Reference Publication: McIlvaine, J., Mallette, M., Parker, D., Callahan, M., Lapujade, P., Floyd, D., Schrum, L., Stedman, T., Cumming, B., Maxwell, L., Salamon, M., "Energy-Efficient Design for Florida Educational Facilities," Prepared for the Florida Department of Education, Tallahassee, FL., September, 2000.
Disclaimer: The views and opinions expressed in this article are solely those of the authors and are not intended to represent the views and opinions of the Florida Solar Energy Center.
Energy-Efficient Design for Florida Educational Facilities
Janet McIlvaine, Michele Mallette, Danny Parker, Michael
Callahan, Philippe Lapujade, David Floyd, Lynn Schrum,
Ted Stedman, Brian Cumming, Larry Maxwell, Milt Salamon
Solar Energy Center (FSEC), R. Douglas Stone Associates, Inc.,
Spacecoast Architects, Technical Editor
Section III Continued
Objective: Provide superior conditioned environment with low energy consumption.
Considerations: Selection and design of heating, ventilating, and cooling (HVAC) systems embodies complex relationships, many of which relate directly to the comfort and well being of occupants. When dealing with humidity control and ventilation, health must always take precedent over energy issues. In addition to the urgency of indoor air quality needs, engineers are faced with a wide variety of options.
Issues and ECM Options:
a. Dehumidification Technologies
b. Ventilation To Offset Cooling Requirements
4. Cooling Systems
a. Central Chillers
b. Condensing Systems
c. Packaged Terminal Air Conditioners
d. Unitary Heat Pump
e. Incremental Heat Pump
f. Room Air Conditioners
g. Rooftop Units
5. Air Distribution Systems
a. Constant Air Volume
b. Variable Temperature Constant Volume
c. Variable Air Volume
d. Low Temperature Air Systems
e. Dual Duct
f. Duct Energy Losses
g. Air-and-Water Systems
a. Cooling Systems
HVAC systems for educational facilities should be designed considering that various portions of educational facilities may be used according to varying seasonal schedules. This, in turn, may suggest adaptation of HVAC equipment selection to meet the differing annual schedules. For instance, libraries generally require year round space conditioning and effective humidity removal even during vacant periods. On the other hand, the schedule of classrooms may be quite seasonal. Other spaces, such as auditoriums are only intermittently occupied while school administration facilities often operate for a month or more beyond the school year. Since it is not desirable to operate an entire large chiller plant during the summer to solely serve a library or administrative facility, secondary supplemental HVAC systems should be considered in planning for educational facilities used beyond the regular school year.
Another fundamental issue is the decision between large scale chiller systems and unitary equipment. Until recently, the decision has always been to go with chillers except in the smallest facilities. However, analysis of data taken in the 1996 Survey of Florida Educational Facilities showed that schools using unitary (DX) systems often have lower energy use than those relying on chillers. It should be kept in mind that the IPLV curves for chillers are somewhat misleading to consulting mechanical engineers. Although a screw chiller's performance may suggest an EER of 16 or more, but this is considerably degraded once pumps, fans, controls and cooling towers are added. Often such combinations end up similar in performance to unitary equipment after these 'hidden' energy uses are considered. If large amounts of outside air are provided, dedicated dehumidification systems may be appropriate to control interior humidity levels.
Also, unitary equipment provides the considerable advantages of allowing individual zoning and temperature preference within the school system. This may reduce the complaints of individual teachers. Also, facility consumption may be reduced during night education operation when only a limited number of classrooms may be in use. Although maintenance may be more frequent than with chillers, the systems are more amenable to operation by school facilities personnel who often have skills well suited to such equipment. Also, the survey data seemed also to indicate that older chillers are not well maintained within the educational system.
Recently, ASHRAE has interpreted the 15 cfm per student requirement to allow the use of average classroom occupancy rates, rather than design occupancy with the provision that the ventilation rate never be reduced below 50% of the recommended rate (7.5 cfm). Research has shown that ventilation rates of 15 cfm per student are necessary to hold interior CO2 levels below 1,000 ppm (Downing and Bayer, 1993). However, the 1000 ppm level has no health significances. The importance of this level of ventilation and its role in altering indoor air quality is contentious. Moreover, Florida schools have experienced a continuing problem with increases in illnesses and indoor air quality (IAQ) complaints from students and staff. Although reduction of pollutant source strength is fundamental towards addressing this issue, increased levels of ventilation are often seen as desirable to reduce remaining interior pollutant concentrations. However, in hot and humid climates such as Florida, the introduction of this increased fraction of outside air into the building, coupled with high internal latent loads from students, may lead to unacceptable relative humidities if improperly designed and operated.
Although the ASHRAE outdoor air ventilation minimums can be widely observed in office buildings, applications in classrooms have been limited. In addition, higher occupant densities in schools require increases in the outdoor air requirement by up to 30% (Wheeler, 1991). This leads to increased HVAC capacity requirements to meet the increased load, as well as greater difficulties in achieving acceptable interior humidity levels. Although increased ventilation rates will increase energy consumption during the times when the building is occupied and the ventilation is required, proper control of the introduction of outside air depending on space occupancy can significantly reduce the energy-related costs. Many previously designed Florida facilities drew in inadequate amounts of outside ventilation air during occupied periods, but had no option for reduction of ventilation during unoccupied periods.
The use of ventilation controls, based on CO2 sensors can be a very effective method of providing for occupancy based ventilation control. This is often useful in portions of the facility characterized by high occupancy density for relatively short periods of time. Such areas include auditoriums, gynasiums, cafeterias and classrooms. Under CO2 sensor control, these areas receive the design ventilation rate during periods of occupancy, but reduced outside air during unoccupied periods. According to the current interpretation of ASHRAE 62-1989 the floor during unoccupied periods could be 50% of the design ventilation rate. A Florida engineering firm specializing in IAQ mitigation provides a minimum ventilation rate of unoccupied spaces of 0.14 cfm per square foot of conditioned floor area when no special pollutant sources are recognized.
Outside air added to the conditioned space obviously must be exhausted. In general, it is advantageous to design for systems where the exhaust air is approximately 90% of the outside air intake to maintain the building at positive pressure (Morawa, 1993). This ensures that most air entering the building is preconditioned. However, exhaust air fans must be interlocked to associated air handlers and placed on independent schedules to assure that buildings remain at positive pressures.
With all mechanical ventilation systems, care should be taken that supply air is properly filtered. Outside air should be drawn through a mesh filter with supply air passed through a medium/high efficiency filter prior to introduction to the conditioned space. Return air should also be filtered. One Florida engineering firm recommends that a minimum of six air changes per hour be designed into the ventilation rate to properly filter a room. Particular care should be exercised with VAV or other systems where the air flows to spaces are modulated to be certain that face velocities are sufficient to provide effective filtration.
In addition to increased energy use, control of indoor relative humidity (RH) becomes a challenging design problem. It is established that the lack of proper humidity control can lead to increased illness and IAQ complaints (Arundel et al., 1986). ASHRAE Standard 55-92 defines that space humidities should not exceed 60% RH at any temperature.(5) However, this is a difficult objective in the hot and humid climate of Florida. There are numerous strategies to controlling relative humidities while admitting more ventilation air. They are:
a. Dehumidification Technologies
Conventional Systems: Conventional AC systems primarily control the temperature and not the humidity level. As a first step, effective humidity removal using packaged and unitary equipment or fan coils critically depends on not over-sizing the equipment for the load. Another important innovation is the use of variable speed indoor air handlers which offer better humidity removal characteristics. Newer Energy Management Systems (EMS) allow explicit monitoring and control of space relative humidity. These should be considered particularly where interior moisture levels are likely to be of concern (i.e. with higher ventilation rates). However, it should be recognized that EMS controls alone cannot mitigate humidity concerns since humidistat control without means of reheat may over-cool the conditioned space. Even so, there are several operationally related suggestions that can help to reduce moisture levels with conventional systems:
Reduce supply air flow: Reducing supply air flow in direct-expansion (DX) systems will decrease air coil velocity and result in increased moisture removal. It will also, however, somewhat reduce system relative efficiency. This can be partly offset by utilization of multi-speed blowers which can save electricity otherwise used for fan power. However, reducing the air flow to DX coils should only be done with electronic and pressure/temperature regulated coils. DX coils using capillary tube control for metering refrigerant flow are not capable of providing predictable control over a wide range of operating conditions. Also, it should be kept in mind that lowering the air flow may affect air filtration and the adequacy of ventilation air supplied to conditioned spaces.
