A Critical Look at Cold Supply Air Systems.
The following are the primary advantages of a VAV delivery system using colder SAT:
1. A greater differential between SAT and desired room temperature ([DELTA]T) on the airside means lower airflow is needed to be delivered to each room for the same cooling effect, and lower airflow means lower fan energy. In fact, a reduction in building fan energy is typically the primary justification given by those who promote colder SAT systems.
2. Colder SAT can reduce supply air duct size, air-handling unit physical size, and (perhaps) ceiling space, saving building first cost. The reduction in duct and/or ceiling cavity size may be especially advantageous when adding ductwork to an older building with smaller floor-to-floor heights constructed before air-conditioning became commonplace.
3. Greater moisture removal associated with colder SAT (discussed further in the first disadvantage below) results in lower indoor relative humidity, making it possible in some cases to set the room temperature setpoint up a degree or two higher and still keep people comfortable. According to the well-known Comfort Zone graph found in Standard 55,3 a person wearing 1.0 clo can be comfortable at 75[degrees]F (23.9[degrees]C) if the space relative humidity is 60%, but can be comfortable at 76[degrees]F (24.4[degrees]C) if the space relative humidity is 50%. In actual practice, I am not convinced that buildings using colder SAT actually implement a warmer space temperature setpoint, but if they did, there would be an associated modest annual energy savings.
4. Cold SAT may be complementary to systems using ice-based or low-temperature liquid-based thermal storage, if thermal storage is advantageous for other reasons such as time-shifting of peak electrical demand. (A thermal storage system analysis is a separate topic that can't reasonably be addressed within the scope of this article.)
5. Lower fan airflow associated with colder SAT might make the HVAC system quieter.
Disadvantages of Colder Supply Air
The advantages of colder SAT discussed above are primarily centered on reducing fan energy or system first cost. However, fan energy is not the only energy consideration when thinking about colder SAT. The following are some of the inherent disadvantages of using colder SAT in a VAV system:
1. Cold SAT condenses more moisture at the cooling coil, meaning a greater latent cooling load is forced onto the coil and cooling system when dehumidifying outdoor air. This can be a significant energy penalty because each pound of additional water vapor condensed represents an additional 970 Btus (1 MJ). More latent heat removal not only means more chiller/compressor energy, but also more chilled water and condenser water pumping energy, cooling tower fan energy (where applicable), and a higher first cost for the central cooling plant itself.
2. Colder SAT reduces the number of hours per year that a full airside economy cycle is in effect. For example, if we are using 55[degrees]F (13[degrees]C) SAT and it happens to be 55[degrees]F (13[degrees]C) outside, we can use 100% outdoor air without any mechanical cooling. However, if we are using 48[degrees]F (9[degrees]C) SAT and it happens to be 55[degrees]F (13[degrees]C) outside, some mechanical cooling remains necessary.
In some climates, this range can represent a substantial number of hours annually, making the building's annual energy use higher. In my hometown of St. Louis, the temperature is in the range of 50[degrees]F to 54[degrees]F (10[degrees]C to 12[degrees]C) 577 hours annually. In Boston that figure is 766 annual hours; in Seattle, 1,440 hours; and in San Francisco, 2,267 hours per year fall between 50[degrees]F and 54[degrees]F (10[degrees]C to 12[degrees]C)--over 25% of the year! (4)
3. Colder SAT increases reheat energy, because colder SAT causes a lightly loaded room to over-cool faster and need reheat sooner, once the VAV box has reached its minimum allowable airflow. The code-required zone minimum airflow (the point at which reheat may commence) is generally the same regardless of SAT, and reheat of colder air obviously uses more energy than reheat of conventional supply air. Reheat can represent a serious energy cost penalty depending on heating type and heating fuel source.
Reheat can be mitigated somewhat by implementing SAT reset, but the same is true for the conventional SAT system. It is likely that at least some lightly loaded interior-zone VAV boxes may reach their minimum damper position even in the summer when SAT reset is not available. If other interior zones are not impacted by weather or sunshine, and with a nearly constant year-round cooling demand, SAT reset may not be effective in that case either.
