0 title page - submittal(standard) rev2.pdfsimplifying bernoulli’s law, we find the relationship...

Fans

HVAC Clinic

Table Of Contents

Introduction ............................................................................................................................................ 3 Fan Performance .................................................................................................................................... 3 Fan Types ............................................................................................................................................. 13 Fan Capacity Control ........................................................................................................................... 21 Application Considerations................................................................................................................. 27

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Nevada Mechanical Systems Fans of 33

Introduction A fan is a machine used to create flow within a fluid, typically a gas such as air. The focus of this clinic is to discuss fans as they apply to air distribution systems. A fan consists of a rotating arrangement of vanes or blades which act on the air. This system will discuss two types of fans, centrifugal and axial fans (figure 1).

Figure 1. Fans

Fan Performance A fans performance is determined using an apparatus similar to the one shown in figure 2. The fan is connected to a long piece of straight duct with a throttling device at the end. The throttling device is used to change the air resistance of the fan duct system. The fan is operated at a single speed and the power applied to the fan shaft is measured by a device called a dynamometer. A pressure measuring device such as manometer is used to measure the velocity pressure. Velocity pressure is the difference between the total and static pressures. The test is first conducted with the throttling device removed. This operating point is called wide-open airflow. With no resistance to airflow, the pressure generated by the fan is velocity pressure only. It takes some length downstream of the fan for the velocity pressure to convert to static pressure. Because the distance at which the manometer is located is very short, the static pressure is negligible as the duct length is very short. Virtually all of the pressure is velocity pressure. The throttling device is then put in place and progressively moved in precise intervals toward the closed position. The pressures are recorded at each throttling device position. When the throttling device is fully closed, only static pressure is being generated by the fan because it cannot produce airflow. This point is called the blocked-tight static pressure.

Figure 2. Testing Method



Simplifying Bernoulli’s Law, we find the relationship between velocity pressure and velocity is:

The manometer from the fan performance test gives an accurate measure of velicty pressure. Knowing the test condition air density, we can then find the actual airflow. This point can then be plotted on a chart that has static pressure on the vertical axis and airflow on the horizontal axis (figure 3).

Figure 3. Operating Point



The data is measured at each interval from wide-open airflow to blocked-tight static pressure and plotted for a given rotational speed or RPM (figure 4).

Figure 4. Constant Speed

A smooth curve is drawn between the points. The curve graphically illistrates the fan peformance at a constant rotational speed. The curve extends from wide-open airflow to blocked-tight static pressure.

Figure 5. Constant Speed Curve



Fan Laws, also known as Affinity Laws, govern the performance of geometrically similar Centrifugal fans. The fan law that governs the relationship between fan speed and airflow is:

The fan law that governs the relationship between fan speed and static pressure is:

The fan law that governs the relationship between fan speed and fan brake horsepower is:

Next, we apply the fan law to plot the fan performance at several pratical fan speeds (figure 6).

Figure 6. RPM Curves

Finally, using the measurements from the dynamometer and the fan laws, curves can be calculated and plotted to



represent the fan’s power consumption at each operating condition (figure 7).

Figure 7. BHP Curves

When a fan approaches the blocked-tight static pressure condition, a period of instability in encountered. This condition is referred to a surge. Sure occurs when the airflow being moved by the fan falls below an amount required to sustain the static pressure difference between the inlet and outlet of the fan. When this state occurs, the airflow flow backwards through the fan wheel. This surge quickly re-enables the fan to re-establish airflow in the proper direction. This fluctuation of the direction of airflow causes the fan to surge. This surge can result in vibration, excessive noise and potentially damage to the fan.

Figure 8. Surge

Plotting the condition at which surge occurs results in a curve as shown in figure 9. Anything to the left of the curve is the area in which surge occurs and should be avoided during fan selection.



Figure 9. Plotting Surge

Often, fan manufacturer will display the fan data in tabular form (figure 10). This tabular data allows the designer to quickly and accurately determine fan performance.

