- Tech Tips
Almost everyone knows that CO (Carbon Monoxide) is really nasty stuff and nothing to play around with. Humans will often begin exhibiting symptoms of CO poisoning at 200ppm (parts per million) of CO in a space. Studies have also shown that the effects of CO exposure can accumulate over time resulting in health problems that don’t show up all at once in extreme symptoms.
All of this to say that testing for and monitoring CO levels in occupied spaces using stationary and personal protective CO monitors and alarms is a must. It is also a best practice to inspect fuel-burning appliances for combustion safety and efficiency using a combustion analyzer like the Testo 330 shown above.
One of the readings taken in the flue of a fuel-burning appliance is the CO in PPM, but in many cases we refer to it as CO “air free” and many combustion analyzers calculate it automatically.
In essence, CO Air Free is a calculated PPM of CO that takes the “excess air” or air that was not used in combustion and removes it from the equation. This means that the CO Air Free number will always be higher than the uncorrected CO number.
The way air free CO is calculated is by first calculating the difference in the Flue O2 (Oxygen) percentage as compared to the percentage of oxygen in regular air at sea level which is 20.9%.
So the formula is
Air Free CO (in ppm) = Measured CO ppm x (20.9 ÷ (20.9 – O2% in flue gas)
Or set by step
20.9 minus the measured percentage of oxygen in the flue, let’s say we measured 5% so that would equal 15.9%
20.9 ÷ 15.9% = 1.31
We then multiply the measured PPM of CO (let’s say it was 87) times the 1.31 we calculated and we get an air free CO PPM of 87 x 1.31 = 114 PPM of CO adjusted to air free.
Using air free CO helps to create a more level playing ground for comparing one appliance to another by preventing high levels of excess air from diluting the CO and giving an improperly low reading.
While it is widely recognized that 400 PPM CO is the allowable limit for flue CO levels it is generally possible and recommended for flue CO levels to be below 100 PPM air free.
Here is part 2 from Michael Housh from Housh Home Energy in Ohio. Thanks Michael!
This is part two in a series on a deeper look into the Sensible Heat Rate Equations. The first article can be found here if you missed it. In this article, I will dive deeper into the density of “standard air” and water based on a change in temperature. “Standard air” is air with 0% RH and at sea-level (14.7 psia).
As mentioned in the previous article density can be defined as its mass per unit of volume (or weight per unit of volume). Also mentioned before is that the Sensible Heat Rate equations stem from a lower level equation that is about moving a mass of a fluid.
Let’s look at the equation to find the density of air first. The following equation can be used to determine the density of air for a given pressure and temperature. For all of our equations and graphs, we will use the absolute pressure of 14.7 (the pressure at sea-level).
We must first convert our temperature to an absolute temperature (Rankine).
T = Temperature (°F)
Next, we can solve for the density for the given temperature and atmospheric pressure.
D = density of air (lb/ft3)
Pa = absolute pressure of air (psia) = 14.7
Tr = temperature (°R) converted in the previous step
As you can see, we are dealing with pretty minute changes in the density of air based on temperature change when the humidity is 0%.
The above formula can also be used to solve for the density of air at different altitudes, by substituting Pa with the absolute pressure at a given altitude, but that will be left for you to figure out.
Next, let’s look at the density of water equation. This is a little more complex of an equation. It is also only valid for water temperatures between 50°F and 250°F.
D = density of water (lb/ft3)
T = temperature (°F)
As you can see the density of water has a little more of a drastic change based on temperature than that of air.
First off, when we relate these to the Sensible Heat Rate equations (that are about moving pounds of a fluid), we can see that the mass can be affected by a significant amount (at least in the case of water). Secondly, 1 cubic foot of water weighs significantly more than 1 cubic foot of air.
Let’s envision 2 – 1’ x 1’ x 1’ cubes, one is full of air the other is full of water. Both of these are at 50°F. Let’s envision we have a fan or pump moving 1 of these cubes each minute. Next, let’s look at how we would compare the Btu’s carried/rejected from both of these cubes with 20° Delta T.
We are going to use the lower level equation found in the first article to make this a
little easier to understand.