4. Colder SAT may require colder chilled water, depending on how aggressive your cold air target is. Typical chillers making colder chilled water generally use more energy on a kW/ton (COP) basis.
5. Colder supply air ducts are more prone to condensation or "sweating" on bare sheet metal surfaces (and eventually risk of mold growth on ceilings) if ductwork is not perfectly and flawlessly insulated, particularly in humid climates, and particularly upon startup after setback for a night or weekend or where high internal moisture loads exist (such as in kitchens). If using colder SAT in these cases, it is essential that your specifications regarding duct insulation be well-written and even better enforced. Similarly, air-handling units (AHUs) producing colder supply air may be more expensive than conventional AHUs if "thermal break" or "no through-metal" specification clauses are needed to avoid condensation, dripping, and puddling in unconditioned or less-conditioned mechanical rooms.
6. Diffuser selection can become more challenging, and available diffuser choices more limited, to avoid "dumping" of cold air and to ensure a high ADPI (air diffusion performance index) in the space.
7. If attempting cold SAT with a direct-expansion (DX) system, there is a greater chance of building ice on the DX evaporator coil due to reduced airflow per unit of refrigeration, and more moisture being condensed on the coil surface.
Analysis: Fan Energy Savings of Colder SAT
At peak summer conditions, fan energy savings (Advantage 1) for colder SAT is obvious. But VAV systems don't operate at 100% load very often. This fan energy differential may become overstated on an annual basis, depending on how many hours are spent at part load.
Consider a simple example system whose peak cooling load requires 5,000 cfm (2360 L/s) of conventional SAT (at or around 55[degrees]F [13[degrees]C]) but needs only 3,700 cfm (1750 L/s) of colder SAT (at or around 48[degrees]F [9[degrees]C]) to achieve the same space sensible cooling. Let's further suppose that the duct designer or project criteria has chosen a maximum allowable velocity of 2,500 fpm (12.7 m/s) and a maximum allowable duct friction loss rate of 0.3 in. w.g. per 100 ft (2.5 Pa/m).
The conventional system sized this way will need a 22 in. (560 mm) round duct, but a colder SAT system will only need an 18 in. (460 mm) round duct to meet that same criteria. It is logical to assume that the designer of the colder SAT system will similarly take advantage of the reduced airflow to downsize air-handling equipment, filter banks, and so forth, per Advantage 2.
Also, let's suppose for this example that the code-minimum * allowable outdoor airflow for this system happens to be 1,500 cfm (708 L/s), which would generally be the same minimum regardless of SAT. The total outdoor air change rate for a building typically depends upon how many people and how many square feet (square meters) of floor area are in that building. Neither the floor area nor the number of people change when using cold versus conventional SAT, so neither should the minimum outdoor air change rate in absolute terms.
So in this example, the conventional SAT system modulates between 1,500 cfm (708 L/s) and 5,000 cfm (2360 L/s); the colder SAT system modulates between 1.500 cfm (708 L/s) and 3,700 cfm (1750 L/s). At the top end of that range, the cold SAT system is clearly using less fan energy based on lesser airflow alone, even if the two systems are sized for essentially equal friction.
However, at the bottom of that range, the colder air supply fan clearly has to work harder to push 1,500 cfm (708 L/s) through an 18 in. (460 mm) duct, smaller filter bank, etc., sized for colder SAT, versus pushing the same 1.500 cfm (708 L/s) through a 22 in. (560 mm) duct, larger filter bank, etc., sized for the conventional system. So somewhere in that modulating range, the two fan systems actually cross each other in terms of energy use.
