1. Introduction
Air traffic can play an important role in the thermal comfort of man and beast. Wind on a wet summer day can significantly affect thermal comfort. Recent energy efficiency strategies in buildings take into account the cooling effects of air movement from natural ventilation. When the hull is closed for air conditioning, local air movement is maintained below 40 feet / min. This ignores the possibility of increasing air movement to reduce cooling energy in an air-conditioned room. This article discusses the potential for energy savings through the use of indoor air movement.
2. Saving energy cooling in a room with air conditioning with increased air velocity
The current version of the ANSI / ASHRAE Standard 55-2004 “Thermal Environmental Conditions for Man” (ASHRAE, 2004) provides for a limited increase in the temperature parameters of the summer thermostat due to an increase in the local air velocity. Figure 1 is shown in Fig. 5.2.3 in the standard 55-2004.
Curves of equal heat loss from the skin for a combination of working temperature and air movement belong to the upper limit of the comfort zone (PMV = +0.5). The limits of 160 fpm and 5.4ºF are set for sedentary activity, from 1.0 to 1.3 met. Large individual differences in preferred air velocity
requires passengers to have personal airspeed control in 30 ft / min increments.
The Standard states that interpolating between these curves is acceptable. Air velocity is more effective at compensating for temperature increases, when the average radiation temperature is greater than the average air temperature in dry light.
It should be noted that in Figure 5.2.3 of the Standard there are two errors. “18 ° C” should be read “18 ° F”, and there is a scaling error between the fpm and m / s scales.
Five separate curves are provided to match the temperature difference of -18 ° F, -9 ° F, 0.0 ° F, + 9 ° F, and + 18 ° F between the average radiation temperature, tr and the average air temperature in dry light, ta. The authors established a relationship to the plot of curves bounded by a sedentary activity of 160 fpm and 5.4 ° F for 1.0 from 1.3 to 0.5 and 0.7 clo.
The writer also provided a comparison with part of the activity curves outside the sedentary limits. The limits of the cooling effect for these coefficients established for the curves in Fig. 5.2.3 in standard 55-2004, was 300 fpm and 8 ° F
2.1. The curve for tr - ta = 0,0 K
For tr - ta = 0.0 ° F, an air speed of 160 fpm allows the thermostat to increase with a limitation of 4.4 ° F for easy sitting activity (from 1 to 1.3 meters) and from 0.5 to 0.7 clo.
V = 40 + 6.8 "T 1.85 (1)
Where V is the average relative air velocity in fpm, and t is the cooling effect in ° F
In most temperature-controlled air-conditioned rooms, the temperature on the wall, the ceiling, and the floor surface is close to the air temperature. This is tr - ta = 0 ° F. Conditions when tr - ta is not zero include spaces with weakly 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 velocity of 160 fpm allows the thermostat setting to be limited to 5.4 ° F for easy sitting activity (from 1 to 1.3 meters) and from 0.5 to 0.7 clo.
V = 40 + 1.26 "t 2.85 (2)
Where V is the average relative air velocity 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 velocity of 126 fpm allows to increase the thermostat setpoint with a limitation of 5.4 ° F for easy sitting activity (from 1 to 1.3 meters) and from 0.5 to 0.7 clo.
V = 40 + 1.28 "t 2.7 (3)
3. Beyond sedative activity
The standard is unclear with regard to restrictions for proportions of curves to 89 ° F and 300 fpm, beyond the limits established for sedentary activity. Studies have measured the cooling effect of air movement up to 600 fpm in warm climates (Khedari et al., 2000, Tanabe and Kimura, 1994, and Scheatzie et al., 1989). Air traffic above 160 feet per minute is used in air-conditioned gymnasiums and shopping centers to increase passenger cooling. The writer has established relationships to part of the activity curves outside the sedentary limits.
For tr - ta = 0.0 ° F, an air velocity of 300 fpm indicates that an increase in thermostat setpoint may be 6.6 ° F with activity levels above 1.3 meters.
V = 40 + 2.52 "t 2.5 (4)
The limits for equation 4 range from 160 fpm to 300 fpm and from 4.4 F to 6.6 F
For tr - ta = + 9ºF, an air velocity of 276 fpm allows the thermostat setpoint to be increased by 8ºF at activity levels above 1.3 meters.
