective, it can be troublesome and inefficient if the piping system is not properly designed.

Even with the evaporator fans stopped, 50% or more of the heat given up by the refrigerant vapor may be lost to the space. Because this heat transfer rate varies with the temperature difference between coil surface and room air, the temperature/pressure of the refrigerant during defrost should be minimized.

Another reason to maintain the lowest possible defrost temperature/pressure, particularly in freezers, is to keep the coil from steaming. Steam increases refrigeration load, and the resulting icicle or frost formation must be dealt with. Icicles increase maintenance during cleanup, and ice formed during defrost tends to collect at the fan rings and restrict fan operation.

Defrosting takes slightly longer at lower defrost pressures, and defrost efficiency suffers slightly as a result. However, the lower pressure also produces a lower defrosting temperature and raises the overall defrosting efficiency because return vapor flow is reduced.

It is important to minimize HGD flow from leaving the coil before it is condensed and vented to the compressor through the wet return line. Float drainers allow liquid to pass but prevent gas flow, thus retaining the HGD vapor in the evaporator until it transfers its heat to the coil. A reduced amount of vapor can still pass through the drainer, and vapor is also produced when the condensed liquid refrigerant flashes from evaporator conditions down to a lower-pressure part of the system. Energy efficiency improves as the work required to compress this vapor decreases, so, in a two-stage system, evaporator defrost relief should be vented to the intermediate or high-stage compressor. Figure 38 shows a conventional hot-gas defrost system for evaporator coils of 50 kW of refrigeration and below. Note that the wet return is above the evaporator and that a single riser is used.

Figure 38. Conventional Hot-Gas Defrost Cycle (For coils with 50 kW refrigeration capacity and below)

Figure 39. Demand Defrost Cycle (For coils with 50 kW refrigeration capacity and below)

Defrost Control. Because defrosting efficiency is low, frequency and duration of defrosting should be kept to the minimum necessary to keep the coils clean. Less defrosting is required during winter than during hotter, more humid periods, so an effective energy-saving measure is to reset defrost schedules in the winter.

Several methods are used to initiate the defrost cycle. Demand defrost can be initiated by differential pressure sensors measuring air pressure drop across the coil, ice thickness sensors, and other devices. The coil is defrosted automatically only when necessary, which minimizes defrost cycles, duration, and energy usage. Demand initiation combined with a float drainer to send defrost refrigerant condensate to an intermediate vessel is the most efficient defrost system available (Figure 39). The most common method, however, is time-initiated, time-terminated defrost; it includes adjustable defrost duration and an adjustable number of defrost cycles per 24 h period. This function is commonly provided by a defrost timer or control system.

Estimates indicate that the load placed on a refrigeration system by a coil during defrost is up to three times the operating design load. Although estimates indicate that the maximum hot-gas flow can be up to three times the normal refrigeration flow, note that hot-gas flow varies during defrost depending on the amount of ice remaining on the coils: flow is greatest when defrost begins, and decreases as the ice melts and the coil warms. It is therefore not necessary to engineer for the maximum flow, but for some lesser amount. The lower flow imposed by reducing the hot-gas pipe and valve sizes reduces the maximum hot-gas flow rate and makes the system less vulnerable to various shocks. Estimates show that engineering for hot-gas flow rates equal to the normal refrigeration flow rate is adequate and only adds a small amount of time to the overall defrost period to achieve a clean defrost.

Designing Hot-Gas Defrost Systems. There are several approaches to designing hot-gas defrost systems. Figure 39 shows a typical demand defrost system for both upfeed and downfeed coils, which returns defrost liquid to the system’s intermediate-pressure level. A less efficient alternative is to direct defrost liquid into the wet suction. A float drainer or thermostatic trap with a hot-gas regulator installed at the hot-gas inlet to the coil is an alternative to the relief regulator (see Figure 39). When using a condensate drainer, the device must never be allowed to stop flow completely during defrost, because this allows the condensed hot gas remaining in the coil to pool in the lower circuits and become cold. Once this happens, defrosting of the lower circuits ceases. Water still running off the upper circuits refreezes on the lower circuits, resulting in ice build-up over successive defrosts. Any condensate drainer that can cycle closed when condensate flow momentarily stops should be bypassed with a metering valve or an orifice.

Sizing and Designing Hot-Gas Piping. Hot gas is supplied to the evaporators in two ways:

A maximum of one-third of the coils in a system should be defrosted at one time. If a system has 900 kW of refrigeration capacity, the main hot-gas supply pipe could be sized for 300 kW of refrigeration. Hot-gas mains should be sized one pipe size larger than the values given in Table 3 for hot-gas branch lines under 30 m. The outlet pressure-regulating valve should be sized in accordance with the manufacturer’s data.