More importantly, at the low end of the airflow modulating range, we are talking about a small amount of fan energy. The key point is that when you are operating at off-peak airflow, the significance of the fan energy advantage of colder SAT diminishes quickly because fan energy falls off exponentially, making fan energy a smaller and smaller fraction of the overall building energy use, while some other forms of HVAC energy use don't fall off exponentially. The latent energy penalty at the cooling coil is closer to constant for several months of the year in humid climates.
The conventional system cools outdoor air to 55[degrees]F (13[degrees]C) saturated, which is about 64.6 grains of moisture per pound of dry air (142.4 g/kg) [humidity ratio of 0.00923] and an enthalpy of about 23.2 Btu/lb (53.9 kJ/kg); while the colder air system cools outdoor air to 48[degrees]F (9[degrees]C) saturated, which is about 49.7 grains (109.6 g/kg) [humidity ratio 0.0071] and an enthalpy of about 19.2 Btu/lb (44.6 kJ/kg). The sensible load associated with the two systems is the same, but the latent load on the conventional system is less than the colder air system based on enthalpy differential.
For the fan energy differential to dominate the latent energy differential, the system fan static pressure would have to be high; or the system would have to operate at near-peak airflow a large number of hours; or the system would have to be located in a very dry climate where the ambient summer absolute humidity is below about 55 grains (121.3 g/kg) [humidity ratio 0.0079], which is not the case in any of the 'A' climate zones5 (2A, 3A, 4A, 5A, etc.) in the eastern two-thirds of the United States.
Other disadvantages of colder SAT, such as reduced economizer hours and greater reheat, tend to exacerbate the situation. But these impacts are too difficult to manually analyze on an annual basis with any degree of accuracy; even the fan energy differential, to be accurate, needs an energy model reflecting an actual duct system total static pressure and an actual annual load profile to evaluate the true fan energy yearly differential.
Analysis: Energy Modeling
Many of the popular green building rating systems require a full-year energy analysis for a proposed building. So it was an easy task for this author to call up three energy models already programmed for recent projects and simply run them for various SAT setpoints. The energy model software used is a popular commercial branded product that has been tested and validated in compliance with ASHRAE Standard 140-2014.6 Among other features, it allows the energy modeler to specify a SAT, or if none is specified, it will compute the warmest SAT that still meets the space humidity setpoint requirement.
Project 1 is a military dining hall facility located in Dublin, Calif. (Alameda County, Climate Zone 3C) of 32,000 [ft.sup.2] (3000 [m.sup.2]) in a single story. Project 2 is a medical clinic located in Killeen, Texas (Bell County, Climate Zone 2A) of 32,000 [ft.sup.2] (3000 [m.sup.2]) in a single story. Project 3 is also a medical clinic located just outside Columbus, Ga. (Muscogee County, Climate Zone 3A) of 19,400 [ft.sup.2] (1800 [m.sup.2]) in a single story. All three buildings are served by gas-fired condensing boilers, air-cooled scroll chiller, VAV air-handling system(s) with hydronic heating and cooling coils, and zone VAV boxes with hydronic reheat. All three use an airside economizer and none of them include thermal storage.
These three projects were chosen to represent three different climate zones and were selected for their somewhat smaller size, making multiple 8,760-hour runs not too time-consuming. (Larger projects would simply feature more or larger air-handling units, and may include a more-efficient water-cooled chiller plant, but the concepts are similar on a larger scale.) SAT reset was enabled, as was duct static pressure setpoint reset. VAV box control sequence was dual-maximum, meaning that the VAV damper closes from its cooling maximum to its minimum as the space temperature is satisfied, but begins to remodulate open to a heating maximum when reheat is necessary. Each project was modeled using a 55[degrees]F (13[degrees]C) SAT and a 48[degrees]F (9[degrees]C) SAT. The results appear in Table 1.
In all three projects studied, conventional SAT is the better choice. Fan energy savings offered by use of colder supply air does not offset energy increases attributable to increased latent loads and reheat energy. In each case, reheat energy increased significantly, as did cooling energy. While fan energy is indeed lower with the colder supply air, we in our office (see Acknowledgments) were surprised about the small magnitude of the difference.
