Outdoor air is a universal heat source and sink medium for heat pumps and is widely used in residential and light commercial systems. Extended-surface, forced-convection heat transfer coils transfer heat between the air and refrigerant. Typically, the surface area of outdoor coils is 50 to 100% larger than that of indoor coils. The volume of outdoor air handled is also greater than the volume of indoor air handled by about the same percentage. During heating, the temperature of the evaporating refrigerant is generally 10 to 20°F less than the outdoor air temperature. Air heating and cooling coil performance is discussed in more detail in Chapters 23 and 27.

When selecting or designing an air-source heat pump, two factors in particular must be considered: (1) the local outdoor air temperature and (2) frost formation.

As the outdoor temperature decreases, the heating capacity of an air-source heat pump decreases. This makes equipment selection for a given outdoor heating design temperature more critical for an air-source heat pump than for a fuel-fired system. Equipment must be sized for as low a balance point as is practical for heating without having excessive and unnecessary cooling capacity during the summer. A procedure for finding this balance point, which is defined as the outdoor temperature at which heat pump capacity matches heating requirements, is given in Chapter 49.

When the surface temperature of an outdoor air coil is 32°F or less, with a corresponding outdoor air dry-bulb temperature 4 to 10°F higher, frost may form on the coil surface. If allowed to accumulate, frost inhibits heat transfer; therefore, the outdoor coil must be defrosted periodically. The number of defrosting operations is influenced by the climate, air-coil design, and the hours of operation. Experience shows that, generally, little defrosting is required when outdoor air conditions are below 17°F and 60% rh. This can be confirmed by psychrometric analysis using the principles given in Chapter 23. However, under very humid conditions, when small suspended water droplets are present in the air, the rate of frost deposit may be about three times as great as predicted from psychrometric theory and the heat pump may require defrosting after as little as 20 min of operation. The loss of available heating capacity caused by frosting should be considered when sizing an air-source heat pump.

Following commercial refrigeration practice, early designs of air-source heat pumps had relatively wide fin spacing of 4 to 5 fins/in., based on the theory that this would minimize defrosting frequency. However, experience has shown that effective hot-gas defrosting allows much closer fin spacing and reduces the system’s size and bulk. In current practice, fin spacings of 10 to 20 fins/in. are widely used.

In many institutional and commercial buildings, some air must be continuously exhausted year-round. This exhaust air can be used as a heat source, although supplemental heat is generally necessary.

High humidity caused by indoor swimming pools causes condensation on ceiling structural members, walls, windows, and floors and causes discomfort to spectators. Traditionally, outdoor air and dehumidification coils with reheat from a boiler that also heats the pool water are used. This is ideal for air-to-air and air-to-water heat pumps because energy costs can be reduced. Suitable materials must be chosen so that heat pump components are resistant to corrosion from pool treatment chemicals and high humidity.

Water can be a satisfactory heat source, subject to the considerations listed in Table 1. City water is seldom used because of cost and municipal restrictions. Groundwater (well water) is particularly attractive as a heat source because of its relatively high and nearly constant temperature. Water temperature depends on source depth and climate, but, in the United States, generally ranges from 40°F in northern areas to 70°F in southern areas. Frequently, sufficient water is available from wells (water can be reinjected into the aquifer). This use is nonconsumptive and, with proper design, only the water temperature changes. Water quality should be analyzed, and the possibility of scale formation and corrosion should be considered. In some instances, it may be necessary to separate the well fluid from the equipment with an additional heat exchanger. Special consideration must also be given to filtering and settling ponds for specific fluids. Other considerations are the costs of drilling, piping, pumping, and a means for disposal of used water. Information on well water availability, temperature, and chemical and physical analysis is available from U.S. Geological Survey offices in many major cities. For further information on water quality, groundwater heat pumps, or surface water heat pumps, see Chapter 34 in the 2015 ASHRAE Handbook—HVAC Applications, or Kavanaugh and Rafferty (2014).

Heat exchangers may also be submerged in open ponds, lakes, or streams. When surface or stream water is used as a source, the temperature drop across the evaporator in winter may need to be limited to prevent freeze-up.

In industrial applications, waste process water (e.g., spent warm water in laundries, plant effluent, warm condenser water) may be a heat source for heat pump operation.

Sewage, which often has temperatures higher than that of surface or groundwater, may be an acceptable heat source. Secondary effluent (treated sewage) is usually preferred, but untreated sewage may be used successfully with proper heat exchanger design (Phetteplace and Ueda 1989).

Use of water during cooling follows the conventional practice for water-cooled condensers.

Water-to-refrigerant heat exchangers are generally direct-expansion or flooded water coolers, usually shell-and-coil or shell-and-tube. Brazed-plate heat exchangers may also be used. In large applied heat pumps, the water is usually reversed instead of the refrigerant.

The ground is used extensively as a heat source and sink, with heat transfer through buried coils. Soil composition, which varies widely from wet clay to sandy soil, has a predominant effect on thermal properties and expected overall performance. The heat transfer process in soil depends on transient heat flow. Thermal diffusivity is a dominant factor and is difficult to determine without local soil data. Thermal diffusivity is the ratio of thermal conductivity to the product of density and specific heat. The soil’s moisture content influences its thermal conductivity.

There are three primary types of ground-source heat pumps: (1) groundwater, which is discussed in the previous section; (2) direct-expansion, in which the ground-to-refrigerant heat exchanger is buried underground; and (3) ground-coupled (also called closed-loop ground-source), in which a secondary loop with a brine connects the ground-to-water and water-to-refrigerant heat exchangers (see Figure 5).

Ground loops can be placed either horizontally or vertically. A horizontal system consists of single or multiple serpentine heat exchanger pipes buried 3 to 6 ft apart in a horizontal plane at a depth 3 to 6 ft below grade. Pipes may be buried deeper, but excavation costs and temperature must be considered. Horizontal systems can also use coiled loops referred to as slinky coils. A vertical system uses a concentric tube or U-tube heat exchanger. Design of ground-coupled heat exchangers is covered in Chapter 34 of the 2019 ASHRAE Handbook—HVAC Applications and Kavanaugh and Rafferty (2014).

Solar energy may be used either as the primary heat source or in combination with other sources. Air, surface water, shallow groundwater, and shallow ground-source systems all use solar energy indirectly. The principal advantage of using solar energy directly is that, when available, it provides heat at a higher temperature than the indirect sources, increasing the heating coefficient of performance. Compared to solar heating without a heat pump, the collector efficiency and capacity are increased because a lower collector temperature is required.

The direct system places refrigerant evaporator tubes in a solar collector, usually a flat-plate type. A collector without glass cover plates can also extract heat from the outdoor air. The same surface may then serve as a condenser using outdoor air as a heat sink for cooling.

