Tag: airflow

Many installers and service technicians know how to read and use a manufacturer fan table, but this is a quick review with a few extra tips for newer techs. It’s also a good reminder to senior technicians how this easy-to-use practice can also be easily abused.

At installation, it is imperative to the performance and longevity of the appliance to set up airflow properly. A practical way to do this is utilizing the manufacturer-supplied fan tables found in every installation manual. Here’s a review on how to set up airflow on a new system:

  1. Determine your target airflow (The national average is 400cfm/ton. However, in a dry climate, design airflow may be 450-500cfm/ton, and in a humid climate, airflow is typically designed at 350-300cfm/ton.)
  2. Set your fan speed (choose the speed tap, or set the dip switches)
  3. Verify the equipment and duct work is clean, and all packing materials are removed from inside the appliance (yes, this gets missed sometimes)
  4. Run the system in order to achieve the test conditions in which the Fan Table was created (Fan Table airflow readings are only valid if the field conditions match as closely to the lab conditions as possible; i.e. wet coil, dry coil, with or without heat strip kits, etc.)
  5. Measure Total External Static Pressure (see how to measure TESP below)
  6. On the fan table, find the model matching the equipment you have, and locate the speed tap being used
  7. Match the real-time static pressure with the fan table
  8. The point at which both the TESP column and Speed Tap row meet is the corresponding estimated airflow.
  9. Make any adjustments to ductwork or fan speed in order to achieve the target airflow (This is made easy if ductwork is slightly oversized and installed with manual dampers on the supply.)


For servicing, techs may use the fan table method as a quick and dirty way of verifying airflow without extensive and time-consuming testing. This can be acceptable, but only if the following conditions are met:

  1. The equipment and ductwork are clean (This includes making sure the filter has been replaced)
  2. The equipment has been benchmarked once before (Without a reference, the fan table cannot be relied upon as an accurate representation of estimated airflow.)
  3. The equipment is running as closely to the documented lab conditions as possible. (But even then, how wet is “wet”?)

Static pressure readings stand alone as a valuable measurement during a service call, and TESP can inform a technician whether more extensive testing is required. But if the equipment has never been worked on by you, or your company did not install the equipment, the fan tables will not be useful until a full-system commissioning has been completed. 

Carrier FB4CNF Installation Manual

Another important tip is to always keep the return static pressure below 0.4” w.c. According to many manufacturers’ literature, a return static pressure of 0.4’ w.c. or higher can potentially result in water from the primary drain pan being picked up and thrown around inside the cabinet area, and sometimes into the ductwork. 

It is important to understand static pressure measurement is NOT a measurement of airflow. This is where many technicians abuse this method. Static pressure is just that: a measurement of pressure in reference to the space outside the ductwork. Based on lab testing conditions, a manufacturer is able to determine the airflow of a system under a known resistance. Static pressure is used as a proxy to estimate airflow, but this method is only as good as the conditions in which it is applied. Static pressure readings are air density dependent, so zeroing a manometer in a cold, dry attic, then inserting the probes into a humidified, warm duct system will adversely affect the accuracy of your measurements. This method is also heavily dependent on how detailed the manufacturer fan table is. An example of a good fan table would be one that lists the equipment model, if the unit was tested under wet or dry conditions, if heat strips were installed during testing, and any corresponding wattage/rpm determinations under given conditions. 

Carrier FB4CNF Installation Manual

The difficulty with using Fan Tables as a way to measure airflow is realizing the resistance across the equipment is dynamic, and will likely change many times over the course of a test (the coil may get wetter as it is loaded with latent heat, the coil will become dirty over time, etc.) Measuring actual airflow is difficult to do, but static pressure measurements are still very valuable, and are a good way to determine if a problem exists and on which side of the ductwork it exists (supply or return). 

A great product for measuring airflow in the field is the TrueFlow Grid by The Energy Conservatory. For more information on Airflow and Airflow Measurements, TruTechTools has an entire section of literature and webinars on the topic. Here is a video we recorded for them in 2017 regarding Static Pressure and Fan Tables:

— Kaleb

This is VERY in depth look at ACFM vs. SCFM and why it matters to airflow measurement from Steven Mazzoni… Thanks Steve!

