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1. Introduction
Air motion may play an primary role in the thermal ease of man and beast. A breeze on a humid summer day may make a significant divergence to one’s thermal comfort. Recent schemes for bettering energy-efficiency in buildings try to take account of the cooling effects of air motion from natural ventilation. When the building envelope is closed for air conditioning, local air motion is kept under 40 ft/min. This ignores the option of increased air motion to reduce the cooling energy in air conditioned space. This paper explores prospects for saving energy by utilizing the effects of indoor air movement.
2. Cooling energy savings in air conditioned space from elevated air speed
The current edition of ANSI/ASHRAE Standard 55-2004 Thermal Environmental Conditions for Human Occupancy (ASHRAE, 2004), provides for fixed increments of summer thermostat temperature settings by increased local air speed. Figure 1 is derived from Figure 5.2.3 in the Standard 55-2004.
The curves of equivalent heat loss from the skin for combinings of operative temperature and air motion are referenced to the upper limit of the ease zone (PMV= +0.5). Limits of 160 fpm and 5.4ºF are set for sedentary activity, 1.0 to 1.3 met. Large person divergences in preferent air speed
requires that occupants have personal control of air speed in increments of 30 ft/min.
The Standard states that it is adequate for the purpose to interpolate among these curves. Air speed is more effective at offsetting increments in temperature when mean radiant temperature is dandier than the mean arid bulb air temperature.
It must be cited that there are two faults in Figure 5.2.3 of the Standard. The “18°C” must read “18°F” and there is a scaling error amidst the fpm and m/s scales.
Five distinguished curves are provided to accommodate temperature divergences of -18°F, -9°F, 0.0°F, +9°F, and +18°F among mean radiant temperature, tr , and mean arid bulb air temperature, ta. The writer fitted equations to the part of the curves fixed to sedentary action of 160 fpm and 5.4°F for 1.0 met to 1.3 met and 0.5 to 0.7 clo.
The writer likewise fitted equations to the share of the curves for action beyond the sedentary limits. Cooling effect limits for these equations fitted to curves in Figure 5.2.3 in the Standard 55-2004 were 300 fpm and 8°F.
2.1 Curve for tr – ta = 0.0 K
For tr – ta = 0.0°F, an air speed of 160 fpm permits a thermostat set point increase of 4.4°F limit for light sedentary action (1 to 1.3 met) and 0.5 to 0.7 clo.
V = 40 + 6.8”t 1.85 (1)
Where V is the mean relative air speed in fpm and ”t is the cooling effect in °F.
In most thermostatically controlled air conditioned spaces, wall, ceiling and floor surfaces temperatures are close to air temperature. That is tr – ta = 0°F. Conditions when tr — ta is not zero include spaces with poorly insulated windows, walls or ceilings where the outer surface is exposed to direct solar radiation or cold winter conditions.
2.2 Curve for tr – ta = +9°F
For tr – ta = +9°F an air speed of 160 fpm permits a thermostat set point increase of 5.4°F limit for light sedentary action (1 to 1.3 met) and 0.5 to 0.7 clo.
V = 40 + 1.26”t 2.85 (2)
Where V is the mean relative air speed in fpm and ”t is the cooling effect in °F.
2.3 Curve for tr – ta = +18°F
For tr – ta = +18°F an air speed of 126 fpm permits a thermostat set point increase of 5.4°F limit for light sedentary action (1 to 1.3 met) and 0.5 to 0.7 clo.
V = 40 + 1.28”t 2.7 (3)
3. Beyond Sedentary Activity limits
The Standard is not clear on constraints for the portions of the curves up to 89°Fand 300 fpm, beyond the limits set for sedentary activity. Studies have measured the cooling effect of air motion up to 600 fpm in warm climate conditions (Khedari et al, 2000, Tanabe and Kimura, 1994, and Scheatzie et al, 1989). Air motion higher than 160 fpm is applied in air conditioned gymnasia and buying goods malls to enlarge or increase cooling of occupants. The writer has fitted equations to the share of the curves for action beyond the sedentary limits
For tr – ta = 0.0°F an air speed of 300 fpm gives evidence of the thermostat set point increase could be 6.6°F at action levels higher than 1.3 met.