After several checks and reruns to make sure the fan energy was modeled correctly, we concluded that not only does part-load diminish the fan energy differential as discussed in the previous subsection, but that the dual-maximum VAV zone control strategy may result in significant duration where cfm at reheat is actually greater for the colder air than for conventional air.
Certainly, three projects are far too small a sample size to prove a concept across the board. There are scholarly papers, e.g., Bhatia, (7) and well-respected proponents of colder SAT whose energy models are more favorable to cold SAT than are mine. But I will remain rather skeptical about claims that the fan energy savings of colder supply air more than make up for higher chiller/compressor energy use, fewer annual hours of economizer use, and increased reheat energy. The typical project seems more likely to achieve lower overall energy use by using the warmest SAT that still maintains the space humidity setpoint. This is perhaps becoming even more true with dual-maximum VAV box damper controls, which tend to decrease fan energy savings of colder supply air when in heating mode.
For colder supply air to make sense, it seems you would need a project in which fan energy dominates, duct static pressure losses are high, dehumidification loads are low, and/or if you employ cold thermal storage for peak-shifting or similar reasons. If you do use a colder-air system, it is critical to incorporate an intelligent SAT reset algorithm to minimize reheat during off-peak conditions and maximize economizer hours; and use energy modeling as accurately as possible to understand the overall energy difference on an annual basis.
The author gratefully acknowledges the assistance of Annelise M. Smith, Associate Member ASHRAE, Daniel P. Phelan, P.E., Associate Member ASHRAE, and Matthew D. Fisher, P.E., Member ASHRAE, in the energy modeling described in this article.
(1.) Murphy, J. 2011. "High-performance VAV Systems." ASHRAE Journal (10).
(2.) Taylor, S., et al. 2012. "Dual maximum VAV box control logic." ASHRAE Journal (12).
(3.) ANSI/ASHRAE Standard 55-2013, Thermal Environmental Conditions for Human Occupancy, Figure 5.3.1.
(4.) NCDC. 2016. "Engineering Weather Data." National Climatic Data Center, Climate Services Division. http://tinyurl.com/ gmgdul5.
(5.) ANSI/ASHRAE Standard 90.1-2013, Energy Standard for Buildings Except Low-Rise Residential Buildings, Appendix B, Figure B1-1.
(6.) ANSI/ASHRAE Standard 140-2014, Standard Method of Test for the Evaluation of Building Energy Analysis Computer Programs.
(7.) Bhatia, A. 2012. "HVAC Optimization with Cold Air Distribution." Continuing Education and Development, Inc.
BY STEPHEN W. DUDA, P.E., BEAP, HBDP, HFDP, FELLOW ASHRAE
Stephen W. Duda, P.E., is senior mechanical engineer at Ross & Baruzzini, Inc. in St. Louis.
* For example, International Mechanical Code-2015, Section 403.
TABLE 1 Energy modeling results summary. PROJECT 1 1 2 2 3 SUPPLY AIR TEMPERATURE ([degrees]F) 55 48 55 48 55 HEAT/REHEAT (MWH/YR) 302.6 372.9 208.2 480.7 159.9 COOLING PLANT (MWH/YR) 93.7 130.5 125.4 176.6 214.9 FAN ENERGY (MWH/YR) 166.6 162.9 150.3 148.9 158.9 NET TOTAL (MWH/YR) 562.9 666.3 483.9 806.2 533.7 PROJECT 3 SUPPLY AIR TEMPERATURE ([degrees]F) 48 HEAT/REHEAT (MWH/YR) 249.6 COOLING PLANT (MWH/YR) 256.6 FAN ENERGY (MWH/YR) 135.5 NET TOTAL (MWH/YR) 641.7
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|Title Annotation:||COLUMN: ENGINEER'S NOTEBOOK|
|Author:||Duda, Stephen W.|
|Date:||Dec 1, 2016|
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