An indirect system circulates either water or air through the solar collector. When air is used, the collector may be controlled in such a way that (1) the collector can serve as an outdoor air preheater, (2) the outdoor air loop can be closed so that all source heat is derived from the sun, or (3) the collector can be disconnected from the outdoor air serving as the source or sink.

In air-to-air heat pump systems, air circuits can be interchanged by motor-driven or manually operated dampers to obtain either heated or cooled air for the conditioned space. In this system, one heat exchanger coil is always the evaporator, and the other is always the condenser. Conditioned air passes over the evaporator during the cooling cycle, and outdoor air passes over the condenser. Damper positioning causes the change from cooling to heating.

Soil type, moisture content, composition, density, and uniformity close to the surrounding field areas affect the success of this method of heat exchange. With some piping materials, the material of construction for the pipe and the corrosiveness of the local soil and underground water may affect the heat transfer and service life. In a variation of this cycle, all or part of the heat from the evaporator plus the heat of compression are transferred to a water-cooled condenser. This condenser heat is then available for uses such as heating air or domestic hot water.

Figure 5. Heat Pump Types

A hybrid ground-coupled heat pump is a variation that uses a cooling tower or air-cooled condenser to reduce the total annual heat rejection to the ground coupling.

Additional heat pump types include the following:

Air-to-Water Heat Pumps Without Changeover. These heat pumps use both sensible and latent heat in air from either natural sources (e.g., tropical climates) or from indoor air that needs to be dehumidified and cooled for comfort. They are commonly called heat pump water heaters. Heat from the heat pump is used to heat domestic and/or service water. Chapter 49 of this volume and Chapter 50 of the 2019 ASHRAE Handbook—HVAC Applications have more detail on the equipment.

Refrigerant-to-Water Heat Pumps. These condense a refrigerant by the cascade principle. Cascading pumps the heat to a higher temperature, where it is rejected to water or another liquid. This type of heat pump can also serve as a condensing unit to cool almost any fluid or process. More than one heat source can be used to offset those times when insufficient heat is available from the primary source.

The principal types of compressors used in applied heat pump systems are briefly described in this section. For more details on these compressor types and others, refer to Chapter 38.

 Compressor Selection.

 A compressor used for comfort cooling usually has a medium clearance volume ratio (the ratio of gas volume remaining in the cylinder after compression to the total swept volume). For an air-source heat pump, a compressor with a smaller clearance volume ratio is more suitable for low-temperature operation and provides greater refrigerating capacity at lower evaporator temperatures. However, this compressor requires somewhat more power under maximum cooling load than one with medium clearance.

More total heat capacity can be obtained at low outdoor temperatures by deliberately oversizing the compressor. This allows some capacity reduction for operation at higher outdoor temperatures by multispeed or variable-speed drives, cylinder cutouts, or other methods. The disadvantage is that the greater number of operating hours that occur at the higher suction temperatures must be served with the compressor unloaded, which generally lowers efficiency and raises annual operating cost. The additional initial cost of the oversized compressor must be economically justified by the gain in heating capacity.

One method proposed for increasing heating output at low temperatures uses staged compression. For example, one compressor may compress from −30°F saturated suction temperature (SST) to 40°F saturated discharge temperature (SDT), and a second compressor compresses the vapor from 40°F SST to 120°F SDT. In this arrangement, the two compressors may be interconnected in parallel, with both pumping from about 45°F SST to 120°F SDT for cooling. Then, at some predetermined outdoor temperature on heating, they are reconnected in series and compress in two successive stages.

Figure 6 shows the performance of a pair of compressors for units of both medium and low clearance volume. At low suction temperatures, the reconnection in series adds some capacity. However, the motor for this case must be selected for the maximum loading conditions for summer operation, even though the low-stage compressor has a greatly reduced power requirement under the heating condition.

Figure 6. Comparison of Parallel and Staged Operation for Air-Source Heat Pumps

Refrigerant-to-air and refrigerant-to-water heat exchangers are similar to heat exchangers used in air-conditioning refrigeration systems.

Figure 7. Suction Line Separator for Protection Against Liquid Floodback

 Liquid Subcooling Coils.

 A refrigerant subcooling coil can be added to the heat pump cycle (Figure 8) to preheat ventilation air during heating and, at the same time, to lower the liquid refrigerant temperature. Depending on the refrigerant circulation rate and quantity and temperature of ventilation air, heating capacity can be increased as much as 15 to 20%, as indicated in Figure 9.

Figure 8. Liquid Subcooling Coil in Ventilation Air Supply to Increase Heating Effect and Heating COP

Figure 9. Typical Increase in Heating Capacity Resulting from Using Liquid Subcooling Coil

Refrigerant piping, receivers, expansion devices, and refrigeration accessories in heat pumps are usually the same as components found in other types of refrigeration and air-conditioning systems.

A reversing valve changes the system from cooling to heating mode. This changeover requires using a valve(s) in the refrigerant circuit, except where the change occurs in fluid circuits external to the refrigerant circuit. Reversing valves are usually pilot-operated by solenoid valves, which admit compressor head and suction pressures to move the operating elements.

The expansion device for controlling the refrigerant flow is normally a thermostatic expansion valve, which is described in Chapter 11 of the 2014 ASHRAE Handbook—Refrigeration. The control bulb must be located carefully. If the circuit is arranged so that the refrigerant line on which the control bulb is placed can become the compressor discharge line, the resulting pressure developed in the valve power element may be excessive, requiring a special control charge or pressure-limiting element. When a thermostatic expansion valve is applied to an outdoor air coil, a special cross-charge to limit the superheat at low temperatures allows better use of the coil. In its temperature-sensing element, a cross-charged thermostatic expansion valve uses a fluid, or mixture of fluids, that is different from the refrigerant. An electronic expansion valve improves control of refrigerant flow.

A capillary tube (used as a metering device for an air-source heat pump that operates over a wide range of evaporating temperatures) may pass refrigerant at an excessive rate at low back pressures, causing liquid floodback to the compressor. Suction line accumulators or charge-control devices are sometimes added to minimize this effect. Suction line accumulators may also prevent liquid refrigerant that has migrated to the evaporator from entering the compressor on start-up.

When separate metering devices are used for the two heat exchangers, a check valve allows refrigerant to bypass the metering device of the heat exchanger serving as the condenser.

A refrigerant receiver, which is commonly used to store liquid refrigerant, is particularly useful in a heat pump to take care of the unequal refrigerant requirements of heating and cooling. The receiver is usually omitted on heat pumps used for heating only.

Heating needs may exceed the heating capacity available from equipment selected for the cooling load, particularly if outdoor air is used as the heat source. When this occurs, supplemental heating or additional compressor capacity should be considered. The additional compressor capacity or the supplemental heat is generally used only in the severest winter weather and, consequently, has a low usage factor. Both possibilities must be evaluated to determine the most economical selection.