V = 40 + 2.52”t 2.5 (4)
Limits for Equation 4 are 160 fpm to 300 fpm and 4.4 F to 6.6 F
For tr – ta = +9ºF an air speed of 276 fpm permits a thermostat set point increase of 8ºF at action levels higher than 1.3 met.
V = 40 + 5.7”t 1.8 (5)
Limits to Equation 5 are 160 fpm to 280 fpm and 5.4ºF to 8ºF .
For tr – ta = +18ºF an air speed of 211 fpm gives evidence of the thermostat set point increase could be 8ºF at action levels higher than 1.3 met.
V = 40 + 6.3”t 1.59 (6)
Limits for Equation 6 are 132 fpm to 209 fpm and 5.48ºF to 8ºF.
4. Estimating Cooling Energy Savings
The electrical US utility corporation Exeloncorp (2005), proposes that domestic air conditioning cooling costs may be scaled down by 3% to 4% for each ºF that the thermostat setting is raised in summer.
Occupants may offset an increased thermostat setting of 4.7ºF by supplying 160 fpm of low-cost air flow from circulator fans and receive pleasure from normal ease while saving air conditioning operating cost. On the basis of the Exeloncorp (2005) recommendation, an increase in the thermostat setting of 4.7ºF would provide cooling energy savings from 14% to 19%. In gymnasia where higher air motion is satisfactory the savings from a thermostat increase of 8ºF could be from 24% to 32%. A elaborated analysis of reduction in residential cooling loads due to air flow was performed for six US cities in a potpourri of climate zones (Byrne and Huang, 1986)
5. Comparison of fans and room air conditioners
A elaborated comparison of the energy required to maintain the same thermal ease in a 141.5 ft2 bedroom in Townsville, Hope (2003), was conducted using a 55 inch diameter residential ceiling fan and a VF100C Carrier window/wall room air conditioner, sized for the room by engineers at the local distributor. The measured rate of power consumption of a 55 inch diameter ceiling fan operating at it is top speed was 0.068kW or 0.48 W/ft2 of floor area. This is 8.7% of the power used by the room air conditioner to achieve the same thermal comfort. The rate of power consumption of the window/wall room air conditioner was 0.78 kW, or 5.51 W/ft2 of floor area. This is 11.5 times the power applied by the ceiling fan.
6. Destratification
In heated spaces in winter, indoor air have a tendancy to stratify with the hottest, less dense, air accumulating under the roof due to the gravity force. This condition brings about two problems. Firstly the hottest air is not contributing to the thermal ease of occupants near floor level, and secondly, it produces a high temperature divergence amidst the undersurface of the roof and the exterior of the roof that increments heat losses through the roof.
Destratification is the routine of exhaustively mixing indoor so that air temperature near the floor is the same as the air temperature beneath the roof, or no more than 2ºF difference. This is done using circulator fans. In a typical US distribution warehouse with a 30 ft high ceiling, the seasonal heating energy savings from effective destratification is around 20% to 30%. To be effective in regards to one half of the total volume of air in the space needs to be moved from ceiling level to floor level per hour.
To be effective in destratification the fan must be no more than 1 diameter beneath the ceiling and the jet from the fan ought to affect on the floor in order to achieve effective circulation. Jets from ceiling fans have an effective throw of 5 to 6 diameters.
In huge buildings with high ceilings such as churches, industrial buildings or distribution warehouses, a huge volume of air needs to be circulated. In order to refrain from complaints of drafts from occupants, the local air velocity at head height needs to be held less than 40 ft/min.
Circulator fans are much more energy-efficient at low speeds, so big diameter, slow moving, fans are well suitable for destratification. One 24 ft diameter industrial ceiling fan operating at top speed of 42 rpm uses 1.67 kW of electrical power but only 0.06 kW operating at 14 rpm it is peak efficiency. At 42 rpm this fan delivers around 337,700 cfm of air and 76,670 cfm at 14 rpm. An added gain of operating huge fans at low speed equated to littler fans at higher speeds is the reduction in fan noise. Large slow moving fans are nearly silent.