Imagine your job is to figure out how fast baseballs were traveling before they hit a sheet-rock wall. The only method you have is to measure the depth of the dent left in the wall. Suppose at 60 mph, the ball leaves a ¼” deep dent. At 80 mph, it leaves a ½” dent, and so forth. No problem, all you have to do is measure the dents and you can derive the speed (velocity).

But it’s more complicated than that. You discover some of the balls are a bit lighter than others. Otherwise, they are all identical. What does this mean? The lighter balls leave behind a shallower dent than the heavy ones, even if they were traveling at the exact same velocity before hitting the wall. Obviously, more is needed than just the depth of the dents. The weight of the balls must also be factored in. Suppose you are able to weigh the balls in addition to measuring the depth of the dent they leave. You come up with an equation that factors in the ball’s weight and depth of the dent and solves for its velocity.

Something similar to the baseballs is happening when we measure airflow. To determine the airflow

(cfm, or ft3/min) in a duct, all we need to find out is its average velocity (ft/min) and the duct area (ft2). Measuring the air’s velocity (duct traverse) is the tricky part. A pitot tube & manometer measure the speed of the air flowing in a duct. At a faster speed, or velocity, more force is imparted to the column of water in the manometer. The pressure difference (velocity pressure or VP) is used to determine the air’s velocity, in feet/minute.

However, like the baseballs, air’s density isn’t always the same. Thus, the force it imparts to the column of water when traveling at a given velocity changes if it’s density changes. “Heavy” air will lift a column of water to a higher level (velocity pressure, in inches of water) on a manometer than “light” air will, even though moving at the exact same velocity. Thus, the velocity pressure and the air’s density must be factored in before we can determine it’s velocity.

What factors determine air’s density? Mainly its temperature and the barometric pressure. Warm air is lighter (less dense) than cold air. Air at higher barometric pressures near sea level is denser than air at lower pressures (high altitudes). Air’s moisture content also plays a minor role. Moist air (high humidity) at a given temperature is lighter than dry air at the same temperature.

The flow of air (volumetric) is usually expressed in cfm (ft3/min). To be more specific, actual cfm (ACFM) and standard cfm (SCFM) are used. ACFM & SCFM have been defined as follows:

Air is at “standard conditions” when it’s density is @ 0.075 lb/ft3. We can thus conclude a couple of key points. First, if the airflow measurement is taken at or near standard conditions, the ACFM and the SCFM will have the exact same value. Second, if the reading was taken on air at a significantly different density, ACFM and SCFM will have two different values.

Let’s work through an example duct traverse at a high elevation & temperature to show how to determine ACFM & SCFM. Suppose a 4-point duct traverse has been taken at the following conditions. A pitot tube was used to obtain velocity pressures (VP), but these have not yet been converted to velocity (ft/min). Let’s keep it simple and assume a 1.0 ft2 duct.

Elevation:4,000 ft
Barometric pressure:25.84”hg
Duct temperature:120 deg f
Duct area:12” x 12” = 1.0 ft2
Actual air density:0.059 lb/ft3
Standard air density:0.075 lb/ft3
Actual velocity pressure (VP) readings:0.020” wc

0.025” wc
 0.030” wc
 0.035” wc

Now, what do we do with these four velocity pressure readings? We need to convert them to velocity, using one of the equations below. The “4,005” equation is only valid for air at standard density. The “1,096” equation works at any density.

Here is where it gets interesting. Which density should we use to convert the VP readings to velocity, so we can then determine ACFM & SCFM? The actual density (0.059 lb/ft3), or standard density (0.075 lb/ft3)? We’ll explore 2 options.

  • Option 1: Calculate the actual average duct velocity using the actual density of the air measured.

Then multiply average velocity by the duct area in ft2. The result will be in ACFM.