Lower chilled water temperature: Lowering the chilled water supply temperature will increase the relative moisture removal although it will also reduce system chiller efficiency and may require increased need for supply air reheat.
Reheat: Adding sensible (non-latent) heat to a school's interior will increase the fraction of time that the cooling system operates and removes moisture. Previously, the most commonly used strategy to control humidity was to use electric reheat coils to increase the compressor run-time of thermostatically controlled constant volume systems or with Variable Air Volume (VAV) systems when the minimum cfm ratio is reached. Humidistats are often used with constant volume reheat systems. However, using electric reheat is very energy-intensive and must be avoided. In any case, the use of reheat should be minimized by the use of controls because it increases space cooling energy use. Since Florida state law prohibits electric reheat, other reheat sources must be used with hydronic coils. These can include natural gas, solar hot water and reclaimed condenser heat.
Heat Recovery: With chiller systems, two condenser tube bundles can be used to capture some of the rejected heat for heating the water loop used for reheat. One of the condenser tube bundles is piped to the cooling tower with the other piped to the building hot water circuit. The chiller then operates as a straight chiller when there is no call for heat. The other bundle operates as a heat pump to reject heat to the heating circuit up to its maximum capacity. Supplementary heat for the water loop can then be provided by a natural gas boiler.
Hot-gas Bypass: Some packaged systems have provision for hot-gas bypass. At low cooling loads, a valve routes compressed hot refrigerant gas back to the compressor suction inlet, creating lower flow though the main refrigerant loop. This allows the compressor to provide reduced cooling at low loads even though there is no reduction in electrical demand. Additional moisture is removed since the compressor on-cycle time is increased, although at the cost of cooling system performance. Thus, a 10 EER packaged unit operating with 50% hot gas bypass would drop the energy efficiency ratio to only 5 Btu/W. Also, hot gas bypass does not provide additional dehumidification unless the air flow on the coil is reduced proportionately. This means that hot gas bypass will be completely ineffective for constant volume systems. In general hot-gas bypass is inefficient and should be avoided.
Central Fresh Air Units: Outside air is drawn into a central unit which uses a direct expansion (DX) vapor compression machine or chilled water with a low sensible heat ratio (SHR) to remove humidity from the moist air prior to its introduction into the building's interior (these units may be a special case where hot-gas bypass is acceptable). The relative efficiency of the DX equipment may be lower than the normal cooling equipment, but saves energy since the introduced outside air is dehumidified, greatly reducing the need for reheat of supply air under low-load conditions. Generally, DX units are more effective at removing moisture than chilled water systems. Other methods can be used with fresh air units: run around coils, face and by pass dampers, desiccant dehumidification, total energy recovery systems and heat pipes.
Desiccant Dehumidification: Conventional HVAC systems remove humidity by passing the supply air across a cooling coil maintained at a low enough temperature so that water vapor will condense on the coils as liquid. With desiccant dehumidification, a stream of warm dry air from the desiccant dryer is mixed with the cool moist air from the conventional cooling coil. Properly regulated, the combined air is then within the desired comfort range in terms of temperature and relative humidity. Desiccants dry air by condensation and adsorption. In the process of adsorption the latent heat is converted into sensible heat, converting warm moist air to hot dry air. Desiccant dryers generally use solid-desiccant rotating wheel heat exchangers. Silica gel is the main solid desiccant material. Condenser heat, natural gas, solar or other heat sources can be used to regenerate the desiccants. Such a system is often used as a central fresh air unit instead of a low SHR DX system. The installed cost of desiccant dehumidification equipment is approximately $1,400 per ton (R.S. Means, 1992) or $10 per cfm based on a change in outside air enthalpy of 20 Btu/lb. Other than cost, the extensive maintenance requirements and high regenerative energy and added cooling energy expense for desiccant dehumidifiers may make them a poor choice for use with educational facilities.
Total Energy Recovery System (TERS): A variation on desiccant dehumidification system is the TERS or enthalpy recovery system. The TERS consists of an air-to-air heat exchanger with a rotating-wheel heat exchanger assembly (Figure 38). The wheel is coated with a molecular sieve desiccant coating to provide both sensible and latent heat recovery. The desiccants remove moisture from the supply air stream while also cooling the incoming air with the exhaust air stream. The increased outdoor supply air is matched by a nearly similar level of exhaust air. Generally, the supply air from the desiccant wheel is further conditioned with additional DX or chilled water coils prior to being added to the interior space.
Figure 38. Total energy recovery system.
A TERS has the significant advantage of producing real energy savings through the transfer of sensible and latent heat between the exhaust and supply air streams. A realistic heat recovery rate is approximately 50%-75% for both sensible and latent heat depending on the target space temperature and relative humidity. The cost is approximately $400-$1,000/ton installed. TERS is mainly applicable for new buildings. Retrofit may be difficult due to required ducting. Maintenance costs are also likely to be elevated, a fact that should be considered in specifying such systems.
Run Around Coils: Another variation on the heat pipe is the use of "run around coils," usually containing a water-glycol mixture with finned heat exchangers to transfer heat from the return side of the cooling coil to the supply air and thus avoid the need for reheat. Run around coils have the advantage of being more easily retrofit onto existing systems where there may be a fairly large separation between the return and supply air sides of the cooling system. However, unlike passive heat pipes (described below), this configuration require some pumping energy to move the liquid from one coil to the other. Another limitation, similar to that of heat pipes, is that although run-around coils enhance dehumidification under full load conditions, they do not provide improvement to systems with no part load performance.
Heat Pipe Dehumidification: Heat pipes offer an attractive alternative to reheat with other heat sources while greatly increasing the moisture removal capacity of conventional DX and chilled water systems. Unlike, TERS, heat pipes can not reduce the cooling load, but only serve to increase the dehumidification potential. A heat pipe is a refrigerant-charged device with a heat exchanger at either end. One end of the coil removes heat from the incoming air stream as the refrigerant in that end is evaporated (Figure 39). The conventional cooling coil (containing chilled water or the DX evaporator) then has less sensible-cooling load and runs colder, removing more moisture from the pre-cooled air. The heat pipe refrigerant gas then migrates freely to the other end of the coil where it condenses, giving up heat. The condensed refrigerant then passively returns to the evaporator section by gravity or capillary action. The bypassed heat is transported by the heat pipe, condensing at its other end and reheating the chilled air without requiring any additional energy. Because the refrigerant in the heat pipe flows passively in a loop between the pre-cooler coil and the reheater coil, it requires no external power and has no moving parts. The added cost of heat pipe dehumidification equipment is approximately $400 per installed ton of cooling capacity, but provides no reduction in energy use (Dinh, 1992). Heat pipes become less effective as the approach temperature decreases and may not provide adequate reheat during part load operation.