V = 40 + 5.7 "t 1.8 (5)
The limits of Equation 5 range from 160 to 280 fpm and from 5.4ºF to 8ºF.
For tr - ta = + 18ºF, an air velocity of 211 fpm indicates that the increase in thermostat setpoint may be 8ºF at activity levels above 1.3 meters.
V = 40 + 6.3 "T 1.59 (6)
The limits for equation 6 are 132 fpm to 209 fpm and 5.48ºF to 8ºF.
4. Evaluation of cooling energy savings
US electric utility corporation Exeloncorp (2005) suggests that the cost of cooling an indoor air conditioner can be reduced by 3% to 4% for each ºF, that the installation of the thermostat increases in the summer.
Residents can compensate for the increased thermostat setting of 4.7ºF, providing 160 fpm of low-cost airflow from the circulation fans and enjoying normal comfort while maintaining the operating costs of the air conditioner. Based on the recommendation of Exeloncorp (2005), increasing the thermostat setting to 4.7 ° F will save cooling energy from 14% to 19%. In a gymnasium where higher air traffic is acceptable, the savings from increasing the thermostat by 8ºF can be from 24% to 32%. A detailed analysis of the reduction in cooling load due to airflow was performed for six US cities in different climatic zones (Byrne and Huang, 1986)
5. Comparison of fans and room air conditioners
A detailed comparison of the energy needed to maintain the same thermal comfort in a 141.5-foot bedroom in Townsville, Hope (2003) was carried out using a 55-inch ceiling ceiling fan and a VF100C Carrier ventilation / wall-mounted air conditioner designed for room by a local distributor. The measured power consumption of a 55-inch ceiling fan operating at maximum speed was 0.068 kW or 0.48 W / ft2 floor area. This is 8.7% of the power used by the room air conditioner to achieve the same thermal comfort. The energy efficiency ratio of the window / wall air conditioner was 0.78 kW, or 5.51 W / ft2 of floor space. This is 11.5 times the power used by the ceiling fan.
6. Destruction
In heated rooms in winter, indoor air tends to stratify the hottest, less dense air that accumulates under the roof due to gravity. This condition creates two problems. Firstly, the hottest air does not contribute to the thermal comfort of passengers near the floor level, and secondly, it creates a high temperature difference between the lower roof and the exterior of the roof, which increases heat loss through the roof.
Destratification is a process of thorough mixing inside the premises, so that the air temperature near the floor is the same as the air temperature under the roof, or the difference is no more than 2ºF. This is done using fans. In a typical 30-foot-high US warehouse, seasonal energy savings from effective degradation range from 20% to 30%. To be effective, approximately half of the total air volume in the space must be moved from the ceiling level to the floor level per hour.
To be effective in destruction, the fan must be no more than 1 diameter below the ceiling, and the jet from the fan must act on the floor to ensure effective circulation. Aircraft from ceiling fans have an effective throw of 5 to 6 diameters.
Large buildings with high ceilings, such as churches, industrial buildings or distribution warehouses, need to distribute a large amount of air. To avoid complaints about drafts from passengers, the local air velocity at the height of the head should be less than 40 feet / min.
Fans with circulation are much more energy efficient at low speeds, so a large diameter, slow running, fans are well suited for destruction. One industrial ceiling fan with a diameter of 24 feet, operating at a maximum speed of 42 rpm, uses 1.67 kW of electricity, but only 0.06 kW, operating at 14 rpm, its maximum efficiency. At 42 rpm, this fan provides about 337,700 cu. Feet of air and 76,670 cubic feet per minute at 14 rpm. An additional advantage of running large fans at low speeds compared to smaller fans at higher speeds is to reduce fan noise. Large slow moving fans are virtually silent.
7. Estimation of energy saving destruction
The recommended method for estimating the energy savings from degradation is to determine the seasonal heat exchange rate for a building envelope and determine the difference in heat loss before and after destruction (Pignet and Saxena, 2002).
