Display refrigerators are designed to merchandise food to maximum advantage while providing short-term storage. Proper maintenance of product temperature plays a critical role in food safety. An estimated 24 to 81 million people annually become ill from microorganisms in food, resulting in an estimated 10,000 needless deaths every year. As a result, in 1995 the Food and Drug Administration (FDA) Food Code recommended a lower storage temperature for certain refrigerated food products for further prevention of food-borne diseases. The FDA 2001 Food Code requires that the core temperature of meat, poultry, fish, dairy, deli, and cut produce not exceed 41°F throughout packaging, shipping, receiving, loading, and storing (FDA 2009).

Proper maintenance of product temperature relies heavily on the temperature of air discharged into the refrigerator. Table 1 lists discharge air temperatures in vario.us display refrigerators, although compliance with FDA requirements may require different refrigerator air temperatures. Figure 5 depicts a relationship between discharge air, return air, and average product temperatures for an open vertical meat display refrigerator. These profiles were obtained from controlled tests conducted over a 24 h period. Discharge and return air temperatures were measured at the air grille. As shown, all temperatures reach their peak at the end of each of four defrosts (Faramarzi et al. 2001).

Figure 5. Selected Temperatures in Open Vertical Meat Display Refrigerator

Product temperatures inside a display refrigerator may also vary, depending on the location of the product. Figure 6 depicts product temperature profiles and variations for an open vertical meat display refrigerator over a period of 24 h. As shown, the lowest product temperatures are observed at the top shelf near the discharge air grille, and the highest product temperatures are at the bottom shelf near the return air grille (Gas Research Institute 2000).

Display refrigerators are not designed to cool the product, but to maintain product temperature. When put into the refrigerator, merchandise should be at or near the proper temperature. Food placed directly into the refrigerator or into another adequately refrigerated storage space on delivery to the store should come from properly refrigerated trucks. Little or no delay in transferring perishables from storage or trucks to the display refrigerator or storage space should be allowed.

Display refrigerators should be loaded properly. Most manufacturers provide indicators of physical load limits that define the refrigerated zone. Product on display should never be loaded so that it is out of the load limit zone, or be stacked so that circulation of refrigerated air is blocked. The load line recommendations of the manufacturer must be followed to obtain good refrigeration performance. Proper refrigerator design and loading minimize energy use, maximize efficiency of the refrigeration equipment, maximize food safety, and minimize product loss.

Figure 6. Product Temperature Profiles at Four Different Locations Inside Multideck Meat Refrigerator (Average Discharge Air Temperature of 29°F)

In actual applications, however, products may not always be loaded properly. Survey results (Faramarzi 2003) reveal that improper loading of products inside display refrigerators may fall into the following categories:

Improper loading of the products can significantly affect maximum product temperatures, which adversely affects food safety and product loss. Figure 7 depicts the consequences of various improper product-loading scenarios on maximum product temperature of an open vertical meat display refrigerator (Faramarzi 2003).

Additionally, packaging may also affect food temperatures. The surface temperature of a loosely wrapped package of meat with an air space between the film and surface may be 2 to 4°F higher than the surrounding air inside the display refrigerator.

Display fixture performance is affected significantly by the temperature, humidity, and movement of surrounding air. Display refrigerators are designed primarily for supermarkets, virtually all of which are air conditioned.

Table 2 summarizes a study of ambient conditions in retail food stores. Individual store ambient readings showed that only 5% of all readings (including those when the air conditioning was not operating) exceeded 75°F db or 0.0102 lb of moisture per pound of dry air. Based on these data, the industry chose 75°F db and 64°F wb (55% rh, 57.5°F dew point) as summer design conditions. This is the ambient condition at which refrigeration load for food store display refrigerators is normally rated.

Store humidity is one of the most critical variables that can affect performance of display refrigerators and refrigeration systems. Store relative humidity may depend on climatic location, seasonal changes, and, most importantly, on the store dehumidification or HVAC system.

Figure 7. Comparison of Maximum Product Temperature Variations Under Different Improper Product Loading Scenarios in Open Vertical Meat Display Refrigerator

Figure 8. Comparison of Collected Condensate vs. Relative Humidity for Open Vertical Meat, Open Vertical Dairy/Deli, Narrow Island Coffin, and Glass-Door Reach-In Display Refrigerators (Gas Research Institute 2000)

Figure 8 shows an example of the relationship between refrigerator condensate and relative humidity. The increase in frost accumulation on the evaporator coils, and consequent increase in condensate, is more drastic for open vertical display refrigerators. In other words, open vertical fixtures are more vulnerable to humidity variations and remove more moisture from the ambient (or store) air than other types of display refrigerators (Gas Research Institute 2000).

High humidity can lead to product frost in low-temperature equipment, and can cause unwanted condensation on the interior and exterior of refrigerated equipment. Increased frost formation from higher relative humidity increases latent load, which the refrigeration system must remove (Figure 9). Additional defrosts may be needed to maintain the product at its desired temperature.

Figure 9. Percentage of Latent Load to Total Cooling Load at Various Indoor Relative Humidities (Gas Research Institute 2000)

When store ambient relative humidity is different from that at which the refrigerators were rated, the energy requirements for refrigerator operation will vary. Howell (1993a, 1993b) concludes that, compared to operation at 55% store rh, display refrigerator energy savings at 35% rh range from 5% for glass-door reach-in refrigerators to 29% for multideck deli refrigerators. Table 3 lists correction factors for the effect of store relative humidity on display refrigerator refrigeration requirements when the dry-bulb temperature is 70 and 78°F.

Manufacturers sometimes publish ratings for open refrigerators at lower ambient conditions than the standard because the milder conditions may significantly reduce the cooling load on the refrigerators. In addition, lower ambient conditions may allow both reductions in antisweat heaters and fewer defrosts, allowing substantial energy savings on a storewide basis.

The application engineer needs to verify that the year-round store ambient conditions are within the performance ratings of the various refrigerators selected for the store. Because relative humidity varies throughout the year, the dew point for each period should be analyzed. The sum of these refrigerator energy requirements provides the total annual energy consumption. In a store designed for a maximum relative humidity of 55%, the air-conditioning system will dehumidify only when the relative humidity exceeds 55%.

In climates where the outdoor air temperature is low in winter, infiltration of outdoor air and mechanical ventilation can cause store humidity to drop below 55% rh. Separate calculations need to be done for periods during which mechanical dehumidification is used and those when it is not required. For example, in Boston, Massachusetts, mechanical dehumidification is required for only about 3 1/2 months of the year, whereas in Jacksonville, Florida, it is required for almost 7 1/2 months of the year. Also, in Boston, there are 8 1/2 months when the store relative humidity is below 40%, whereas Jacksonville has these conditions for only 4 1/2 months. The engineer must weigh the savings at lower relative humidity against the cost of the mechanical equipment required to maintain relative store humidity levels, for example, below 40% instead of 55%.

Additional savings can be achieved by controlling antisweat heaters and reducing defrost frequency at ambient relative humidities below 55%. Energy savings credit for reduced use of display refrigerator antisweat heaters can only be taken if the display refrigerators are equipped with humidity-sensing controls that reduce the amount of power supplied to the heaters as the store dew point decreases. Also, defrost savings can be considered when defrost frequency or duration is reduced. Controls can reduce the frequency of defrost as store relative humidity decreases (demand defrost). Individual manufacturers give specific antisweat and defrost values for their equipment at stated store conditions. Less defrosting is needed as store dew point temperature or humidity decreases from the design conditions.

Note the condition in which store dry-bulb temperatures are higher than the industry standard, because this raises the refrigeration requirements and consequently the energy demand.

Heat transfer in a display refrigerator involves interactions between the product and the internal environment of the refrigerator, as well as heat from the surroundings that enters the refrigerator. Heat components from the surrounding environment include transmission (or conduction), radiation, and infiltration, whereas heat components from the internal environment include lights and evaporator fan motor(s). In addition, defrost and antisweat heaters also increase the cooling load of a display refrigerator. Conduction, radiation, and infiltration loads from the surroundings into the refrigerator, as well as heat exchanges between the product and parts of the refrigerator, depend on the temperatures of ambient air and air within the refrigerator. Open vertical display refrigerators rely on their air curtains to keep warm ambient air from penetrating into the cold environment inside the refrigerator. An air curtain consists of a stream of air discharged from a series of small nozzles through a honeycombed baffle at the top of the display refrigerator. Air curtains play a significant role in the thermal interaction of the display refrigerator with the surrounding air (see Figure 11).

The cooling load of a typical display refrigerator has both sensible and latent components. In general, the sensible portion consists of heat gain from lights, fan motor(s), defrost (electric and hot gas), antisweat heater, conduction, radiation, infiltration, and product pulldown load. The latent portion consists of infiltration and product latent heat of respiration.

Conduction Load. The conduction load refers to the heat transmission through the physical envelope of the display refrigerator. The temperature difference between air in the room and air inside the refrigerator is the main driving force for this heat transfer.

Radiation Load. Heat gain of the display refrigerator through radiation is a function of conditions inside the refrigerator, including surface temperature, surface emissivity, surface area, view factor with respect to the surrounding (store) walls/objects, floor, ceiling, and their corresponding emissivities and areas.

Infiltration Load. The infiltration load of the display refrigerator refers to the net entrainment of warm, moist air through the air curtain into the refrigerated space. The infiltration load has two components: sensible (i.e., direct heat added by the temperature difference between cold air in the refrigerator and warm room air drawn into the refrigerator) and latent (i.e., heat content of moisture added to the refrigerator by the room air drawn into the refrigerator). The total performance of the air curtain and the amount of heat transferred across it may depend on several factors, including

Internal Loads. The internal load includes heat energy from refrigerator components and from any necessary product pulldown. Examples include the following:

According to a test report by Gas Research Institute (2000), the major contributor to the total cooling load of open display refrigerators are infiltration and radiation (Figure 10). Infiltration constitutes approximately 80% of the cooling load of a typical medium-temperature open vertical display refrigerator. The relative role of infiltration diminishes for low-temperature open coffin (or tub) refrigerators, and is supplanted by radiation. Infiltration and radiation constitute roughly 24 and 43%, respectively, of the cooling load of a typical open coffin refrigerator.

Multideck open refrigerator shelves are an integral part of the air curtain and airstream. Without shelves, there will be substantial air distribution problems. An air deflector may be required when shelves are removed. As shown in Figure 10, infiltration through the air curtain plays a significant role in the cooling load of open vertical display refrigerators (Faramarzi 1999). Figure 11 depicts the air curtain velocity streamlines of an 8 ft open vertical meat display refrigerator. These velocity streamlines represent the actual airflow patterns using digital particle image velocimetry. As shown, warm air is entrained into the display refrigerator at several locations along the plane of the air curtain. Based on the law of conservation of mass, an equal (and substantial) amount of cold air from the display refrigerator spills into the room near the return air grille of the fixture.

Figure 10. Components of Refrigeration Load for Several Display Refrigerator Designs at 75°F db and 55% rh

Commercial refrigerators for market installations are usually of the endless construction type, which allows a continuous display as refrigerators are joined. Clear plastic panels are often used to separate refrigerator interiors when adjacent refrigerators are connected to different refrigeration circuits. Separate end sections are provided for the first and last units in a continuous display. Methods of joining self-service refrigerators vary, but they are usually bolted or cam-locked together.

Figure 11. Velocity Streamlines of Single-Band Air Curtain in Open Vertical Meat Display Refrigerator, Captured Using Digital Particle Image Velocimetry Technique

All refrigerators are constructed with surface zones of transition between the refrigerated area and the room atmosphere. Thermal breaks of various designs separate the zones to minimize the amount of refrigerator surface that is below the dew point. Surfaces that may be below the dew point include (1) in front of discharge air nozzles, (2) the nose of the shelving, and (3) front rails or center flue of the refrigerator. In glass-door reach-in freezers or medium-temperature refrigerators, the frame jambs and glass can be below the dew point. In these locations, resistance heat is used effectively to raise the exterior surface temperature above the dew point to prevent accumulation of condensation.

With the current emphasis on energy efficiency, designers have developed means other than resistance heat to raise the surface temperatures above the dew point. However, when no other technique is known, resistance heating becomes necessary. Control by cycling and/or proportional controllers to vary heat with store ambient changes can reduce energy consumption.

Store designers can do a great deal to promote energy efficiency. Not only does controlling the atmosphere within a store reduce refrigeration requirements, it also reduces the need to heat the surfaces of refrigerators. This heat not only consumes energy, but also places added demand on the refrigeration load.

Evaporators and air distribution systems for display refrigerators are highly specialized and are usually fitted precisely into the particular display refrigerator. As a result, they are inherent in the fixture and are not standard independent evaporators. The design of the air circuit system, evaporator, and means of defrosting are the result of extensive testing to produce the particular display results desired.

Because the evaporator coil is the most difficult part to clean, consider the judicious use of high-pressure, low-liquid-volume sanitizing equipment. This type of equipment enables personnel to spray cleaning and sanitizing solutions into the duct, grille, coil, and waste outlet areas with minimum disassembly and maximum effectiveness. However, this equipment must be used carefully because the high-pressure stream can easily displace sealing and caulking materials. High-pressure streams should not be directed toward electrical devices. Hot liquid can also break the glass on models with glass fronts and on closed-service fixtures.