Face and Bypass Dampers: Facilities using chilled water and constant volume systems can provide face and bypass dampers to reduce the need for reheat while providing effective humidity control. With such a configuration air dampers channel a portion of the return air (which would normally pass through the cooling coil) around the coiling coil (Figure 40). This results in a suitably high supply air temperature along with effective dehumidification. The system uses a draw-through air handling unit to take advantage of the fan heat as added reheat. A mixing box is used to add building air for ventilation and room pressure control. The face and bypass section has cross-linked dampers that give it the ability to pass all the air through the coil or divert any needed amount around the unit. The air passage around the coil is engineered to have the same pressure drop as the coil so the total discharge air volume does not change. When dehumidification is needed the cooling coil temperature is lowered to remove additional moisture; a portion of the entering air bypasses the cooling coil to hold the discharge temperature needed to maintain the required room temperature. The result is improved temperature and humidity conditions with less need for reheat.
Monitoring of face and bypass controlled chilled water systems in schools shows that classroom humidities can generally be controlled to around 60% RH except during low-load periods when supplemental supply air reheat will be necessary to avoid higher humidity levels. Low loads also prevail during after-hours operation when face and bypass control can lead to high moisture conditions under constant fan operation. However, through the use of CO2 sensors with an energy management system, face and bypass dampers can control the humidity below 50% without reheat, by changing the routine to fan cycling during after hours periods.
One improvement on the overall concept is to use dual face and bypass dampers, one controlling the space return air and the other to pre-conditioning the outside air. This provides better control under low load conditions. By controlling the outside air with face and bypass dampers as the first stage of the system, better control of interior humidity should be possible with the need for reheat obviated except during the most adverse part-load conditions.
b. Ventilation To Offset Cooling Requirements
Natural Ventilation: Natural ventilation was commonplace in Florida schools in the 1950s prior to the advent of mechanical air conditioning. It has become less common in recent times. However, natural ventilation is an option to mechanical air conditioning as per the Florida Code. Detailed guidelines have been developed showing how educational facilities can be naturally ventilated (Chandra, 1985). However, if the building is mechanically conditioned there is no specific requirement for natural ventilation. Whether natural ventilation is advisable in mechanically conditioned educational facilities is a subject of debate. Generally, the high humidity levels in Florida's outdoor environment serves to exact very stringent limits on the utility of natural ventilation if interior moisture levels are of concern. Also, our detailed survey of Florida schools found that perceptions of poor indoor air quality are strongly linked to perceived humidity problems.
For this project, we conducted an analysis using Typical Meteorological Year (TMY) data to examine when natural ventilation might be useful to offset mechanical cooling. We examined the weather data for Orlando from September to May between 8 AM and 4 PM to establish the number of hours when ventilation might be feasible for educational facilities. Of the 2,184 hours during this period, some 906 hours (41% of the time) had an outside air temperature less than 72oF -- a seemingly promising value. However, if outside relative humidity is considered, there are only 408 hours (19% of the time) when the temperature is less than 72oF and the outside humidity is less than 65%. Realistically, a classroom environment with its heavy internal heat production will require outside air at 68oF or below to produce feasible natural cooling.
Disregarding humidity, we found 29% of hours were below 68oF, while only 12% of total hours were below 68oF and had outside relative humidity less than 65%. Moreover, although these values would seem to suggest that 12 - 19% of the air conditioning season might be eliminated by natural ventilation, a much smaller fraction of the overall space cooling energy would be eliminated, since the hours when ventilation is feasible are precisely those times when building loads are low and cooling system efficiency is highest.
Although not attractive as the main building cooling system, the capability of having operable windows for natural ventilation does have other potentially important justifications. In some circumstances it may be very desirable to extensively ventilate a space. Such situations include the need to fully ventilate during cleaning or painting operations or other times when a complete purge of the room air is needed. Also, the availability of operable windows does provide some cooling and ventilation during times when the space conditioning system is inoperable or in the process of undergoing repair or replacement. Against these advantages, however, a facilities planner must weigh the increased first cost of operable windows as well as minimizing potential for vandalism or unauthorized access.
Economizer Cycles: An economizer cycle is a mechanical version of natural ventilation and can be used with single and dual duct ventilation systems. Economizers consist of three sets of dampers with linked controls. An exhaust damper relieves system return air to offset ventilation air brought in. An outside air damper controls the quantity of ventilation air brought into the system, and a return damper balances the return and outside air portions of the economizer. At low temperatures (below 60oF) the economizer dampers adjust to the minimum ventilation setting. This reduces the cooling load while providing necessary ventilation air. At high ambient temperatures (above 68oF) the dampers return to this minimum position to provide for ventilation requirements. Between these temperatures the economizer dampers modulate from minimum ventilation air to 100% outside air to meet cooling requirements. Integrated economizers allow simultaneous economizer and mechanical cooling; non-integrated economizers do not.
Economizers can be set to operate solely based on outside temperature (operate when the temperature of the outside air is below the temperature of the return air). Enthalpy economizers compare the enthalpy of the outside air and the return air stream to determine which air has the lower heat content. Enthalpy economizers should always be specified if considered in Florida's humid climate. We simulated the economizer cycle to be activated at a time when the temperature fell below 68oF and relative humidity was lower than 55% (internal latent heat gain would drive space levels to approximately 60%). Unfortunately, the same limitations which exist for natural ventilation also hold true for the economizer cycle. Generally, the humidity is too high and hours within acceptable temperature limits too short for such systems to be cost effective in Florida's conditions.
Ceiling Fans: The use of ceiling fans to provide air motion and improve comfort and allow higher thermostat set-point is well established. Research suggests that air motion permits comfort at temperatures from 2 - 6oF higher than would be possible without ceiling fans. Indeed, our survey and analysis of utility records in Florida found evidence of lower energy use in classrooms using ceiling fans (Callahan et al. 1997). For our analysis we chose conservative assumptions. We assumed that use of ceiling fans would allow a 2oF increase in the thermostat setting in educational facilities. However, we also increased the building equipment loads to reflect a 40-watt average electricity use for each fan. Ceiling fans were analyzed as an option for the classroom and administrative buildings. Their use was found unpractical in the multi-use facility. Our analysis found moderate savings in the administrative facility, but very substantial savings potential in the classroom building. Regardless of these results, we do not necessarily endorse the use of ceiling fans in Florida educational facilities since fans may create problems: the effects of air motion on papers, and the potential distraction from moving shadows caused by fan blades moving below recessed ceiling luminaires. There have even been reports of students experiencing motion sickness from the strobing effect of the fan blades across the lighting fixtures. There are also concerns associated with vandalism.
|45 15 CFM per Person
46 TERS 5 CFM per Person
47 TERS 15 CFM per Person (compared to #45)
48 Energy Management System
49 Enthalpy Economizer
50 Economizer Cycle
51 Reheat Constant Volume
Figure 41. Ventilation and dehumidification options.
Furthermore, for the savings to be realized, the thermostat must be set higher. Since the occupants often do not control the air temperature setting, the theoretically available savings may not be achieved. It is noteworthy, however, that project staff members have observed ceiling fans in use in some older Florida schools. It is thus possible that ceiling fans may be appropriate in some situations.
a. Central Chillers
Central chilled water systems are generally larger than standard split or packaged DX systems and may be used to provide cooling for a large educational facility. These systems typically consist of a central reciprocating, centrifugal scroll or screw chiller to cool water which is then distributed by the appropriate fan coil or air handling system. Chiller efficiency is often rated at kW/ton of cooling at standard conditions (evaporator=40oF, condenser=105oF, suction gas =55oF, no sub-cooling). Most schools will need chillers in the 50 - 300 ton range depending on facility size and characteristics. In choosing a chiller, select a unit with the best cooling efficiency and lowest life cycle cost. This is often given in kW/ton of cooling or alternately as cooling Coefficient of Performance (COP), (IPLV), or Energy Efficiency Ratio (EER). Look for chillers with the lowest kW/ton and conversely with the highest COPs, IPLVs, and EERs.(6) As pointed out earlier, however, it is important to keep in mind that air handlers, pumps and controls will reduce the comparative cooling performance of central chillers relative to unitary equipment. Particularly, when considering larger equipment, consider the part-load performance of cooling systems, since the facility will seldom be operating under full load conditions. A retrofit project at Fellsmere Elementary in Indian River County demonstrated a 175 kWh/day (19%) reduction in HVAC energy use with the new, more efficient reciprocating and screw chillers.