The seasonal seasonal heat transfer rate for a building envelope in Watts can be calculated using:
A x U = qbd / (ti-to) (7)
Where: A - the surface area of the envelope of the building in ft2; U is the bevelled heat transfer coefficient for the envelope of the building in Btu / ft2.h.ºF; qbd is the rate of heat loss through the envelope of a building in Btu / h before destruction; and ti -to is the average heating season indoors with the difference in outdoor temperature in ºF.
The total heat lost from a building is the sum of heat emitted from the furnaces plus heat generated in space from other sources such as lighting, people, cars, or production processes. The heat released from the furnaces can be determined from the fuel bills for the season, the calorific value of the fuel for heating and the efficiency of the system. The calorific value of natural gas is about 1000 Btu / ft3. The time used in these calculations is the heating season associated with the measured fuel consumption.
Forced air furnaces with chimneys have an efficiency of about 0.7. Radiation heaters without flues have an efficiency of 0.8. Electric heaters have an efficiency of 1.0. Heat from other sources is estimated in the usual way, as indicated in the HVAC directories (ASHRAE, 2005).
With a total heat loss U x A for the heating season before destruction, it is confirmed that the decrease in heat after destruction, qad can be determined from:
qad = U x A x (tibd - tiad) (8)
Where: qad = Reduced heat load after destruction in BTU / hour; U = time-corrected average heat loss rate for a building envelope in Btu / hr.ft2.ºF; A = Surface area of a building envelope, ft2; tibd = average air temperature in the heating season before destruction, ° F ;; It depends on the vertical temperature profile. This should be measured on site, as the shape of the temperature profile can vary significantly depending on the type of heater, its height above the floor and how the ventilation is provided; tiad = The average air temperature in the room in the room after destruction, ° F. This is taken as the thermostat setpoint, since the air temperature inside the room is close to homogeneous after the destruction.
The reduced heat load due to destruction can be converted to the amount of fuel, taking into account the efficiency of the heating system and the calorific value of the fuel. The savings in heating costs, typically from 20% to 30%, are calculated using the unit cost of fuel.
8. Thermal comfort in a room without air.
The ANSI / ASHRAE Standard 55-2004 offers a method for determining an acceptable range of operating room temperature in human-controlled rooms with natural air conditioning. A window controlled by a person, naturally conditioned by space, is defined as a space where the thermal conditions of a space are regulated mainly by residents through opening and closing windows. These are rooms without cooling, cooling, cooling or drainage. Fans can be used when natural ventilation does not provide sufficient air movement.
In such premises, residents have different expectations for thermal comfort and accept wider ranges of thermal conditions in both winter and summer than air conditioners. This method is intended for a climate where the average monthly air temperature falls in the range of 50 ° F to 92 ° F. This method is usually described as an adaptive model (de Dear and Schiller, 2001).
Using an adaptive approach, the first step is to determine the average monthly temperature for each month of the cooling season for a location. In non-air-conditioned ventilated buildings, the temperature for operational comfort toc is based on the monthly monthly outdoor temperature and can be calculated using the following equation (ASHRAE, 2005).
toc = 66 + 0.255 (tout - 32) (9)
The range of the comfort zone of the working temperature to satisfy 80% of the acclimatized people can be read on the graph in the standard or by adding and subtracting 6.3 ºF to the working comfortable temperature.
With an average daily air temperature of 83.6ºF in the city of Houston in July, toc = 66 + 0.255 (83.6 -32) = 79.2 ºF. Thermal comfort zone to meet 80% of people in July is 72.9ºF to 85.5ºF.
Considering the long-term average daily outdoor temperature for Houston-Texas in July is 83.6ºF, this represents the average need for a cooling effect of air movement in January from 83.6ºF to 79.2ºF or 4.4ºF to restore normal operating standards. The question is, how much air is required to achieve a cooling effect of 4.4 ° F? Using data from Khedari et al (2000), a warm humid climate with a relative humidity of 75% requires 87 feet per minute for cooling 4.4 ° F.
9. The cooling effects of air movement in naturally air-conditioned rooms.
The US Naval Medical Command (1988), in a chapter on removing heat stress, published data on the relative cooling effect of air movement. Figure 7. These data do not provide a quantitative cooling effect, but are useful in that they indicate that the maximum cooling effect occurs with air movement at about 1500 feet per minute.