Dairy Display. Dairy products include items with significant sales volume, such as fresh milk, butter, eggs, and margarine. They also include a myriad of small items such as fresh (and sometimes processed) cheeses, special above-freezing pastries, and other perishables. Available display equipment includes the following:

Meat Display. Most meat is sold prepackaged. Some of this product is cut and packaged on the store premises. Control of temperature, time, and sanitation from the truck to the checkout counter is important. Meat surface temperatures over 40°F shorten its salable life significantly and increase the rate of discoloration.

Figure 12. Multideck Dairy Display Refrigerator

Figure 13. Typical Walk-In Cooler Installation

Figure 14. Vertical Rear-Load Dairy (or Produce) Refrigerator with Roll-In Capability

The design of open fresh meat display refrigerators, either tub-type single-deck or vertical multideck, is limited by the freezing point of meat. Ideally, refrigerators are set to operate as cold as possible without freezing the meat. Temperatures are maintained with minimal fluctuations (with the exception of defrost) to ensure the coldest possible stable internal and surface meat temperatures.

Sanitation is also important. If all else is kept equal, good sanitation can increase the salable life of meat in a display refrigerator. In this chapter, sanitation includes limiting the amount of time meat is exposed to temperatures above 40°F. If meat has been handled in a sanitary manner before being placed in the display refrigerator, elevated temperatures can be more tolerable. When meat surfaces are contaminated by dirty knives, meat saws, table tops, etc., even optimum display temperatures will not prevent premature discoloration and subsequent deterioration of the meat. See the section on Meat Processing Rooms for information about the refrigeration requirements of the meat-wrapping area.

Along with molds and natural chemical changes, bacteria discolor meat. With good control of sanitation and refrigeration, experiments in stores have produced meat shelf life of one week and more. Bacterial population is greatest on the exposed surface of displayed meat because the surface is warmer than the interior. Although cold airflow refrigerates each package, the surface temperature (and thus bacterial growth) is cumulatively increased by

Improper control of these factors may cause meat surface temperatures to rise above values allowed by food-handling codes. Great care in every building and equipment detail, as well as in refrigerator loading, is necessary to maintain meat surface temperature below 40°F. However, the required diligence is rewarded by excellent shelf life, improved product integrity, higher sales volume, and less scrap or spoilage.

Surface temperatures rise during defrost. Tests compared matched samples of meat: one went through normal defrost, and the other was removed from the refrigerator during its defrosting cycles. Although defrosting characteristics of refrigerators vary, such tests showed that the effects on shelf life of properly handled defrosts are negligible. Tests for a given installation can easily be run to prove the effects of defrosting on shelf life for that specific set of conditions.

Self-Service Meat Refrigerators. Self-service meat products are displayed in packaged form. Processed meat can be displayed in similar refrigerators as fresh packaged meat, but at slightly higher temperatures. The meat department planner can select from a wide variety of available meat display possibilities:

All these refrigerators are available with a variety of lighting, superstructures, shelving, and other accessories tailored to special merchandising needs. Storage compartments are rarely used in self-service meat refrigerators.

Figure 15. Single-Deck Meat Display Refrigerator

Figure 16. Multideck Meat Refrigerator

Closed-Service Meat or Deli Refrigerators. Service meat products are generally displayed in bulk, unwrapped. Generally, closed refrigerators can be grouped in one of the following categories:

Closed-service meat display refrigerators are offered in a variety of configurations. Their fronts may be nearly vertical or angled up to 20° from vertical in flat or curved glass panels, either fixed or hinged, and they are available with gravity or forced-convection coils. Gravity coils are usually preferred for more critical products, but forced-air coil models using various forms of humidification systems are also common.

Figure 17. Closed-Service Display Refrigerator (Gravity Coil Model with Curved Front Glass)

These service refrigerators typically have sliding rear access doors, which are sometimes removed during busy periods. This practice is not recommended by manufacturers, however, because it affects the internal product display zone temperature and humidity.

Produce Display. Wrapped and unwrapped produce is often intermixed in the same display refrigerator. Ideally, unwrapped produce should have low-velocity refrigerated air forced up through the loose product. Water is usually also sprayed, either by manually operated spray hoses or by automatic misting systems, on leafy vegetables to retain their crispness and freshness. Produce is often displayed on a bed of ice for visual appeal. However, packaging prevents air from circulating through wrapped produce and requires higher-velocity air. Equipment available for displaying both packaged and unpackaged produce is usually a compromise between these two desired features and is suitable for both types of product. Available equipment includes the following:

Produce equipment is generally available with a variety of merchandising and other accessories, including bag compartments, sprayers for wetting the produce, night covers, scale racks, sliding mirrors, and other display shelving and apparatus.

Figure 18. Multideck Produce Refrigerator

Figure 19. Single-Deck Tub-Type Frozen Food Refrigerator

Figure 20. Single-Deck Island Frozen Food Refrigerator

Frozen Food and Ice Cream Display. To display frozen foods most effectively (depending on varied need), many types of display refrigerators are available. These include the following:

In a self-service meat market, cutting, wrapping, sealing, weighing, and labeling operations involve precise production control and scheduling to meet varying sales demands. The faster the processing, the less critical the temperature and corresponding refrigeration demand.

Figure 21. Multideck Frozen Food Refrigerator

Figure 22. Glass-Door, Medium-Temperature and Frozen Food Reach-In Refrigerator

The wrapping room should not be too dry, but condensation on the meat, which provides a medium for bacterial growth, should be avoided by maintaining a dew-point temperature within a few degrees of the sensible temperature. Fan-coil units should be selected with a maximum of 10°F temperature difference (TD) between the entering air and the evaporator temperature. Low-velocity fan-coil units are generally used to reduce the drying effect on exposed meat. Gravity coils are also available and have the advantage of lower room air velocities.

The meat wrapping area is generally cooled to about 45 to 55°F, which is desirable for workers but not low enough for meat storage. Thus, meat should be held in that room only for cutting and packaging; then, as soon as possible, it should be moved to a packaged product storage cooler held at 28 to 32°F. The meat wrapping room may be a refrigerated room adjacent to the meat storage cooler or one compartment of a two-compartment cooler. In such a cooler, one compartment is refrigerated at about 28 to 32°F and used as a meat storage cooler, and the second compartment is refrigerated at 45 to 55°F and used as a cutting and packaging room. Best results are attained when meat is cut and wrapped to minimize exposure to temperatures above 28 to 32°F.

At some point between the wrapping room and display refrigerator, refrigerated storage for the wrapped cuts of meat must be provided. Without this space, a balance cannot be maintained between the cutting/packaging rate and the selling rate for each particular cut of meat. Display refrigerators with refrigerated bottom storage compartments, equipped with racks for holding trays of meats, offer one solution to this problem. However, the amount of stored meat is not visible, and the inventory cannot be controlled at a glance.

A second option is a pass-through, reach-in cabinet. This cabinet has both front and rear insulated glass doors and is located between the wrapping room and the display refrigerators. After wrapping, the meats are passed into the cabinet for temporary storage at 28 to 32°F and then are withdrawn from the other side for restocking the display refrigerator. Because these pass-through cabinets have glass doors, the inventory of wrapped meats is visible and therefore controllable.

The third and most common option involves a section of the back room walk-in meat storage cooler or a completely separate packaged meat storage cooler. The cooler is usually equipped with rolling racks holding slide-in trays of meat. This method also offers visible inventory control and provides convenient access to both the wrapping room and the display refrigerators.

The overriding philosophy in successful meat wrapping and merchandising can be summarized thus: keep it clean, keep it cold, and keep it moving.

Each category of displayed food product that requires refrigeration for preservation is usually backed up by storage in the back room. This storage usually consists of refrigerated rooms with sectional walls and ceilings equipped with the necessary storage racks for a particular food product. Walk-in coolers are required for storage of meat, some fresh produce, dairy products, frozen food, and ice cream. Medium and large stores have separate produce and dairy coolers, usually in the 35 to 40°F range. Meat coolers are used in all food stores, with storage conditions between 28 and 32°F. Unwrapped meat, fish, and poultry should each be stored in separate coolers to prevent odor transfer. Walk-in coolers, which serve the dual purpose of storage and display, are equipped with either sliding or hinged glass doors on the front. These door sections are often prefabricated and set into an opening in the front of the cooler. In computing refrigeration load, allow for the extra service load.

Moisture conditions must be confined to a relatively narrow range because excessive humidity encourages bacteria and mold growth, which leads to sliming. Too little moisture leads to excessive dehydration.

Air circulation must be maintained at all times to prevent stagnation, but it should not be so rapid as to cause drying of an unwrapped product. Forced-air blasts must not be allowed to strike products; therefore, low-velocity coils are recommended.

For optimum humidity control, unit coolers should be selected at about a 10°F TD, with TD defined as the temperature difference between evaporator entering air temperature and saturation temperature of the refrigerant flowing through the evaporator. Although this definition holds for azeotropic refrigerant blends, it becomes a little more complicated with zeotropic refrigerant blends. A zeotropic refrigerant blend is a mixture made up of two or more refrigerants, each with its own distinct (and different) pressure/temperature characteristics. At a given pressure, the individual components will boil at their respective saturation temperature. The blend’s average saturation temperature depends on the individual components’ saturation temperatures and on the percentage of each component comprising the blend. As the blend begins to change state in the evaporator, each component boils at a different rate, based on saturation temperature at a given pressure: components with lower saturation temperatures boil at a faster rate than those with higher values.

The process of each component changing states at different rates is called fractionation. During fractionation, the component with the highest saturation temperature (and thus the last to boil) becomes increasingly concentrated in the evaporator, thereby raising the blend’s average saturation temperature. The blend’s saturation temperature as the first molecules start to change state from liquid to vapor is called bubble point, and the saturation temperature of the blend as the last few molecules of liquid change state from liquid to vapor is called the dew point. The difference between the bubble point and dew point is called the glide. For example, R-407A at 50 psig has a bubble point of 16°F, and a dew point of 26.3°F. The glide is 10.3°F (i.e., 26.3 − 16).

This presents a problem in determining what TD to use for evaporator selection. There is no industry standard, but one common approach is using the mean saturation temperature [(bubble point + dew point)/2] as the basis for TD. This approach avoids some of the issues involved in selecting TD based on a saturation temperature that varies throughout the evaporator piping run. However, the mean temperature changes depending on liquid temperature, and sufficient TD must be designed into the system to allow for superheat when measured from the dew point. Follow manufacturers’ ratings when designing the system.

Note that published ratings of commercial unit coolers do not reflect the effect of frost accumulation on the evaporator. The unit cooler manufacturer can determine the correct frost derating factor for its published capacity ratings. From experience, a minimum correction multiplier of 0.80 is typical.

A low-temperature storage capacity equivalent to the total volume of the low-temperature display equipment in the store is satisfactory. Storage capacity requirements can be reduced by frequent deliveries.

Generally, forced-air coils are selected for low-temperature coolers where humidity is not critical for packaged products. For low-temperature coolers, gas or electric defrost is required. Off-cycle defrosts are used in produce and dairy coolers. Straight time or time-initiated, time- or temperature-terminated gas or electric defrosts are generally used for meat coolers. For more details, see the section on Walk-In Coolers/Freezers in Chapter 16.

Food stores sell all types of perishable foods that require a variety of refrigeration systems to best preserve and most effectively display each product. Moreover, the refrigerating system must be highly reliable because it must operate 24 h per day for 10 or more years, to protect the large investment in highly perishable foods. Temperature controls vary greatly, from a produce preparation room (which may operate with a wet coil) requiring no defrost to the ice cream refrigerator requiring induced heat to defrost the coil periodically.

When selecting refrigeration equipment to operate display refrigerators and storage rooms for food stores, consider (1) cost/space limitations, (2) reliability, (3) maintainability and complexity, and (4) operating efficiency. Solutions span from the very simple (one compressor and associated controls on one refrigerator) to the complex (central refrigeration plant operating all refrigerators in a store).

Suction Groups. Various refrigerators have different evaporator pressure/SST temperature requirements. Produce preparation areas and meat wrapping rooms have the highest design temperature requirements of all the refrigeration loads in a typical supermarket, and may approach the suction pressures used in air-conditioning applications. Ice cream display units have the lowest design temperature requirements. Open vertical ice cream display freezers may have suction pressures corresponding to temperatures as low as −40°F, but they are only found in older stores in need of remodel. New store designs use either vertical glass door or single-deck island display freezers for ice cream display. These two styles have suction pressures corresponding to approximately −22°F. All other refrigerators and coolers fall between these extremes.

Refrigeration Loads. Refrigerator requirements are often given as refrigeration load per unit length. A lower value is sometimes allowed for more complex parallel systems, because peak loads are smaller with the reserve capacity present in multiple-compressor rack systems, making refrigerator temperature recovery after defrost less of a strain than on a single-compressor system.

Published refrigerator load requirements allow for extra capacity for temperature pulldown after defrost, per ASHRAE Standard 72. Ratings for open multideck refrigerators are typically for shelves in the flat position. Door cases follow the door opening sequence found in Standard 72. Heavily shopped stores may use more energy. The industry considers a standard store ambient condition to be 75°F and 55% rh, which should be maintained with air conditioning. Part of this air-conditioning load is carried by the open refrigerators, and credit for heat removed by them should be considered in sizing the air-conditioning system.