When supplemental heaters are used, the elements should always be located in the air or water circuit downstream from the heat pump condenser. This allows the system to operate at a lower condensing temperature, increasing heating capacity and improving the COP. Controls should sequence the heaters so that they are energized after all heat pump compressors are fully loaded. An outdoor thermostat is recommended to limit or prevent energizing heater elements during mild weather when they are not needed. Where 100% supplemental heat is provided for emergency operation, it may be desirable to keep one or more stages of the heaters locked out whenever the compressor is running. In this way, the cost of electrical service to the building is reduced by limiting the maximum coincidental demand.

A flow switch should be used to prevent operation of the heating elements and heat pump when there is no air or water flow.

Closed-cycle systems use a suitable working fluid, usually a refrigerant in a sealed system. They can use either absorption or vapor compression Heat is transferred to and from the system through heat exchangers similar to refrigeration system heat exchangers. Closed-cycle heat pump systems are often classed with industrial refrigeration systems except that they operate at higher temperatures.

Heat exchangers used must comply with federal and local codes. For example, a double separation between potable water and refrigerant is usually required. However, some applications need local interpretation of potable versus service definitions, such as water used in food preparation (e.g., poultry scalders, equipment cleanup).

Heat exchangers must also be resistant to corrosion and fouling conditions of the source and sink fluids. The refrigerant and oil must be (1) compatible with component materials and (2) mutually compatible at the expected operating temperatures. In addition, the viscosity and foaming characteristics of the oil and refrigerant mixtures must be consistent with the lubrication requirements at the specific mechanical load imposed on the equipment. Proper oil return and heat transfer at the evaporators and condensers must also be considered.

The specific application of a closed-cycle heat pump frequently dictates the selection considerations. In this section, the different types of closed-cycle systems are reviewed, and factors important to the selection process are addressed.

Air-to-air heat pumps or dehumidification heat pumps (Figure 10) are most frequently used in industrial operations to dry or cure products. For example, dehumidification kilns are used to dry lumber to improve its value. Compared to conventional steam kilns, the heat pump provides two major benefits: improved product quality and reduced percent degrade. With dehumidification, lumber can be dried at a lower temperature, which reduces warping, cracking, checking, and discoloration. The system must be selected according to the type of wood (i.e., hard or soft) and the required dry time. Dehumidification heat pumps can also be used to dry agricultural products; poultry, fish, and meat; textiles; and other products.

Figure 10. Dehumidification Heat Pump

Air-to-water heat pumps or heat pump water heaters (Figure 11) are a special application of closed-cycle systems that are usually used to dehumidify indoor air where humidity is generated from some activity or process. Typical applications are indoor pools, large commercial kitchens, and industrial laundries. In many tropical climates, the ambient air can also be used to heat water economically where power systems are costly and local natural gas is scarce (e.g., on islands). In most of these applications, the air coil is usually wet and may require coated coils or special metallurgy. See Chapter 49 of this volume and Chapter 50 of the 2019 ASHRAE Handbook—HVAC Applications for more information.

Figure 11. Air-to-Water Heat Pump

Water-to-water heat pumps may have the most widespread application in industry. They can use cooling tower water, effluent streams, and even chilled-water makeup streams as heat sources. The output hot water can be used for product rinse tanks, equipment cleanup water systems, and product preheaters. The water-to-water heat pump system may be simple, such as that shown in Figure 12, which recovers heat from a process cooling tower to heat water for another process. Figure 12 also shows the integration of a heat exchanger for preheating the process water. Typically, the heat pump COP is in the range of 4 to 6, and the system COP reaches 8 to 15.

Water-to-water heat pump systems may also be complicated, such as the cascaded HRHP system (Figure 13) for a textile dyeing and finishing plant. Heat recovered from various process effluent streams is used to preheat makeup water for the processes. The effluent streams may contain materials, such as lint and yarn, and highly corrosive chemicals that may foul the heat exchanger; therefore, special materials and antifouling devices may be needed for the heat exchanger for a successful design.

Figure 12. Cooling Tower Heat Recovery Heat Pump

Most food-processing plants, which use more water than a desuperheater or a combination desuperheater/condenser can provide, can use a water-to-water heat pump (Figure 14). A water-cooled condenser recovers both sensible and latent heat from the high-pressure refrigerant. Water heated in the condenser is split into two streams: one as a heat source for the water-to-water heat pump, the other to preheat the makeup, process, or cleanup water. Preheated water is blended with hot water from the storage tank to limit the temperature difference across the heat pump condenser to about 20°F, which is standard for many chiller applications. The heated water is then piped back to the storage tank, which is typically sized for 1.5 to 5 h of holding capacity. Because the refrigeration load may be insufficient to provide all the water heating required, existing water heaters (usually steam) provide additional heat and control for process water at the point of use.

Figure 13. Effluent Heat Recovery Heat Pump

Three major tasks must be addressed when adding heat recovery to existing plant refrigeration systems: (1) forcing the hot-gas refrigerant to flow to the desired heat exchanger, (2) scheduling the refrigeration processes to provide an adequate heat source over time while still meeting process requirements, and (3) integrating water-cooled condensers with evaporative condensers. The refrigerant direction can be controlled either by series piping of the two condenser systems or by three pressure-regulating valves (PRVs): one to the hot-gas defrost (lowest-pressure setting), one to the recovery system (medium-pressure setting), and one to the evaporative condensers (highest-pressure setting). PRVs offer good control but can be mechanically complicated. Series piping is simple but can cost more because of the pipe size required for all the hot gas to pass through the water-cooled condenser.

Figure 14. Refrigeration Heat Recovery Heat Pump

Each refrigeration load should be reviewed for its required output and production requirements. For example, ice production can frequently be scheduled during cleanup periods, and blast freezing for the end of shifts.

In multisystem integration, as in expanded system integration, equalization lines, liquid lines, receivers, and so forth must be designed according to standard refrigeration practices.

Ammonia compressors in a condensing (Figure 15) or cascaded system sometimes can also be used where plant refrigerant gas is compressed further to generate the needed temperature. This option is strongly influenced by the plant refrigeration and compressor systems, because pressure requirements may require high-pressure compressors. Consult the ammonia compressor manufacturer if considering this option.

Process-fluid-to-process-fluid heat pumps can be applied to evaporation, concentration, crystallization, and distillation process fluids that contain chemicals that would destroy a steam compressor. A closed-cycle vapor compression system (Figure 16) is used to separate a solid and a liquid in an evaporator system and to separate two liquids in a distillation system. Both systems frequently have COPs of 8 to 10 and have the added benefit of cooling tower elimination. These systems can frequently be specified and supplied by the column or evaporator manufacturer as a value-added system.