Reciprocating Chillers: Reciprocating compressors compress refrigerant using pistons that are driven directly through a connecting rod from the drive crankshaft. This is the default chiller type used in the simulations. There are two main types: open and hermetic reciprocating chillers. In hermetic systems, the electric motor and compressor encapsulated in the refrigerant stream, whereas the motor is external in the open-drive configuration. The open-drive types are generally preferred due to their greater efficiency since the motor heat is not released into the refrigerant loop. One advantage to reciprocating chillers is that they are available in small sizes. Open-drive types are also amenable to use of gas or diesel powered engines to drive the compressors. Reciprocating chiller sizes vary from 5 tons to over 300 tons. Reciprocating chillers generally have lower efficiency than most other types. In these sizes a typical reciprocating chiller would have an efficiency of approximately 0.90 kW/ton or a COP of approximately 3.8 (ASHRAE, 1991). They are also noisy and cause vibration. Consider scroll compressors as an alternative.
Centrifugal Chillers: Centrifugal chillers are a common larger chiller type. They often have very high efficiencies (COP~5.0) but generally have lower part-load efficiencies. Centrifugal compressors are usually large; sizes typically range from 100-11,500 tons. Systems are available in open-drive types which are amenable to use with mechanical drives. Electric, gas or diesel engines may be used to drive such open-drive chillers. The best centrifugal chillers have efficiencies approaching 0.55 kW/ton (COP= 6.4) at full load.
Screw Chiller: Helical rotary or screw chillers generally have a slightly lower efficiency (COP ~5) compared with centrifugal chillers, but have a higher part-load efficiency at low cooling load ratios. They are also available in smaller sizes from approximately 75 tons to 750 tons. The best helical rotary chillers have efficiencies approaching 0.60 kW/ton (COP = 5.9). Due to their excellent part-load performance, screw compressors are often configured in pairs with each unit able to meet the building load under full-load conditions. Units are then operated in a staged fashion, either singly or in combination so that they operate near optimum part-load efficiency. Such a configuration also imparts greater redundancy to the central chiller system since a single chiller can meet load while the other unit is serviced. Screw chillers are characteristically noisy and require acoustical isolation from occupied spaces and surrounding neighborhoods.
Absorption System: This is a special case where most of the cooling energy is really provided by the heating plant. Unlike an electric chiller, a chemical process is used to vaporize and condense the refrigerant. Thermal energy from a gas burner is used in the generator where high-pressure refrigerant is liberated from the absorber (Figure 42). The high-pressure refrigerant is then cooled and condensed in the condenser, where the heat is rejected to a cooling tower. The condensed refrigerant passes to the evaporator, where the concentrated absorbent causes some of the refrigerant to evaporate which in turn cools the chilled water in the evaporator coil. The efficiency of absorption systems is fairly low, although the fact that they use a low-cost source of energy may make them competitive. For instance, an absorption chiller will generally use 60% more source energy than a centrifugal chiller. Absorption chillers are water-cooled, typically using a lithium bromide/water or water/ammonia cycle. They are available in several sub-types, although direct-fired systems are most popular. They are more compact that conventional chillers, have few moving parts and produce less noise than conventional systems. Absorption systems should be considered for large central cooling systems where there is a large difference between the applicable price for natural gas and electricity.
Figure 42. Gas fired absorption chiller.
Gas-powered Chillers: Natural gas-powered engines can be used as the drive for various type of open reciprocating, centrifugal and screw chillers. According to the ASHRAE Systems Handbook, the specific gas consumption for natural gas engines is from 8 - 13 cubic feet of natural gas per horsepower hour (ASHRAE, 1992). Since one horsepower equals 0.746 kW, such engines will use 6 - 10 cubic feet per kW of input power to chillers. The natural gas consumption per ton of cooling can then be estimated by knowledge of the specific chiller performance. For instance, a gas engine driving a screw chiller producing 0.70 kW/ton of cooling would require 4.2 to 7 cubic feet of gas to produce a ton of cooling depending on the engine characteristics. It should be noted, however, that the performance efficiency curve for gas-driven engines is fairly flat down to approximately 40% of the design capacity. Below this point, the efficiency drops off sharply where cylinders are unloaded to further reduce capacity below idle speed. Also, first cost and maintenance of natural gas-fired engines are both generally greater than with electrical drives.
b. Condensing Systems
Cooling Towers: With cooling towers, water is cooled by contact with the air to reject heat from the water-cooled condensers of the air-conditioning systems. Either natural draft or mechanical draft systems perform the cooling. Mechanical draft towers are most common because they do not depend on wind to function properly. Towers are usually specified in multiples to provide redundancy, ease of maintenance and the capability to be run at reduced capacity where their efficiency is highest. If considering a central chilled water system that will use a cooling tower, the designer should ask project engineers for an incremental analysis of the efficiency improvements available from increasing the size of the projected cooling tower. Increasing the cooling tower size will reduce heat exchanger approach temperatures and improve chiller system efficiency.
Air Cooled Condensers: Air-cooled condensers pass outdoor air over a dry coil to condense the refrigerant. This results in a higher condensing temperature and lower performance under peak conditions. It is preferred, however, for packaged systems and unitary heat pumps due to its simplicity and low maintenance requirements. It should be avoided with larger central chilled water systems since chiller efficiency is impaired (see Figure 43 for chiller and air distribution options).
|35 Centrifugal Chiller
36 Screw Chiller
37 Gas Absorption Chiller
38 Ceiling Fans Vol.
39 Variable Speed Pumps
|40 Non-Variable Speed Fans
41 Variable Temperature Constant Volume
42 Multizone Constant Volume
43 Dual Duct Constant Volume or Variable
44 Four Pipe Fan Coil
Figure 43. Chiller and air distribution options.
Evaporative Condensers: Evaporative condensers pass air over coils sprayed with water taking advantage of the latent heat of vaporization to reduce the condenser temperature.
These condensing systems are the most efficient, although such systems increase water use and have extensive maintenance requirements. As with, water cooled cooling towers, these systems must have freeze protection and close control of water treatment to function successfully. However, because of their greater efficiency, projects should consider their use.
c. Packaged Terminal Air Conditioners
A packaged terminal air conditioning (PTAC) system is a combination of a cooling unit (using air-cooled direct expansion with integral compressor, evaporator and condenser), an optional heating unit (usually electric resistance or "strip heat") and an optional minimum ventilation intake. Zone temperature is controlled by off-cycling of the unit. Packaged heat pumps are also used, and are generally identical to the PTACs although with greater heating energy efficiency. Such packaged systems are very popular in older Florida schools and for cooling manufactured classroom buildings (portables). However, they offer unacceptable humidity control at a ventilation rate of 15 cfm/person if fans are run continuously while the compressor cycles. If specified, higher efficiency units (SEER, EER) should be chosen while selecting the lowest SHR (sensible heat ratios) possible.
If multiple units are used to serve a space, their operation should be staged to provide better humidity control. In general, PTACs, unitary heat pumps, and roof-top units will not provide sufficient moisture removal to achieve humidities less than 60% even at 5 cfm per person. FSEC and engineering firms in Florida have often seen that packaged AC systems are significantly oversized for the sensible loads in Florida educational facilities. Along with their poor part load performance, such oversizing has grave consequences for effective humidity removal. In general, units should never be sized larger than that indicated by approved sizing methods.
d. Unitary Heat Pump
Multiple packaged unit heat pumps are justifiably popular in Florida schools. Each packaged heat pump has a self-contained direct-expansion cooling system, an outside condenser, and associated controls (Figure 44). They have the advantages of being relatively low cost, modular in installation, with simple controls and maintenance. Also, when air handlers and pumping are considered, unitary heat pumps are not generally as efficient as central chiller systems with VAV air distribution. However, they will often be considerably more efficient than constant-volume systems, with smaller inefficient reciprocating compressors. More specifically, these units will not control humidity if the fan remains on while the compressor cycles. However, newer variable speed air handlers offer both higher efficiency, quiet operation and improved humidity control. These systems are recommended.