Equipment Selection. The designer matches the load requirements of the refrigerator lineups to the capacity of the chosen refrigeration system. Manufacturers publish load ratings to help match the proper refrigeration system with the fixture loads. For single-compressor applications only, the ratings can be stated (for selection convenience) as the capacity the condensing unit must deliver at an arbitrary suction pressure (evaporator temperature). In general, manufacturers of display refrigerators use ASHRAE Standard 72, which specifies standard methods of testing open and closed refrigerators for food stores. These standards establish refrigeration load requirements at rated ambient conditions of 75°F and 55% rh in the sales area with specific door-opening patterns. Display refrigerators for similar applications are commercially available from many manufacturers. Manufacturers’ recommendations must be followed to achieve proper results in both efficiency and product integrity.

Appropriate equipment selection depends on several factors.

Life-Cycle Cost. The total cost elements of the refrigeration system include not only the purchase price but also the operating cost (energy), cost of installation and commissioning, cost of maintenance and service, and the environmental cost.

Space Limitations. Store size, location, and price per square foot play a role in determining the type and location of equipment. Locations can include an equipment room at the back of the store, on a mezzanine, in a machine house on the roof, or distributed throughout or on top of the store.

Refrigerants and Fluids. Selection of a suitable refrigerant for food stores has been affected by international concern about the potential negative effects on the environment, particularly refrigerants’ ozone depletion potential (ODP) and global warming potential (GWP). Under international treaty, chlorofluorocarbon refrigerants are no longer produced, and hydrochlorofluorocarbon (HCFC) refrigerants (e.g., R-22) are being phased out under schedules that vary by country.

In the United States, R-22 and its blends are still used in existing equipment, but manufacture and import for use in new equipment stopped in 2010, and manufacture and import for servicing existing equipment will be banned in 2020.

In the United States and most other countries, the majority of new systems manufactured after January 2010 used some form of hydrofluorocarbon (HFC) refrigerants, such as R-404A, R-134a, and R-507.

At the time of writing, there have been several significant changes and reversals in the United States regarding the status Environmental Protection Agency (EPA) regulations affecting the use of certain high-GWP refrigerants that have been commonly used in commercial refrigeration applications.

Background and Timeline. A historical overview of actions over the past few years follows.

June 25, 2013: Presidential Climate Action Plan released, which specifically referenced the future of commonly used HFC refrigerants. The plan supported a phasedown (not phaseout) of production and consumption of HFC refrigerants in the context of the Montreal Protocol over the next 30 years. Additionally, it called for immediate efforts to move away from HFCs to more environmentally friendly alternatives. This was to be implemented by the EPA using the Significant New Alternatives Policy (SNAP).

August 5, 2014: EPA original proposal for changing status of certain refrigerants used in specific applications. The change of status was from acceptable for use (listed) to unacceptable for use (de-listed) in certain and specific applications.

July 20, 2015: The EPA issued its final rule changing the status of certain refrigerants used in specific applications, using Section 612 of the Clean Air Act as the source of its statutory authority to implement these rules and changes. Its changes included the following:

In February 2017, a suit was brought against the EPA in the U.S. Court of Appeals (D.C. Circuit), stating that section 612 of the Clean Air Act did not grant the EPA the statutory authority to delist refrigerants with global warming potential. In August 2017, the court decided for the petitioner and against the EPA, resulting in a nullification of all regulations involving the delisting of HFC refrigerants.

A petition was filed with the D.C. Circuit Court of Appeals for a rehearing of the decision, but in January 2018, the court announced its decision to deny this petition.

As a result of the proposed regulatory changes, commercial refrigeration equipment manufacturers dedicated significant budgetary expenditures for research and development of equipment compatible with, and rated for, lower-GWP alternative refrigerants. The available nonnatural refrigerant choices have been limited to HFC blends or HFC/HFO blends with lower GWPs. Although the list of available HFC blends with lower GWP is long, R-407A has become the predominant choice in new supermarket applications (GWP approximately 50% of R-404A and R-507). Newer HFC/HFO blends such as R-448A, R-449A, and R-449B (with GWPs approximately 32% of R-404A and R-507) are seeing interest in some new supermarket applications.

Future Legislation. A bill (The American Innovation and Manufacturing Act) has been introduced in the U.S. Senate that, if passed, would entitle the EPA to phase down HFCs used in refrigeration and air conditioning, in consultation with the HVAC&R industry. The proposed bill would operate in accordance with the guidelines set forth in the 2015 Kigali Amendment to the Montreal Protocol. Although the United States has not ratified the Kigali Amendment at the time of this writing, the U.S. State Department issued a statement (November 23, 2017) that is had initiated the process to consider ratification of the amendment.

Use in the Field. Some natural refrigerants see limited use in the marketplace (ASHRAE 2009). Carbon dioxide is used in transcritical and cascade systems, and as the secondary refrigerant in secondary refrigerant systems. R-290 (propane) has been used in self-contained refrigerated display cases since 2012. In some European countries, R-290 and R-717 (ammonia) are used as primary refrigerants in secondary systems. For details on these and other refrigerants, see ASHRAE Standard 34.

Retrofit recommendations have been developed by equipment and refrigerant manufacturers to guide stores in converting from ozone-depleting and/or high-GWP substances to more environmentally friendly alternatives. Many of the new refrigerants produce discharge temperatures in higher ranges that could lead to possible chemical decomposition. System manufacturers may require compressor options such as liquid injection or demand cooling to keep discharge temperatures within acceptable levels. Close consultation with equipment manufacturers is necessary to stay current on this issue.

Compressor performance and material compatibility were two major concerns in selecting new refrigerants to replace chlorinated refrigerants. Neoprene seals swell in the presence of refrigerants, and at higher rates when chlorinated refrigerants are present. Thus, after converting a system from a CFC or HCFC to an HFC blend or HFC/HFO blend, neoprene seals will effectively shrink (swell less). Industry recommendations require replacing all neoprene seals when undertaking a refrigerant conversion.

Secondary loop systems and fluid-cooled, self-contained systems require appropriate secondary fluids, because viscosity and heat transfer properties directly affect system performance. Secondary fluids are often divided into two categories: single phase and two phase. Single-phase secondary fluids absorb heat by means of sensible heat transfer, meaning that they change temperature as heat is added. Examples include water; propylene and ethylene glycols; and salt solutions or brines such as potassium formate, potassium acetate, sodium chloride, and calcium chloride. Two-phase secondary fluids absorb heat by means of changing phase rather than changing temperature. The phase change may be from solid to liquid, or from liquid to vapor. Examples include ice slurries (solid to liquid) and CO₂ (liquid to vapor). Ice slurries typically comprise two materials, either water/glycol or water/ethanol mixtures; application of these materials has been limited. CO₂ has seen wider use in both commercial and industrial applications. For further information, see the sections on specific operations and applications; additional information on selection of single-phase secondary fluids can be found in Chapter 13. More information on refrigerants, secondary fluids, and their properties can be found in the 2017 ASHRAE Handbook—Fundamentals.

Refrigerant Lines. Sizing liquid and suction refrigerant lines is critical in most refrigeration installations, because of the typically long horizontal runs and frequent use of vertical risers. Correct liquid-line sizes are essential to eliminate unnecessary frictional pressure loss, which could hinder the system’s ability to provide a full feed of liquid to the expansion valve. In addition, oversizing must be avoided so the system does not contain unnecessary refrigerant charge; not only will this reduce the carbon footprint of the equipment, but it ensures that the liquid receiver provided with the equipment will be of sufficient capacity to allow for a complete system pumpdown.

Proper suction-line sizing is required to ensure adequate oil return to the compressor without excessive pressure drop. Oil separates in the evaporator and moves toward the compressor more slowly than the refrigerant. Unless the suction line is properly sized and installed, oil can accumulate in low places, causing problems such as compressor damage from oil slugging or insufficient lubrication. Undersized suction lines result in excessive pressure drop between the refrigerated fixture and the compressor(s), requiring the compressor(s) to operate at a suction pressure lower than design to maintain the design refrigerated fixture temperature, thereby reducing reduced compressor (and system) capacity. To prevent these problems, horizontal suction lines must pitch down as gas flows toward the compressor, the bottoms of all suction risers must be trapped, and refrigerant speed in suction risers must be maintained according to piping practices described in Chapters 1 and 2. To overcome the larger pressure drop necessary in suction risers, suction lines may be oversized on long horizontal runs; however, they still must pitch down toward the compressor for good oil return.

The modern supermarket is a complex system, and will be subject to periods where the actual load is lower than design. In particular, because use of glass door refrigerators is widespread, limited or nonexistent customer activity during off-peak hours may result in extended periods where the doors remain closed and result in below-design load conditions. Open cases that use a night curtain also experience low loads during off hours. These circumstances could result in the suction riser being oversized at low load, impeding oil return (because of reduced refrigerant velocity). A double suction riser is an option in applications where low/reduced load might impede oil return.

Manufacturers’ recommendations and appropriate line sizing charts should be followed to avoid adding heat to either suction or liquid lines. In large stores, both types of lines can be insulated profitably, particularly if subcooling is used.

Refrigeration systems in use today can generally be categorized into one of the following types: single (a single compressor connected to one or more evaporator loads), multiplex (or parallel compressor) rack, loop, distributed, and secondary refrigerant. Each type has distinct advantages and disadvantages, and may be chosen based on the weight a designer assigns to the different components of equipment life-cycle cost.

The most common compressors used in a typical supermarket refrigeration system include reciprocating, scroll, and screw compressors, which are discussed in Chapter 38 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment. Planning load management and sizing the compressors are very important to a successful refrigeration installation. Factory assembly of the necessary compressor systems with either a direct air-cooled condenser or any style of remote condenser is common practice. Both single and parallel systems can be housed, prepiped, and prewired at the factory. The complete unit is then delivered to the job site for placement on the roof or beside the store.

Many supermarket designers choose to have compressor equipment installed in factory-prefabricated housing, commonly called a mechanical center, to reduce real estate costs for the building. The time requirements for installation of piping and wiring may also be reduced with prefabrication. Most of the rooms are modular and prewired and include some refrigeration piping. Their fabrication in a factory setting should offer good quality control of the assembly. They are usually put into operation quickly upon arrival at the site.

Single System. A single-compressor/single-evaporator system is sometimes referred to as a conventional system. Each compressor may be piped to an individual condenser, or several single compressors may be piped to a larger condenser with multiple circuits. Some single-compressor systems are connected to two or more evaporator systems, in which case each evaporator system uses its own liquid and suction lines and is controlled independently.

A solid-state pressure control for single systems can help control excess capacity when ambient temperature drops. The control senses the pressure and adjusts the cutout point to eliminate short cycling, which ruins many compressors in low-load conditions. This control also saves energy by maintaining a higher suction pressure than would otherwise be possible and by reducing overall running time. Some applications might benefit from the addition some form of compressor capacity control, to more closely match the compressor capacity to the actual load at any given point in time. Some of the more common methods are (1) mechanical compressor unloading, (2) digital unloading for semihermetic or scroll compressors, and (3) variable-frequency drive (VFD).

Multiplex System. Another common refrigeration technique couples two or more compressors in parallel (multiplex compressor rack), piped together with a common suction manifold (header) and common discharge manifold (header). The compressors share a common oil management system and usually operate connected to one or more large condensers. The condensers are usually remote air cooled or evaporative cooled, but they can also be built as part of the compressor rack assembly. The multiplex rack system has several evaporator systems, each individually controlled and individually piped to the compressor rack’s common liquid and suction line.

Multiple-evaporator systems are usually designed such that each evaporator system operates at a different saturated suction temperature (pressure). Because all of the evaporator systems are connected to one common suction manifold, the compressors are forced to operate at the saturated suction temperature (pressure) required to maintain the design temperature of the coldest operating evaporator system. The obvious result is a sacrifice in compressor efficiency. Running all the equipment at the low suction pressure required for ice cream (on low-temperature systems) or for fresh meat (on medium-temperature systems) causes all the other compressors to operate at lower suction pressures than are necessary. To overcome this inefficiency, large parallel systems frequently isolate the suction piping from the lowest operating suction group (such as ice cream on low-temperature compressor racks, or fresh meat on medium-temperature compressor racks) and connect it to an independent satellite compressor. The satellite compressor has its own independent suction connection to the specified load, but shares the rack system’s common discharge piping and oil management system. Similarly, split-suction manifolds are often used for larger loads where two different suction pressure groups are combined and connected to a single compressor rack. The compressors suction groups are isolated, and piped to the suction header of their respective compressor group, but all compressors discharge into a common header and share the oil management system.

Consult manufacturers to determine the appropriate suction pressure (temperature) at the fixture and the load that each system adds to the total. The multiplex rack system must then be designed to deliver the total of all the loads at a common suction pressure no higher than the lowest system pressure requirement less the suction line pressure drop. Individual evaporator systems designed to operate at suction pressures higher than the common must use some means of suction line regulation to prevent higher-temperature evaporators from operating at temperatures below what is necessary to maintain product temperatures.