Calculate ACFM Using Option 1:

0.020” wc =638 ft/min
0.025” wc =713 ft/min
0.030” wc =782 ft/min
0.035” wc =844 ft/min
Avg =744 ft/min

  • Determine SCFM for our example using one of these 2 methods:
    • Method A: Determine mass flow rate of the ACFM. From that, determine what volumetric flow at standard conditions would result in the same mass flow. The result will be in SCFM.
    • Method B: Multiply the ACFM by the ratio of the actual density to standard density. The result will be in SCFM.
      • Method A & B both result in @ 585 SCFM.
      • Option 2: Even though we realize the actual density at the traverse was not standard, calculate using standard density. Multiply by the area in ft2. Then take the result and apply a correction factor to determine ACFM & SCFM. o Calculate velocity & flow using the same VP’s from the non-standard density traverse, but using the standard density 4,005 formula:
0.020” wc =566 ft/min
0.025” wc =633 ft/min
0.030” wc =694 ft/min
0.035” wc =749 ft/min
Avg =661 ft/min

  • Is this 661 “cfm” the ACFM? No. Is it the SCFM? No. Obviously, it falls in between the 744 ACFM and 585 SCFM we calculated above. What is it then? It is a value that, when corrected, can get us to the true ACFM & SCFM.
    • Determine a unique correction factor for our example as follows. Notice the square root function:
    • Now what? Use this correction factor to convert the “uncorrected” 661 cfm to ACFM as follows:
    • Next, use the same correction factor to convert the “uncorrected” 661 cfm to SCFM as follows:

Conclusions: · Consider the type of instrument you are using to measure the differential pressure coming from a pitot tube. Velocity pressure readings from inclined manometers and simple differential pressure instruments will need the correct math applied. Electronic ones may be able to correct for local density and display the actual velocity.

  • Both Option 1 & 2 resulted in the same ACFM & SCFM values.
  • In Option 1, we used the actual local density to determine the actual average duct velocity and the ACFM. From the ACFM, we calculated the SCFM based on either the mass flow (Method A) or the ratio of actual density to standard density (Method B).
  • In Option 2, standard density was used to calculate a “reference cfm”. This reference cfm did not reflect reality, but was used to calculate ACFM & SCFM. A correction factor had to be calculated (square root of the ratio of the two densities) and used to convert the reference cfm to ACFM and SCFM. This method is similar to assuming all the baseballs are the heavy ones and calculating a reference speed based on that incorrect premise. Then the result must be corrected based on the actual weight of the baseball.
  • To avoid confusion, it seems best to use Option 1 along with Method B when working with air at non-standard conditions. At least then, the calculation gives you the ACFM directly, and SCFM can be calculated easily based on the ratio of the two densities. No other correction factors are needed.

Steven Mazzoni

HVAC/R Instructor

When you start talking airflow, it can get pretty in-depth pretty quick. There is a big gap between what is useful for the average tech to apply every day and the whole story so let’s start with the simplest part to understand, Static Pressure.

Static pressure is simply the force exerted in all directions within any contained fluid, or in this case air. This means it’s not the directional force of air moving or blowing (that is called velocity pressure), it is simply to force pushing out on the positive side of the air system and pulling in on the negative side.

In other words, it’s energy exerted or inward in all directions instead of in one direction like velocity.

Measuring static pressure helps a tech know whether or not the system has excessive resistance to air flow overall or at a particular point.

Static pressure is measured in inches of water column (“WC) and is the amount of pressure needed to displace one inch of water in a water manometer.


A Magnehelic is a brand name for a high-quality Dwyer analog pressure gauge that comes in many different scales. Many techs will already have a high-quality digital differential manometer (like the  Fieldpiece SDMN5) for reading gas pressure, which makes getting a separate Magnehelic largely unnecessary.

When using a manometer or a Magnehelic, you will first zero it out to room pressure (for a Magnehelic make sure it is level). Next place the negative side probe in the return side of the unit after the filter but before the blower and place the positive probe in the supply duct. Keep the negative side probe away from the side of the blower and insert the probes in as straight and square as possible. It is advised to use a static pressure tip like the one shown below to prevent air velocity pressure or air currents from interfering with the static pressure reading.

With a static pressure tip point the tip against the direction of airflow (points opposite the airflow) in both the return and supply.