Figure 10. Fan Curve

Ultimately, the goal is to determine the interaction of the fan within a ducted system. For a fixed quantity of airflow, an air distribution system imposes a certain amount of frictional resistance to the passage of the air within the system. This resistance is the sum of the pressure losses experienced as the air passes through the ductwork, dampers, filters, coils, supply air diffusers, return air grilles, etc (figure 11). This is the resistance that the fan must overcome to move the air within the system.



Figure 11. System Resistance

For example, assume that a system’s design static pressure drop is 1.6” w.g. at 720 CFM. If we plot the point, it will result in a point as shown in figure 12.

Figure 12. Plotting Operating Point

For any given system resistance, there is an interaction between the airflow and system static pressure given by the relationship:



Solving for airflow, we find:

If we plot airfow as a function of static pressure given the relationship above, a curve representing the system resistance can be determined (figure 13). This curve is called the system resitance curve.

Figure 13. System Resistance Curve

Finally, we can overlay the system resitance curve on a fan curve (figure 14). The intersection predicts the airflow and static pressure at which the fan and system will balance (point A). If the installed system resistance is different from that assumed during the design process, the fan and system operating point will not be as intended during design. This results in the fan and system balance point shifting to a point that has a higher or lower airflow, static pressure, and input power.



Figure 14. Overlay of System Resistance Curve with Fan Curve

First, lets analyze a case where the air resistance through the system is greater than anicipated. In this scenario, the system resistance curve shifts from operating at point A, to a situation where the system resistance curve shifts to point B. This results in a lower system airflow. To compensate for the loss of airflow, the fan speed should be increased to point C. This results in an increase in airflow and a consequential higher delivered fan static pressure. The fan must generate more static pressure to deliver the original design airflow, requiring more power than origainally intended.

Figure 15. Increased System Resistance



Conversely, if the system resitance is lower than anticipated, the system resitance curve drops to point B. This resutls in a corresponding increase in airfow. To compensate for the increase in airflow, the speed should be reduced to point C. Dropping the speed to point C decreases the airflow back to the intended design airflow. The fan delivers less static pressure and a lower power than oriingally projected.

Figure 16. Decreased System Resistance

In these examples, it was possible to overcome for the inaccuracies in estimated system resistance through fan speed adjustment. However, the actual fan operating points fell at conditions other than designed. It may be wise to re-evaluate the fan selection. It is possible that another fan size would perform more efficiently at the amended system conditions. In order to properly evaluate fan selections, it is necessary to assess fan efficiency. One such method used to express the efficiency of a fan is static efficiency. Static efficiency expresses the percentage of input power that is achieved as useful work in terms of static pressure. Fan static efficiency is:

For air, the equation above becomes:

The equation above remains unchanged with changes in air density.



Fan Types Several types of air movement fans are commercially available today. A centrifugal fan the airflow enters the center of the fan from the side and follows a radial path through the fan wheel (figure 17). There are three primary types of centrifugal fans, each distinguished by the type of fan wheel used. Those three types are forward curved (FC), backward inclined (BI), and airfoil (AF) type fan wheels.

Figure 17. Centrifugal Fan

A forward curved (FC) fan wheel has blades that are curved in the direction of airflow. FC fans are operated at relatively low speeds and are used to deliver large volumes of air against relatively low static pressures (figure 18). The inherently light construction of the forward curved fan wheels do not permit this wheel to be operated at the speeds needed to generate high static pressures.

Figure 18. FC Fan

The curved shape of the FC blade imparts a forward motion to the air as it leaves the blade tip. This, together with the speed of wheel rotation, causes the air to leave at a relatively high velocity (figure 19). The static pressure produced by a fan is a function of the forward motion of the air at the blade tip. The FC fan can perform, within its airflow and static pressure range, at lower rotational speeds compared to other fan types.



Figure 19. FC Fan Blade

The application rage of the FC fan is from 30% to 80% of wide-open airflow (figure 20). Selecting the fan to the left of 30% wide open airflow may place the fan in surge or instability. Above 80% wide open airflow typically produces noise and inefficiency. A FC fan generally experiences a maximum static efficiency of 50-65% and that selection occurs just to the right of the maximum static pressure point on the fan curve.