Suction pressure can be regulated in one of two ways:

In conventionally piped multiplex systems (individual liquid and suction lines for each evaporator system), EPRs, EEPRs, and liquid solenoid valves can be installed either at the rack header of near the refrigerator lineups.

The suction gas temperature leaving display fixtures should be superheated to ensure that only vapor enters the compressor suction intake. Be careful to avoid excessive temperature of suction vapor returning to the compressor, because this yields a less dense vapor and reduces the compressor’s mass flow pumping capacity. This is particularly crucial on low-temperature fixtures, because the suction line vapor temperature increase from heat gained from the store ambient (or outdoor ambient, for remote roof piping runs) can be substantial, not only adversely affecting compressor capacity, but also elevating compressor discharge vapor temperature. This must be considered for system design.

One method of ensuring that suction vapor temperature is not abnormally high is to simply set the system thermostatic expansion valves (TEVs) to their design superheat set point. If the TEV is underfeeding, the resulting high superheat at the evaporator outlet contributes to higher-than-normal suction vapor temperature at the compressor inlet. In addition, using sufficiently sized insulation on the suction piping between the refrigerated fixture outlet and the compressor inlet can help keep the suction vapor temperature at acceptable temperatures. If the suction vapor temperature is too high after these two practices are followed, then a desuperheating expansion valve can be used to inject a quantity of saturated liquid refrigerant into the suction line upstream of the compressor inlet, to lower the suction vapor temperature. This method for controlling liquid injection can maintain either a constant suction vapor temperature or a constant discharge vapor temperature.

To ensure proper thermostatic expansion valve operation, the engineer should verify that liquid entering the fixture is subcooled. For systems without a mechanical subcooler, some amount of natural subcooling is required to ensure that the liquid entering the thermostatic expansion valve is vapor free. For system designs with mechanical subcooling, liquid-line insulation is required to ensure that the liquid entering the thermostatic expansion valve is at the design liquid temperature condition.

Parallel operation is also applied in two-stage or compound systems for low-temperature applications. Two-stage compression includes interstage gas cooling before the second stage of compression to avoid excessive discharge temperatures. A multiplex rack system with multiple compressors of equivalent capacity is called an even parallel system; with compressors of different capacities, it is called an uneven parallel system.

Parallel compressor systems must be designed to maintain proper refrigerator temperatures under peak summer load. During the rest of the year, store conditions can be easily maintained at a more ideal condition, and refrigeration load is lower. In the past, refrigeration systems operated at 90°F condensing conditions or above to maintain enough high-side pressure to feed the refrigerated display fixture expansion valves properly. When outdoor ambient conditions allow, current technology allows the condensing temperature to follow the ambient down to about 70°F or less. When proper liquid-line piping practices and thermostatic expansion valve (TEV) or electric expansion valve (EEV) selection guidelines are followed, the expansion valves feed the evaporators properly under these varying condensing pressures (temperatures) between the summer and winter conditions.

TEV/EEV capacity is based on (1) the available liquid refrigerant ΔP across the valve’s port, with the lower condensing temperature (pressure) in the winter reducing valve capacity, and the higher condensing temperature (pressure) in the summer increasing valve capacity; and (2) the temperature of the liquid refrigerant entering the valve. In applications without mechanical subcooling, the reduced liquid refrigerant temperature during lower-condensing-temperature operation increases valve capacity, closely offsetting the capacity reduction caused by the lower condensing pressure. Likewise, the higher liquid refrigerant temperature during higher-condensing-temperature operation decreases valve capacity, closely offsetting the capacity increase caused by the higher condensing pressure. In applications where a mechanical subcooler is used, the constant liquid temperature under both lower- and higher-condensing-temperature operation does not counteract the varying condensing pressure in these two ambient extremes. As such, the lower condensing pressure during lower-condensing-temperature operation results in a substantial capacity loss to the TEV or EEV. TEV/EEV capacity should be verified in both summer and winter design conditions to ensure that adequate capacity is available in the winter (lower condensing) condition, and capacity is not excessive in the summer (higher condensing) condition.

Multiple compressors may be controlled or staged based on a drop in system suction pressure. If the compressors are equal in size, a mechanical device can turn off one compressor at a time until only one is running. The suction pressure will be perhaps 5 psi or more below optimum. Microprocessors are the standard control for multiple compressor systems and offer the option of remote control and system operation for all types of compressors, managing compressor cycling and run time for each compressor, and ensuring the common suction pressure is optimized. Satellite compressors can be controlled accurately with one control that also monitors other components, such as oil pressure and alarm functions. To match changing evaporator loads, rack capacity can be varied by cycling compressors, varying the speed of one or more compressors, and/or unloading compressor cylinders by closing valves or moving ports on screw compressors.

Unequally sized compressors can be staged to obtain more steps of capacity than is possible with the same number of equally sized compressors. Figure 23 shows seven stages of capacity from a 5, 7, and 10 hp compressor parallel arrangement. In addition, using a lead compressor with digital unloading capability further enhances the microprocessor controller’s ability to maintain precisely consistent common suction pressure.

In the multiplex system design, a typical supermarket includes one or more medium-temperature parallel compressor systems for meat, deli, dairy, and produce refrigerators and medium-temperature walk-in coolers. The system may have a satellite compressor for meat or deli refrigerators, or all units may have a single compressor. Energy efficiency ratios (EERs) typically range from 8 to 9 Btu/h per watt for the main load. Low-temperature refrigerators and coolers are grouped on one or more parallel systems, with ice cream refrigerators on a satellite or on a single compressor. EERs range from 4 to 5 Btu/h per watt for frozen-food units to as low as 3.5 to 4.0 Btu/h per watt for ice cream units. Cutting and preparation rooms are most economically placed on a single unit because the refrigeration EER is nearly 10 Btu/h per watt. Air-conditioning compressors are also separate because their EERs can range up to 11 Btu/h per watt (Figure 24).

Figure 23. Stages with Mixed Compressors

Figure 24. Typical Single-Stage Compressor Efficiency

Controls. Refrigerant systems increasingly use electronic microprocessor control systems. The microprocessor controls all of the major components and valves in the refrigeration system. Algorithms control evaporator temperatures, defrosts, condensers, and compressors, and optimize performance and reduce energy. Many systems allow remote computer access and data logging. Systems can be designed with a central process or with a supervisor that monitors individual control boards for the components. Typically, the individual control boards can maintain operation if the supervisor fails.

Loop Systems. A loop system is simply multiplex rack system with a variation in the piping layout. Rather than each evaporator system (or circuit) having its own individual liquid and suction line connecting it to the compressor rack, a single suction and liquid loop is piped out to the store for each common suction pressure group. Individual branch circuits are then connected to the loop near the fixtures. If EPRs, EEPRs, and solenoid valves are used, they must be installed in the branch circuit, near the refrigerator lineups for easy access.

Secondary Loop Systems. In secondary coolant systems, heat is removed from refrigerated spaces and display cabinets by circulating a chilled fluid in a secondary loop cooled by a primary refrigeration system. Fluid circulation is typically provided by a centrifugal pump(s) designed for the flow rate and pressure drop required by the system load and piping arrangement.

Performance Characteristics. Secondary coolant systems have several advantages. Because primary refrigeration piping is located almost wholly within the machine room, the amount of piping and refrigerant required can be reduced by as much as 80 to 90%. Because field piping of the primary system is typically limited to only a few joints, the majority of the primary system piping joints are factory installed. Factory-installed joints are generally of higher quality than field-installed joints, because they are formed in controlled conditions by skilled labor, using nitrogen and a variety of pressure-testing and leak-identification methods. Higher-quality joints combined with a lower refrigerant charge can significantly lower refrigerant leakage rates, reducing the environmental effects associated with the primary refrigerant. The compressors and evaporator are close coupled, so suction line pressure losses and heat gains are minimized, enhancing system performance. Secondary coolant systems are inherently less complex than direct-expansion types, requiring fewer and less complicated valves and control devices. Less expensive nonmetallic piping systems and components can also be used for the secondary fluid piping circuit if that system’s operating pressure is low (typically less than 60 psig). Service of the refrigeration system is basically limited to the machine room area, and maintenance costs can be reduced. Because a fluid loop is used, thermal storage may be applied to reduce peak power demands and take advantage of lower off-peak utility rates. Ambient or free cooling may be used in areas with colder climates. Secondary systems have reduced refrigerant charge and can use primary refrigerants not typically suitable for direct-expansion systems with large refrigerant charge requirements, including ammonia and hydrocarbons (where local codes allow).

Disadvantages of secondary systems include thermodynamic loss inherent in the additional step of heat transfer in the chiller, as well as the energy consumed by the fluid pump and the heat it transfers to the circulating fluid. Insulation must also be applied to both coolant supply and return lines to minimize heat gain.

Secondary Fluid Selection. The secondary fluid is critical to system efficiency because viscosity and heat transfer properties directly affect system performance. Characteristics of secondary fluids are discussed in the Design Considerations section, in the subsection on Refrigerant and Fluids. Additional information regarding selection of single-phase secondary fluids can be found in Chapter 4 of this volume, and physical properties of selected fluids can be found in Chapter 31 of the 2017 ASHRAE Handbook—Fundamentals.

For medium-temperature commercial refrigeration applications, the chosen secondary fluid is most often the single-phase fluid inhibited propylene glycol. Propylene glycol systems operate at low pressure and are designed in the same manner as hydronic systems for air-conditioning applications. Flow rates are controlled using various flow control valves and/or pump control strategies. System components and heat exchangers in the refrigerated spaces must be carefully designed to avoid both high flow rates and high pressure drops, which can significantly affect pumping power and system energy consumption. Figures 25 and 26 illustrate basic schematics of a conventional direct-expansion system and a system designed for a single-phase secondary coolant such as propylene glycol.

Figure 25. Direct-Expansion System

Low-temperature secondary coolant systems using potassium-based brines have been used since the 1990s; however, concerns remain about material compatibility and corrosion with materials common to commercial refrigeration equipment construction. Recently, plastic piping networks have been used to reduce corrosion issues. CO₂ is also used as a secondary coolant, operating at a higher pressure; concerns about material compatibility have been eliminated and energy consumption has been improved because of the two-phase operation of CO₂ and the associated dramatically lower mass flow rates, which reduce both pumping power and the distribution line sizes. In these systems, the CO₂ is pumped from the refrigeration system to the evaporators as a pure liquid and is only partially evaporated in heat exchangers inside the display cases and walk-ins. The liquid/vapor mixture returns to the refrigeration system. Figure 27 shows a basic schematic of secondary system designed for use with CO₂.

Figure 26. Secondary Coolant System For Single-Phase Fluids

Heat Exchangers. Heat can be removed from the secondary fluid using a chiller of any design, but often a plate type is used for highest efficiency. Coils engineered to remove heat effectively from refrigerated spaces are generally designed differently from those for volatile refrigerants. For single-phase applications, liquid should enter the bottom of the coil, leave at the top, and be circuited to avoid trapping air. Drain and vent valves must also be equipped to assist air removal and service.

Figure 27. CO₂ Secondary Coolant System

System Application. Typically, the entire refrigeration system for supermarkets is divided into two temperature groups: low (frozen food, ice cream) and medium (meat, dairy, produce, preparation rooms). To increase efficiency, the systems may be further subdivided, though often at a higher capital cost. Temperature is controlled by regulating flow using a balance valve, or cycling flow around a set point using a solenoid valve. Piping may be in circuited or loop arrangement, or a combination of the two. Circuited systems have the advantage of containing most of the control valves in a central location, but at the cost of a greater amount of installed piping.

Charge and Leak Rate Reduction. Secondary coolant systems have several advantages, the most prominent being reduction of primary refrigerant charge and resulting leakage rates. Because primary refrigerant piping is located almost wholly within the machine room, the amount of refrigerant and refrigeration piping required can be reduced by 60 to 90%. Because field piping of the primary system is typically limited to only a few joints, most of the primary system piping joints are factory installed; these joints are generally of higher quality than field-installed joints, because they are formed in controlled conditions by skilled labor, using nitrogen and a variety of pressure-testing and leak-identification methods. Higher-quality joints combined with a lower refrigerant charge can significantly lower refrigerant leakage rates, which reduces the environmental effects associated with the primary refrigerant.

Performance Characteristics. The compressors and evaporator are close coupled, so suction line pressure losses and heat gains are minimized, enhancing system performance. Secondary coolant systems are inherently less complex than direct-expansion types, requiring fewer and less complicated valves and control devices. Less expensive nonmetallic piping systems and components can also be used, because the system operating pressure is low, typically less than 60 psig. Service of the refrigeration system is basically limited to the machine room area, and maintenance costs can be reduced. Because a fluid loop is used, thermal storage may be applied to reduce peak power demands and take advantage of lower off-peak utility rates. Ambient or free cooling may be applied in areas with colder climates. Secondary systems also can use primary refrigerants not typically suitable for direct-expansion systems, including ammonia and hydrocarbons.

Disadvantages of secondary systems include thermodynamic loss inherent in the additional step of heat transfer in the chiller, as well as the energy consumed by the fluid pump and the heat it transfers to the circulating fluid. Insulation must also be applied to both coolant supply and return lines to minimize heat gain. System first cost is higher, though this can be offset by refrigerant charge reductions and reduced system maintenance.