Figure 44. Heat pump in heating mode (top) and cooling mode (bottom).
A variation on the packaged system is the split direct-expansion system where the evaporator and condenser components are separated. Otherwise, performance and other considerations are very similar to the packaged heat pump. There is some latitude, however, in specification of specific fan coil units for match with given condensing units. In any case it is important the equipment not be oversized since effective humidity removal with such system will critically depend on extended run-time fraction. Also, when several through-the-wall packaged units are used for a building, it is preferable to run the fewest units that will satisfy the building load rather than to run all units. Running all units in an unstaged fashion will lead to short-duty cycles, lower efficiency and ineffective humidity removal. Generally, equipment should be chosen with the greatest seasonal energy-efficiency-ratio (EER) for cooling and heating season performance factor (HSPF) for heating. SEERs should be greater than 10.0 Btu/W; HSPF should be greater than 7.0 Btu/W. For a given efficiency, choose a unit with the lowest sensible heat ratio (SHR) to provide better humidity removal.
e. Incremental Heat Pump
This system consists of water-to-air heat pumps in each zone interconnected by a water loop, which is sometimes in line with a large storage tank (Figure 45). The temperature of the water loop is kept within a limited range by a combination of a cooling tower and a heating plant. The heating plant is usually a boiler, though any other heating plant could be used. When all zones require cooling, the heat pumps reject heat to the water loop, which in turn transports the heat to a cooling tower where it is rejected to the atmosphere. Depending on the size of the storage tank and the duration and intensity of the heating demand, the water temperature may drop sufficiently to require heating from the boiler. During intermediate seasons, zones in the heating mode use the heat rejected by other zones in the cooling mode. The specific zone temperature is controlled by the on-off cycling of the unit. Additionally, the water loop can be used to produce reheat when needed.
Figure 45. Heat pumps in each space are connected by water loop.
f. Room Air Conditioners
Room air conditioners are packaged systems with a combined compressor, condenser and evaporated all in one unit. These have often been used in older schools or as add-on units to condition individual classrooms. Cooling capacities typically vary from 5,500 Btu/hour all the way up to 20,000 Btu/hour. The units' energy efficiency ratios (EER or Btu/W of input power at the ARI Standard Rating condition) vary considerably from one unit to the next, from 8.0 Btu/W all the way up to 12.6. Chosen units should have EERs greater than 10.0 Btu/W.
g. Rooftop Units
Roof-top units (RTUs) are a packaged form of a Variable Temperature Constant Volume (VTCV) system. They have been popular with some school boards due to their low expense and the ease with which they can be added to buildings in a modular fashion. However, many school districts prohibit their use due to maintenance issues. Sizes vary from 6 to 20 tons of capacity but their efficiency is often low. Units are commonly rated at their EER at standard conditions. Chosen units should have EERs greater than 9.5 Btu/W with the lowest possible SHRs. A buyer's guide is available from E-Source (303-440-8500) that lists the most efficient packaged roof-top systems by their cooling capacity. Dehumidification performance is poor; particularly under part load conditions. However, to enhance moisture removal, these units should have dual split-row coils with the lead coil on bottom for part-load dehumidification.
Although there are numerous variations, there are two primary types of air distribution systems for non-packaged units: constant air volume distribution systems and variable air volume systems. Within these types, single-duct systems tend to be less costly and less expensive than dual-duct and multizone systems, but reheat is generally required. VAV systems are generally more efficient.
a. Constant Air Volume Systems
A Constant Volume Terminal Reheat (CVTR) system cools a constant flow of mixed air to design minimum cold supply conditions, sometimes modified by an outside air reset or a discriminator. Reheat coils near each zone air diffuser provide zone temperature control. This system is simple and inexpensive to install but it is expensive to operate due to the excessive cooling of the design and because so much of the heating energy is required for reheating cold supply air. Proper operation of this system requires that heating and cooling equipment be energized simultaneously for all but the most extreme seasons. Generally, outdoor air reset or discriminator control should not be used in Florida's humid conditions (see Figure 46 for packaged options).
|52 Unitary heat pumps
53 Packaged single zone variable temp DX unit
54 Packaged multizone DX unit
55 Packaged terminal AC/heat pump
Figure 46. Packaged HVAC options.
b. Variable Temperature Constant Volume (VTCV)
The variable Temperature Constant Volume (VTCV) system cools or heats a constant flow of air. Zone temperature control is achieved by modulating cooling or heating coils sequentially (i.e., never at the same time as with a CVTR system). A central forced-air residential system with a gas furnace and a central air conditioner is a very simple version of a VTCV system. However, because it provides only heating or cooling and only at the cumulative rate demanded by all zones, the VTCV system is relatively energy efficient, but provides poor temperature control for individual zones. The return air thermostat of a VTCV system responds only to the aggregate needs of the building. If more zones require heating than cooling, heat is supplied. Dehumidification performance can be very poor; these systems should be avoided in Florida's climate.
c. Variable Air Volume Systems
A Variable Air Volume (VAVS) (Figure 47) controls temperature within a space by varying the quantity of supply air rather than the supply air temperature. However, as the cooling demand in a zone decreases, the air flow to it is reduced to a "minimum fraction" typically on the order of 15 - 30% of the full flow rate. The minimum fraction is generally the lowest acceptable air flow to limit maximum humidity and provide required ventilation air. Reheat occurs only when this point is reached and is generally lower than with single-zone reheat systems. As the heating load increases, the air flow may increase to meet the load. Compared to a terminal reheat system, a VAV system saves energy in three ways: (1) reheat energy consumption is minimized; (2) fan energy consumption is decreased at low volumes; (3) cooling coil consumption decreases significantly as less volume of mixed air must be cooled. Energy is saved with the VAV system by reduced need for simultaneous heating and cooling of spaces as with reheat, as well as reduced fan power. Dampered VAV boxes should be preferred over fan-powered boxes which typically have inefficient motors. One useful case for fan-powered boxes, however, is with a Power Induction Unit. A power induction unit is used with a VAV system to keep zone air temperatures at a sufficiently high level without paying the price of higher energy consumption associated with higher air volumes and reheat. As system air volume is decreased to the minimum stop at lower cooling loads, a fan between the VAV box and the zone air inlet (the power induction unit) starts adding return plenum air to the supply air. Heat from light fixtures rejected to the return air plenum is thus recirculated into the zone. Conventional reheat is added only when this recirculated air is inadequate to maintain zone temperature. However, plenum return air systems are very energy inefficient and can lead to conditions that promote air quality problems. They should be avoided in favor of ducted return systems.
Figure 47. Variable Air Volume System (VAVS) refers to an HVAC system which varies the amount of conditioned air supplied to a zone in accordance with the need in that zone.
There are several potential problems to note, however, with VAV systems:
1. Reduced air flows at lower part loads
can provide low air turnover rates and filtration for the space.
2. VAV systems can fail to effectively dehumidify spaces at low part loads. Theoretically, when the students are in the classroom and the system reduces the air flow under low load conditions, the effective air dehumidification is reduced while the latent loads in the room remain fixed. Resetting the supply air temperatures can further aggravate the problem.
3. Fan powered boxes, can lack effective filtration equipment which can load the ductwork downstream with dust and dirt leading to potential IAQ problems.