Figure 20. FC Fan

FC fan input power lines generally slope to a much higher degree than the fan speed curves. Because of the slope of the lines, if the system resistance were to drop (moving the system resistance from point A to point B), the fans input power would rise (figure 21). This could possibly overload the motor. Because of this characteristic of a forward curved fan, the FC fan is referred to as an overloading type fan.



As with all fan types, the FC fan can exhibit unstable operation, or surge. However, since FC fans are used typically in low speed and low static-pressure applications, many small FC fans can operate in surge without noticeable noise and vibration. This generally makes FC fans more forgiving with regards fan selection.

Figure 21. Overloading Characteristics

The second type of centrifugal fan is the backward inclined (BI) fan. BI fans utilize blades that are slanted away from the direction of wheel rotation. BI fans operate at higher speeds than FC fans. BI fans are generally a more rugged fan compared to FC fans. This rugged construction makes them suitable for moving large volumes of air in higher static pressure applications (figure 22).

Figure 22. BI Fan

The angle of the backward inclined blade causes the air that is leaving the wheel to be bent back against the direction of



rotation. However, the speed of wheel rotation causes the air to assume a velocity and direction as shown in figure 23.

Figure 23. BI Fan Blade

Comparing the operation of a BI fan to that of an FC Fan at identical rotational velocities (RPM), the BI fan results in a lower comparable leaving air velocity due to the difference in slope of the air leaving the blade tip relative to the rotation of the wheel. This difference in relative slope in the direction of wheel rotation decreases the overall air producing capacity of the BI fan wheel (figure 24). Despite the decrease in relative airflow between the two wheel types, the input power requirement of the BI fan is less, making it a more efficient selection. This higher speed demands that the BI fan be built with a larger shaft and bearing. In addition, it places more importance on proper fan balance.

Figure 24. FC versus BI Fan Blade

The application rage of the BI fan is from 40% to 85% of wide-open airflow (figure 25). Selecting the fan to the left of 40% wide open airflow may place the fan in surge or instability. Above 85% wide open airflow typically produces noise and inefficiency. A BI fan generally encounters a maximum static efficiency of 60-75% and that selection occurs at approximately 50% wide-open airflow. Generally speaking, the magnitude of the surge experienced by BI fans is greater than that experienced by a similar FC fan. For this reason, it is very important not to allow a BI fan to operate in the surge reason or fan and system damage may result.



Figure 25. BI Fan Application Range

In contrast to the FC fan, the fan brake horsepower lines are nearly parallel to the fan speed curves. If the system resistance were to drop moving the system curve from point A to point B, the fan brake horsepower would increase by relatively small increment (figure 26). Because of this characteristic of a backward inclined fan, a BI fan is referred to as a non-overloading type fan.

Figure 26. Non-Overloading Fan



A enhancement of the BI fan types involves changing the shape of the blade from that of a flat plate to that of an airfoil shape, similar to that of an airplane wing. The airfoil blade improves the aerodynamics and reduces the turbulence of the fan blade. In addition, an airfoil blade actually creates lift similar to that of an airplane wing (utilizing Bernoulli’s principle), increasing the overall efficiency of the fan wheel. This increases the static efficiency and decreases the overall sound levels produced by the fan. Airfoil (AF) fans display performance characteristics that are fundamentally the same as those of the flat-bladed BI fan.

Figure 27. AF Fan

The application rage of the AF fan is from 50% to 85% of wide-open airflow (figure 28). Selecting the fan to the left of 50% wide open airflow may place the fan in surge or instability. Note that the surge region for an AF fan is considerably larger than that of a comparable FC fan. Thus, considerable care must be taken when selecting AF fans to avoid surge. Like BI fans, AF fans are considered non-overloading type fans. Static efficiencies as high as 86 percent can be realized with airfoil fans. Because surge occurs at a higher airflow, the magnitude of the surge characteristics of the airfoil fan is greater than that of the FC or BI fans. In addition, care should be taken when considering the minimum operating airflow on VAV boxes when selecting AF fans. If the minimum set points are too low, the fan may encounter surge at lower airflows.