Systems Using CO₂ as a Refrigerant. Carbon dioxide has gained renewed use as a refrigerant in commercial refrigeration, domestic and commercial water heating, and mobile air conditioning. Widely applied in European supermarket systems, and to a lesser degree in North America, CO₂ can be used in a number of system types and configurations. Primary drivers for use of CO₂ in refrigeration systems include reduced emission of global warming gases, higher volumetric cooling capacity (which results in reduced compressor displacement and smaller pipe diameters), and excellent heat transfer properties. In addition, as a two-phase heat transfer secondary fluid, it allows smaller pumps and piping. System operating pressure has been considered the primary barrier to implementation of this natural refrigerant; however, the introduction of R-410A systems and alternative materials has largely overcome this barrier.

CO₂ systems are generally divided into two categories: subcritical and transcritical. The critical point is the point on a phase diagram at which the liquid and vapor phases of a substance have the same density (i.e., they are indistinguishable). The critical temperature is the maximum temperature at which a vapor can be converted into a liquid by increasing the vapor’s pressure. If the vapor is above the critical temperature, it cannot exist in the liquid state. Although every refrigerant has a critical point, the commonly used refrigerants (other than CO₂) never come close to conditions where the system is operating near the critical point. For example, the critical temperature for R-22 is 205.1°F, with a corresponding pressure of 723.7 psia. This condition will never exist in the normal operation of an R-22 system. In contrast, the critical temperature for CO₂ is 87.9°F, with critical pressure at about 1070 psia. Therefore, a CO₂ system operating with an air-cooled condenser, selected with a 10°F TD, will reach its critical temperature when the outdoor ambient temperature reaches 77.9°F. This property of CO₂ does not allow many locations to use CO₂ without planning for transcritical operation.

Most subcritical systems use cascading, with CO₂ in the lower cascade and an HFC, HFO/HFC blend, or a natural refrigerant (e.g., HC, ammonia) in the upper cascade to reject heat to ambient. Transcritical systems typically use CO₂ as the only refrigerant contained in the system, and reject heat directly to ambient, albeit at much higher pressures. CO₂ systems can be configured in a wide array of combinations with other systems, depending on application, including single-phase and CO₂ secondary systems.

Cascade Systems. CO₂ cascade systems are most often subcritical systems which provide low-temperature (LT) refrigeration and reject heat to an upper cascade. DX cascade-subcritical CO₂ systems are similar to a typical vapor compression cycle, with CO₂ as the refrigerant; the difference is that the heat transfer capacity necessary to condense the high-temperature/high-pressure CO₂ vapor leaving the compressor into a high-pressure/high-temperature liquid is provided by an HFC refrigeration system, with the heat transfer occurring in a cascade evaporator/condenser. Electric expansions valves (EEVs) are the preferred method of expansion device used in all CO₂ DX systems. The upper cascade system may be dedicated to the low-temperature (LT) CO₂ system only, as shown in Figure 28, or it may provide medium-temperature refrigeration to other loads. A common example of a combined system is one where the CO₂ receiver of the subcritical system also functions as a liquid/vapor separator for a medium-temperature CO₂ secondary system, though many variations and combinations have been successfully used.

The system in Figure 29 combines CO₂ secondary liquid overfeed and CO₂ LT DX cascade-subcritical CO₂. The MT liquid overfeed portion requires the receiver to be of the liquid/vapor separator design. Liquid from the separator is pumped to both the LT and MT evaporator systems. The MT evaporator systems operate as flooded evaporators, using a thermostat and liquid solenoid valve to maintain fixture temperature. The LT evaporator systems are fitted with EEVs, and operate as DX evaporators. Temperature control can be accomplished with a thermostat and liquid solenoid valve, or with an EEPR. The liquid/vapor mixture leaving the MT evaporators is piped into the top of the liquid/vapor separator, where the liquid mixes with the existing liquid in the vessel, and the vapor is eventually drawn to the colder surface of the condenser/evaporators. Discharge vapor leaving the LT compressors is piped to the inlet of the condenser-evaporators, where the vapor changes state into a liquid. The liquid then enters the liquid/vapor receiver.

Figure 28. LT DX Cascade-Subcritical CO₂ System with CO₂ Direct Expansion as Lower Cascade Circuit

Advantages of cascade systems include smaller line sizes and slight improvements in system efficiency compared to conventional technologies. Also, compared to transcritical CO₂ systems, cascade systems operate at much lower CO₂ pressures. System first cost is generally considered the primary disadvantage.

Figure 29. Combination MT Cascade-CO₂ Secondary and LT DX Cascade-Subcritical CO₂ System

Transcritical Systems. CO₂ transcritical systems operate both above and below the critical pressure and are used for refrigerating medium-temperature systems and rejecting heat directly to ambient. Low-temperature refrigeration can be accomplished by applying the transcritical system as the upper cascade in a cascade system, or by applying two-stage (compound) systems. Two-stage compressors are also available.

Figure 30 shows a two-stage (compound) compression system that compresses the CO₂ vapor from the low-temperature evaporator systems in two stages, allowing for greater compressor efficiency by reducing the operating compression ratio of all compressors. Vapor from the LT evaporator systems enters the LT compressors, which raise the pressure from low to intermediate, corresponding with the MT vapor pressure. Discharge vapor from the LT compressors flows through an oil separator and then enters the suction manifold to the MT compressors, where it joins vapor leaving the MT evaporator systems, and the entire system mass flow then enters the MT compressors.

Discharge vapor leaving the MT compressors flows through an oil separator before entering the gas cooler. If the ambient conditions are such that the saturated condensing temperature (SCT) is less than 87.9°F, the CO₂ vapor will condense into a liquid. In this condition, the system will operate similarly to a typical vapor compression cycle, with the liquid leaving the gas cooler flowing into the receiver, and supplying the liquid header as needed. When the SCT is above 87.9°F, the gas cooler transfers heat from the discharge vapor without a change of state taking place. The refrigerant then enters the flash tank, where the flash tank high-pressure control valve vents the high pressure to the suction manifold of the MT compressors. This allows sufficient reduction in pressure/temperature to move the refrigerant below the critical point, and allows the CO₂ to revert to a saturated condition, with both liquid and vapor phases present. From here, the liquid is supplied to the liquid header, supplying liquid to the various liquid feeds to each evaporator system as needed.

Figure 30. Two-Stage (Compound) Transcritical CO₂ System

Gas coolers rejecting heat to ambient can operate at pressures of 1200 to 1600 psia and must be made of higher-strength materials such as steel or stainless steel, or use microchannel technologies suitable for high-pressure operation. Recently available copper-iron alloy piping might offer alternatives in the future. Gas coolers are also well suited for water heating applications and have been widely used for this purpose in Asia.

Advantages of transcritical systems include eliminating the need for HFC systems or other natural refrigerants that are toxic and/or flammable, and increased availability of heat for heat reclaim systems. Disadvantages include the following:

Work continues on improving efficiency of transcritical operation, including using parallel compressor systems for flash compression, and using ejectors to transfer medium-temperature suction gas to the flash tank without using a compressor.

Distributed Systems. Distributed systems eliminate the long lengths of piping needed to connect display fixtures with compressor racks in back-room parallel compressor systems. The compressors are located in cabinets, close-coupled to the display refrigerator lineups, placed either at the end of the refrigerator lineup or, more often, behind the refrigerators around the store’s perimeter.

Distributed systems are typically located in the store to provide refrigeration to a particular food department, such as meat, dairy, or frozen food. With this arrangement, the saturated suction temperature (SST) for each rack closely matches the evaporator temperature of the display refrigerators and walk-in coolers. This is not always the case for parallel-rack direct-exchange (DX) systems, because a single rack often serves display refrigerators with three or four different evaporator temperatures, and the parallel-rack DX system must operate at an SST that satisfies the requirements of the lowest-temperature one connected. Better evaporator temperature matching with distributed systems can decrease the system’s overall energy consumption.

Distributed systems typically require a much lower refrigerant charge than parallel-rack DX systems, because of the former’s shorter suction and liquid lines to display refrigerators. Refrigerant piping to remote condensers can be eliminated by using a closed-loop water-cooled system.

Close-coupling the refrigerated display fixtures to the common suction manifold of the distributed system has other ramifications for energy consumption. Shorter suction lines mean that pressure drop between evaporators and the compressor suction manifold is less than with parallel-rack DX systems, so the SST of distributed systems is closer to the display refrigerator evaporator temperature: about 1 to 2°F less than refrigerator evaporator temperature, compared to 2 to 4°F difference with a parallel-rack DX. Shorter suction lines also mean less heat gain.

For a closed-loop water-cooled system, a central pump station contains the circulation pump and all valves needed to control fluid flow between the parallel compressor cabinets and fluid cooler. Inlet and outlet pipes sized for the entire system flow are provided to and from the fluid cooler and pump station. Flow to each distributed system is branched from these central supply and return pipes at a continuous rate; flow to each distributed compressor system is controlled by manual balancing valves set at installation to ensure proper flow to each cabinet.

Liquid-Cooled, Self-Contained Systems. In these systems, refrigeration condensing units connected (underneath, behind, above, or in a nearby enclosure) to one or more refrigerators are located in the display area of the supermarket. A low-temperature fluid or coolant, typically a brine or glycol solution, is pumped through a refrigerant-to-liquid heat exchanger, which serves as the condenser. The heated coolant then flows to a remote refrigeration system or chiller, which removes heat and then pumps the coolant back out to the refrigerator.

As with other systems, there are advantages and disadvantages. As much as 80% less refrigerant charge is needed, there is less potential for refrigerant loss by leakage, and initial equipment costs may be lower. In addition, refrigerators can be performance tested before they are shipped from the factory, and installations may be less labor intensive.

As with secondary cooling systems, the biggest disadvantage is the increased energy requirement from the additional step of heat transfer and the secondary fluid pumps. Noise levels can also be higher, and compressor service must be done in the display area of the supermarket. Advances in compressor technology leading to quieter, more compact, and energy-efficient systems would allow liquid-cooled, self-contained systems to become more feasible low-charge alternatives for widespread applications.

Air-Cooled. The remote condenser may be placed outdoors or indoors (where it can be used as a heat reclaim condenser to heat portions of the building in winter). The air-cooled condenser may be either single circuit or multiple circuit. The manufacturer’s heat rejection factors should be followed to ensure that the desired TD is accommodated. Generally accepted TDs for remote air-cooled refrigeration condenser sizing are 15°F for medium-temperature systems and 10°F for low-temperature systems. Remote air-cooled condensers are popular for use with parallel compressors. Figure 33 shows a basic parallel system with an air-cooled condenser and heat recovery coil.

Pressure must be controlled on most outdoor condensers. Fan-cycle controls work well down to 50°F on condensers with single or parallel groups of compressors. Below 50°F, condenser flooding (using system refrigerant) can be used alone or with fan-cycle controls. Flooding requires a larger refrigerant charge and liquid receiver. In conjunction, splitting condensers with solenoid valves in the hot-gas lines can reduce the condenser surface during cold weather, thereby minimizing the additional refrigerant charge. Natural subcooling can be integrated into the design to save energy.

To maintain the desired minimum condensing pressure (temperature), condenser fan motors can either be cycled on/off or allowed to operate at variable speed. To minimize temperature fluctuations at the condenser header, which might lead to tubing expansion/contraction and potential condenser tubing leaks, fan motor(s) nearest the header typically are not cycled on/off.

Conventional fan motors can be cycled on/off using (1) pressure controls, (2) liquid-line thermostats, or (3) ambient thermostats. Ambient control, where the condenser fan motors are cycled on/off based on ambient conditions, is common, but may not give the degree of condensing pressure (temperature) control required in systems designed for high-efficiency gain, and is not recommended except in mild climates with daily low temperatures below 50°F. Sometimes, pressure switches work with gravity louvers to cycle the condenser fans; this system type requires no refrigerant flooding charge.

Conventional fan motors can be used in conjunction with a variable-frequency drive (VFD) to vary motor speed, if the motors are VFD rated. This setup allows modification of motor speed (and thus fan rotation speed) as required to maintain the minimum condensing pressure (temperature).

Electrically commutated motors (ECMs) also vary motor speed, but do not require adding an aftermarket component such as a VFD. ECMs are also more efficient than a VFD-controlled motor. Recent advances in motor design have resulted in ECMs available in the same dimensions, motor frame, and shaft size as their standard-motor counterparts; ECM cost has also gone down significantly, resulting in increased use. When using ECMs, the typical strategy calls for all condenser fan motors to operate at the same speed. Thus, all motors would increase/decrease in speed simultaneously.

Remote condenser installations, particularly when associated with heat recovery, have substantially higher internal high-side volume than other types of systems, resulting in higher refrigerant charges. As system load demands fluctuate, receiver refrigerant level also fluctuates. Consequently, the liquid receiver must be sized carefully (especially for remote condenser applications) to ensure it has sufficient capacity for system pumpdown when necessary for service work.