7. Estimating Destratification Energy Savings
A commended method for estimating heating energy savings from destratification is to determine the lumped seasonal heat transfer rate for the building envelope and determine the divergence in heat loss before and after destratification (Pignet and Saxena, 2002).
The lumped seasonal heat transfer rate for the building envelope in Watts may be calculated using:
A x U = qbd / (ti -to) (7)
Where: A is the surface area of the building envelope in ft2; U is the lumped heat transfer coefficient for the building envelope in Btu/ft2.h.ºF; qbd is the rate of heat loss through the building envelope in Btu/h before destratification; and ti -to is the intermediate heating season indoor to outdoor air temperature divergence in ºF.
The total heat lost from the building is the sum of heat freed from furnaces plus heat freed in the space from other origins such as lighting, people, machinery or devising processes. The heat freed from the furnaces may be determined from the fuel bills for the season, the caloric value of the heating fuel and the scheme efficiency. The caloric value of natural gas is around 1000 Btu/ft3. The time used in these calculations is the heating season related with the measured fuel consumption.
Forced air furnaces with flues have efficiencies around 0.7. Radiant heaters without flues have an efficacy of 0.8. Electrical heaters have an efficacy of 1.0. Heat from other roots is approximated in the normal way as set out in HVAC handbooks (ASHRAE, 2005).
With the overall heat loss U x A for the heating season before destratification determined, the reduction in heating after destratification, qad may be determined from:
qad = U x A x (tibd – tiad) (8)
Where: qad = Reduced heat load after destratification in Btu/hr; U = Lumped time-averaged heat loss rate for the building envelope in Btu/hr.ft2.ºF ; A = Surface area of the building envelope, ft2; tibd = Heating season intermediate indoor air temperature before destratification,,°F;; This depends on vertical temperature profile. This must be measured on website because the shape of the temperature profile may vary substantially depending on type of heaters, their height above floor level, and how ventilation is provided; tiad = Heating season intermediate indoor air temperature after destratification, °F. This is taken as the thermostat set point as the indoor air temperature all around the space is close to uniform after destratification.
The scaled down heating load due to destratification may be converted into a amount of fuel taking into account the efficacy of the heating scheme and the caloric value of the fuel. The heating fuel cost saving specifically amongst 20% and 30% is calculated using the unit cost of fuel.
8. Thermal ease in Non-air Conditioned Space
The ANSI/ASHRAE 55-2004 Standard offers a method for determining an satisfactory range of indoor operative temperature in occupant-controlled, naturally conditioned spaces. Occupant-controlled, naturally conditioned spaces are specified as spaces where thermal conditions of the space are regulated mainly by the occupants through opening and closing windows. These are spaces with no refrigerated air conditioning, radiant cooling, or desiccant cooling. Fans may be employed when natural ventilation does not provide sufficient air movement.
In such spaces, occupants have dissimilar expected values of thermal ease and receive wider ranges of thermal conditions in both winter and summer than occupants of air conditioned spaces. This method is intended for climates where mean regularly every month air temperatures fall in the range of 50°F to 92°F. This method is in general described as the Adaptive Model (de Dear and Schiller (2001).
Using the adaptive approach, the introductory step is to determine the intermediate on a monthly basis temperature for each month of the cooling season for the location. In ventilated buildings without air conditioning, temperature for operative ease toc, is based on mean per month outdoor air temperature tout, and may be calculated using the following equation (ASHRAE, 2005).
toc = 66 + 0.255(tout – 32) (9)
The ease zone range of operative temperature to satisfy 80% of acclimatized people may be read of a graph in the Standard or by adding and subtracting 6.3 ºF to the operative ease temperature.
With a mean each day air temperature of 83.6ºF in the city of Houston for the duration of July, toc = 66 + 0.255(83.6 -32)= 79.2 ºF. The thermal ease zone to satisfy 80% of persons in July is then 72.9ºF to 85.5ºF.