Figure 28. AF Fan Application Range



A variation of the AF fan is the un-housed plenum or plug fan (figure 29). A plug fan utilizes an un-house airfoil fan wheel. Air is drawn into an inlet cone and discharged into an open plenum. The fan wheel pressurizes the plenum surrounding the fan. This allows the air to discharge in multiple directions without concern for outlet system effects (discussed later in the clinic). This fan save space by eliminating straight sections of duct (required for balancing) or turns later in the ductwork by allowing for multiple ducts in different directions.

Figure 29. House Plenum Fan

In contrast to a centrifugal fan, an axial fan passes airflow straight through the fan in the direction of the fan shaft. Three axial fan types exist; propeller, tubeaxial and vanaxial fans. Propeller fans are suited for high volume, very low static pressure generating capability. Compared to centrifugal fans, airfoil fans experience a reduced variation in airflow capacity with changes in air density. Tubeaxial and vaneaxial fans are essentially propeller fans mounted in a cylinder. The main difference from a propeller fans is that they utilize vane type straighteners to remove much of the swirl from the air and straighten the airflow path. These vanes improve efficiency, reduce turbulence and reduce the sound generated from the outlet of the fan. The higher efficiency associated with vaneaxial fans make them suitable for larger volumes of air at higher static pressures.

Figure 30. Vaneaxial Fan



The application rage of the vaneaxial fans is from 60% to 90% of wide-open airflow (figure 28). Selecting the fan to the left of 60% wide open airflow may place the fan in surge or instability. Vaneaxial fan brake horsepower lines are virtually parallel to the fan speed curves. This makes a vaneaxial fan a non-overloading type fan. A vaneaxial fan generally encounters a maximum static efficiency of 70-80%. The airflow and static pressure ranges for vaneaxial fans are very similar to BI and AF fans. Compared to centrifugal fans, vaneaxial fans typically have lower sound levels at the higher octave bands, making them an attractive alternative for sound sensitive applications. Higher frequency sounds are easier to attenuator than low frequency sounds.

Figure 31. Vaneaxial Fan Range

The selection of fan type for a particular application should be based on system size, space, application consideration (constant volume vs VAV), sound and availability. The FC fan is best suited for small applied systems requiring less than 20,000 CFM and 4” w.g. total static pressure. FC fans are also very well suited to VAV applications requiring discharge dampers. Because a VAV system can often operate tolerably within the surge region, discharge dampers allow the fan a way to unload while providing energy savings. Often, FC fans with discharge dampers are used for space static pressure control. For sound sensitive systems or systems requiring more than 20,000 CFM or 4” w.g. total static pressure, BI or AF fans should be utilized. The inherent design which utilizes larger fan sizes, more efficient fan blades and larger motors can result in increased system efficiency and considerable energy savings. Finally, for applications in which space is the primary consideration, the vaneaxial fan may be the best solution. The straight through airflow design of this fan permits it to be installed in limited space applications. In addition, if sound is a critical design concern, vanaxial fans may be the best solution.

Forward curved (FC) o Lower airflow, lower static pressure, lower first cost o VAV Discharge Dampers (space pressure control)

Backward inclined (BI) or airfoil (AF) o Higher airflow, higher static pressure, higher efficiency o Lower Sound

Vaneaxial o Limited space o Low Sound



Fan Capacity Control The majority of the discussion regarding fan operation focused on the fan performance at a single point. This was found to be the intersection of the system curve with the intersection of the fan performance curve for a constant volume system. A constant volume system, by definition, is a system that provides a constant volume of airflow to the space while varying the discharge temperature as a function of load. In contrast, a variable volume system (VAV) controls the environment by varying the volume of airflow delivered to the space while maintaining a fixed discharge air temperature (figure 32). VAV systems place additional demands on the fan performance with regards to fan capacity control. In a VAV system, the quantity of air being delivered to the space is controlled by an airflow modulating device (a VAV damper) that is contained within a terminal unit. The device is controlled by a thermostat which varies the quantity of air required to condition and meet the space load. As the device modulates, the system resistance changes.