As a general guideline, roof-mounted condensers should have at least 3 ft of space between the roof deck and bottom of the condenser slab to minimize the radiant heat load from the roof deck to the condenser surface. Also, free airflow to the condenser should not be restricted. Remote condensers should be placed at least 4 ft from any wall, parapet, or other airflow restriction. Side-by-side condensers should be placed at least 8 ft from each other. Consult equipment manufacturers for their specific recommendations, and see Chapter 24 of the 2017 ASHRAE Handbook—Fundamentals for discussion of the problems of locating equipment for proper airflow.

Single-unit compressors with integral air-cooled condenser systems can be mounted in racks up to three high to save space. These units may have condensers sized so that the TD is in the 10 to 25°F range. Optionally available next-larger-size condensers are often used to achieve lower TDs and higher energy efficiency ratios (EERs) in some supermarkets, convenience stores, and other applications. Single compressors with heated crankcases and heated insulated receivers and other suitable outdoor controls are assembled into weatherproof racks for outdoor installations. Sizes range from 0.5 to 30 hp.

Air-Cooled Machine Room. Standard air-cooled condensing units in a separate air-cooled machine room are still used in some supermarkets. Dampers, which may be powered or gravity operated, supply air to the room; fans or blowers controlled by room temperature with a thermostat are used to exhaust the air.

A complete indoor air-cooled condensing unit requires ample, well-distributed ventilation. Ventilation requirements vary, depending on maximum summer conditions and evaporator temperature, but 750 to 1000 cfm per condensing unit horsepower has given proper results. Exhaust fans should be spaced for an even distribution of air (Figure 31).

Rooftop air intake units should be sized for 750 fpm velocity or less to keep airborne moisture from entering the room. When condensing units are stacked (as shown in Figure 31), the ambient air design should provide upper units with adequate ventilation. Rooftop intakes are preferred because they are not as sensitive to wind as side wall intakes, especially in winter in cold climates. Butterfly dampers installed in upflow exhaust fans, which are controlled by a thermostat in the compressor room, exhaust warm air from the space.

Figure 31. Typical Air-Cooled Machine Room Layout

The air baffle helps prevent intake air from short-circuiting to the exhaust fans (Figure 31). Because air recirculation is needed around the condensers for proper winter control, intake air should not be baffled to flow only through the condensers.

Ventilation fans for air-cooled machine rooms normally do not have a capacity equal to the total of all the individual condenser fans. Therefore, if air is baffled to flow only through the condenser during maximum ambient temperatures, the condensers will not receive full free air volume when all or nearly all condensing units are in operation. Also, during winter operation, tight baffling of the air-cooled condenser prevents recirculation of condenser air, which is essential to maintaining sufficiently high room temperature for proper refrigeration system performance.

Machine rooms that are part of the building need to be airtight so that air from the store is not drawn by the exhaust fans into the machine room. Additional load is placed on the store air-conditioning system if the compressor machine room, with its large circulation of outdoor air, is not isolated from the rest of the store.

Evaporative. Evaporative condensers are equipped with a fan, circulating water-spray pump, and a coil. The circulating pump takes water from the condenser sump and sprays it over the surface of the coil, while the fan introduces an ambient airstream that comes into contact with the wet coil surface. Heat is transferred from condensing refrigerant inside the coil to the external wet surface and then into the moving airstream, principally by evaporation. Where the wet-bulb temperature is about 30°F below the dry-bulb, the condensing temperature can be 10 to 30°F above the wet-bulb temperature. This lower condensing temperature saves energy, and one evaporative condenser can be installed for the entire store. Chapter 39 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment gives more details.

Evaporative condensers are also available as single- or multiple-circuit condensers. Manufacturer conversion factors for operating at a given condensing and wet-bulb temperature must be applied to determine the required size of the evaporative condenser.

In cold climates, the condenser must be installed to guard against freezing during winter. Evaporative condensers demand a regular program of maintenance and water treatment to ensure uninterrupted operation. The receiver tank should be capable of storing the extra liquid refrigerant during warm months. Line sizing must be considered to help minimize tank size.

The extremely high temperature of the entering discharge gas is the prime cause of evaporative condenser deterioration. The severity of deterioration can be substantially reduced by using the closed water condensing arrangement. How much deterioration is reduced depends on how much the difference is reduced between the high discharge gas temperatures (experienced even with generously sized evaporative condensers) and the design entering water temperature for the closed water circuit.

Hybrid Adiabatic Condenser. Hybrid condensers evaporatively precool ambient air before it enters the condenser fin-tube coil. This allows the air-cooled condenser to operate more efficiently in elevated ambient temperatures, because air supplied to the condenser coil is below the ambient temperature. Unlike an evaporative condenser, the hybrid adiabatic condenser does not require a water sump and water pump.

Water-Cooled. Water-cooled conventional compressor units range in size from 0.5 to 30 hp and are best for hot, dry climates where air-cooled condensers will not operate at elevated discharge pressures, and evaporative condensers are not economically feasible (excessive water consumption). Water-cooled condensers can also be applied to parallel-compressor systems. A city-water-cooled condensing unit that dumps hot water to a drain is usually not economical because of the high cost of water and sewer fees. Cooling towers or evaporative fluid coolers, which cool water for all compressor systems in a single loop, are used instead. If open cooling towers are used to remove heat from condensing water, shell-and-tube heat exchangers must be used, and brazed-plate heat exchangers avoided.

Water flow in the closed water circuit can be balanced between multiple condensers on the same evaporative fluid cooler circuit with water-regulating valves. Usually, low condensing temperatures are prevented by temperature control of the closed water circuit. Three-way valves provide satisfactory water distribution control between condensers.

Fluid Cooler. In a closed-loop water condenser/evaporative cooler arrangement, an evaporative fluid cooler removes heat from water instead of refrigerant. This water flows in a closed, chemically stabilized circuit through a regular water-cooled condenser (a two-stage heat transfer system). Heat from condensing refrigerant transfers to the closed water loop in the regular water-cooled condenser. The warmed water then passes to the evaporative cooler.

The water-cooled condenser and evaporative cooler must be selected considering the temperature differences between the (1) refrigerant and circulating water and (2) circulating water and available wet-bulb temperature. The double temperature difference results in a higher condensing temperature than when the refrigerant is condensed in the evaporative condenser. On the other hand, this arrangement causes no corrosion inside the refrigerant condenser itself because the water flows in a closed circuit and is chemically stabilized.

Cooling Tower Arrangements. Few supermarkets use water-cooled condensing units; the trend is instead toward air cooling. Nearly all water-cooled condensing units are installed with a water-saving cooling tower because of the high cost of water and sewage disposal.

Designing water cooling towers for perishable foods is different than for air conditioning because (1) the hours of operation are much greater than for space conditioning; (2) refrigeration is required year-round; and (3) in some applications, cooling towers must survive severe winters. A thermostat must control the tower fan for year-round control of the condensing pressure. The control is set to turn off the fan when the water temperature drops to a point that produces the lowest desired condensing pressure. Water-regulating valves are sometimes used in a conventional manner. Dual-speed fan control is also used.

Some engineers use balancing valves for water flow control between condensers and rely on water temperature control to avoid low condensing temperature. Proper bleedoff is required to ensure satisfactory performance and full life of the cooling towers, condensers, water pumps, and piping. Consult water treatment specialists, because each locality has different water and atmospheric conditions. A regular program of water treatment is mandatory.

Air-cooled condensing units located outdoors, either as single units with weather covers or grouped in prefabricated machine rooms, produce sounds that must be evaluated. The largest source of noise is usually propeller-type condenser fans. Other sources are compressor and fan motors, high-velocity refrigerant gas, general vibration, and amplification of sound where vibration is transmitted to mounting structures (most critical in roof-mounted units).

A fan-speed or fan-cycle control helps control fan air noise by ensuring that only the amount of air necessary to maintain proper condensing temperature is generated. Take care not to restrict discharge air; when possible, it should be discharged vertically upward. ECMs also yield lower noise levels than conventional condenser fan motors.

Resilient mountings for fan motors and small compressors and isolation pads for larger motors and compressors help to reduce noise transmission. Proper discharge line sizing and mufflers are the best solution for high-velocity gas noise. Lining enclosures with sound-absorbing material is of minimal value. Isolation pads can help on roof-mounted units, but even more important is choosing the right location with respect to the supporting structure, so that structural vibration does not amplify the noise.

If sound levels are still excessive after these controls have been implemented, location becomes the greatest single factor. Distance from a sensitive area is most important in choosing a location; each time the distance from the source is doubled, the noise level is halved. Direction is also important. Condenser air intakes should face parking lots, open fields, or streets zoned for commercial use. In sensitive areas, ground-level installation close to building walls should be avoided because walls reflect sound.

When it is impossible to meet requirements by adjusting location and direction, barriers can be used. Although a masonry wall is effective, it may be objectionable because of cost and weight. If a barrier is used, it must be sealed at the bottom because any opening could allow sound to escape. Barriers also must not restrict condenser entering air. Keep the open area at the top and sides at least equal to the condenser face area.

When noise is a consideration, (1) purchase equipment designed to operate as quietly as possible (e.g., 850 rpm condenser fan motors instead of higher-speed motors), (2) choose the location carefully, and (3) use barriers when the first two steps do not meet requirements.

See Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications for information on outdoor sound criteria, equipment sound levels, sound control for outdoor equipment, and vibration isolation.

Off-Cycle or Off-Time Air Defrost. In applications with an evaporator system piped to a single compressor, defrosting is achieved by shutting the compressor off during the defrost cycle. Typically, the defrost cycle is initiated as the defrost time clock deenergizes the liquid refrigerant solenoid valve. The compressor continues to run, pumping refrigerant out of the low side of the system, until the system suction pressure reaches the low-pressure cutout. At this point, the low-pressure switch opens, breaking the compressor starter control circuit and shutting off the compressor. The compressor remains off until the defrost cycle terminates.

For an evaporator system piped to a multicompressor rack, the defrost time clock simply deenergizes the system’s liquid refrigerant solenoid valve, shutting off the supply of liquid refrigerant to the circuit.

In both single- and multicompressor systems, the evaporator reaches a temperature that allows defrosting and gives ample time for condensate drainage. Because this method obtains its defrost heat from air circulating in the display fixture, it is slow and limited to open fixtures operating at 34°F or above.

In open display refrigerators, air defrost moves ambient air from the store into the refrigerator. Multideck refrigerators with glass doors use air contained within the fixture plus air that enters as doors are opened to facilitate defrosting. Various systems are used; some use supplemental electric heat to ensure reliability. The heat content of the store’s wintertime ambient air is critical for good results from this method.

Electric Defrost. Electric defrost methods apply heat externally to the evaporator and require 20 to 60 min to properly defrost (time varies by manufacturer and application). This can be up to 1.5 times longer than gas defrost requires. The heating element may be in direct contact with the evaporator, relying on conduction for defrost, or may be located between the evaporator fans and the evaporator, relying on convection or a combination of conduction and convection for defrost. In both instances, the manufacturer generally installs a temperature-limiting device on or near the evaporator to prevent excessive temperature rise if any controlling device fails to operate.

Electric defrost simplifies installation of low-temperature fixtures. Controls used to automate the cycle usually include one or more of the following: (1) defrost timer, (2) solenoid valve, (3) electrical contactor, (4) evaporator fan delay switch, and (5) defrost termination control. The typical application begins with the defrost time clock initiating a defrost cycle. For an evaporator system piped to a single compressor, this deenergizes the liquid refrigerant solenoid valve, allowing the circuit to pump down. Once the compressor cycles off, the defrost heater contactor energizes, supplying power to the defrost heaters. For an evaporator system piped to a multicompressor rack, the time clock deenergizes the liquid refrigerant solenoid valve. After allowing sufficient time delay for the refrigeration circuit to pump down, the defrost heater contactor is energized, supplying power to the defrost heaters. The time clock is set for a maximum fail-safe amount of time, but a defrost termination thermostat terminates the defrost cycle once its setpoint has been achieved, to prevent an unnecessarily long defrost cycle when frost buildup is light. After the defrost cycle terminates, the time clock energizes the liquid refrigerant solenoid valve, allowing (1) the suction pressure to rise to its cut-in point on single compressor applications, completing the compressor starter control circuit and bringing the compressor starter circuit back on; or (2) energizing the liquid refrigerant solenoid valve, bringing the refrigeration circuit back on. In both applications, the fan delay thermostat temporarily prevents the evaporator fans from energizing, allowing the remaining condensate on the evaporator surface to freeze and prevent its being blown into the refrigerated space. The fan delay thermostat also allows the compressor to pull down to the design operating condition more quickly. Some applications of open low-temperature refrigerators may operate the fans during the defrost cycle.

Gas Defrost. Gas defrost requires careful design consideration and the use of additional differential valves to prevent liquid refrigerant from accumulating in the defrosting evaporator coils. One rule of thumb for gas defrost is that no more than 25% of the circuits can call for defrost at one time, to ensure that enough heat is available from circuits still in refrigeration mode to supply the gas necessary for those in defrost. Given the size of many modern supermarket refrigerator lineups, it is often practical to sequence gas defrosts such that no two circuits are in defrost at the same time. One advantage of gas defrost is that frost/ice buildup is melted more quickly than with other methods, allowing minimal disruption in maintaining the refrigerated product at the design temperature.