Q = M * C * T
Q = BTU’s absorbed or rejected from the fluid (BTU/hr)
M = mass / density of the fluid (lb/ft3)
C = specific heat (BTU/lb) : Air = .24 & water = 1.0
T = temperature change dry bulb (°F)
When we calculate the density of a fluid that has a temperature change to it, we want to use the density for the average temperature. So for our experiment since we are starting at 50° and we end at 70°, we will use the Density for water and for air at 60°.
|Air (Density @ 60° = .076,
Specific Heat = .24)
|Water (Density @ 60° = 62.37,
Specific Heat = 1.0)
|Q = .076 * .24 * 20 * 60 (min)||Q = 62.37 * 1.0 * 20 * 60 (min)|
|Q = 21.89 BTU/h||Q = 74,844 BTU/h|
At these parameters, we’d have to move 3,419 more cubic feet of air to achieve the same as 1 cubic foot of water.
So let’s see how we can use this information to build a little bit of a better equation than the default “standard air” and “standard water” Sensible Heat Rate Equations.
Since the density of air is pretty tricky and is affected by both altitude and humidity I would recommend using the online calculator at HVACR School. With water, I would use the above equation to solve for density or graph above. Remember for both of these we want to use the average temperature to solve for the density.
Now, let’s rewrite the Sensible Heat Rate equations into a slightly more accurate way to calculate the BTU/h of both air and water.
Let’s run through a little comparison of these equations from the “standard air” and “standard water” equations. I’ll start with the air-side first. Let’s imagine we have a 70° return air temperature and a 130° supply air temperature (a 60° Delta T). The relative humidity is 40% and I used the elevation of my home which is approximately 800 ft. above sea level. By using the online calculator I figured the density of the average temperature of the air at these parameters is 0.06857. Now let’s assume we have 800 CFM moving across our appliance, so now we can solve for the BTU transfer of our appliance into the air.
|Standard Air||Improved Air|
|Q=1.08 x 800 x 60 = 51,840 btu/hr||Q = .06857 x 14.4 x 800 x 60 = 47,395 btu/hr|
That is about a 10% difference by correcting for the density of the air.
Next, for the water-side, let’s assume that we have an incoming water temperature of 170° and outlet water temperature of 190° across a boiler (a 20° Delta T). Our pump is moving 5 gallons per minute. The density of water at the average temperature of 180° is 60.59814, so now we can solve for the BTU transfer from the boiler into the water.
|Standard Water||Improved Water|
|Q = 500 x 5 x 20 = 50,000 btu/hr||Q = (60.59814 / 7.48) x 60 x 5 x 20 = 48,605 btu/hr|
This is about a 5% difference by correcting for the density of water.
For some, the difference in accuracy may not be worth it in the field, but it depends on what you are trying to solve. Stay tuned for more details on better Sensible Heat Rate Equations.
One of the most common questions we get from techs is about using a voltmeter to diagnose a high voltage circuit. It’s especially tricky when a tech is used to working on Low voltage or 120V circuit where there is a clear “hot” side of the circuit and a clear “grounded” side of the circuit. In 120V you have one hot leg and the other side is neutral which is actually connected to ground back in the panel. Most (but not all) 24v transformers have one hot leg and the other leg is grounded. A car has one 12VDC Hot and the other side is grounded to the chassis.
All of these cases cause techs to get used to putting one meter lead to ground and “walking” the other lead through the circuit, looking for where the voltage is lost. While this is still not the best idea even on these circuits, it usually works.
in 240v or 3 phase diagnosis it doesn’t work. Here is why –
The other “side” of the completed circuit is not grounded at all. So when you check to ground, you are checking to a point that has literally NOTHING to do with the completed circuit you are diagnosing. Even more important is the fact that you will often read “120v to ground” even when the leg you of power you are attempting to diagnose is open.
Here’s an example
Let’s say you are trying to see if the IFR contact is open. So you put your meter from L1 to ground. Good news you have 120v. So now you are feeling confident and you read from IFR terminal 2 to ground and you still have 120v. So now you think, “The IFR terminals are closed because I have 120v on each side”…
You will have 120V to ground on IFR terminal 2 regardless of whether the contacts are open or closed. If they are open you will be reading 120v backfed through the motor from L2, if they are closed then you will read L1.
In other words, it’s a pointless test.
Take a deep breath…
This next part is gonna take some focus to understand. If you don’t intend to pay careful attention to these next paragraphs you won’t benefit.
Instead read from L1 to L2 and confirm 240V then read from IFR1 to L2 and from IFR2 to L2. If you have 240v on IFR1 and not IFR2 then you know IFR is open…
An alternate method if you are DEAD SET on reading to ground is to check IFR1 to ground. If you have 120V then check from IFR1 to IFR2. If you read anything across the contacts you would then know they were open.