Figure 32. VAV System

As the system resistance changes, the system curve shifts (figure 33). Thus, in a VAV system, the system no longer operates at a single point, but at a range of operating points.

Figure 33. System Resistance Changes



The first and simplest method of fan capacity control is called “riding the fan curve.” This method does not involve any direct method of capacity control at the fan. Rather, as the VAV system experiences a change in resistance at the VAV terminal (from point A), the fan modulates along a fixed sped curve to the new operating point (point B). Because the fan “rides up” on the fixed speed curve, the system can balance at a new fan operating point. The new operating point B is at a new lower airflow and higher static pressure.

Figure 34. Riding The Fan Curve

While this method of fan modulation can technically be used with any fan type, it is most applicable using a FC type fan. First, an FC type fan has a wider operating range than either a BI, AF or vaneaxial fan. Second, an FC type can operate in the surge region. Finally, an FC type fan will operate at a lower fan brake horsepower as the system “rides up” on the fan curve. This is because an FC type fan is an overloading type fan.

Figure 35. Power Reduction



If an FC type fan is selected for a VAV system and is to operate by “riding the fan curve,” it is advisable to select the fan diameter two to three sizes smaller than one would for a constant volume system. This allows the fan additional room to operate before surge in encountered. FC fan are also often chosen for packaged rooftop units which employ discharge dampers to control space pressure. A FC fan which utilizes discharge operates in a similar manner to a FC fan utilized in a VAV system. Finally, VAV dampers can be applied to constant speed air handling or packaged unit system which employs FC fans. As a rule, so long as the total volume being delivered by the VAV terminal units does not exceed 20-30% of the total system airflow, the constant volume zones will see only a slight change in overall airflow at the system modulates in response to the VAV terminal zones. In addition, the FC fan should be able to run over the range without adverse effects such as excessive vibration or noise. However, “riding the fan curve” is used most successfully when the systems airflow modulation range is relatively small. If the fan is required to modulate over a wider range of operating conditions, the increased pressure experienced at the airflow terminal units may cause over-pressurization. This will result in a greater than anticipated space airflow and potential noise problems. Due to this issue, some form of system static pressure control is generally required. A VAV system must overcome the fixed pressure required for proper operation of the VAV terminals and diffusers. In addition, a VAV system must overcome a second variable component. The variable component is the amount of static pressure required to overcome the system pressure losses due to the filters, ducts, fittings, coils, dampers, etc. These variable components must be overcome at various airflows as these losses are a function of airflow. In order to ensure that these losses are overcome in a VAV system, a control loop is employed. The input to the control loop is the static pressure being sensed at some point in the system (traditionally 2/3’s of the way down the longest duct run). The controller conditions the input variable by comparing that value to a known setpoint and outputs a signal to modulate the capacity of the fan (figure 36). The fan capacity is adjusted and the set point is maintained at the location of the system sensor.

Figure 36. VAV Fan Control

To demonstrate the operation of a VAV system, refer to figure 37. As the system curve changes from point A to point B, the system resistance changes. If we were to draw a smooth curve from the sensor setpoint (which represents the setpoint at zero airflow) to the original point A, we get a fan modulation curve. The intersection of the fan modulation curve with the new operating system curve at point B represents the change in airflow required for the system to overcome the static pressure required to offset the variable system losses (point C)



Figure 37. Fan Modulation Curve

The equation for the system modulation curve is:

Where: CFMd = airflow at design SPd = static pressure at design SPc = sensor control setpoint Next, we will discuss the three methods used to control the capacity of a fan. They are discharge dampers, inlet vanes, and fan speed control.

Methods Of Fan Capacity Control

Discharge Dampers

Inlets Vanes

Fan Speed Control Discharge dampers match the airflow and static pressure supplied by the fan with the airflow and static pressure required by the system. The do this by adding a static pressure loss imparted by the discharge dampers just downstream of the fan (figure 38).