DO NOT confuse a static pressure tip with a pitot tube. A pitot tube is designed to measure velocity pressure or total pressure (velocity + static = total)  NOT static pressure, and it will have an open end and two connection points.

Total external static pressure is return plus supply, positive plus negative and in general, you would like to see it be 0.5″ or less…

If you see 0.8″ or higher that is when you start to see trouble on most newer residential systems, but as always, each piece of equipment is different depending mostly on motor design. Whenever possible design your equipment / duct system so the result is 0.4″ – 0.6″ of total static (Once again talking general residential / light commercial here).

If you do find it to be high, then read the return and supply separately to see which is higher which is just a matter of removing the hoses to your manometer or Magnehelic alternately. Whichever reads higher is the greater cause of the issue.

I could keep going on this, but instead, I will just link to some more in-depth articles if you want to do more reading.

— Bryan

Epic airflow write up from Dwyer 

Measuring Airflow from TruTech

Troubleshooting Ductwork by ACHR News


Let’s use a bit of imagination for a minute.

Imagine you have two totally identical 3-ton systems. One of them is completely normal and the other has no fins at all on the evaporator coil.

They both have the same charge, airflow and compressor capacity. What will be different in terms of readings and performance of the one with no fins? Think about it yourself before moving on.

The fins attached to the tubing increase the contact time and turbulence of the air over the lower temperature coil surface

This increased air to metal contact time and surface area allows more heat to leave the air and enter the refrigerant. The pressure and temperature of the refrigerant in the evaporator coil (assuming it is changing state from liquid to vapor) is a function of the amount of refrigerant being moved through the coil (by the compressor) and the amount of heat entering the refrigerant from the air.

If the evaporator coil had no fins we would see these results in the system running with no evaporator fins compared to the normal system –

  • Lower suction & coil pressure because less heat is entering the coil
  • Lower Coil (saturated suction) temperature because of the lower pressure
  • Lower Delta T because the air has less contact time on the metal as it rushes right past
  • The Compressor moves less refrigerant because the suction vapor is less dense and the compression ratio increases
  • The system is less efficient because the compressor is moving less BTUs per Watt due to the higher compression ratios
  • Very little humidity is removed because even though the coil is very cold there is less contact between the metal and the air moving over it

Eventually the suction pressure / temperature would stabilize once the amount of heat being picked up in the coil and the reduced compressor capacity reached an equilibrium but by the time the tubing would likely be freezing over and the system would be VERY inefficient.

A coil with no fins has a high BYPASS FACTOR and a low CONTACT FACTOR

As a technician you will likely never calculate bypass or contact factors but you will see their impacts all the time like in the extreme thought experiment above. It is easier to wrap your head around the impact of bypass factor by considering the extreme edges of design with a coil with no fins being an extreme example of high bypass and a massively over-sized evaporator with very dense and deep fins being the opposite side of the spectrum.

If you have more surface contact between air and metal and when the air is moving at lower velocity (slower) the air will impart more heat to the coil and will get closer to the coil temperature.

If you have less surface contact (like no fins) and the air is moving at a higher velocity (faster) the air will impart less heat into the evaporator coil and will stay warmer when compared to the coil temperature.

The part that can make this tricky to imagine is once you realize that changing the bypass factor simultaneously impacts the difference between the air temperature and coil temperature AND changes the coil temperature itself because the coil temperature is partially dictated by the amount of heat entering the coil FROM the air.

Look back up at the bullet list of impacts that occur with a coil that has no fins. Now flip the results for an over-sized evaporator coil (assuming the air flow remains constant)

  • Higher suction & coil pressure because more heat is entering the coil
  • Higher Coil (saturated suction) temperature because of the higher pressure
  • Higher Delta T because the air has more contact time on the metal as it moves more slowly over the larger surface
  • The Compressor moves more refrigerant because the suction vapor is more dense and the compression ratio decreases
  • The system is more efficient because the compressor is moving more BTUs per Watt due to the lower compression ratios
  • More humidity is removed because even though the coil is warmer there is more contact between the metal and the air moving over it***

Now this last point about the humidity has a caveat, the coil still needs to be well below dew-point so this correlation between increased coil size and increased dehumidification  isn’t absolute. When you couple a larger coil with LOWER airflow you can get the best of both worlds by having lower bypass factors without driving the coil temperature up too far resulting in improved efficiency and latent capacity.