Figure 15. Condensing Ammonia Heat Pump
(Courtesy Vilter Manufacturing and Emerson Climate Technologies)

Figure 16. Closed-Cycle Vapor Compression System

The benefits of water-to-water, refrigerant-to-water, and other fluid-to-fluid systems may include the following:

Open- and semi-open-cycle heat pump systems use process fluid to raise the temperature of available heat energy by vapor compression, thus eliminating the need for chemical refrigerants. The most important class of applications is steam recompression. Compression can be provided with a mechanical compressor or by a thermocompression ejector driven by the required quantity of high-pressure steam. The three main controlling factors for this class of systems are vapor quality, boiling point elevation, and chemical makeup.

Recompression of boiler-generated process steam (Figure 17) has two major applications: (1) large facilities with substantial steam pressure drop due to line losses and (2) facilities with a considerable imbalance between steam requirements at low, medium, and high levels. Boiler-generated process steam usually conforms to cleanliness standards that ensure corrosion-free operation of the compression equipment. Evaluation of energy costs and steam value can be complicated with these applications, dictating the use of analytical tools such as pinch technology.

Application of open-cycle heat pumps to evaporation processes is exceptionally important. Single-effect evaporators (Figure 18) are common when relatively small volumes of water and solid need to be separated. The most frequent applications are in the food and dairy industries. The overhead vapors of the evaporator are compressed, and thus heated, and piped to the system heater (i.e., calandria). The heat is transferred to the dilute solution, which is then piped to the evaporator body. Flashing occurs upon entry to the evaporator body, sending concentrate to the bottom and vapors to the top. System COPs reach 10 to 20, and long-term performance of the evaporator improves as less of the product (and thus less build-up) occurs in the calandria because of lower operating temperatures. Multiple-effect evaporators (Figure 19), usually applied in the paper and chemical industries, use the same principles, only on a much larger scale. The multiple evaporator bodies can be piped in series or in parallel. System COPs can exceed 30.

Figure 17. Recompression of Boiler-Generated Process Steam

Using open-cycle heat pumps in distillation processes (Figure 20) is similar to evaporation. However, compression of flammable gaseous compounds can be dangerous, so great care must be taken. The overhead vapors are compressed to a higher pressure and temperature, and then condensed in the reboiler. This eliminates the need for boiler steam in the reboiler and reduces overall energy consumption. As the pressure of the condensed vapor is reduced through the expansion valve, some of the liquid at a lower temperature is returned to the column as reflux, and the balance forms the overhead product, both as liquid and as vapor. Vapor is recycled as necessary. System COPs can reach 30 or 40.

Figure 18. Single-Effect Heat Pump Evaporator

Process emission-to-steam heat pump systems can be used for cooking, curing, and drying systems that operate with low-pressure saturated or superheated steam. The vapors from a cooking process, such as at a rendering plant, can be recovered to generate the steam required for cooking (Figure 21). The vapors (cooking exhaust) are compressed with a screw compressor because noncondensable materials removed from the process by the vapor could erode or damage reciprocating or centrifugal compressors. Compressed steam is then supplied to the cooker, but may require some makeup water to desuperheat the steam. The steam condenses in the cooking process, and the condensate is then used to heat process water using a heat exchanger for additional heat recovery. The system condensate cannot be recirculated to the compressor because of contaminants; instead, it is sent to the system scrubbers or cleanup system where noncondensables are usually scrubbed or incinerated, and the water is treated or discharged to a sewer.

Figure 19. Multiple-Effect Heat Pump Evaporator

Contaminants in some processes require the use of a semi-open-cycle heat pump (Figure 22). This system uses a heat exchanger, frequently called a reboiler, to recover heat from the stack gases. The reboiler produces low-pressure steam, which is compressed to the desired pressure and temperature. A clean-in-place (CIP) system may be used if the volume of contaminants is substantial.

Figure 20. Distillation Heat Pump System

Variable-speed drives can be specified with all these systems for closer, more efficient capacity control. Additional capacity for emergencies can be made available by temporarily overspeeding the drive, while sizing for nominal operating conditions to retain optimal efficiency. Proper integration of heat pump controls and system controls is essential.

Apply the following basic principles when designing heat recovery systems:

ASHRAE research project RP-807 (Caneta Research 1998) produced guidelines for evaluating environmental benefits (e.g., energy, water, plant emission reductions) of heat recovery heat pumps. Calculation procedures for energy and water savings and plant emission reductions were outlined. Other, less easily quantified benefits were also documented. Evaluation guidelines were provided for a heat pump water heater in a hospital kitchen; a heat recovery heat pump at a resort with a spa, pool, and laundry service building; heat recovery from refrigeration compressor superheat for makeup water heating; and various drying processes using heat pumps and concentration processes. These guidelines can be used to quantify energy, water, and emission reductions and help justify using heat recovery systems in commercial, institutional, and industrial buildings.

Figure 21. Heat Recovery Heat Pump System in a Rendering Plant

Figure 22. Semi-Open-Cycle Heat Pump in a Textile Plant

In many large buildings, internal heat gains require year-round chiller operation. The chiller condenser water heat is often wasted through a cooling tower. An increasingly common approach to using this waste heat is adding a water-to-water heat pump in the chilled-water system to boost the quality of this waste heat to create hot water. This arrangement uses the otherwise wasted heat to provide the higher temperatures required for one or more of the following:

This section provides an overview of what engineers should consider before adding a heat recovery heat pump into a new or existing chilled-water system to offset less efficient means of heat production.

Note that, to ensure correct application of an HRHP, proper second-law, pinch, or other thermodynamic analysis methods should be used before applying a heat recovery heat pump in a waste heat stream. Furthermore, the ability to exchange heat before using a heat pump should always be considered first in any heat exchange process. For example, if the only area of consideration is hot water produced by a boiler, it might be more economical to investigate a heat exchanger in the cooling tower loop to preheat feed water rather than a full mechanical system to replace the full boiler operation. This section uses the term heat recovery heat pump (HRHP) to designate a chiller that provides usable hot water as at least one of its intended roles.

Heat recovery heat pumps for waste heat recovery should be considered during any initial design, retrofit, or changeout of an existing chiller. The exact reasons for deciding on a heat recovery chiller vary, but economic justification of an HRHP is generally based on one or a combination of the following (Schwedler and Brunsvold 2011):

Other benefits that may be considered include (Dorgan et al. 1999)

Though it may be possible to apply an HRHP to most hot-water needs, the specific temperature requirement of the hot water must be fully understood because it affects the overall system efficiency, availability of applicable equipment, and operating conditions of the HRHP.

In applications where the heat produced by the chiller plant is not enough to provide the desired amount of heat for the heat recovery system (e.g., when a load-shedding economizer reduces the operation of a chiller plant yet there is still a perimeter heating load), consider an alternative source of “free” heat. Examples include

Figure 23 shows the potential locations of an HRHP in a simplified chiller loop. Note these locations are for illustration only; this figure does not show any bypass valves, pumps, control points, or auxiliary heat exchangers that may be needed in the actual layout.