The fundamental response to these potential problems lies in insuring that the minimum stop ("minimum fraction") for the VAV boxes is no less than the minimum ventilation rate for the space. For classrooms, this can be a much higher level (~50% of the design flow rate) than commonly specified for less densely occupied spaces, such as offices. When lower flows are required to control temperature, reheat coils should then be used to control space conditions (see Dehumidification Technologies).
Effective and efficient filtration equipment is a must for the overall air handling system. In general, fan powered VAV boxes should be avoided, both because of their lack of suitable filtration as well as their added impact on energy use. Obviously, the test and balance of VAV systems to insure adequate air flows and ventilation to building spaces is vital to successful performance.
Also, with VAV systems, proper specification of air diffusers is vital to avoid "dead zones" from poor air circulation in rooms. In constant volume and particularly in VAV systems, it is critical to specify louvered type ceiling diffusers or sidewall grilles that meet the ASHRAE recommended throw ratios at design conditions. These types of diffusers offer excellent part load throws that maintain good air motion. Do not oversize air devices in VAV systems. If oversized, they will operate at lower than design airflows during most periods.
d. Low Temperature Air Systems
A recent development and potentially controversial innovation for air distribution systems is the use of low-temperature air. Instead of 55oF air, air at a temperature of approximately 40oF is circulated in the system. The low-temperature air system allows the use of considerably smaller duct work in air handlers. Chiller efficiency may be considerably reduced due to the low operating temperature of the water-glycol solution, however, such systems require less fan power and provide enhanced dehumidification due to the low evaporator or chilled water temperatures involved. Low-temperature air systems are often used in concert with thermal storage systems, but can be utilized in conventional configurations. First costs can be reduced because of the lower expense of the smaller motors, fans and duct work. However, high-induction air diffusers and well insulated ducts are necessary with low-temperature air systems to ensure that diffuser condensation and localized overcooling does not become a problem.
From a practical stand point low temperature air systems may be viable for administrative and library buildings. However, at 15 cfm per student, low temperature air becomes a less viable alternative for classroom buildings due to the high concentration of occupants. Typically, classroom spaces have about 30 occupants and require between 900 and 1200 cfm of 55o air for cooling. At the new ventilation rates, the space would require 450 cfm of outside air which is between 40 and 50% of the total cooling supply air at 55oF. If the supply air were reduced to 40oF, the typical classroom air volume would be between 514 cfm and 685 cfm while the ventilation airflow would remain at 450 cfm, or between 66 and 88% of the total airflow. Cooling humid outside air to these low temperatures may be impractical from a energy and operational standpoint.
However, the described limitation is obviously affected by the required ventilation rate as well as the potential to reduce the fraction of outside air to total airflow through the use of a central fresh air unit for the overall facility.
Lastly, and most importantly, it should be pointed out that schools using such systems in our survey showed both the highest normalized energy use and costs of any group analyzed (Callahan et al. 1997). These systems were also associated with increased complaints of IAQ and moisture problems. This suggests that without careful engineering, commission and maintenance they may not be a good match for educational facilities.
e. Dual-Duct Systems
Single-duct systems are typically more energy efficient. The dual-duct and multizone systems described below should be generally avoided:
Dual-Duct and Multizone System: The Dual Duct System (DDS) has central supply and return fans and two sets of ducts, one with a cooling coil, the cold deck, and the other with a heating coil, the hot deck. Mixing boxes are located in each zone for supply air temperature control while the flow of air to each zone is fixed. At maximum cooling demand, only air from the cold deck is supplied to the zone. As cooling demand drops, more air from the hot deck is added. Similarly, during peak heating needs, only air from the hot deck is added. The Multizone (MTZ) system is identical to the Dual Duct System except that the mixing boxes are located at the central air handling unit with only one duct per zone to carry supply air. The hot deck can carry preconditioned ventilation air at a constant supply flow. Ventilation air temperature can then be modulated to maintain the temperature in the most critical space.
Dual-Duct Variable Air Volume: A Dual-Duct Variable Air Volume (DDVAV) system can be thought of as a dual-duct system with variable volume central fans and variable volume mixing boxes at each zone. At maximum cooling demand the DDVAV system supplies the maximum volume of cold air to the zone. As cooling demand diminishes, this volume is reduced until reaching the "minimum stop." As the zone begins to need heating, the hot deck starts supplying air volumes at gradually increasing rates up to the maximum design air volume.
f. Duct Energy Losses
Energy waste due to air leakage in duct work and terminal devices can be considerable. All air handling systems and associated duct work should be carefully sealed. ASHRAE (1991) estimates that ductwork installed in many commercial buildings can have leakage rates of 20% or more. Projects should strive to:
Ensure that all system air handling equipment and duct work is located within the insulated building envelope.
Specify in the project construction specifications that the duct work for the building be commissioned prior to occupancy and pressure tested to yield air leakage no greater than 30 cfm at 50 Pa pressure per 1,000 square feet of conditioned floor area.
Also, to the extent possible duct systems should be installed inside the conditioned space to prevent low delivery efficiency.
g. Air-and-Water Systems
Fan Coil Systems: Fan coil (FC) systems consist of zonal fan units connected by one or two water loops with circulation pumps to a central cooling and heating plant. Optionally, fan coil units can provide minimum outside air for ventilation, but are not particularly appropriate in humid conditions. Two-pipe FC units can provide heating or cooling through the same pair of pipes, but not at the same time. The change from the heating to the cooling mode is made seasonally. These two-pipe units are not recommended for Florida's climate. A four-pipe FC unit can provide heating and cooling simultaneously. Each individual zone can be either in heating or cooling mode, but not both. Neighboring zones may be in a different mode, however. Although initially expensive, fan coil systems provide high performance levels since the water loops more effectively transfer heat than air. Also, pumping power requirements for the working fluid is generally lower than that for the fans in air handling systems. However, the zone-located fan coil units also tend to have greater operating sound levels than conventional air systems; this should be carefully evaluated for the specific application. Additionally, when these units need service, maintenance personnel may have to disrupt class or work time. As with packaged systems, proper sizing (never oversized) is critical to the dehumidification performance of fan coil systems.
There are two basic types of control used with fan coil systems:
1. Modulating fan coil: The chilled water flow is throttled depending on the space cooling needs. However, at a reduction of 40% of the water flow, some 55% of the dehumidification potential at full flow is lost. Even small reductions to chilled water flow can severely reduce humidity removal potential.
2. On/Off coil: The coil operates like a standard PTAC air conditioner. The coil is only powered when the thermostat calls for cooling. However, again, under part load conditions the effective moisture removal of the system will be degraded due to the pronounced off-cycle times.
Small fan coil units seldom have sufficient latent capacity to handle the high latent conditions in classrooms. Classroom unit ventilators, a higher quality form of fan coil, have an option for face and bypass control which does provide more effective humidity removal. However, all such fan coil systems, should be used only with dedicated outside air systems that precondition 100% of ventilation air.
Variable speed drives can provide important efficiency improvements to variable air volume systems by varying fan motor speeds rather than using variable inlet vanes or inlet cones with constant speed systems to modulate delivered air volume. Within unitary systems, they can also provide enhanced dehumidification. Variable speed fan motors are inverter-controlled so that fan speed is adjusted dynamically as VAV boxes call for changing air volume. Variable speed fan drives have the added advantage of limiting the current inrush for the start-up of large motors ("soft start" feature). Variable speed drives for VAV systems save energy and should be specified with such systems. In choosing variable speed fan drives, the designer should be aware that such inverter-controlled devices can introduce undesirable harmonics in the power supply. The designer should work with the drive manufacturer to ensure that such problems are minimized. The cost effectiveness of variable frequency (VF) drives vary with motor size and conditioning system use schedules. Generally, it is cost-effective to use VF drives for motors of 20 hp or more and inlet vanes for smaller units. Keep in mind that inlet vanes increase fan power by 5% simply due to their presence.