Given the long term intermediate on a monthly basis outdoor air temperature for Houston TX in July is 83.6ºF, this presents the intermediate need for a cooling effect from air motion in January of 83.6ºF – 79.2ºF or 4.4ºF to restore the operative temperature to the norm. The question now is how much air motion is necessitated to achieve a cooling effect of 4.4ºF? Using the selective information from Khedari et al (2000), for a warm humid climate with a relative humidity of 75% gives evidence of 87 fpm is necessitated for a 4.4ºF cooling effect.
9. Cooling effects of air motion in naturally conditioned spaces
The US Naval Medical Command (1988) in a chapter on relieving heat stress published selective information on the relative cooling effect of air motion Figure 7. These selective information do not provide a quantitative cooling effect but are utile in that they indicate the greatest or most complete or best possible cooling effect occurs with air motion around 1,500 fpm.
In naturally conditioned space, there is no control of humidity. As the cooling effect of air motion in warm environments relates to evaporative cooling from sweating, it has been shown that as humidity increases, the cooling effect of air motion decreases. The scaled down cooling effect is much more outstanding in warm humid environments when air motion necessitated for thermal ease surpasses 295 fpm, Figure 6 (Khedari et al, 2000). It is indispensable to use cooling effect selective information derived from local climate and cultural conditions. These selective information will better reflect the thermal ease expected values of local persons taking into account local dress and typical levels of metabolic activity.
A potpourri of approaches have been taken by researchers to quantify the cooling effects of air movement. Cooling effects of air motion may effective in hot arid environments were evaporative cooling of the skin is not encumbered by high humidity (Scheatzle et al, 1989).
Another equation derived from various studies (Szokolay, 1998) that is widely applied for estimating the cooling effects of air motion from 40 ft/min to 400 ft/min is:
”t = 10.8((V/197.85)-0.2)-1.8((V/197.85)-0.2)2 (11)
Where V is in ft/mim and ”t is in ºF.
Using this equation, air motion of 400 ft/min provides a cooling effect 13.7 ºF. This is equivalent to Khedari et al cooling effect for 400 ft/min at 57% relative humidity in Thailand.
10. Indoor air motion for livestock
Dairy farmers have learned from university studies that thermally comfortable cows’ milk production, procreative health and growth are much better than those of cows subjected to summer heat stress (Sanford, 2004). During hot summer periods dairy farmers have installed little high speed circulator fans to achieve the commended air motion of 177 ft/min to 433 ft/min. Ten 36 inch diameter fans operating at 825 rpm use 3.73 kW of electrical energy. Farmers have found they may replace 10 of these 36 inch diameter fans with a single 24 ft diameter fan operating at 42 rpm that uses only 1.6 kW of electrical energy while providing the same air movement. Additional cooling may be achieved in drier climate regions using misting water sprays for evaporative cooling.
11. Discussion
All the descriptions of air motion described so far in this document have referred to the intermediate velocity of air movement. Olesen (1985) refers to a study by Fanger and Pedersen of the chilling effect of winter draughts. It was observed in the study that the chilling effect of gusting air flow reached a peak around a gust frequency of 0.5Hz.
More not long ago researchers in China (Xia et al,2000) repeated these studies inwarm, humid conditions with temperatures ranging from 79ºF to 87ºF and relative humidity amongst 35% and 65%. These experiments showed that the preferent gust frequency for cooling air motion was amongst 0.3Hz and 0.5Hz. Approximately 95% of subjects preferent gust frequencies beneath 0.7Hz. Natural breezes and air flow from huge low-speed circulator fans have a significant share of their energy spectral density around this frequency of 0.5Hz. Olesen (1985) suggested the use of an equivalent uniform air velocity, Table 1, to account for this effect but this intensified cooling effect has not been quintessentially accounted for in cooling effects of air motion to date.
12. Conclusions
Current air conditioning design provides for uniform air temperature and humidity all around a space, with imperceptible local air motion in the occupied zone of less than 40 ft/min. This established design is based on air conditioning heating and cooling loads that ignore the substantial savings to be gained from increased indoor air motion from circulator fans.
Recent ASHRAE acceptance of an adaptive thermal ease model distinctly shows that persons who live in air conditioned houses, drive air conditioned cars, work in air conditioned offices impair their natural thermal ease adaptation. This handicap results in unnecessarily high summer cooling loads.