 Parallel Application (Location 1 in Figure 23).

 This is generally a high-lift chiller because it must provide a temperature lift from chilled-water ranges to usable hot-water temperatures. However, this location can provide the highest overall system savings in a new design. This unit can be the only chiller in the plant, equal-percentage loaded, or preferentially loaded. However, this application does not allow for control of the leaving hot-water temperature without a variable evaporator flow (Schwedler and Brunsvold 2011).

Figure 23. Possible Heat Recovery Heat Pump Locations

Using an HRHP on the condenser side of a chilled-water system (locations 3, 4, and 5) may allow for a smaller and/or lower lift system. However, there are specific benefits and challenges to consider in the overall effect on the building’s systems.

Benefits to the cooling tower loop include the ability to design bypass piping and to bypass the cooling tower when the HRHP can absorb the total heat rejection of the condenser loop to which it is attached. Take this offloading of the cooling tower and the associated runtime reduction into account in the overall economic analysis of the HRHP system. Under low-temperature heating needs, condenser-side heat recovery can also allow use of the existing chiller with minor changes through an increase in the operating pressure ratio of an existing chiller.

Challenges associated with the application of condenser-side heat recovery include the dependency of heat recovery on the operation of the main chiller. If the loads are not coincident, the heat recovery system cannot operate without the main chiller system operating. This could significantly impact the economics of the HRHP. Additionally, fouling of the HRHP must be considered when using the condenser-loop, and isolation of the fluid streams may be required.

Furthermore, take into account the effect of the heat recovery system on the condenser loop temperature, to ensure it does not negatively affect the existing chiller system’s operating ability. Although the potential to bypass the cooling tower may be a benefit, it can also be a challenge in the control of the overall system, and requires a cooling tower bypass or the ability to appropriately offload the HRHP to maintain condenser loop temperatures entering the existing chiller. However, this increased capacity on the condenser side could provide an increase in the operational efficiency of the existing chiller.

Just as there are both benefits and challenges of condenser-side recovery, there are similar analogues with evaporator-side heat recovery (locations 1 and 2) that must be considered in the overall operational scheme and impact on the overall system.

Notable benefits of evaporator-side heat recovery include the ability to preferentially load the HRHP and provide a maximum amount of recovered heat while offloading other chillers in the system. Furthermore, the same magnitude of cooling tower savings is achieved using evaporator-side heat recovery, because the recovered heat is never sent to the cooling tower loop. The concern for fouling is also eliminated. Evaporator-side heat recovery should be analyzed in applications where expanded chiller capacity is required, if a beneficial sink of the recovered heat is available, such as those listed in the Applications of Waste Heat Recovery section.

Notable challenges associated with evaporator-side heat recovery include its effect on the overall chiller plant. Because the HRHP must have maximum runtime for payback considerations, the turndown ability of other chillers in the system must be considered. Furthermore, the low-temperature source water entering the HRHP could require a high-lift or multi-stage HRHP system to provide the required sink temperatures, which results in a chiller efficiency impact that must be considered in the economic evaluation of the HRHP system.

Double-bundle (or dual-condenser) systems offer a hybrid application of evaporator-side and condenser-side heat recovery with benefits of both application types. The ability to reject heat to either the cooling tower loop or the heat recovery loop enables flexibility in the application of an HRHP system. Though the challenges mentioned above should still be considered. The double-bundle application is not specifically covered in the applications presented in this chapter.

It is extremely important to understand the loading and unloading ability of the type of machine selected for energy recovery capability and the impact this ability has on overall system and equipment efficiency. Both dynamic (or turbomachinery) and positive displacement compressors must be considered and their impact thoroughly noted. During system and compressor selection, consideration should be given to

Chapters 38 and 43 discuss compressors in detail and covers many of the considerations listed above.

A compressor should not be allowed to enter the surge region on any machine, but specifically for an HRHP, a system should be checked for its ability to unload at the expected condenser conditions (Schwedler and Brunsvold 2011).

Multiple options exist for control points on condenser water flow, including variable or constant flow. The chilled-water flow equally has options, including primary/secondary variable flow (Figure 24) or variable primary (Figure 25) schemes. Each of these flow conditions will greatly impact the operation of the HRHP system and its ability to reach its desired heating conditions. Descriptions of these configurations are given in the System Flow Design section of Chapter 3.

Figure 24. Primary/Secondary, Equal Loading
(Schwedler and Brunsvold 2011)

For HRHP condenser water flow, it is important to understand the turndown and unloading characteristics of the HRHP system. A constant-flow condenser system can be controlled based on a constant leaving-water temperature or a constant entering-water temperature, each of which has a distinctive impact on the unloading of the HRHP (Schwedler and Brunsvold 2011).

Figure 25. Variable Primary Flow Example
(Schwedler and Brunsvold 2011)

Once potential uses and their associated temperature needs are known, steps can be taken to evaluate the heat recovery potential of the system. Higher leaving-water temperatures may reduce the cost of airside equipment, but the higher lift required by the HRHP will have associated equipment and operating cost implications. Therefore, the application must be analyzed from a system scale, both before and after application of an HRHP, to ensure the optimal capital and operating cost balance is met. Note that it is always more beneficial to exchange heat rather than apply a heat pump system. Therefore, before sizing an HRHP, search for opportunities to recover heat through heat exchange rather than moving heat with heat pumps.

An HRHP must be able to reject all of its condenser heat at all times, and must therefore be able to either always reject 100% of its load or have appropriate turndown. Therefore, it is imperative to fully analyze annual source and sink loads to ensure the unit is not oversized.

Before sizing an HRHP, an annual hourly inventory of loads should be collected, including

Dorgan et al. (1999) outline four steps for applying an HRHP, as shown in Figure 26.

 Step 1.

 Evaluate heating and cooling loads, including all loads identified in the inventory. Multiple methods are recommended for evaluation of the building loads, including bin and full hourly analysis. However, the loads should be used to create a heat balance chart similar to Figures 27 and 28. Figure 27 illustrates a building with significantly different occupied and unoccupied periods and unbalanced loads, which requires deeper analysis to ensure the proper size of the HRHP. Figure 28 shows a more balanced load condition and the point at which the HRHP rejects all of its heat produced.

Figure 26. HRHP Application Flowchart
(Dorgan et al. 1999)

Figure 27. Heat Balance Chart
(Dorgan et al. 1999)

In a situation where the available heat source outweighs the heat sink loads, the HRHP can be sized to the peak heating load, as long as the HRHP’s turndown can match the heating load’s variability. However, in a simultaneous heating and cooling situation, the ideal goal is to turn off the chiller in the winter and turn off the boiler in the summer, as shown in Figure 29. Note that, in Figure 29, supplemental cooling and heating are required where the HRHP [or water-to-water heat pump (WTWHP) here] cannot provide the required capacities. This case assumes the HRHP is preferentially loaded on the chilled-water loop while also providing heat recovery benefit with the condenser. In cold temperatures, as the heating load rises, the remaining heat load must be covered by an auxiliary heat source. This is the case with any HRHP system.