Figure 38. Discharge Dampers

A VAV system with discharge dampers is demonstrated in figure 39. The system design operating point is denoted by point A. As the system resistance changes, the fan begins to ride up the fan curve toward point B. Point B is the intersection of the new system resistance curve with the fan speed curve. The system tries to balance this increase in static pressure and modulates the discharge damper along the modulation curve to point C (the intersection of the new system curve with the modulation curve. The operating point at point C represents the static pressure after the discharge dampers. The actual static pressure at the outlet of the fan is represented by point D. That is because the fan simply rides the fan curve. The difference in static pressure between point D and point C is the static pressure drop across the dampers. Because the fan employs the method of “riding the fan curve,” discharge dampers should only be used with FC type fans.

Figure 39. Discharge Dampers

The second method of fan capacity control is inlet guide vanes. Inlet guide vanes (figure 40) pre-swirl the inlet air in the direction of wheel rotation as it enters the fan. This pre-swilling effect decreases the “bite” of the fan into the incoming airflow, thus decreasing its velocity imparting component upon the airstream. Inlet guide vanes essentially re-configure the shape of the fan speed curves. By re-shaping the fan speed curves, the system is able to unload. As the cost of variable frequency drive has continued to decrease, inlet guide vanes have become increasingly rare.



Figure 40. Inlet Vanes

Finally, fan speed control modulates fan capacity by varying the speed of the wheel rotation. This is commonly accomplished using a variable speed device on the fan motor, such as a variable-frequency drive. To demonstrate the operation of a fan speed control, see figure 41. Like discharge dampers, the controller senses a change in static pressure, it climbs the fan speed curve (point B), and the system adjusts to the new operating point C at the intersection of the new system curve with the fan speed curve. However, because the fan can adjust the fan speed curve, point C is the actual static pressure at the discharge of the fan. Not only does this inherently save fan energy, but the modulation range is vastly increased at part load. This is because the modulation curve crosses the fan surge curve at a much lower airflow.

Figure 41



Application Considerations Fan capacity control always requires a minimum of two elements. The first is a static pressure sensor located somewhere in the system. The second is a method of fan capacity control (VFD, discharge dampers, etc). A controller compares the static pressure sensor to its setpoint and adjusts the capacity of the fan to maintain the setpoint. The most common method of fan capacity control involves mounting the sensor two-thirds of the way down the longest duct run (figure 42). The controller is set to maintain the pressure corresponding to that location in the duct system at design airflow conditions. In larger system where the location of the sensor may not be obvious, multiple sensors may be required.

Figure 42 Two Thirds Method

However, this method is arbitrary and often leads to duct over pressurization. For example, it is impossible to tell if we are providing enough static pressure to ensure adequate airflow to all the zones. Conversely, we may be providing too much pressure, over pressurizing the ductwork and wasting valuable fan energy. The latter scenario is the more common. In addition, the fan modulation curve will still cross into the surge region at lower airflow. The higher the static pressure control setpoint, the minimum airflow at which surge occurs will increase. One clever solution which enables designers to overcome the limitations imposed with the “two-thirds” method is a controls strategy called critical zone reset. The critical zone reset controls strategy is employed by polling all of the VAV terminal units in the system at regular intervals. The VAV terminal with the highest damper position (% wide open) is tagged as the critical zone. The critical zone damper position becomes the controlled variable. A static pressure sensor, located at the discharge of the supply fan, is reset until the critical zone damper position reaches a setpoint of 95% (figure 43). In order to implement this control scheme, the building automation system must be able to poll and read the damper positions at all the VAV terminal units. This method is far less arbitrary than the two-thirds method. The control scheme ensures that none of the terminal units ever reach 100% wide open. If they did, we know that we are starving the zones for airflow and that the static pressure in the system needs to be increased. In addition, we are optimizing the energy consumption of the fan. We are never over pressurizing the system, as is common problem associated with the two-thirds method. The critical zone reset method can save as much as 15% of the total system fan energy consumption. In addition, critical zone reset can virtually eliminate the tolerance at which surge occurs. Because the static pressure is continually reset, at lower airflows the static pressure is generally reduced. This reduction in the static pressure control setpoint at lower airflow resets the fan modulation curve lower. Lowering the fan modulation curve in turn reduces or eliminates the possibility of encountering surge.