Optimal start is an option available with most energy management systems (EMS). At the beginning of the day, the energy management system determines the time at which the cooling or heating system should be activated to bring the building to comfort conditions by the time it is occupied. The time that chiller, boilers and fans are activated is based on a calculation within the EMS based on outside and inside temperatures and historic data (which the EMS accumulates) to determine the best time to activate the conditioning system. Several indoor and outdoor sensors are used. Optimal start saves energy by reducing the system operation to the minimum time necessary to provide comfort conditions. More sophisticated systems allow the building to remain below the heating set-point in the morning hours if it is anticipated that the building will soon need cooling by mid-day. Our analysis generally showed optimum start to be a desirable design feature for Florida educational facilities. Most new educational facilities will have an energy management system. Specification of an EMS with optimal start capability is desirable since paybacks should be very short.
Limited temperature controls should be provided for classrooms. Activity levels in the classroom, particularly for elementary schools, vary to the extent that some level of individual control is desirable. In addition, students come to class attired differently in the various seasons, requiring different space temperatures for comfort.(7) Thermostats are now available, which give the occupant limited control range, allowing them to raise or lower their space temperature within pre-programmed temperature limits. The psychological effect of providing instructors with some degree of control is of great advantage in reducing comfort complaints. Additionally, computer classrooms, with a high saturation of electronic equipment, will require lower temperatures for effective comfort when the equipment is in use.
Thermostats specified for educational facilities should be of the electronic type. These respond much more rapidly to changes in temperature than the old bi-metallic thermostats and provide finer comfort control with less variation in interior conditions. One common oversight to be avoided: the location of control thermostats should be given careful consideration. In general, the devices should be located on interior walls away from direct air movement from diffusers. Since most thermostats operate under PID control, improperly located units may lead to over-controlled short cycling and associated lack of effective dehumidification.
One useful feature for EMS systems in schools is that of a pre-programmed thermostat override. Electronic thermostats are available with an override button on the face plate that allows the occupant to signal the EMS system that a space is occupied and needs conditioning during periods when the space is normally vacant. Thus, a teacher or janitorial staff, staying late can achieve temporary comfort without disabling the EMS's energy saving thermostat set-ups for long-term system operation.
One area that cannot be over emphasized with the EMS is the need to properly set-up and commission the equipment. Improper set-up or specification can lead to a lack of flexibility in control, disabling of valuable functions, under control and excess energy use. EMS functions for humidity control should be carefully tested and commissioned, since this is a common area for problems. The project bid specification should also include provision for the proper training of the maintenance personnel at the facility to insure understanding of the systems and ability to effectively use controls. Unnecessary features should be avoided, with the project mechanical engineer providing input to the level of EMS required for the facility rather than the financially motivated vendors.
These EMS systems have demonstrated significant savings in the past. The FLASTAR (Florida Alliance to Save Taxes And Resources) project at Fellsmere Elementary had an EMS system installed and this showed savings of 355 kWh/day (17%) after the retrofit and it had better temperature control and more consistent scheduling of equipment.
Variable speed pump drives can potentially reduce pump motor electricity consumption in large central chilled water systems with four or more cooling coils. This is accomplished by matching pump horsepower output relative to the actual flow requirements. A differential pressure sensor on the VAV chilled water line modulates the pump speed based on the pressure across a valve located two thirds of the distance from the pumps to the furthest coil. The valve includes a bypass line so that as the space loads decrease, the control valves on the coil close. The increasing pressuring across the bypass valve is used to slow the pump speed. The pump flow rates are throttled by varying the pump output with a frequency inverter so that it becomes a variable speed drive. Variable speed pumps can be used only on systems that use two-way control valves. Three way values should be placed at the ends of the system to minimize the time necessary to get chilled water to the active coils. Also, savings potential on systems with supply temperature reset controls is significantly reduced. Both chillers and boilers typically have minimum flow rates which limit the potential minimum turndown of variable speed pumps. Reduction of harmonic line distortion is an issue in choosing variable speed pumps as with all inverter-controlled drives.
The many HVAC options available to schools reduce to a few options which have superior performance. These are:
a. Cooling Systems:
Large facilities: Central chiller for facilities with total loads greater than 150 tons. Screw or centrifugal chillers should be specified with minimum kW/ton. Chiller drives can be electric or natural gas depending on relative price of fuels. Absorption chillers may also be considered based on a similar analysis. The decision between air, water-cooled and evaporative condensers should consider the appropriate trade-offs between first cost and system performance. With water-cooled condensers, cooling tower size should be subjected to a careful analysis. Primary/secondary pumping with variable speed pumps should be specified for chilled water systems.
Small or medium sized facilities: Select high-efficiency packaged or split systems for educational facilities with total loads less than 150 tons. Packaged VAV systems may also be considered. However, pre-treatment and dehumidification of introduced outside air is a necessity with packaged equipment. Select the highest system cooling COP, EER, IPLV, or SEER. For heat sources consider using heat pumps or natural gas with straight cooling systems. Hot-gas bypass and re-heat should be avoided. Improved dehumidification can be achieved by choosing variable speed air handlers for equipment within a given efficiency level and never oversizing packaged or unitary equipment used to serve space conditions. Dedicated dehumidification systems may be appropriate when high classroom ventilation rates are called for.
Air Handling: Variable air volume systems, four-pipe fan coil, or constant volume systems with face and bypass dampers should be specified for projects not using packaged systems. Variable speed fans should be used with VAV systems for larger motors. Outside air should be added to such systems with a central fresh air unit, preferably with heat recovery from the exhaust air stream either using heat pipes or a heat recovery ventilation system. Fan-powered VAV boxes should be avoided. If reheat is found necessary for low load operation. It should be provided by non-electric sources, either natural gas, solar or condenser heat recovery.
Duct systems should be well sealed and pressure tested prior to occupancy. The duct system should be located within the envelope insulation. The use of low-temperature air distribution systems should be considered early in the design process with consideration of feasibility against required ventilation rates. Systems should be controlled by an energy management system (EMS) with optimal start capability. CO2 sensor ventilation control should be considered for intermittently used facilities. Air handling equipment should be balanced and commissioned prior to acceptance.
ASHRAE Standard 62 recommendation of 15 cfm per person should be considered as the outside air ventilation rate for classrooms, modified by average versus design occupancy. The later is an important point as it will frequently allow lower ventilation. However, natural ventilation and economizer cycles were found to provide only very small savings benefits (<2% reduction in energy use) due to the high humidity levels in Florida and are likely not cost-effective. In any case, enthalpy economizers should be used in any installation considering the use of economizer cycles. The high first and maintenance costs should be carefully weighed before specifying.
However, other means of dehumidifying ventilation air is very effective: ERVs, heat pipes etc.
Dehumidification of outside ventilation air will be required if a 15 cfm/person ventilation rate is adopted for the facility. Electric or baseboard reheat must be avoided. Hot-gas bypass should also be avoided with packaged systems. A preferred alternative for packaged equipment is to choose high efficiency units with variable speed air handlers.
Optimum system design will precondition
the outside air prior to its being introduced to the indoor environment.
For this, a central fresh air unit is advisable using one of the following
technologies: a dedicated DX system, heat pipes, run-around coils, or a
total energy recovery system. Face and bypass dampers can be used with
constant volume systems. Second-stage air conditioning and humidity removal
of the outside air is performed prior to its introduction to the conditioned
environment, using the conventional system either with chilled water or
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Tennessee Valley Authority (TVA), and Sizemore Floyd, Architects and Energy Planners, 1985. Energy Design Guidelines for Schools, TVA, Division of Conservation and Energy Management, Commercial and Industrial Branch, Chattanooga, Tennessee, March 1985.
U.S. Congress, Office of Technology Assessment, Building Energy Efficiency, 1992, OTA-E-518, U.S. Government Printing Office, Washington, DC.