I’m not saying you should go out and put larger coils on everything, but as a design consideration you may want to take a good look at manufacturer expanded performance data when choosing a coil match and you may find an up-size gives you better performance even if you dial back the airflow a bit.

I know this may be a little bit of a brain mush concept but keep going back to the impact of the evaporator with no fins to help set your mind straight. It certainly has helped me.

— Bryan


Recommended Duct Velocities (FPM)

Duct TypeResidentialCommercial / InstitutionalIndustrial
Main Ducts700 – 9001000 – 13001200 – 1800
Branch Ducts600 – 700600 – 900800 – 1000

As a service technician, we are often expected to understand a bit about design to fully diagnose a problem. Duct velocity has many ramifications in a system including

  • High air velocity at supply registers and return grilles resulting in air noise
  • Low velocity in certain ducts resulting in unnecessary gains and losses
  • Low velocity at supply registers resulting in poor “throw” and therefore room temperature control
  • High air velocity inside fan coils and over cased coils resulting in higher bypass factor and lower latent heat removal
  • High TESP (Total External Static Pressure) due to high duct velocity

Duct FPM can be measured using a pitot tube and a sensitive manometer, induct vane anemometers like the Testo 416  or a hot wire anemometer like the Testo 425. Measuring grille/register face velocity is much easier and can be done with any quality vane anemometer, with my favorite being the Testo 417 large vane anemometer

First, you must realize that residential, commercial and industrial spaces tend to run very different design duct velocities. If you have ever sat in a theater, mall or auditorium and been hit in the face with an airstream from a vent 20 feet away you have experienced HIGH designed velocity. When spaces are large, high face velocities are required to throw across greater distances and circulate the air properly.

In residential applications, you will want to see 700 to 900 FPM velocity in duct trunks and 600 to 700 FPM in branch ducts to maintain a good balance of low static pressure and good flow, preventing unneeded duct gains and losses.

Return grilles themselves should be sized as large as possible to reduce face velocity to 500 FPM or lower. This helps greatly reduce total system static pressure as well as return grille noise.

Supply grilles and diffusers should be sized for the appropriate CFM and throw based on the manufacturer’s grille specs like the ones from Hart & Cooley shown above. Keep in mind that the higher the FPM the further the air will throw but also the noisier the grille will be.

— Bryan

I am in the midst of testing the accuracy and repeatability of different types of airflow measurements for techs in search of the most practical methods for different applications.

One commonly taught method for measuring airflow is the temperature rise method where you use a heat source that produces a set # of btu/h such as heat strips and using the sensible heat air equation you can “easily” calculate the CFM being produced by the equipment. Let’s use this specific example to illustrate the challenges in getting a truly reliable measurement.

The Typical Equation used for a fan coil with electric heat is –

CFM = (Volts x Amps x 3.41) / (1.08 x Delta T)

  1. Measured Heat Strip Volts x Amps gives you the Watts (This will generally be a reasonably accurate measurement depending on the accuracy of the meter)
  2. 3.41 is a constant watt to BTU conversion
  3. The 1.08 is a combination of factors 1 CFM x 60 minutes per hour = 60 CFH x .075 pounds per cubic foot (Standard Air) = 4.5 x 0.24 Specific heat of air = 1.08
  4. The Delta T is only as accurate as the thermometer used and it’s placement in the air stream

The first trouble comes in when you realize that almost no air is “standard air” which is 70°, 0% RH air at sea level. This leaves us with a .075 lbs per cubic foot standard air # that isn’t very accurate at all. To see how far off it can be, take a look at this calculator

In the case of my test, I found that a 1.05 multiplier was more accurate than 1.08 based on my indoor air conditions during the test.

Before the test, I used a TEC TrueFlow meter to confirm the actual system airflow. The system was a 2-ton Carrier FV ECM air handler and both the fan charts and the TrueFlow confirmed a system airflow of 700 & 718 CFM respectively.