Reducing defrost hot-gas pressure in the equipment room has advantages, notably that less liquid condenses in the hot-gas line as the condensing temperature drops to 11 to 18°C. A typical equipment room hot-gas pressure control system is shown in Figure 40. If hot-gas lines in the system are trapped, a condensate drainer must be installed at each trap and at the low point in the hot-gas line (Figure 41). Defrost condensate liquid return piping from coils where a float or thermostatic valve is used should be one size larger than the liquid feed piping to the coil.

Hot-gas defrost systems can be subject to hydraulic shock. See the section on Avoiding Hydraulic Shock, under Safety Considerations.

Soft Hot-Gas Defrost System. This system is particularly well suited to large evaporators and should be used on all coils of 50 kW of refrigeration or over. It eliminates the valve clatter, pipe movements, and some of the noise associated with large coils during hot-gas defrost. Soft hot-gas defrost can be used for upfeed or downfeed coils; however, the piping systems differ (Figure 42). Coils operated in the horizontal plane with vertical headers must be orificed. Vertical coils with horizontal headers that usually are crossfed are also orificed.

Figure 40. Equipment Room Hot-Gas Pressure Control System

Soft hot-gas defrost is designed to increase coil pressure gradually as defrost begins. This is accomplished by beginning the HGD flow at approximately 25 to 30% of full flow by use of an additional small solenoid valve, a multiple flow level solenoid valve or modulating control valve such that evaporator pressure rises to about 275 kPa (gage) in 3 to 5 min. HGD flow increases to full design value after this delay and evaporator defrost occurs (see Sequence of Operation in Figure 42). After defrost, similar logic opens a small suction-line solenoid first so that the coil can be brought down to operation pressure gradually before liquid is introduced, the main suction control valve is opened and the fans started. Evaporator pressure should be monitored by a pressure sensor to prevent premature opening of the main suction control valve to the suction line. Note that control valves are available to provide the soft-gas feature in combination with the main hot-gas valve capacity. There are also combination suction valves to provide pressure bleed down at the end of the defrost cycle.

Figure 41. Hot-Gas Condensate Return Drainer

Figure 42. Soft Hot-Gas Defrost Cycle (For coils with 50 kW refrigeration capacity or above)

The following additional features can make a soft hot-gas defrost system operate more smoothly and help avoid shocks to the system:

Static pressure penalty is the pressure/temperature loss associated with a refrigerant vapor stream bubbling through a liquid bath in a wet return riser. If speed in the riser is high enough, it will carry over a certain amount of liquid, thus reducing the penalty. For example, at −40°C, ammonia has a density of 689.9 kg/m³, which is equivalent to a pressure of 689.9(9.807 m/s²)/1000 = 6.77 kPa per metre of depth. Thus, a 5 m riser has a column of liquid that exerts 5 × 6.77 = 33.9 kPa. At −40°C, ammonia has a saturation pressure of 71.7 kPa. At the bottom of the riser then, the pressure is 33.9 + 71.7 = 105.6 kPa, which is the saturation pressure of ammonia at −33°C. This 7 K difference amounts to a 1.4 K penalty per metre of riser. If a riser were oversized to the point that the vapor did not carry liquid to the wet return, the evaporator would be at −33°C instead of −40°C. This problem can be solved in several ways:

If a single riser is sized for minimum pressure drop at full load, the static pressure penalty is excessive at part load, and lubricant return could be a problem. If the single riser is sized for minimum load, then riser pressure drop is excessive and counterproductive.

Figure 43. Recirculated Liquid Return System

Double risers solve these problems (Miller 1979). Figure 44 shows that, when maximum load occurs, both risers return vapor and liquid to the wet suction. At minimum load, the large riser is sealed by liquid ammonia in the large trap, and refrigerant vapor flows through the small riser. A small trap on the small riser ensures that some lubricant and liquid return to the wet suction.

Risers should be sized so that pressure drop, calculated on a dry-gas basis, is at least 70 Pa/m. The larger riser is designed for approximately 65 to 75% of the flow and the small one for the remainder. This design results in a velocity of approximately 25 m/s or higher. Some coils may require three risers (large, medium, and small).

Over the years, freezer capacity has grown. As freezers became larger, so did the evaporators (coils). Where these freezers are in line and the product to be frozen is wet, the defrost cycle can be every 4 or 8 h. Many production lines limit defrost duration to 30 min. If coils are large (some coils have a refrigeration capacity of 700 to 1000 kW), it is difficult to design a hot-gas defrost system that can complete a safe defrost in 30 min. Sequential defrost systems, where coils are defrosted alternately during production, are feasible but require special treatment.

Cold liquid refrigerant should not be confined between closed valves in a pipe where the liquid can warm and expand to burst piping components.