Figure 28. Selection of Simultaneous Heating and Cooling HRHP
(Campbell et al. 2012)

 Step 2.

 Once the required loads are understood, including the temperature requirement of the heat recovery loop and the available heat to be recovered, the system type can be selected. Special attention should be paid to the compressor type and turndown ability. Perhaps most importantly, attention should be paid to the corrected capacity and efficiency impacts on the HRHP and other chillers in the system as a result of application of the HRHP system. Condenser- and evaporator-side temperature changes affect capacities of all chillers in the system, and their effects on overall efficiency must be known for economic analysis.

Once an HRHP system is selected, the appropriate auxiliary source should be selected or analyzed if auxiliary heating will be required at any time throughout the year. If no heating system is present, as in a new design, select one with proper consideration of capacity and inlet water temperatures. For an existing plant, analyze the existing heating source for applicability to the altered system, including analysis of turndown, inlet water temperature, operating hours, etc.

Figure 29. Operating Areas for Simultaneous HRHP
(Campbell et al. 2012)

A detailed example of HRHP application, including load analysis, equipment selection, and economic evaluation, can be found in Dorgan et al. (1999).

A water-loop heat pump (WLHP) system combines load transfer characteristics with multiple water-to-air heat pump units (Figure 30). Each zone, or space, has one or more water-to-air heat pumps. The units in both the building core and perimeter areas are connected hydronically with a common two-pipe system; however, one-pipe approaches have been proven to be highly effective if designed to take advantage of the less complex system layout. Each unit cools conventionally, supplying air to the individual zone and rejecting the heat removed to the two-pipe system through its integral condenser. Excess heat gathered by the two-pipe system is expelled through a common heat rejection device, which often includes a closed-circuit evaporative cooling tower with an integral spray pump. If and when some of the zones, particularly on the northern side of the building, require heat, the individual units switch (by means of reversing refrigerant valves) into the heating cycle. These units then extract heat from the two-pipe water loop, a relatively high-temperature source that is totally or partially maintained by heat rejected from the condensers of the units that provide cooling to other zones. When only heating is required, all units are in the heating cycle and, consequently, an external heat input to the loop is needed to maintain the loop temperature. The water-loop temperature has traditionally been maintained in the range of 60 to 90°F and, therefore, seldom requires piping insulation. However, extended-range water-loop heat pumps are becoming more common because they provide high efficiency over a wider operating range. The expanded operating range (45 to 110°F) reduces boiler and cooling tower operating costs. The lower operating temperatures may require insulating the main supply and return lines. These units typically come equipped with insulated water-to-refrigerant heat exchangers, loss-of-flow protection, and thermoexpansion valves instead of the traditional capillary tube expansion device.

Figure 30. Heat Recovery System Using Water-to-Air Heat Pumps in a Closed Loop

Any number of water-to-air heat pumps may be installed in such a system. Water circulates through each unit via the closed loop.

The water circuit can include two circulating pumps (one pump is 100% standby) or may use variable-speed drives and a means for adding and rejecting heat to and from the loop. Each heat pump can either heat or cool to maintain the comfort level in each zone. Detailed information on the piping and pumping design for closed-loop water-to-air heat pump systems can be found in Chapter 34 of the 2019 ASHRAE Handbook—Applications and Chapter 6 of Kavanaugh and Rafferty (2014).

Units in heating mode extract heat from the circulated water, whereas those in cooling mode reject heat to the water. Thus, the system recovers and redistributes heat, where needed. Unlike air-source heat pumps, heating output for this system does not depend on outdoor temperature. The water loop conveys rejected heat, but a secondary heat source, typically a boiler, is usually provided.

Another WLHP version uses a coil buried in the ground as a heat source and sink. This ground-coupled system does not normally need the boiler and cooling tower incorporated in conventional WLHP systems to keep circulating water within acceptable temperature limits. However, ground-coupled heat pumps may operate at lower entering water (or antifreeze solution) temperatures. Some applications may require a heat pump tolerant of entering water temperatures ranging from 25 to 110°F. In climates where air conditioning dominates, a cooling tower is sometimes combined with ground coupling to reduce overall installed costs. Such hybrid systems are covered in more detail in Chapter 34 of the 2019 ASHRAE Handbook—Applications and Chapter 6 of Kavanaugh and Rafferty (2014).

Figure 31 shows a system with a storage tank and solar collectors. The storage tank in the condenser circuit can store excess heat during occupied hours and provide heat to the loop during unoccupied hours. During this process, solar heat may also be added within the limitations of the temperatures of the condenser system. The solar collectors are more effective and efficient under these circumstances because of the temperature ranges involved.

For further heat reclaim, a water-to-water heat pump can be added in the closed water loop before the heat rejection device. This heat pump provides domestic hot water or elevates loop water temperatures in a storage tank so the water can be bled back into the loop when needed during the heating cycle. Figure 32 shows such a system.

Figure 31. Closed-Loop Heat Pump System with Thermal Storage and Optional Solar-Assist Collectors

Many facilities require large cooling loads (e.g., for interior zones, lights, people, business machines, computers, switchgear, and production machinery) that result in net loop heat rejection during all or most of the year, particularly during occupied hours. This waste heat rejection often occurs while other heating loads in the facility (e.g., ventilation air, reheat, domestic hot water) are using external purchased energy to supply heat. By including a secondary water-to-water heat pump (as in Figure 32), additional balanced heat recovery can be economically achieved. The secondary heat pump can effectively reclaim this otherwise rejected heat, raise its temperature, and use it to serve other heating loads, thus minimizing use of purchased energy.

Figure 32. Secondary Heat Recovery from WLHP System
(Adapted from Marketing the Industrial Heat Pump, Edison Electric Institute 1989)

In another WLHP system variation, the building sprinkler system is used as part of the loop water distribution system.

Some aspects of systems described in this section may be proprietary and should not be used without appropriate investigation.

Water-loop heat pump (WLHP) systems are used in many types of multiroom buildings. A popular application is in office buildings, where heat gains from interior zones can be redistributed to the perimeter during winter. Other applications include hotels and motels, schools, apartment buildings, nursing homes, manufacturing facilities, and hospitals. Operating costs for these systems are most favorable in applications with simultaneous heating and cooling requirements.

Accurate design tools are needed to model and predict WLHP energy use and electrical demand. ASHRAE research project RP-620 (Cane et al. 1993) evaluated computer models for water-loop heat pump systems and compared the accuracy of the test models with actual monitored data from buildings with similar systems. Although the computer programs predicted whole-building energy use within 1 to 15% of measured total energy use over a 12-month period, much larger variations were found at the HVAC system and component levels.

