Industrial Cool and Comfort

When cooling large industrial facilities that have high heat loads, substantial gains in efficiency can often be achieved by taking a new look at existing technologies (see "Slower Fans for Lower Costs," p. 36) and strategies. For example, lowering the entering condenser-water temperature (ECWT) to water-cooled chillers is a well-known strategy for reducing energy use.

By Donald Nurisso, P.E., Member and Jon Wintermeyer, P.E., Treasurer, ASHRAE Golden Gate Chapter San Francisco June 1, 2002

When cooling large industrial facilities that have high heat loads, substantial gains in efficiency can often be achieved by taking a new look at existing technologies (see “Slower Fans for Lower Costs,” p. 36) and strategies.

For example, lowering the entering condenser-water temperature (ECWT) to water-cooled chillers is a well-known strategy for reducing energy use. But successful implementation of this approach depends on a determination of the exact reset curve.

The Air-Conditioning and Refrigeration Institute (ARI) Standard 550/590-98 allows for lowering the ECWT—as a function of reduction in load on the chiller—by approximately 3°F per 10% reduction in chiller load, down to a minimum of 65°F. While this standard offers a good framework, it does not always supply the most energy-efficient solution.

The recent retrofit of a large industrial facility, for example, required an examination of the ARI reset schedule to determine whether it was actually the most energy efficient. The facility had a constant internal chiller heat load of 500 tons, which peaked at 2,600 tons in the summer. Existing capacity was multiple 1,400-ton chillers with inlet-vane control.

Because there were significant hours of chiller-plant operation at the 500-ton load point, the chiller manufacturer proposed a change to one of the chillers—providing a VFD-controlled 750-ton chiller, and reusing the existing 1,400-ton condenser barrel and chiller barrel. Obviously, a chiller operating at two-thirds load—a 750-ton chiller operating at a 500-ton load—is more efficient than a chiller operating at 35% load: i.e., a 1,400-ton chiller operating at the 500-ton load point.

In addition to the chiller conversion, the retrofit included fitting the cooling tower fans with variable-frequency drive (VFD) control. The chiller and cooling tower manufacturers—utilizing a design ECWT of 80

Creating these mix-match curves required some coordination: For each load point, the chiller manufacturer determined the heat rejection and kilowatts-per-ton efficiency of the chiller. Armed with these data, mechanical system designers were then able to request the cooling tower vendor to determine the speed of the VFD that would handle this heat-rejection load. From the load calculations, the designing engineers knew the mean-coincident wet-bulb temperatures for each load point—information that was required by the cooling tower vendor to determine performance of the tower.

In addition to 80

Analyzing the setpoints

For the 1,400-ton chiller—when at the 1,200-ton load point and design outside-air (OA) wet-bulb temperature—the cooling tower was unable to provide water colder than 80°F, the original design point. When the same chiller was operated at the 700-ton load point, it was most efficient at an ECWT of 70°F. There is an asymptotic shape to the curve, with a minimum point of power consumption at 70°F. This power consumption is the sum of chiller motor energy plus cooling tower motor energy.

On the other hand, when the 1,400-ton chiller was operated at the 400-ton load point, the most efficient ECWT was 60°F. This curve has not developed an asymptote; power draw is still pointed downward at 60°F, implying that colder ECWT will decrease energy consumption. However, the chiller manufacturer would not allow operation of the chiller below 60°F ECWT.

Also, when the 750-ton VFD-equipped chiller was operated at the 700-ton load point, it was most efficient at an ECWT of 70°F—another asymptotic curve. At the 400-ton load point, it was most efficient at a load point of 400 tons but has not yet reached an asymptote curve.

Plotting sensitivity

The sensitivity of the two chillers’ energy use was plotted as a function of ECWT and load point of the chiller. Data for both chiller sizes is shown graphically in Figure 2. Note the differing shape of the curves, implying that the VFD-controlled 750-ton chiller is much more efficient at part-load operation points. For each 1°F decrease in ECWT, the power draw of the VFD-controlled chiller decreases 2%, while power draw of the other chiller decreases 1%.

Finally, all data was summarized to generate the three reset schedules given in Figure 3. The curve that is marked “most efficient” shows an increase in the ECWT at the 72.5°F load point, because one goes from two 1,400-ton chillers operating at 52% load to one 1,400-ton chiller operating at 90% load. The reset schedule based on OA temperature requires only information on OA temperature, whereas the “most efficient” reset schedule requires load feedback on each chiller’s load point, a more expensive controls modification.