Figure 43. Critical Zone Reset

Fan arrays (figure 44) are becoming an increasingly popular option available to designers considering built up and custom air handling systems. Fan arrays offer redundancy, improved acoustical characteristics and improved performance at part load.

Figure 44. Multiple Fan Array

In temperature critical applications, if a single fan fails, it generally takes down the entire system it serves. By offering as few as two delivery fans, the system designer will have as much as 76% of the system design airflow if a single fan should fail (figure 45). A system with six fans will offer as much as 96% of design airflow in the event a single fan fails. Should a loss of airflow be unacceptable should a fan fail, it is relatively easy to slightly oversize the fans in the arrays such that a loss of airflow is not experienced. Each fan array can be fitted with either blockoff plates or inlet isolation dampers for fan isolation in the event a fan fails.

Figure 45. Airflow Redundancy

Additionally, VAV systems employing non-reset static pressure control methods (2/3rd longest duct run pressure control for example), fans can be turned off as the array approaches surge. Thus, air handlers employing fan arrays can be



turned down to a far greater effect than with a single fan alone, allowing for greater controllability. However, should a static pressure reset strategy be employed (similar to that shown in figure 43), fan should always be unloaded in such a way takes advantage of the fan affinity laws. Recall as a fan speed is reduced by 50%, brake horsepower is reduced by 87.5%. However, turning off fans unloads fan power in a much more linear fashion. Consider a system with (2) 5 HP fans. At 50% airflow, by unloading both fans to 50% speed, we reduce the fan consumption to 1.25 hp (figure 46).

Figure 46. Unloading Multiple Fans In An Array

Alternately, if turn a fan off, the fan affinity laws only apply to a single fan 5 hp fan. Recall that a single fan operates at about 20% greater airflow than with a system operating with two fans in series. Thus applying the fan affinity laws to a single fan, our new fan power consumption is more than double than that with unloading the entire array simultaneously (figure 47).

Figure 47. Turning Fans Off In An Array

Keep in mind, with any fan array, there becomes a point of diminishing returns. Recall that with a fan array consisting of six fans, should a single fan fail, the system produces 95-96% of its design airflow. In modern refrigeration systems, the goal is generally to reduce the number of moving parts and thus reduce the chance for failure. Thus systems with greater than six fans generally simply increases the chances of system failure and maintenance. Additionally, larger quantity fan arrays generally use smaller nominal hp motors. Yet, the motor efficiencies associated with smaller horsepower motors are typically several percent lower than their larger equivalents. Earlier in the clinic, we discussed the effect of the air resistance through the system being greater than anticipated. System effects are a common cause of just such a scenario. System effects can be attributed to the turbulence at the fan inlet, fan outlet or non-uniform air distribution in the downstream ductwork. This system effect has an effect very similar to that of shifting the system resistance curve up and to the left (figure 48). Instead of the system operating at point A, the system resistance curve shifts to point B.

Figure 48. System Effects



System effects related to the fan inlet (figure 49) or fan outlet (figure 50) generally occurs when the air is not allowed to establish a uniform velocity profile.

Figure 49. Inlet Disturbances

Figure 50. Outlet Disturbances

The third source for system effects occur when there is not enough straight duct at the inlet our outlet of the fan. It takes 2.5 to 10 duct diameters for the air velocity profile, as it discharges the fan, to become uniform (figure 51). If the velocity profile entering or leaving the fan is not uniform it will add a system effect.

Figure 51. Developing a Uniform Velocity Profile



Fan noise is generally one of the largest contributors to adverse noise in an indoor environment. Several design guidelines should be considered when designing fans in air transport systems with regards to noise. Those guidelines are:

Optimize fan and air-handler selection for lowest overall sound

Select fan to operate away from surge

Minimize system effects

Use low pressure drop duct fittings (follow SMACNA recommendations)

Avoid rectangular sound traps, if possible. If required, keep static pressure below 0.15” w.c.