Cane et al. (1993) concluded that the models could be improved by

 Heat Rejector Selection.

 A closed-loop circuit requires a heat rejector that is either a heat exchanger (loop water to cooling tower water), a closed-circuit evaporative cooler, or a ground coil. The heat rejector is selected in accordance with manufacturer’s selection curves, using the following parameters:

• Water flow rates. Manufacturers’ recommendations on water flow rates vary between 2 and 3 gpm per ton of installed cooling capacity. Lower flow rates are generally preferred in regions with a relatively low summer outdoor design wet-bulb temperature. In more humid climates, a higher flow rate allows a higher water temperature to be supplied from the heat rejector to the heat pumps, without a corresponding increase in temperature leaving the heat pumps. Thus, cooling tower or evaporative cooler size and cost are minimized without penalizing the performance of the heat pumps.

• Water temperature range. Range (the difference between leaving and entering water temperatures at the heat rejector) is affected by heat pump cooling efficiency, water flow rate, and diversity. It is typically between 10 and 15°F.

• Approach. Approach is the difference between the water temperature leaving the cooler and the wet-bulb temperature of the outdoor air. The maximum water temperature expected in the loop supply is a function of the design wet-bulb temperature.

• Diversity. Diversity is the maximum instantaneous cooling load of the building divided by the installed cooling capacity:

  (1)

where

D	=	diversity
Qm	=	maximum instantaneous cooling load
Qi	=	total installed cooling capacity

Diversity times the average range of the heat pumps is the applied range of the total system (the rise through all units and the drop through the heat rejector). For systems with a constant pumping rate regardless of load,

(2)

where

Rs	=	range of system
Rp	=	average range of heat pumps

For systems with a variable pumping rate and with each pump equipped with a solenoid valve to start and stop water flow through the unit with compressor operation,

(3)

The average leaving water temperature of the heat pumps is the entering water temperature of the heat rejector. The leaving water temperature of the heat rejector is the entering water temperature of the heat pumps.

• Winterization. For buildings with some potential year-round cooling (e.g., office buildings), loop water may be continuously pumped through the heat rejector. This control procedure reduces the danger of freezing. However, it is important to winterize the heat rejector to minimize the heat loss.

In northern climates, the most important winterization step for evaporative coolers is installing a discharge air plenum with positive-closure, motorized, ice-proof dampers. The entire casing that houses the tube bundle and discharge plenum may be insulated. The sump, if outside the heated space, should be equipped with electric heaters. The heat pump equipment manufacturer’s instructions will help in the selection and control of the heat rejector.

If sections of the water circuit will be exposed to freezing temperatures, consider adding an antifreeze solution. In a serpentine pipe circuit with no automatic valves that could totally isolate individual components, an antifreeze solution prevents bursting pipes, with minimal effect on system performance.

An open cooling tower with a separate heat exchanger is a practical alternative to the closed water cooler (Figure 33). An additional pump is required to circulate the tower water through the heat exchanger. In such an installation, where no tubes are exposed to the atmosphere, it may not be necessary to provide freeze protection on the tower. The sump may be indoors or, if outdoors, may be heated to keep the water from freezing. This arrangement allows the use of small, remotely located towers. Temperature control necessary for tower operation is maintained by a sensor in the water-loop system controlling operation of the tower fan(s).

The combination of an open tower, heat exchanger, and tower pump frequently has a lower first cost than an evaporative cooler. In addition, operating costs are lower because no heat is lost from the loop in winter and, frequently, less power is required for the cooling tower fans.

 Ductwork Layout.

 Often, a WLHP system has ceiling-concealed units, and the ceiling area is used as the return plenum. Troffered light fixtures are a popular means of returning air to the ceiling plenum.

Figure 33. Cooling Tower with Heat Exchanger

Air supply from the heat pumps should be designed for quiet operation. Heat pumps connected to ductwork must be capable of overcoming external static pressure. Heat pump manufacturers’ recommendations should be consulted for the maximum and minimum external static pressure allowable with each piece of equipment. However, note that using smaller units and less ductwork is generally preferable, to maximize efficiency.

The WLHP system has simpler controls than totally central systems. Each heat pump is controlled by a thermostat in the zone. There are only two centralized temperature control points: one to add heat when the water temperature approaches a prescribed lower temperature (45 to 60°F), and the other to reject heat when the water temperature approaches a prescribed upper temperature (typically about 90 to 110°F).

The boiler controls should be checked to be sure that outlet water is controlled at 60°F, because controls normally supplied with boilers are in a much higher range.

An evaporative cooler should be controlled by increasing or decreasing heat rejection capacity in response to the loop water temperature leaving the cooler. A reset schedule that operates the system at a lower water temperature (to take advantage of lower outdoor wet-bulb temperatures) can save energy when heat from the loop storage is not likely to be used.

Abnormal condition alarms typically operate as follows:

Outdoor ambient control should be provided to prevent operation of the cooling tower sump pump at freezing temperatures.

Optional system control arrangements include the following:

In an ideal heat recovery system, all components work year-round to recover all the internal heat before adding external heat. Any excess heat is either stored or rejected. Such an idealized goal is identified as a balanced heat recovery system. Note that “balanced” refers to operation of the heat recovery system; loads will rarely match perfectly year round.

When the outdoor temperature drops significantly, or when the building is shut down (e.g., on nights and weekends), internal heat gain may be insufficient to meet the space heating requirements. Then, a balanced system provides heat from storage or an external source. When internal heat is again generated, the external heat is automatically reduced to maintain proper temperature in the space. There is a time delay before equilibrium is reached. The size of the equipment and the external heat source can be reduced in a balanced system that includes storage. Regardless of the system, a heat balance analysis establishes the merits of balanced heat recovery at various outdoor temperatures.

Outdoor air less than 55 to 65°F may be used to cool building spaces with an air economizer cycle. When considering this method of cooling, the space required by ducts, air shafts, and fans, as well as the increased filtering requirements to remove contaminants and the hazard of possible freeze-up of dampers and coils must be weighed against alternatives such as using deep row coils with antifreeze fluids and efficient heat exchange. Innovative use of heat pump principles may give considerable energy savings and more satisfactory human comfort than an air economizer. In any case, hot and cold air should not be mixed (if avoidable) to control zone temperatures because it wastes energy.

Many buildings, especially those with computers or large interior areas, generate more heat than can be used for most of the year. Operating cost is minimized when the system changes over from net heating to net cooling at the break-even outdoor temperature at which the building heat loss equals the internal heat load. If heat is unnecessarily rejected or added to the space, the changeover temperature varies from the natural break-even temperature, and operating costs increase. Heating costs can be reduced or eliminated if excess heat is stored for later distribution.