There is no humidity control in the natural air-conditioned space. Since the cooling effect of air movement in warm conditions is associated with evaporative cooling from perspiration, it was shown that as the humidity increases, the cooling effect of air movement decreases. The reduced cooling effect is much higher in warm, humid environments when the air movement required for thermal comfort exceeds 295 fpm, Figure 6 (Khedari et al., 2000). It is important to use cooling effect data from local climate and cultural conditions. These data better reflect the expectations of thermal comfort of local residents, taking into account local clothing and typical levels of metabolic activity.
The researchers investigated various approaches to quantifying the cooling effects of air movement. The cooling effects of air movement can be effective in heat-resistant environments, evaporative cooling of the skin is not burdened with high humidity (Scheatzle et al., 1989).
Another equation derived from several studies (Szokolay, 1998), which is widely used to evaluate the cooling effects of air movement from 40 ft / min to 400 ft / min:
"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, 400 feet / min of air movement provides a cooling effect of 13.7 ºF. This is equivalent to the cooling effect of Khedari et al. At 400 ft / min at 57% relative humidity in Thailand.
10. Air traffic in livestock facilities
Dairy farmers have learned from university studies that thermally convenient cow milk production, reproductive health and growth are much better than cows predicted by summer heat stress (Sanford, 2004). During hot summer periods, dairy farmers installed small high-speed circulating fans to achieve recommended air movement from 177 feet / min to 433 feet / min. Ten 36-inch fans operating at 825 rpm use 3.73 kW of electricity. Farmers have found that they can replace 10 of these 36-inch-diameter fans with a single 24-foot diameter fan operating at 42 rpm, which uses only 1.6 kW of electrical energy while ensuring air movement. Additional cooling can be achieved in drier climatic zones using aerosols to spray water for evaporative cooling.
11. Discussion
All descriptions of the movement of air, described so far in this document, concerned the average speed of movement of air. Olesen (Olesen, 1985) refers to the study of Fanger and Pedersen about the cold effect of winter precipitation. В исследовании наблюдалось, что охлаждающий эффект порывного потока воздуха достиг пика вокруг частоты порыва 0,5 Гц.
Более поздние исследователи в Китае (Xia et al., 2000) повторили эти исследования в теплых, влажных условиях с температурой от 79ºF до 87ºF и относительной влажностью от 35% до 65%. Эти эксперименты показали, что предпочтительная частота порывов для движения охлаждающего воздуха была между 0,3 Гц и 0,5 Гц. Примерно 95% испытуемых предпочитают частоты порывов ниже 0,7 Гц. Естественный бриз и поток воздуха от больших низкоскоростных циркуляционных вентиляторов имеют значительную долю их энергетической спектральной плотности вокруг этой частоты 0,5 Гц. Олесен (Olesen, 1985) предложил использовать эквивалентную равномерную скорость воздуха, табл. 1, чтобы учесть этот эффект, но этот улучшенный охлаждающий эффект специально не учитывался в охлаждающих эффектах движения воздуха на сегодняшний день.
12. Выводы
Современная конструкция кондиционера обеспечивает равномерную температуру и влажность воздуха в пространстве, с незаметным местным движением воздуха в зоне захвата менее 40 фут / мин. Эта традиционная конструкция основана на нагрузках на отопление и охлаждение кондиционера, которые игнорируют существенную экономию, получаемую от увеличения движения воздуха внутри помещений от вентиляторов циркуляции.
Недавнее признание ASHRAE адаптивной модели теплового комфорта ясно показывает, что люди, которые живут в кондиционированных домах, управляют автомобилями с кондиционерами, работают в кондиционированных офисах и имеют естественный температурный комфорт. Это неблагоприятное явление приводит к излишне высокой летней охлаждающей нагрузке.
Там, где естественные кондиционированные здания приемлемы, внутренний тепловой комфорт может быть достигнут благодаря значительной экономии энергии за счет лучшего использования движения воздуха в помещении.