There has been a trend toward operating the system at lower-than-design condensing temperatures when ambient temperatures are colder, to improve compressor efficiency and reduce electrical consumption. During such periods, it might be necessary to include a control strategy to temporarily elevate the condensing pressure (temperature) during defrost to ensure that adequate heat content is available in the defrost vapor to properly defrost evaporators in the refrigerated fixtures.

Hot-gas defrost uses heat from the compressor’s discharge gas to defrost the evaporators. Reverse-flow gas defrost, the most common method, uses the existing circuit suction line to introduce discharge gas in a backward-flow direction through the evaporator system. Discharge gas from the common discharge line is piped to a defrost header, which then supplies defrost gas to each individual evaporator circuit through a defrost solenoid valve. Defrost solenoid valves are typically sized for evaporator loads twice the size of the non-defrost evaporator load. To accomplish reverse flow, a circuit suction line regulator must be installed upstream of the common suction header and electrically closed during the defrost cycle; otherwise, the discharge vapor would take the path of least resistance and flow to the compressor inlet. The suction regulator may be a suction solenoid valve, EPR, or EEPR. The high-pressure discharge gas flows backward through the lower-pressure evaporator and eventually exits to the liquid header through check valves around the TEV and liquid-line solenoid valve.

A valve that can provide a pressure differential must be installed somewhere between the defrost header and the liquid header. This valve is known as the defrost differential valve, and is typically placed in either the discharge line (downstream from the oil separator) or the liquid line (upstream of the liquid header).

Using a liquid-line differential pressure regulator has the disadvantage of generating flash gas for the system if the liquid is not subcooled. This system also returns the mixture of liquid and cooled gas to the liquid header. This mixture is sent out to other operating evaporators, reducing distributor orifice and expansion valve capacity as well as non-defrosting evaporator efficiency during the defrost. Additional load is placed on the system after defrost to recover temperatures and compress returning noncondensed hot gas passing through the liquid lines.

When using a discharge line differential valve, flash gas is not added to the liquid lines and the mixture of liquid and cooled gas can be returned to a discrete header, which allows the liquid to return to the receiver and the gas to return to the condenser. This system raises compressor discharge pressure during defrost, which reduces compressor efficiency.

Saturated gas defrost is similar to hot-gas defrost but is piped differently: the gas for defrosting is the saturated gas, which occupies the upper portion of the liquid receiver. In addition, the defrost differential valve must be installed in the main liquid line upstream of the liquid header, to allow for a proper differential to help the saturated vapor reenter the liquid header after defrosting the evaporator.

Occasionally, supplemental electric refrigerator heaters are added to ensure rapid and reliable defrosting. Temperature generally terminates the defrost cycle, although timers are used as a backup.

Fixture piping must have equal pressure drops to allow even distribution of hot gas to the evaporators and of liquid/cool gas from the evaporators, for all evaporators that are piped in the circuit. Consistent pressure drop can be accomplished by using oversized liquid-line headers. All pipe takeoffs from the header should be the same (e.g., all tees with no elbows at the end of the run) to maintain consistent pressure drop.

When system design or store layout makes it difficult to maintain equal pressure drop, each evaporator may use a defrost termination solenoid. The solenoid is controlled by a temperature sensor at the evaporator. When the temperature reaches the termination point, the solenoid closes, forcing hot gas to the evaporators that remain in defrost. Solenoids continue to close until all coils are defrosted.

Two defrost methods, time-off and warm fluid, are most commonly applied to secondary systems. Time-off defrost can be used in some medium-temperature applications. However, the most effective method is warm-fluid defrost, which is used for all low-temperature applications and in selected medium-temperature refrigerators where product temperatures are critical or time-off defrost is not practical. Fluid for defrost is typically heated using refrigerant discharge gas, but system efficiency can be increased by heat exchange with liquid refrigerant. Warm fluid temperatures vary and must be optimized for the coil application; however, typical values are 50 to 60°F for medium-temperature systems and 70 to 80°F for low-temperature systems. Warm-fluid defrost is most often terminated by the fluid temperature exiting the coil and is preferable to time-off because of the small change in temperature imparted on the products, resulting in lower post-defrost pulldown loads.

Defrost control methods include (1) suction pressure control (no time clock required), (2) time clock initiation and termination, (3) time clock initiation and suction pressure termination, (4) time clock initiation and temperature termination, and (5) demand defrost or proportional defrost.

Defrosting is usually controlled by a variety of clocks, which are often part of a compressor controller system. Electronic sensor control is the most accurate and can also provide a temperature alarm to prevent food loss. Electronic systems often have communication capabilities outside the store.

Liquid and/or suction line solenoid valves can be used to control the circuits for defrosting. Often, a suction-stop EPR is used to allow a single valve to isolate the defrosting circuit from the suction manifold and allow introduction of defrost gas upstream of the valve. Individual circuit defrosts are typically controlled by the rack’s energy management system, or rack controller.

Suction Pressure Control. This control is adjusted for a cut-in pressure high enough to allow defrosting during the off cycle. This method is usually used in fixtures maintaining temperatures from 36 to 43°F. When evaporator pressure is lowered to the cutout point of the control, the control initiates a defrost cycle to clear the evaporator.

However, condensing units and/or suction lines may, at times, be subjected to ambient temperatures below the evaporator’s temperature. This prevents build-up of suction pressure to the cut-in point, and the condensing unit remains off for prolonged periods. In such instances, fixture temperatures may become excessively high, and displayed product temperatures will increase.

A similar situation can exist if the suction line from a fixture is installed in a trench or conduit with many other cold lines. The other cold lines may prevent the suction pressure from building to the cut-in point of the control.

Initiation and Termination. Methods (2), (3), and (4) control defrost using time clocks to break the electrical circuit to the condensing unit, initiating a defrost cycle. The difference lies in the manner in which the defrost period is terminated.

Time Initiation and Termination. A timer initiates and terminates the defrost cycle after the selected time interval. The length of the defrost cycle must be determined and the clock set accordingly. Time-only termination is the least efficient and least reliable method, because it does not account for variations in required amount of defrost.

Time Initiation and Suction Pressure Termination. This method is similar to the first method, except that suction pressure terminates the defrost cycle. The length of the defrost cycle is automatically adjusted to the condition of the evaporator, as far as frost and ice are concerned. However, to overcome the previously described problem of the suction pressure not rising because of the defrost cut-in pressure, the timer has a fail-safe time interval to terminate the defrost cycle after a preset time, regardless of suction pressure.

Time Initiation and Temperature Termination. This method is also similar to the time initiation and termination method, except that temperature terminates the defrost cycle. The length of the defrost cycle varies depending on the amount of frost on the evaporator or in the airstream leaving the evaporator, as detected by a temperature sensor in either location. The temperature-sensing location must correspond to the last region of the evaporator to defrost, so that the defrost cycle does not terminate prematurely. The timer also has a fail-safe setting in its circuit to terminate the defrost cycle after a preset time, regardless of the temperature.

Adaptive (Demand) Defrost or Proportional Defrost. This system initiates defrost based on demand (need) or in proportion to humidity or dew point. Techniques vary from measuring change in the temperature spread between the air entering and leaving the coil, to changing the defrost frequency based on store relative humidity. Other systems use a device that senses the frost level on the coil.

Total Equivalent Warming Impact (TEWI). The concept of total equivalent warming impact combines (1) the global warming effects corresponding with the CO₂ released through energy use over the lifetime of the system (indirect effect) with (2) the effects resulting from lifetime refrigerant emissions (direct effect). The direct effect depends on the global warming potential (GWP) of the refrigerant. GWP values for refrigerants may be found in Chapter 29 of the 2017 ASHRAE Handbook—Fundamentals. The total equivalent warming impact accounts for both direct and indirect effects of refrigeration systems on global warming potential:

Refrigeration systems contribute to global climate change through direct (refrigerant emissions) and indirect effects (energy consumption).

The TEWI for a given application or system can be computed by estimating the total amounts of refrigerant and blowing agent (if applicable) released and the total energy used by the equipment over its useful lifetime. TEWI is calculated by

The factor α is used to convert the annual energy use E and equipment lifetime L to the corresponding CO₂ emissions (Baxter et al. 1998).

In response to concerns about depletion of the stratospheric ozone layer, the U.S. Environmental Protection Agency developed regulations in the early 1990s to minimize refrigerant emissions and prohibit intentional venting of all refrigerants, including the hydrofluorocarbon substitutes. Additional regulations apply to chlorine-containing refrigerants such as R-22.

Further concerns about climate change (global warming) have resulted in additional EPA regulations such as the delisting of specific high-global-warming-potential (GWP) refrigerants from use in retrofit or new supermarket and commercial refrigeration applications:

Consequently, nonnatural refrigerant choices are limited to HFC blends or HFC-HFO blends with lower GWPs. Although the list of appropriate HFC blends is long, R-407A has become the predominant choice in new supermarket applications: its GWP is approximately half that of R-404A and R-507. Newer HFC-HFO blends, such as R-448A, R-449A, and R449B (with GWPs approximately 32% of R-404A and R-507) are seeing interest in some new supermarket applications.

Current regulations state that if systems that contain more than 50 lb of refrigerant leak at an annual rate exceeding 35%, equipment repairs are required. Certain servicing and record-keeping practices are also required (EPA 1990).

As of January 1, 2019, the following regulatory changes will be in effect:

The environmental benefit of advanced low-charge refrigeration systems is a significant reduction in the amount of halogenated refrigerants now used in supermarkets, which presently use as much as 3000 lb of refrigerant. A significant number of facilities use R-22 (HCFC), which has an ozone depletion potential (ODP) of 0.055 and a global warming potential (GWP) of 1700. The latest replacement refrigerants are HFCs, such as R-134a, R-404A, and R-507, which have ODPs of 0 but have high GWP values (1300, 3260, and 3300, respectively).

Refrigerant emissions can be minimized through application of two approaches: reducing the refrigerant charge and maximizing refrigerant containment. There is some variation in charge requirement, depending on the type of heat rejection. The lowest charge is required by systems using a fluid loop for heat rejection. The charge requirement for close-coupled distributed and secondary loop systems is less because of reduced suction-side piping.

Low-Charge Systems. Commercial refrigeration is one of the largest consumers of refrigerant worldwide, and special attention has been devoted to minimizing use of refrigerant in existing and new sites. Present multiplex supermarket systems can require as much as 3000 lb of refrigerant to charge the system, but over the last decade, refrigeration system configurations with lower refrigerant charges have been used to mitigate the resulting environmental issues. Low-charge system types discussed in this chapter include secondary loop, distributed, and liquid-cooled self-contained. Table 5 shows the relative charge sizes for some of these systems.

Refrigerant Containment. Retail refrigeration is more susceptible to refrigerant leaks than domestic refrigeration, which has smaller charges and tends to be assembled in the factory rather than the field. Figure 32 shows an example of magnitude of these leaks (Bivens and Gage 2004). The annual refrigerant losses reported in this figure cover about 1200 refrigerant systems located in 223 supermarkets in California. A few studies have tried to identify the locations and causes of these high losses. Some of the larger refrigerant losses were associated with five subsystem types: condenser, hot-gas defrost, expansion valves, liquid-line solenoid valves, and service valves/stem packings (Troy 1997). Refrigerant losses have also been reported from almost all components, including evaporator coils, compressor fittings and oil lines, receiver values, suction- and liquid-line piping, sight glasses, braided hoses, and filter-driers. Some designers elect to eliminate 0.25 in. male flare access valves with removable valve cores valves and flare fittings to minimize potential leaks.

To comply with regulations and save money incurred by refrigerant loss, retail refrigeration has identified opportunities to better contain the refrigerant. These can be grouped into three areas: design and construction, operations and maintenance, and corporate policies and practices (Gage and Troy 1998). Additional information on refrigerant containment can be found in Chapter 9.

Figure 32. Average Annual Refrigerant Losses in Retail Refrigeration Systems (Bivens and Gage 2004)

Design and Construction. As discussed previously, low-charge systems are a design option that can have smaller refrigerant losses. In particular, both secondary loop and self-contained systems can be designed as factory-assembled and sealed systems. As shown in Figure 33, factory assembly provides a significant benefit over field erection for optimizing leak reduction controls. Other design and construction options include using vertical receivers over horizontal, substituting sweat connections for flare or packed angle valves for threaded-core valves, and installation of isolation values around large or frequently serviced components.

Operations and Maintenance. Changes in operation and maintenance can help minimize refrigerant losses. Establishing a database to monitor equipment maintenance helps to identify stores and components with the highest leak rates. Proactive leak detection and repair on a regular schedule can identify leaks before refrigerant losses become so large that they require emergency repair calls. Maintenance can be coordinated so that when case cleaning is scheduled, the control valves and evaporator coils can be checked for leaks.

Figure 33. Effects of Leak-Reduction Techniques Applied in Factory (Kazachki and Hinde 2006)

Corporate Policies and Practices. Corporate policies play a key role in reducing emissions and moving towards sustainability. Examples include investing in training for the repair contractors to apply the procedures and schedules tailored for the company’s systems, and establishing a corporate commitment to replace consistently leaky equipment with state-of-the art, leak-resistant equipment.