Various Issues, 1990-1992 American School and University. North American Publishing Company, Philadelphia, PA.
Various Issues, 1991-1993 Specifier Reports. Lighting Research Center, Rensselaer Polytechnic Institute, Troy, NY.
Wheeler, A.E., 1991. "Meeting the Objectives of Energy Conservation
and Ventilation for Acceptable Indoor Air Quality in the Classroom."
Indoor Air '91, Healthy Buildings, American Society of Heating,
Refrigeration and Air Conditioning Engineers, Atlanta, GA, p. 215-219.
Altitude -- The position of the sun in the sky relative to the horizon, in degrees. At sunrise and sunset, the solar altitude is zero degrees. If the sun were directly overhead (possible only in the tropical latitudes), the altitude would be 90 degrees.
Azimuth -- The position of the sun in the sky relative to true south, in degrees. At solar noon, the sun is at zero degrees azimuth, and changes 15 degrees every hour. This difference is generally expressed in positive numbers for both east and west variations from south. Thus, both 11 a.m. and 1 p.m. are said to have a solar azimuth of 15 degrees.
Ballast -- A device used with fluorescent and other types of gaseous discharge lamps to aid starting, limit current flow, and to provide voltage control at proper design levels. Can be magnetic or electronic.
British Thermal Unit (Btu) -- Equal to the amount of heat energy necessary to raise the temperature of one pound of water one degree Fahrenheit. One Btu is about equal to the amount of heat given off by a wooden match.
Coefficient of Heat Transmission (U-Value) -- The overall U-value is the amount of heat transmittal from air-to-air in one hour per square foot of wall, floor, roof or ceiling for a 1oF temperature difference between the inside air and the outside air of the wall, floor, roof or ceiling. The lower the U-value, the less heat is transferred.
Coefficient of Performance (COP) -- The ratio of useful refrigeration obtained to the net work input. The ratio of the heat absorbed in the condenser to the difference of the heat rejected in the condenser and the heat absorbed in the evaporator [i.e., Btu (out)/Btu (in)].
Cooling Load -- The amount of cooling (Btu/h) required to offset the rate of heat gain to the building at a steady- state condition when indoor and outdoor temperatures are at their selected design levels, solar gain is at its maximum for the building configuration and orientation, and heat gains due to equipment, infiltration, ventilation, lights, and people are present.
Degree Day -- The degree day value for any given day is the difference between 65oF and the mean daily temperature. Example: for a mean daily temperature of 50oF, the degree days are 65 minus 50, or 15 degree days.
Demand Charge -- Charge for electrical service based upon customer's demand. It is a charge computed separately and distinctly from the energy charge.
Efficacy -- See Luminous Efficacy.
Emissivity -- The ability of a material to radiate absorbed heat.
Energy Efficiency Ratio (EER) -- The ratio of net cooling capacity in Btu/h to total rate of electric input in watts under designated operating conditions.
Economizer -- Conditioning system in which outside air is used
to meet HVAC needs when it falls within required temperature and humidity
Exfiltration -- refers to conditioned air being pushed through small passages in the building envelope to the outside.
Fenestration -- The design and placement of windows in a building.
Footcandle (fc) -- Energy of light at a distance of one foot from standard candle.
Heat Capacity -- The capacity of a given amount of a substance to absorb and store heat while experiencing a given temperature change. There are three standard measurements for heat capacity: Specific Heat, which relates to mass (Btu/lboF); Volumetric Heat Capacity, which relates to volume (Btu/cu.ftoF); and Thermal Mass, which relates to a specific building element of known mass or volume (Btu/oF).
Heating Load -- The amount of heat instantaneously added or removed by the HVAC system. The rate of heat loss from the building at steady state conditions when the indoor and outdoor temperatures are at their selected design levels (design criteria). The heating load always includes infiltration and may include ventilation loss and heat gain credits for lights, equipment, and people.
Transfer -- The measurement of heat from hot spaces to cooler spaces.
There are three principal modes of heat transfer: Conduction, heat
moves through a solid; Convection, heat moves by motion of a fluid
or gas, usually air; and Radiation, heat moves from one body to
another by heat waves without heating the air between the bodies.
Infiltration -- refers to the outside air drawn through small passages in the building envelope to the interior.
Latent Heat -- The amount of heat necessary to change a give quantity of liquid to vapor at constant barometric pressure.
Lumen -- Unit of light energy or output (luminous flux).
Peak-Hour Billing -- A utility
billing scheme that rewards energy users who limit consumption during that
part of the day during which the utility experiences its greatest demand.
Generally, this is implemented by differential billing rates called demand
charges. By limiting peak demand on its generation systems, a utility can
forestall the construction of further generation capacity and operate its
present plants at a more constant and efficient rate.
Perimeter -- Spaces with at least one exterior wall, as opposed to core spaces which have no exterior.
Relative Humidity (RH) -- A measurement, expressed as percentage, indicating the amount of water vapor in the air compared to the amount that the air could contain if it were completely saturated with moisture at the same temperature and pressure.
Sensible Heat Ratio (SHR) -- For HVAC equipment, the SHR is the fraction of the unit's cooling output (Btu/hr or tons) which is sensible. The lower the number, the better the moisture removal capacity of the equipment.
Shading Coefficient (SC) -- The percentage of available full-spectrum solar radiation that passes through a transparent or translucent object. A simple single-pane window assembly might have a SC of 1.0; a reflective window might have a SC of 0.5. The shading coefficient is an important glazing feature to consider when attempting to minimize heat gain through windows.
Simple Payback (SPB) -- Time required for an investment to pay for itself. The cost of the ECM divided by the annual energy cost savings in $/year.
Solar Reflectivity -- The fraction of incident solar energy (0.28-2.8 microns) which is reflected at the surface of the material.
Thermal Bridge -- An area of lower heat flow resistance within a larger area of greater resistance. An example would be an uninsulated 4-inch water pipe in a stud wall filled with R-11 insulation. A great amount of heat can flow through the highly conductive metal and water, defeating much of the benefit of the insulation on both sides.
Thermal Energy Storage Systems -- A refrigeration or heating system that produces and stores a cold or hot sink using off-peak electricity to reduce cooling or heating requirements during on-peak times.
Thermal Resistance (R-Value) -- The total thermal resistance is equal to the reciprocal of the overall coefficient of heat transmission (U-value).
Ventilation Air -- That portion of supply air which comes from outside the building, as opposed to recirculated supply air.
Visible Light Transmission (VT) -- The percentage of available visual-spectrum light that passes through a transparent or translucent object. This number is generally similar to the shading coefficient for a given glazing type, but can be either higher or lower depending on the specific properties of the glass. For daylighting purposes, a glass with a higher VT than SC is desired because it admits visible light while blocking more of the solar heat gain.
Zoning -- The division of a building by expected or actual conditioning demand differences. For example, on a cold morning, the perimeter of a building may need heating, while at the same time the interior areas have built up heat from people and equipment and need cooling. The simplest zoning cases have five zones: north, east, south, west, and interior. Each zone is often treated differently in the scheduling and operation of its HVAC components.
5. One Florida engineering firm sees the 60% RH as an absolute upper limit. With relative humidities around 60% they found that active teachers required air temperatures below 74oF or reported feeling uncomfortably warm. In addition, classrooms with large amounts of electronic equipment also require lower thermostat temperatures due to localized warm spots around the computer equipment. Thus, maintenance of relative humidities below 55% may reduce occupant complaints and allow operation of systems at higher internal temperatures.
6. A chiller with a performance of 0.8 kW/ton would indicate an EER of 15.0 Btu/W (1/8 12,000 Btu / 3,413 Btu) and a COP of 4.4 (15.0/3.413 Btu/W).
7. A reasonable argument can be made that student dress codes should support lighter levels of dress during the warmer seasons to promote higher thermostat set temperatures.