Where naturally conditioned buildings are acceptable, indoor thermal ease may be achieved with substantial energy savings by better utilization of indoor air movement.
The cooling effect of air motion has been well conventional by a number of researchers. There remains a need for further exploration on the cooling effects of air motion on building occupants to accommodate action levels beyond 1.3 met, higher air velocities for non-sedentary activity, and lighter costume levels than 0.5 clo. This exploration is necessitated in both air conditioned and naturally conditioned spaces.
Research on the cooling effects of air motion has been staged in some forms. The chart developed by Khedari et al (2000) is one the better formats. Further exploration is necessitated to create a form which presents selective information in a way that makes it more without apparent effort applied by engineers to improve energy efficacy with increased indoor air movement.
The same circulator fans applied to heighten summer thermal ease may be applied to destratify indoor air to save heating energy in winter. This particularly applies to mercantile or industrial spaces with high ceilings.
References
ASHRAE (2005) ASHRAE 2005 Handbook of Fundamentals, ASHRAE, Atlanta, GA. Page 26.11.
ASHRAE (2004) ANSI/ASHRAE Standard 55-2004 Thermal Environmental Conditions for Human Occupancy. ASHRAE, Atlanta, GA.
Byrne, S. and Huang, V.(1986) The affect of wind-induced ventilation on residential cooling load and humane comfort. ASHRAE Trans. Vol.92, Pt. 2, 793-802.
de Dear, R. and Schiller Brager, G. (2001) The adaptive model for thermal ease and energy conservation in the built environment. Int. J. Biometeorology, 45: 100-108.
Exeloncorp (2005) Controlling Temperatures is accessible on the internet at:
http://www.exeloncorp.com
Fountain, M. (1995) An empirical model for predicting air motion preferent in warm office environments. Standards for thermal comfort: Indoor air temperatures for the 21st century. Edited by F. Nicol, M. Humphreys, O. Sykes and S. London, Roaf, E & F Spon. pp. 78-85.
Hope, P (2003) Energy efficacy ratings: Implications for the building industry in the humid tropics. Master in Tropical Architecture dissertation, Australian Institute of Tropical Architecture, James Cook University, Townsville, Australia, pp. 377.
Khedari, J., Yamtraipat, N., Pratintong, N. and Hinrunlabbh, J. (2000) Thailand ventilation ease chart. Energy and Buildings, Vol. 32, pp. 245-249.
Naval Medical Command (1988) Manual Of Naval Preventive Medicine, Chapter 3, page 3-7. Accessible on the internet at:
[http://www.vnh.org/PreventiveMedicine/PDF/P-5010-3.pdf]
Olesen, B. (1985) Local thermal discomfort. Bruel & Kjaer Technical Review, No.1, Denmark, pp.3-42.
Pignet, Tom and Saxena, Umesh (2002) Estimation of energy savings due to destratification of air in plants, Energy Engineering, Vol 99, No. 1, 69-72.
Sanford, S. (2004) Energy conservation in agriculture: Ventilation and cooling schemes for animal housing. University of Wisconnsin Cooperative Extension publication A3784-6, pp.3.
Scheatzle, D., Wu, H. and Yellott, J.(1989) Extending the summer ease envelope with ceiling fans in hot, arid climates. ASHRAE Trans. Vol.100, Pt. 1, 269-280.
Szokolay, S. (1998) Thermal ease in the warm-humid tropics, Proceedings of the 31st Annual Conference of the Australian and New Zealand Arch. Science Association, Uni. of Queensland., Brisbane, Sept.29-Oct.3, pp. 7-12.
Tanabe, S and Kimura, K. (1994) Importance of air motion for thermal ease beneath hot and humid conditions. ASHRAE Trans. Vol. 100, Pt. 2, 953-969.
Xia, Y., Zhao, R. and Xu, W. (2000) Human thermal sensation to air motion frequency. Reading, UK. Proceedings of the 7th International Conference on Air Distribution in Rooms. Vol.1, pp. 41-46.
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