I turned on the 5KW electric heat and ran the system with electric heat only for 5 minutes and then calculated the BTU/h of the heat strips at 12,158. I took a Delta T between the return riser and the supply plenum 24″ above the fan coil with the same pocket thermometer, The delta T was fairly stable at 1°.

It doesn’t take a math major to figure out that a 1° delta T = WAY MORE AIRFLOW THAN 718 CFM

12,158 / 1 x 1.05 = 12,766 CFM

So what went wrong?

In this particular case, we realized that the evaporator coil was still lower temperature than the return air and there was still moisture on the coil and in the drain pan because it had been running in cooling mode before which was decreasing the temperature of the air inside the air handler. Once the system ran in heat another 15 – 20 minutes the delta T came up to 6°.

Still WAY too low!

12,158 / 6 x 1.05 = 1,930 CFM

Next, I moved the thermometer to the side and I was getting a 10° delta T, then to the back and all of a sudden I was reading a 17° delta T which was putting us right in the range, a little low actually (681 CFM). With the pocket thermometer, we were seeing an 11° variation depending on where I placed the probe!

So why was this occurring? Even at 24″ above the air handler the air had not fully mixed and there were areas defined areas of high temperature and low-temperature air in the supply plenum.

I gave up on the pocket thermometer and used a longer Testo probe and a K-type bead probe to get closer to the center of the air stream. I then took three measurements. One in the front, one on the left side and one in the back, added them together and then divided by 3, this gave an average of the supply air temperature from multiple points and the result was a calculation of 755 CFM which is still on the high side but definitely a usable figure.

I then tried it at different blower settings to see if this new averaging method worked under different air flow conditions, sure enough, it was within 10% of the factory fan charts and the TrueFlow.

Out of curiosity, I ran the system in cooling mode to see if I would get a similar level of variation and while I did see a few degrees difference it was only 2° due to the mixing in the blower after the coil.

For those of you who work on gas furnaces will coils on top you know how much the supply air temperature measurement can vary based on where you place your probe due to poor air mixing and radiant cooling near the coil, this is the same effect I was seeing above the heat strips.

Here are some good rules for accurate measurement in critical test circumstances –

  • Probe placement matters. When possible take readings 6′ away from the appliance and out of “line of sight” from the cooling/heating surface (Heat strips, heat exchanger, coil)
  • Use the same probe to take the return and supply readings to improve accuracy if the device isn’t highly accurate (more than +/- 1°)  
  • The speed of the reading matters, a K-type bead is not the most accurate measurement but the quick read they will help reduce the impact of radiant heat on the probe if you are “line of sight” to the heat source. 
  • When performing any temperature rise calculation the entire inside of the appliance must be the same temperature as the return air and completely dry.  
  • When a temperature measurement is critical it is best to take multiple measurements and average them vs. trusting a single measurement.  

In this case we are looking for an accurate differential temperature, the absolute values don’t matter as much. I found that the quick read and insertion depth of a K-type bead probe (Shown below) was a good tool for this measurement, even though it isn’t the MOST accurate reading.

On another note –

I also neglected to add in additional BTUs of heat to the equation to compensate for what is added by the blower if we wanted to get really crazy.

Motor heat added is

Volts x amps = watts

Watts x (1 – efficiency) = watts of heat added

Watts x 3.412 = BTU of heat added

Just in case you wondered

— Bryan

Measuring airflow is easy… measuring airflow accurately is quite a bit more difficult. In many cases when we as technicians measure airflow we are trying to get to the almighty CFM (Cubic Feet per Minute) volume measurement. You can take CFM readings fairly easily with a hood like the Testo 420 shown above, but even a hood has some limitations when the goal is to measure total system CFM vs. register / grille CFM.

In this series of videos Bill Spohn from Trutech tools demonstrates all of the tools you can use to measure airflow from hot wire and rotating vane anemometers, to flow hoods, to smart grids and pitot tubes, all the way down to using a GARBAGE BAG.

I had the privilege of seeing this presentation in person (I am the one behind the camera) and I wanted to share it with you. It is well worth your time.

— Bryan

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