Use adequate vibration isolation First, the correct fan should be selected for the application. In lower airflow, lower static pressure applications, an FC fan might be the correct application. In higher airflow, higher static pressure, a BI or AF fan might be appropriate. For tight space applications, a vaneaxial fan might be the right choice. Also consider the nature of the sound data. Lower frequency octave bands are more difficult to attenuate that higher frequency octave bands. Select the fan to operate safely away from surge. The one exception to this rule is when the FC fans are the best fit for the application. FC fans can operate somewhat safely within surge. Always minimize system effects. Keep at least 2.5-5 duct diameters downstream of the fan before making bends or takeoffs. Use low pressure drop fittings as recommended by SMACNA. Keeping the pressure in the system minimized will reduce the input power and thus noise produced by friction in the system. Any friction losses in the system must be overcome by the fan. Some percentage of the energy input by the fan to overcome those losses will be converted to noise. It is for this reason that sound traps are commonly misapplied. The additional input energy and thus noise produced to overcome the sound trap often does not overcome the noise removed by the trap. Unless traps are kept to lower pressure drops, they should be avoided. Lower pressure drop traps, if applied correctly, can improve the acoustical performance of systems which would otherwise be acoustically unacceptable. Finally, use adequate vibration isolation for the fans. Proper isolation help prevent vibration transmission directly to the supporting structure. Vibration to the supporting structure is often conveyed into the conditioned space. The Air Movement and Control Association (AMCA) establishes testing procedures and rating standards for air-moving devices. AMCA also certifies performance and labels equipment through programs that involve random testing of a manufacturer’s equipment to verify published performance (figure 52). The overall objective of AMCA Standard 210 (also known as ASHRAE Standard 51), titled “Laboratory Methods of Testing Fans for Rating”, is to promote consistent testing methods for fans. The Air Conditioning & Refrigeration Institute (ARI) has similar standards for fans that are a part of equipment such as fan coils, central-station air handlers, and unit ventilators.

Figure 52. AMCA

• The Air Movement and Control Association is a trade association for the fan industry

– Providing assurance & reliability of manufacturer’s published performance – Providing buyers with information on testing procedures – Verifying manufacturers performance ratings – Standardizing test methods



Fan performance data is published at standard conditions, or atmospheric pressure at sea level (0.075 lb/ft3). However, site conditions often vary from published data. For example, in Reno Nevada, the atmospheric pressure is 0.064 lb/ft3). When selecting fans at conditions other than sea level, we must properly account for site conditions. The air density ratio is defined as:

Where: ρa = desnisty at actual conditions ρs = desnisty at actual conditions Before entering a fan performance chart or table, we should calculate the equivalent pressure at standard conditions. The pressure at standard conditions is:

Where: ρa = density at actual conditions ρs = density at standard conditions Pa = Pressure at actual conditions Ps = Pressure at standard conditions Having determined the static pressure at sea level, we can determine airflow, rotational velocity (RPM) and brake horse power (bhp) at standard conditions from seal level fan tables. Rotational velocity does not change with changes in density. Thus, rotational velocity is:

Where: RPMa = rotational velocity at actual conditions RPMs = rotational velocity at standard conditions Like rotation velocity, airflow should be considered constant with changes in density. Thus, airflow is:

Where: CFMa = airflow at actual conditions CFMs = airflow at standard conditions Finally, BHP can be calculated at actual conditions. Applying the fan laws:

Where: BHPa = brake horsepower at actual conditions BHPs = brake horsepower at standard conditions



ρa = density at actual conditions ρs = density at standard conditions The equations above allow us to take the tabulated data at sea level conditions and re-rate for actual conditions. These equations become very useful for designers that must Engineer projects at non-standard conditions.

0 title page - submittal(standard) rev2.pdfsimplifying bernoulli’s law, we find the relationship...

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