You will read potential (voltage) so long as a path and difference in charges exist, across a load and across an open switch. You will not read potential (voltage) across a closed switch because a closed switch has no potential difference across it.
Final notes –
You are encouraged to check both legs to the ground for safety purposes to confirm the disconnect is actually off and open.
Checking to the equipment ground can be a way to check the ground itself, although in that case, a de-energized ohm or megohm test can often be a better test.
We had a situation a few months back where we needed to monitor amperage on a grocery store panel over a period of time. The trouble was, we needed data logging capability as well as accurate measurement at 600+ amps.
Finding an all in one solution proved to be quite expensive. Luckily my friend Jim Bergmann happens to own an instrument company (Redfish Instruments) and he had a simple solution.
Use the data logging capability of the Redfish multimeter with a Fluke I800 amperage to mA (milliamp) clamp to get the job done.
This particular clamp can be used with any good quality meter that reads in the milliamp range (not to be confused with uA which is microamps often used for testing flame rectification).
Like shown below, for this clamp you connect the meter leads to the correct jacks and select the mA scale. In this particular case 1mA = 1A so the image below is showing a 320 amp measurement.
There are other accessory clamps that use the mV (Millivolt) scale rather than mA like the Redfish IDVM333 shown below.
With this clamp the output scale can be adjusted with the selector on the clamp to either 10mV AC or 1mV AC.
Again you need to make sure the leads are in the right spots based on the type of clamp reading and make sure to get in in the right scale, but once you understand how to it, it’s surprisingly simple and you can use a wide variety of clamps and meters.
Note: My brother Nathan wrote this a few years back and I only did some minor editing
A pool heat pump is essentially a water-cooled air conditioner in reverse with a large air evaporator on the outside that looks like a condenser coil and a heat exchanger (usually tube in tube) on the inside.
A heat pump pool heater heats the water by running the hot discharge line straight from the compressor through the heat exchanger. The heat exchanger works as the condenser for the refrigeration circuit of the pool heater and the refrigerant flows ion the opposite direction as the water.
Leaving the heat exchanger, the copper line should now be a subcooled liquid line that feeds over to some form of metering device. This metering device will then feed the evaporator coil, which is the large coil around the outside of the heater (this can cause confusion because this coil looks like what we would call a condenser coil in an air conditioning application ). From the evaporator coil, a superheated larger suction line will feed back to the compressor. Keep in mind when checking the charge on a heat pump pool heater, that it will be subject to a wider range of temperatures and loads than a normal air conditioner. Higher water/air temperatures tend to produce higher pressure readings, while colder water will tend to lower pressures.
Remember, outdoor air temperature on a pool heater has the same effect on readings, as indoor air temperature would on an AC in cool mode. Often, when the outdoor air is below 60 degrees you may have a suction saturation below freezing, and if the pool is ice cold this will further lower your readings.
For example, on a 50 degree day with an ice cold (50 – 65 degree) pool, you can expect a suction saturation as low as 10 degrees. You can still check superheat and if possible, subcool, If both of these readings appear normal, then your charge is most likely correct (check manufacturer specs first if there are any). If your pressure / saturation readings are high, always suspect water flow, as low water flow is difficult to gauge in these systems. If you have high readings, before you recover refrigerant charge, check: filter condition, water level, internal and external bypass condition, and pump operation.
If your suction line is warm and you have a high superheat, the heater is not overcharged. The electrical circuit and components on a heat pump will vary in style. However, the function of the components remains constant. There will be some form of user controls which report back to a board that takes this input, as well as a reading from a water temperature sensor to determine if the heater should run or not.
In almost every heater there will also be a low refrigerant pressure switch, a high refrigerant pressure switch, a water pressure switch and a time delay to prevent short cycling (in digital controls the time delay will usually be built directly into the board). There are occasionally, defrost controls, or low ambient controls built into these heaters that will allow the fan to continue running while keeping the compressor off. The best way to reduce freezing on a pool heater coil (evaporator) is to not run them at night when it’s coldest and nobody is swimming anyway. Keep in mind that most pool heaters do not have any form of true defrost, so effectively these controls are just leaving the heater off until the air temperature has defrosted the heater… assuming the air temperature is over 32… and if it isn’t WHY IS ANYONE SWIMMING!
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)
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 –
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