The concept of ideal heat balance in an overall building project or a single space requires that one of the following takes place on demand:

In small air-conditioning projects serving only one space, either cooling or heating satisfies the thermostat demand. If humidity control is not required, operation is simple. Assuming both heating and cooling are available, automatic controls will respond to the thermostat to supply either. A system should not heat and cool the same space simultaneously.

Multiroom buildings commonly require heating in some rooms and cooling in others. Optimum design considers the building as a whole and transfers excess internal heat from one area to another, as required, without introducing external heat that would require waste heat disposal at the same time. The heat balance concept is violated when this occurs.

Figure 34. Major Load Components

Humidity control must also be considered. Any system should add or remove only enough heat to maintain the desired temperature and control the humidity. Large percentages of outdoor air with high wet-bulb temperatures, as well as certain types of humidity control, may require reheat, which could upset the desirable balance. Usually, humidity control can be obtained without upsetting the balance. When reheat is unavoidable, internally transferred heat from heat recovery should always be used to the extent it is available, before using an external heat source such as a boiler. However, the effect of the added reheat must be analyzed, because it affects the heat balance and may have to be treated as a variable internal load.

When a building requires heat and the refrigeration plant is not in use, dehumidification is not usually required and the outdoor air is dry enough to compensate for any internal moisture gains. This should be carefully reviewed for each design.

The following examples illustrate situations that can occur in nonrecovery and unbalanced heat recovery situations. Figure 34 shows the major components of a building that comprise the total air-conditioning load. Values above the zero line are cooling loads, and values below the zero line are heating loads. On an individual basis, the ventilation and conduction loads cross the zero line, which indicates that these loads can be a heating or a cooling load, depending on outdoor temperature. Solar and internal loads are always a cooling load and are, therefore, above the zero line.

Figure 35 combines all the loads shown in Figure 34. The graph is obtained by plotting the conduction load of a building at various outdoor temperatures, and then adding or subtracting the other loads at each temperature. The project load lines, with and without solar effect, cross the zero line at 16 and 30°F, respectively. These are the outdoor temperatures for the plotted conditions when the naturally created internal load exactly balances the loss.

Figure 35. Composite Plot of Loads in Figure 34 (Adjust for Internal Motor Heat)

As plotted, this heat balance diagram includes only the building loads with no allowance for additional external heat from a boiler or other source. If external heat is necessary because of system design, the diagram should include the additional heat.

Figure 36 shows what happens when heat recovery is not used. It assumes that, at a temperature of 70°F, heat from an external source is added to balance conduction through the building’s skin in increasing amounts down to the minimum outdoor temperature winter design condition. Figure 36 also adds the heat required for the outdoor air intake. The outdoor air comprising part or all of the supply air must be heated from outdoor temperature to room temperature. Only the temperature range above the room temperature is effective for heating to balance the perimeter conduction loss.

Figure 36. Non-Heat-Recovery System

These loads are plotted at the minimum outdoor winter design temperature, resulting in a new line passing through points A, D, and E. This line crosses the zero line at −35°F, which becomes the artificially created break-even temperature rather than 30°F, when not allowing for solar effect. When the sun shines, the added solar heat at the minimum design temperature would further drop the −35°F break-even temperature. Such a design adds more heat than the overall project requires and does not use balanced heat recovery to use the available internal heat. This problem is most evident during mild weather on systems not designed to take full advantage of internally generated heat year-round.

The following are two examples of situations that can be shown in a heat balance study:

A properly applied heat reclaim system automatically responds to make a balanced heat recovery. An example is a reciprocating water chiller with a hot-gas diverting valve and both a water-cooled and an air-cooled condenser. Hot gas from the compressor is rejected to the water-cooled condenser. This hot water provides internal heat as long as it is needed. At a predetermined temperature, the hot gas is diverted to the air-cooled condenser, rejecting excess heat from the total building system. Larger projects with centrifugal compressors use double-condenser chiller units, which are available from many manufacturers. For typical buildings, chillers normally provide hot water for space heating at 105 to 110°F.

Many buildings that run chillers all or most of the year reclaim some of the condenser heat to provide domestic hot water.

Designers should include a source of external heat for back-up. The control system should ensure that back-up heat is not injected unless all internal heat has been used. For example, if electric back-up coils are in series with hot-water coils fed from a hot-water storage tank, they may automatically start when the system restarts after the building temperature has dropped to a night low-limit setting. An adjustable time delay in the control circuit gives the stored hot water time to warm the building before energizing the electric heat.

This type of heat reclaim system is readily adaptable to smaller projects using a reciprocating chiller with numerous air terminal units or a common multizone air handler. The multizone air handler should have individual zone duct heating coils and controls arranged to prevent simultaneous heating and cooling in the same zone.

Properly applied heat reclaim systems not only meet all space heating needs, but also provide hot water required for showers, food service facilities, and reheat in conjunction with dehumidification cycles.

Heat reclaim chillers or heat pumps should not be used with air-handling systems that have modulating damper economizer control. This free cooling may result in a higher annual operating cost than a minimum fresh air system with a heat reclaim chiller. Careful study shows whether the economizer cycle violates the heat balance concept.

Heat reclaim chillers or heat pumps are available in many sizes and configurations. Combinations include (1) centrifugal, reciprocating, and screw compressors; (2) single- and double-bundle condensers; (3) cascade design for higher temperatures (up to 220°F); and (4) air- or water-cooled, or both.

The designer can make the best selection after both a heating and cooling load calculation and a preliminary economic analysis, and with an understanding of the building, processes, operating patterns, and available energy sources.

ASHRAE research project RP-620 (Cane et al. 1993) evaluated computer models for heat recovery chillers. Whole-building energy use predictions were 5 to 15% of actual monitored data for one building used in the evaluation. Heat recovery chiller estimates were within 10 to 15% of measured values in the same building.

Applications of heat reclaim chillers or heat pumps range from simple systems with few control modes to complex systems having many control modes and incorporating two- or four-pipe circulating systems. Some systems using double- and single-bundle condensers coupled with exterior closed-circuit coolers have been patented. Potential patent infringements should be checked early in the planning stage.

Successful heat recovery design depends on the performance of the total system, not just the chiller or heat pump. Careful, thorough analysis is often time-consuming and requires more design time than a nonrecovery system. The balanced heat recovery concept should guide all phases of planning and design, and the effects of economic compromise should be studied. There may be little difference between the initial (installed) cost of a heat recovery system and a nonrecovery system, especially in larger projects. Also, in view of energy costs, life-cycle analysis usually shows dramatic savings when using balanced heat recovery.

A multiple-building complex is particularly suited to heat recovery. Variations in occupancy and functions provide an abundance of heat sources and uses. Applying the balanced heat concept to a large multibuilding complex can save substantial energy. Each building captures its own total heat by interchange. Heat rejected from one building could possibly heat adjacent buildings.