Environmental Protection Agency (EPA) Resources. The EPA’s GreenChill program (www.epa.gov/greenchill) partners with food retailers to reduce refrigerant emissions and decrease their impact on the ozone layer and climate change. The program’s website provides guidelines for design, installation, and maintenance of refrigeration systems, as well as other useful information.

In refrigeration systems, energy use is minimized by three approaches: refrigeration load reduction, use of energy-efficient components and operations, and design features. The Department of Energy has issued maximum daily energy consumption levels for nearly all retail food store refrigerated display equipment. The general design and energy use of the overall refrigeration system are not regulated. Some of the following options are mature and tested in the industry, whereas others are emerging technologies. Designers must balance energy savings against customer requirements, manufacturing cost, system performance, reliability, and maintenance costs.

Load Reduction. Reducing the amount of heat that needs to be removed from a space leads to instant savings in energy consumption. Display refrigerators should be located to minimize drafts or air curtain disturbance from ventilation ducts, and away from heat sources or direct sunlight. The cooling load of a typical refrigerator depends on infiltration, conduction, and radiation from surroundings, as well as heat dissipation from internal components.

Infiltration. Research indicates that infiltration of warm and moist air from the sales area into an open vertical display refrigerator accounts for 70 to 80% of the display refrigerator total cooling load (Faramarzi 1999). Infiltrated air not only raises product temperatures, but moisture in the air also becomes frost on the evaporator coil, reducing its heat transfer abilities and forcing the fan to work harder to circulate air through the refrigerator. There are several ways to reduce the amount of infiltration into refrigerators:

Thermal Conduction. Improving the R-value of insulation, whether by using materials with low thermal conductivity or simply increasing insulation thickness, reduces conduction heat transfer through walls of the refrigerated space. Conduction accounts for less than 5% of cooling load of medium-temperature refrigerators, but almost 10% for low-temperature refrigerators (see Figure 10).

Thermal Radiation. Warm objects near the display refrigerator radiate heat into the refrigerated space. Night covers protect against radiation heat transfer.

Internal Loads. Reducing power use of individual components also reduces the cooling load for components located inside the refrigerated space. For example, fan motors, lighting, and antisweat heaters produce less heat when high-efficiency versions are chosen, thereby reducing the load on the refrigeration equipment.

Another internal load is the heat added during defrost, which can raise product temperatures. Defrost methods should be chosen so that the minimum amount of heat is added to the refrigerator. For example, hot-gas defrost can be considered an improved technique.

Adaptive (demand) defrost technologies can sense frost formation on the coil, enabling a controller to determine exactly when the refrigerator should begin its defrost cycle. Unnecessary defrosts and excessive frost formation leading to coil blockage can be eliminated. Care is required when selecting a demand defrost system: if the system malfunctions, the refrigerators will require service, and there is the potential for product loss.

Sensors may also be used to verify the end of defrost cycles (intelligent defrost termination). Typically, the refrigerator is allowed to defrost for a set amount of time or until the air temperature leaving the coil reaches a specified level. This usually results in a defrost cycle that runs longer than necessary, allowing more heat to enter the refrigerator and raising product temperatures. Intelligent defrost termination sensors can determine when the coil is free of frost and immediately terminate defrost and restart the refrigeration system. An intelligent defrost termination sensor can be a simple electromechanical thermostat, a solid-state sensor, or other device.

Energy Efficient Components: Careful selection of components based on proper application, energy efficiency attributes, and correct sizing can play a significant role in increasing overall system efficiency.

Evaporator. Evaporator coil design can significantly affect refrigerator performance. Efficient evaporator coils allow the refrigerator to maintain its target discharge air temperature while operating at a higher evaporator temperature. Higher evaporator temperature (or suction pressure) has the benefit of increasing its refrigeration effect; however, it also hampers refrigeration system performance by increasing the density of refrigerant entering the compressor, thus increasing compressor work. Evaporator coil characteristics can be improved in the following ways:

Door Antisweat Heaters. Antisweat heaters (ASHs) with a low watt-per-door rating should be used whenever possible. In addition to using less energy at the antisweat heater level, less heat is introduced into the refrigerated space, thus indirectly reducing the cooling load.

Some controllers can recognize the antisweat heat needs of the door and ensure that the heaters only operate when needed. They adjust their operation accordingly, through pulsation or other mechanisms. Condensate sensors on reach-in glass doors activate ASHs when droplets are detected; RH-based controllers sense the psychrometric properties of air and activate ASHs when needed.

New methods of glass door construction have brought products that require little or no antisweat heat to maintain customer-friendly fog-free panes. This performance is achieved by either using advanced glass types or special door frames, both of which greatly reduce or eliminate the amount of glass heating necessary to resist condensation.

Alternative Expansion Valves. Dual-port thermostatic expansion valves (TXVs) have capacity modulation capabilities not seen in other expansion valves. When the refrigerator emerges from defrost, there is typically a much higher load because of increased product temperatures. In this case, the large port of the expansion valve opens, allowing the system to operate at a higher capacity to account for the increased pulldown load.

Superheat can be most easily controlled by electronic expansion valves (EEVs), which have a much faster response time than bulb-sensing TXVs. EEVs also provide repeatedly precise control, allowing lower superheat settings without the danger of floodback. This allows more efficient use of the entire evaporator surface, which can then accommodate operating the evaporator system at a higher SST. Manufacturers should test the valve and controller to ensure it maintains stable control at targeted superheats.

Sophisticated Refrigerator Controls. All components of a display refrigerator should be linked to one master control system, which can optimally control the operation of individual components.

Motors. Energy-efficient evaporator fan motors such as electronically commutated motors (ECMs) and permanent split capacitor (PSC) motors consume about half the power of standard shaded-pole motors (Faramarzi and Kemp 1999). These motors also can incorporate variable-speed controls to slow fans as the cooling load is satisfied.

Display Case Lights. Currently, standard lighting is with energy-efficient light-emitting diode (LED) bulbs.

Condensers. Condenser design can significantly affect refrigeration equipment performance. The characteristics of condensers can be improved in the following ways:

Condenser Controls. Allowing discharge pressure to float lower during low-ambient periods of can save considerable energy compared to fixed-pressure systems. Careful system design consideration is needed to ensure proper operation of the expansion valve and refrigerant feed to the evaporator coil during lower ambient conditions. Balanced-port thermostatic expansion valves and electronic expansion valves enhance the opportunity for floating pressures down with varying ambient temperatures.

Energy-Efficient Operations: Energy efficiency must be approached from a total-store perspective. Building envelope, lighting, HVAC, refrigeration, antisweat circuits, indoor air quality (IAQ), human comfort, and local utility cost all must be considered in the store design. Once the store is built and operational, effective commissioning, recommissioning, and maintenance practices are critical to keeping energy cost at a minimum.

Design Features. Additional energy-saving features can be designed into the retail refrigeration configuration. Typical opportunities include liquid subcooling strategies and heat recovery.

Heat reclaim condensers and related controls operate as alternatives to or in series with the normal refrigeration condensers. They can be used in winter to return most of the refrigeration and compressor heat to the store. They may also be used in mild spring and fall weather when some heating is needed to overcome the cooling effect of the refrigeration system itself. Another use is for cooling coil reheat for humidity control in spring, summer, and fall. Excess humidity in the store can increase the display refrigerator refrigeration load as much as 20% at the same dry-bulb temperature, so it must be avoided.

Figure 35. Basic Series System with Remote Air-Cooled Condenser and Heat Recovery

In this application, a heat reclaim coil is placed in the store’s air handler where heat from the discharge vapor can be transferred to the ventilating supply air for store heat. When reclaiming the heat content of the discharge vapor to heat the store, one of two strategies is normally used: series or parallel.

In series heat reclaim, the heat transfer coil is properly called a heat reclaim coil; it is not a heat reclaim condenser. Its purpose is to desuperheat the discharge vapor only, and therefore its thermal capacity is much less than that of the outdoor condenser. A solenoid-operated three-way heat reclaim valve is energized when heat is required. This shifts the three-way valve into heat reclaim mode, allowing discharge vapor to flow from the compressor to the heat reclaim coil, and finally to the condenser (Figure 35). Heat from the discharge vapor is transferred to the ventilating air, which is subsequently distributed through the ducting system, providing heat for the store. The normal condenser completes the process of converting the superheated discharge vapor into a liquid. The heat reclaim coil is designed to provide enough capacity to desuperheat the discharge vapor only. The heat recovery coil can be sized for a 30 to 50°F TD, depending on the capacity in cool weather. When operating a parallel compressor system at lower condensing temperatures for energy conservation, less heat will be available in the refrigerant for recovery, unless the system is designed properly: simple controls can create the higher condensing temperature needed for heat recovery. Compared with the cost of auxiliary gas or electric heat, the higher energy consumption of the compressor system may be offset by the value of the heat gained.

Parallel Heat Reclaim. In some applications, such as colder areas where using the entire heat of rejection for store heating would reduce the capacity requirement for traditional heating equipment, or in milder climates where using the entire heat of rejection might eliminate the need for additional heating equipment, a parallel-piped heat reclaim system may offer some benefits. This is an either/or method that uses two condensers of equal capacity: one outdoors, and the other in the ventilation system. Both are sized for the entire heat of rejection load. In the normal condensing mode, the outdoor condenser provides the heat transfer necessary to handle the entire heat of rejection; this full-sized condenser completely condenses the refrigerant, unlike in series applications, where the smaller coil only desuperheats the refrigerant. The entire heat of rejection load is transferred to circulating ventilation air. This yields a larger heating capacity for the building, possibly eliminating the need for a second stage of heat.

The very thing that gives the parallel method a greater heating capacity offers its greatest challenge: being an either/or application. During the outdoor condensing mode, the heat reclaim condenser is idle, receiving no refrigerant flow and containing only a minimal amount of refrigerant vapor and liquid. When the need for heat requires the three-way valve to shift, refrigerant flow to the normal condenser ceases and flow to the reclaim condenser begins. There is a time lag before the relatively empty reclaim condenser contains enough refrigerant to resume the steady supply of liquid to the receiver. Because of this lag, it is desirable to be able to quickly pump out the idle condenser, preferably using a dedicated pump-out solenoid valve for each condenser. A restriction should be used in the pump-out line to prevent liquid floodback. Even with this arrangement, the compressor(s) can fill the now-active reclaim condenser much more quickly than the pump-out solenoid can drain the idle normal condenser. While transitioning from one condenser to the other, the level of refrigerant in the receiver is drained more quickly than the functioning condenser can replenish it. Without a sufficient refrigerant charge, the receiver may lose its liquid seal before system equilibrium is restored, temporarily compromising the ability to provide vapor-free liquid to the TEVs.

Secondary Loop Systems. Another design option for secondary loop systems (Minea 2007) uses rejected heat on the low-temperature refrigeration systems to support heat pumps in northern climates. This integrated concept uses the heat pumps to boost waste heat rejected by the compressors to provide efficient space heating in the store. Similarly, waste heat from the medium-temperature refrigeration system is captured by heat pumps to heat or cool outdoor air before it enters the store. With the increased heating capacity of the heat pumps, there is no need for back-up natural gas or oil heating.

Heat reclamation can also be used to heat water for store use. Recovery tanks are typically piped in series with the normal condenser and sized based on the refrigerant pressure drop through the tank and on the water temperature requirements.

On a large, single unit, water can be heated by a desuperheater; on two-stage or compound R-22 parallel systems, water is commonly heated by the interstage desuperheater.

Constant Volume with Heat Reclaim Coils. This is typically done with one or two large HVAC units. The conditioned air must then be ducted throughout the store.

Multiple Zones. This is typically done with many smaller packaged rooftop units (RTUs), which reduces ductwork but increases electrical and gas infrastructure. Off-the-shelf RTUs do not typically accommodate heat reclaim coils, which is an energy disadvantage in both heating and dehumidification modes.

Open refrigerator display equipment often extracts enough heat from the store’s ambient air to reduce the air temperature in customer aisles by as much as 16°F. This cooled air, along with air curtain spillage at the return grille, settles to the floor and can cause discomfort. Lack of attention to this can substantially reduce sales in these areas. This cool-air spillage from refrigerated cases is commonly referred to as case credits. More information on display case effects and associated HVAC design considerations can be found in the section on Supermarkets in Chapter 2 of the 2019 ASHRAE Handbook—HVAC Applications or in Pitzer and Malone (2005).

In food stores, supply fans operate 100% of the time the store is open, at a volumetric flow of 0.6 to 1 cfm per square foot of sales area. Some chains may have multiple-speed fans, or operate the fan with variable-speed drives (VSDs). Fan speed variation can be based on a number of variables (e.g., store temperature, hood operation, building pressurization, CO₂ level), with the primary objective of minimizing fan energy usage.

Air supply and return grilles must be located so they do not disturb the air in open display refrigerators and negatively affect refrigerator performance. Directional diffusers are helpful in directing air away from cases. Return air can also be positioned to pull treated air into areas with many open refrigerated cases, thus avoiding the higher air speeds created by diffusers.

Thermostats or other HVAC controllers may include options for night setback in cool climates and night setup in warm climates. This feature may save energy by modifying the nighttime store temperature several degrees above or below the daytime set-point temperature. However, store warm-up practices impose an energy use penalty to the display refrigeration systems and affect display case performance, particularly open models.