Ряд исследователей хорошо зарекомендовал себя охлаждающий эффект движения воздуха. По-прежнему существует потребность в дальнейших исследованиях охлаждающих эффектов движения воздуха на жилых зданиях для размещения уровней активности выше 1,3 метра, более высоких скоростей воздуха для не сидячей активности и более легких уровней одежды, чем 0,5 кло. Это исследование необходимо как в кондиционированных, так и в естественных условиях.
Исследования по охлаждающим эффектам движения воздуха были представлены во многих формах. Диаграмма, созданная Khedari et al (2000), является одним из лучших форматов. Дальнейшие исследования необходимы для разработки формы, которая представляет данные таким образом, чтобы они более легко использовались инженерами для повышения энергоэффективности при увеличении движения воздуха внутри помещений.
Те же самые циркуляционные вентиляторы, используемые для повышения летнего теплового комфорта, могут использоваться для разрушения внутреннего воздуха, чтобы сэкономить тепловую энергию зимой. Это особенно относится к коммерческим или промышленным пространствам с высокими потолками.
Recommendations
ASHRAE (2005) ASHRAE 2005 Справочник по основам, ASHRAE, Атланта, Джорджия. Page 26.11.
ASHRAE (2004) ANSI / ASHRAE Стандарт 55-2004 Термические условия окружающей среды для человека. ASHRAE, Атланта, Джорджия.
Byrne, S. and Huang, V. (1986) Влияние ветровой вентиляции на жилую охлаждающую нагрузку и комфорт для человека. ASHRAE Trans. Vol.92, Pt. 2, 793-802.
de Dear, R. and Schiller Brager, G. (2001) Адаптивная модель теплового комфорта и энергосбережения в построенной среде. Int. J. Biometeorology, 45: 100-108.
Exeloncorp (2005) Контроль температуры доступен в Интернете по адресу:
http://www.exeloncorp.com
Fountain, M. (1995) Эмпирическая модель для прогнозирования движения воздуха предпочтительнее в теплых помещениях офиса. Стандарты теплового комфорта: температура воздуха в помещении в XXI веке. Под редакцией Ф. Николь, М. Хамфрис, О. Сайкс и С. Лондон, Руф, E & F Spon. С. 78-85.
Hope, P (2003) Рейтинги энергоэффективности: последствия для строительной отрасли во влажных тропиках. Магистр в области тропической архитектуры, Австралийский институт тропической архитектуры, Университет Джеймса Кука, Таунсвилл, Австралия, с. 377.
Khedari, J., Yamtraipat, N., Pratintong, N. и Hinrunlabbh, J. (2000). Энергия и здания, т. 32, pp. 245-249.
Морское медицинское командование (1988 год) Руководство военно-морской превентивной медицины, глава 3, стр. 3-7. Доступно в Интернете по адресу:
[http://www.vnh.org/PreventiveMedicine/PDF/P-5010-3.pdf]
Olesen, B. (1985) Локальный тепловой дискомфорт. Технический обзор Bruel & Kjaer, № 1, Дания, стр.3-42.
Pignet, Tom and Saxena, Umesh (2002) Оценка экономии энергии за счет разрушения воздуха в установках, Energy Engineering, том 99, № 1, 69-72.
Sanford, S. (2004) Энергосбережение в сельском хозяйстве: системы вентиляции и охлаждения для животноводства. Публикация Совместного продюсера университета Wisconnsin A3784-6, стр. 3.
Scheatzle, D., Wu, H. and Yellott, J. (1989) Расширение летней комфортной оболочки с потолочными вентиляторами в жарком и засушливом климате. ASHRAE Trans. Vol.100, Pt. 1, 269-280.
Szokolay, S. (1998) Тепловой комфорт в теплых влажных тропиках, Труды 31-й ежегодной конференции австралийской и новозеландской арки. Научная ассоциация, Юни. из Квинсленда., Брисбен, 29 сентября - 3 октября, с. 7-12.
Tanabe, S и Kimura, K. (1994). Важность движения воздуха для теплового комфорта в жарких и влажных условиях. ASHRAE Trans. Том 100, Pt. 2, 953-969.
Xia, Y., Zhao, R. and Xu, W. (2000) Тепловая чувствительность человека к частоте движения воздуха. Чтение, Великобритания. Материалы 7-й Международной конференции по распределению воздуха в помещениях. Том 41-46.