Hydraulic shock, also known as water hammer, occurs in two-phase systems experiencing pressure changes. Most engineers are familiar with single-phase water hammer, as experienced in water systems or occasionally in the liquid lines of refrigeration systems. These shocks, though noisy, are not widely known to cause damage in refrigeration systems. Damaging hydraulic shock events are almost always of the condensation-induced type. They occur most frequently in low-temperature ammonia systems and are often associated with the onset or termination of hot-gas defrosting. Failed system components are frequently evaporators, hot-gas inlet piping components associated with the evaporators, or two-phase suction piping and headers exiting the evaporators. Although hydraulic shock piping failures occur suddenly, there are usually reports of previous noise at the location of the failed component associated with hot-gas defrosting.

ASHRAE research project RP-970 (Martin et al. 2008) found that condensation-induced hydraulic shocks are the result of liquid slugs in two-phase sections of the piping or equipment. The slugs normally do not occur during the refrigeration cycle or the hot-gas defrost cycle, but during the transition from refrigeration to hot gas or back. During the transitions, pressure in the evaporator rises at the beginning of the cycle (i.e., gas from the system’s high side rushes into the low side), and is relieved at the end (i.e., gas rushes out into the suction side). At the beginning of these transitions, pressure imbalances are at their maximums, generating the highest gas flows. If the gas flows are sufficiently large, they scoop up liquid from traps or the bottom of two-phase pipes. Once the slug forms, it begins to compress the gas in front of it. If this gas is pushed into a partially filled evaporator or a section of piping without an exit (e.g., the end of a suction header), it will compress even more. Compression raises the saturation temperature of the gas to a point where it starts to condense on the cold piping and cold liquid ammonia. Martin et al. (2008) found that this condensation maintained a reasonably fixed pressure difference across the slug, and that the slug maintained a reasonably constant speed along the 6 m of straight test pipe. In tests where slugs occurred, pressure differentials across the slugs varied from about 35 to 70 kPa, and slug speeds from about 6 to 17 m/s. These slugs caused hydraulic shock peak pressures of as much as 5.2 MPa (gage).

Conditions that are most conducive to development of hydraulic shock in ammonia systems are suction pressures below 35 kPa (gage) and defrost pressures of 480 kPa (gage) or more. During the transition from refrigeration to defrost, liquid slugs can form in the hot-gas piping. If the evaporator or its inlet hot-gas piping are not thoroughly drained before defrosting begins, the slugs will impact the standing liquid in the undrained evaporator and cause shocks, possibly damaging the evaporator or its hot-gas inlet piping. During the transition from defrost back to refrigeration, the 480+ kPa (gage) gas in the evaporator is released into the suction piping. Liquid slugs can come from traps in the suction piping or by picking up slower-moving liquid in wet suction piping. These slugs can be dissipated at suction line surge vessels, but if the suction piping arrangement is such that an inlet to a dead-end section of piping becomes sealed, and the dead-end section is sufficiently long compared to its diameter, then a shock can occur as gas in the dead-end section condenses and draw liquid into the section behind it. The shock occurs when the gas is all condensed and the liquid hits the closure (e.g., an end cap or a valve in the off position). This type of shock has been known to occur in piping as large as 400 mm.

Low-temperature double pumper drum and low-temperature gas-powered transfer systems can also be prone to hydraulic shocks, because these systems use hot gas to move low-temperature liquid. If slugs form in the gas lines or gas is pumped into the liquid lines, then there is potential for hydraulic shock: trapped gas can condense, causing the liquid to impact a closed valve or other piping element.

To decrease the possibility of hydraulic shocks in ammonia systems, adhere to the following engineering guidelines:

Rusting pipes and vessels in older systems containing ammonia can create a safety hazard. Oblique x-ray photographs of welded pipe joints and ultrasonic inspection of vessels may be used to disclose defects. Only vendor-certified parts for pipe, valving, and pressure-containing components according to designated assembly drawings should be used to reduce hazards.

Most service problems are caused by inadequate precautions during design, construction, and installation (ASHRAE Standard 15; IIAR Standard 2). Ammonia is a powerful solvent that removes dirt, scale, sand, or moisture remaining in the pipes, valves, and fittings during installation. These substances are swept along with the suction gas to the compressor, where they are a menace to the bearings, pistons, cylinder walls, valves, and lubricant. Most compressors are equipped with suction strainers and/or additional disposable strainer liners for the large quantity of debris that can be present at initial start-up.

Moving parts are often scored when a compressor is run for the first time. Damage starts with minor scratches, which increase progressively until they seriously affect compressor operation or render it inoperative.

A system that has been carefully and properly installed with no foreign matter or liquid entering the compressor will operate satisfactorily for a long time. As piping is installed, it should be power rotary wire brushed and blown out with compressed air. The piping system should be blown out again with compressed air or nitrogen before evacuation and charging. See ASHRAE Standard 15 for system piping test pressure.