bypassed with a metering valve or an orifice.
Most defrost systems installed today (Figure 42) use a time clock to initiate defrost; the demand defrost system shown in Figure 43 uses a low-differential-pressure switch to sense the air pressure drop across the coil and actuate the defrost. A thermostat terminates the defrost cycle. A timer is used as a back-up to ensure the defrost terminates.
Sizing and Designing Hot-Gas Piping. Hot gas is supplied to the evaporators in two ways:
• The preferred method is to install a pressure regulator set at approximately 700 kPa (gage) in the equipment room at the hot- gas takeoff and size the piping accordingly.
• The alternative is to install a pressure regulator at each evaporator or group of evaporators and size the piping for minimum design condensing pressure, which should be set such that the pressure at the outlet of the coil is approximately 480 kPa (gage). This nor- mally requires the regulator installed at the coil inlet to be set to about 620 kPa (gage).
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 capac- ity, 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 equip- ment room hot-gas pressure control system is shown in Figure 44. 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 45). Defrost condensate liquid return piping from coils where a float
or thermostatic valve is used should be one size larger than the liq- uid feed piping to the coil.
Hot-gas defrost systems can be subject to hydraulic shock. See the section on Avoiding Hydraulic Shock, under Safety Consider- ations.
Demand Defrost. The following are advantages and features of demand defrost:
• It uses the least energy for defrost.
• It increases total system efficiency because coils are off-line for a minimum amount of time.
• It imposes less stress on the piping system because there are fewer defrost cycles.
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 move- ments, 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 46). Coils operated in the horizontal plane with vertical headers must be orificed. Ver- tical coils with horizontal headers that usually are crossfed are also orificed.
Soft hot-gas defrost is designed to increase coil pressure gradu- ally as defrost begins. This is accomplished by a small hot-gas feed having a capacity of about 25 to 30% of the estimated duty with a solenoid and a hand expansion valve adjusted to bring the pressure up to about 275 kPa (gage) in 3 to 5 min. (See Sequence of Opera- tion in Figure 46.) After defrost, a small suction-line solenoid is opened so that the coil can be brought down to operation pressure gradually before liquid is introduced and the fans started. The sys- tem can be initiated by a pressure switch; however, for large coils in spiral or individual quick freezing systems, manual initiation is pre- ferred. Note that control valves are available to provide the soft-gas feature in combination with the main hot-gas valve capacity. There
Fig. 42 Conventional Hot-Gas Defrost Cycle
Fig. 42 Conventional Hot-Gas Defrost Cycle
are also combination suction valves to provide pressure bleeddown at the end of the defrost cycle.
The following additional features can make a soft hot-gas defrost system operate more smoothly and help avoid shocks to the system: • Regulating hot gas to approximately 725 kPa (gage) in the equip- ment room gives the gas less chance of condensing in supply pip- ing. Liquid in hot-gas systems may cause problems because of the hydraulic shock created when the liquid is accelerated into an evaporator (coil). Coil headers and pan coils may rupture as a result.
• Draining condensate formed during the defrost period with a float or thermostatic drainer eliminates hot-gas blowby normally asso- ciated with pressure-regulating valves installed around the wet suction return line pilot-operated check valve.
• Returning liquid ammonia to the intercooler or high-stage recir- culator saves considerable energy. A 70 kW refrigeration coil defrosting for 12 min can condense up to 11 kg/min of ammonia, or 132 kg total. The enthalpy difference between returning to the
low-stage recirculator (–40°C) and the intermediate recirculator (–7°C) is 148 kJ/kg, for 19.5 MJ total or 27 kW of refrigeration removed from the –40°C booster for 12 min. This assumes that only liquid is drained and is the saving when liquid is drained to the intermediate point, not the total cost to defrost. If a pressure- regulating valve is used around the pilot-operated check valve, this rate could double or triple because hot gas flows through these valves in greater quantities.
Soft hot-gas defrost systems reduce the probability of experienc- ing hydraulic shock. See the section on Avoiding Hydraulic Shock, under Safety Considerations.
This system eliminates check valve chatter and most, if not all, liquid hammer (i.e., hydraulic problems in the piping). In addition, the last three features listed in the section on Demand Defrost apply to soft hot-gas defrost.