With this load distribution, an operating penalty of less than 3% energy use was incurred, when comparing the most efficient reset schedule with the OA-temperature reset schedule. The ARI reset schedule fell in the middle. One can conclude that at lower OA temperatures, the “most efficient” schedule is substantially more efficient than the ARI schedule.

Figures 4 and 5 show the outside-air temperature frequency chart and a chart of plant cooling load vs. outside air temperatures, respectively.

VFD and part load

The VFD-controlled chillers are substantially more efficient at part-load operation. For chillers required to operate substantial hours at low outdoor-air wet temperatures at part-load, it is worthwhile to explore chiller operational data with the chiller manufacturer. Chillers are very sensitive to part-load conditions. One can do better than ARI reset, especially when there are many hours of operation at low load conditions.

Slower Fans for Lower Costs

Ceiling-mounted fans are an essential means of providing air movement for cooling industrial facilities, and increasing the diameter and reducing the rotation speed of these fans can result in significant energy savings. But energy efficiency isn’t the only benefit. Using larger, slower fans in factories and warehouses can also reduce fan noise by decreasing the fan blade-tip speed.

Evidence for this conclusion comes from two sources: research conducted by Richard Aynsley, Ph.D., an engineering professor at Southern Polytechnic State University, Marietta, Ga.; and application of the fan laws —a set of mathematical relationships that define the operating principles of fans.

A standard industrial ceiling fan has three blades, a 5-ft. diameter and maximum speed of approximately 315 revolutions per minute (rpm). This fan draws 160 watts of electricity and provides airflow greater than 39 feet per minute (fpm) over an unobstructed floor area 80 ft. in diameter—about 5,025 sq. ft. The fan’s blade-tip speed at 315 rpm is 4,920 fpm; the sound level is 85 decibels.

On the other hand, a ten-blade, 24-ft.-diameter ceiling fan—operating at 50 rpm and drawing 370 watts—can provide airflow greater than 39 fpm over an unobstructed floor area approximately 160 ft. in diameter, or 20,110 sq. ft. The blade-tip speed at 50 rpm is 3,760 fpm, with about a 75-decibel sound level.

As a result, the large-diameter fan is 41.7% more efficient per unit of floor area than the typical 5-ft. fan: [100(0.34 – 0.198)]/0.34 = 41.7%.

The comparison assumes, of course, few airflow obstructions at floor level. Where there are significant obstructions—such as machinery—the effective floor area served will be less. In these situations, location of the fan, relative to floor-level obstructions, can be an important consideration.

Air turbulence

Essentially, airflow from a ceiling fan is a circular column of turbulent air, with maximum downward velocities near the column’s outer wall. As the downward-moving column approaches floor level, some of its velocity pressure is converted to static pressure that provides the energy for the radial outflow of air across the floor. The highly turbulent nature of the airflow results in energy losses proportional to the square of the mean velocity. This reduction in airflow velocities increases with height above floor level.

However, airflow adjacent to the floor experiences less energy loss due to the coanda effect. Energy losses due to skin friction are proportional to the square mean velocity. This means that airflow adjacent to the floor extends further from the fan than air above the floor.

To take advantage of this floor-clinging airflow, obstructions to airflow should be raised approximately 8 to 12 in. above floor level. At workstations, deflectors at floor level can be used to redirect airflow upward around the workers.

The height of ceiling fans above the floor also affects airflow near floor level. For example, when fans are mounted at a height of 20 ft., the mean airflow velocity four inches above the floor—at radial distances beyond the tips of the fan blades—is 210 fpm. With fans mounting 40 ft. high, floor-level mean airflow velocities—at 157 fpm—are 33.7% lower.

Applying Fan Laws

The fan laws demonstrate the benefits of large, slow-moving fans as follows:

Keeping the fan size constant, the total pressure is increased proportionally by increasing fan speed. Flow volume (q), is directly proportional to the fan’s rotational speed (N) in revolutions per second: q 1 /q 2 = N 1 /N 2 . But increasing the fan speed has its drawbacks: greater power consumption and increased fan noise.

Keeping the tip speed of fan blades constant, the flow volume (q) is increased by increasing the fan diameter. Flow volume and air power (A), in watts, are directly proportional to the square of the fan size (D): q 1 /q 2 = A 1 /A 2 = (D 1 /D 2 )

Keeping the speed of rotation constant, flow volume (q) increases in proportion to the cube of the ratio of the fan diameter increase. Flow volume is directly proportional to the cube of the fan size (D): q 1 /q 2 = (D 1 /D 2 )

These are only some of the possibilities. Knowledge of ceiling-fan airflow characteristics and the fan laws can provide further opportunities to increase efficiency.

By Bill Buell, Marketing Director, HVLS Fan Company, Lexington, Ky.