Fig. 43 Demand Defrost Cycle
Fig. 43 Demand Defrost Cycle
(For coils with 50 kW refrigeration capacity and below)
Fig. 44 Equipment Room Hot-Gas Pressure Control System
Fig. 44 Equipment Room Hot-Gas Pressure Control System
Fig. 45 Hot-Gas Condensate Return Drainer
Fig. 45 Hot-Gas Condensate Return Drainer
Double Riser Designs for Large Evaporator Coils
Static pressure penalty is the pressure/temperature loss associ- ated with a refrigerant vapor stream bubbling through a liquid bath. 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 sat- uration 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 evapo- rator would be at –33°C instead of –40°C. This problem can be solved in several ways:
• Install the low-temperature recirculated suction (LTRS) line be- low the evaporator. This method is very effective for downfeed evaporators. Suction from the coil should not be trapped. This arrangement also ensures lubricant return to the recirculator. • Where the LTRS is above the evaporator, install a liquid return sys-
tem below the evaporator (Figure 47). This arrangement eliminates static penalty, which is particularly advantageous for plate, individ- ual quick freeze, and spiral freezers.
• Use double risers from the evaporator to the LTRS (Figure 48). 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.
Double risers solve these problems (Miller 1979). Figure 48 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).
Fig. 46 Soft Hot-Gas Defrost Cycle
Fig. 46 Soft Hot-Gas Defrost Cycle
(For coils with 50 kW refrigeration capacity or above)
Fig. 47 Recirculated Liquid Return System
Fig. 47 Recirculated Liquid Return System
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.
SAFETY CONSIDERATIONS
Ammonia is an economical choice for industrial systems. Al- though ammonia has superior thermodynamic properties, it is con- sidered toxic at low concentration levels of 35 to 50 mg/kg. Large quantities of ammonia should not be vented to enclosed areas near open flames or heavy sparks. Ammonia at 16 to 25% by volume burns and can explode in air in the presence of an open flame.
The importance of ammonia piping is sometimes minimized when the main emphasis is on selecting major equipment pieces. Liquid and suction mains should be sized generously to provide low pressure drop and avoid capacity or power penalties caused by inad- equate piping. Hot-gas mains, on the other hand, should be sized conservatively to control the peak flow rates. In a large system with many evaporators, not all of them defrost simultaneously, so mains should only be engineered to provide sufficient hot gas for the num- ber and size of coils that will defrost concurrently. Slight undersiz- ing of the hot-gas piping is generally not a concern because the period of peak flow is short and the defrost cycles of different coils can be staggered. The benefit of smaller hot-gas piping is that the mass of any slugs that form in the piping is smaller.
Avoiding Hydraulic Shock
Cold liquid refrigerant should not be confined between closed valves in a pipe where the liquid can warm and expand to burst pip- ing 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 asso- ciated 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, the pressure imbalances are at their maximums, generating the highest gas flows. If the gas flows are sufficiently large, they will 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. Compres- sion raises the saturation temperature of the gas to a point where it starts to condense on the cold piping and cold liquid ammonia. Mar- tin et al. (2008) found that this condensation maintained a reason- ably 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. Liq- uid slugs can come from traps in the suction piping or by picking up slower-moving liquid in wet suction piping. These slugs can be dis- sipated 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 com- pared to its diameter, then a shock can occur as gas in the dead-end section condenses and draws 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 con- dense, causing the liquid to impact a closed valve or other piping element.
To decrease the possibility of hydraulic shocks in ammonia sys- tems, adhere to the following engineering guidelines:
• Hot-gas piping should include no liquid traps. If traps are unavoidable, they should be equipped with liquid drainers. • If hot-gas piping is installed in cold areas of the plant or outdoors,
the hot-gas condensate that forms in the piping should be drained and prevented from affecting the evaporator when the hot-gas valve opens.
• The evaporator must be fully drained before opening the hot-gas valve, giving any liquid slugs in the hot gas free flow through the evaporator to the suction piping. If the liquid slugs encounter
Fig. 48 Double Low-Temperature Suction Risers
Fig. 48 Double Low-Temperature Suction Risers