Month: April 2018

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It was December 5th, 1952, only seven years after the end of WWII in London England when the “great smog” settled across the city.

It was a meteorological anomaly combined with unchecked industrial pollution that led the great smog, but even once it settled on the city, bringing business to a screeching halt, everyone thought it pass quickly.

It didn’t.

Four days later the smog finally lifted, and the death toll rose to 4,000. Many modern statisticians place the number of deaths caused by the smog at closer to 10,000. While the world knew the dangers of particles in the air before, there was a keener understanding of what pollution in our air can do to us since that time.

Nowadays you will hear the terms PM2.5 and PM10 thrown around and you may never stop to think about what they mean.

PM10 (Particulate Matter < 10 Microns)

Particles with a diameter less than 10 microns across. For some perspective, a human hair is about 50 – 70 microns across and a micron often called a micrometer is one-millionth of a meter. While a PM10 particle is tiny it still pretty coarse by particle standards and will often contain things like dust, dander, and pollen. These are allergens to be sure but generally, are not the most dangerous particles.

PM2.5 (Particulate Matter < 2.5 Microns)

These are particles less than 2.5 microns and often contain things like Ammonia, Carbon, Lead, and mold. Both chemically and biologically toxic particles generally fall into this smaller 2.5 microns or less size range.

Zhao D, Azimi P, Stephens B – Int J Environ Res Public Health (2015)

In the graph shown above, you can see the results of the % of PM 2.5 sized particles that passed through some various tested filters. The study showed that not only that typical MERV 6 – 8 filters failed to do a good job of capturing a large percentage of these particles. It also showed that the efficiency of the filters varied from brand to brand as shown in MERV12(#1) and MERV 12 (#2).

Indoor / Outdoor

There is a clear link between pollution outdoors and pollution indoors. It stands to reason that if the air outside is dirty, then the air inside will also tend to more particles in it which means that a comparison of outdoor PM2.5 to indoor PM2.5 comparison can be a better way to compare one space to another.

The EPA has tested and found that indoor air can be significantly more polluted with VOCs (Volatile Organic Compounds) often 2 to 5 times more than outdoors.

But hang on…

VOCs are gaseous and are not the same as PM2.5 particles and shouldn’t be confused with them.

From a technicians perspective, you need to be aware that typical air filtration will only effectively deal with the PM10 and larger particles. For the PM2.5 you will need high MERV filtration or other types of active filtration such as PCO or Ionization strategies.

For VOC’s you can ventilate, use carbon filtration or look into products like the Air Oasis bi-polar ionizer that has been shown to reduce VOCs in studies.

Testing for what exactly is in the air is a trickier business then I initially thought and get’s tougher the smaller the particles get. Even today it is virtually impossible to “test” for specific live fungus, bacteria, and VOCs in the field. You can count particles and get in the ballpark of the issue but nailing down all the details can be challenging without sending samples off to a lab.

For most of us, a good particle counter is the best we are going to get out in the field, and we are best off following some good solid practices to help our customers.

  • Encourage your customers to use good quality filters. When possible install MERV12 or better
  • Make sure the unit isn’t pulling air around the filter or through the panels
  • Make sure that ducts are well sealed so that the return isn’t drawing in air from an unconditioned space
  • Keep the indoor relative humidity between 30% – 55%
  • Address any localized moisture issues that may be resulting in organic growth
  • Add in well-filtered ventilation air for VOC reduction
  • Make sure the evaporator, blower wheel, and drain pan are kept clean
  • Research quality IAQ products and recommend them based on your customer’s needs and desires
  • Bring in fewer VOCs by using indoor products that are more natural and have had more time to off-gas

The most dangerous stuff in the air are chemicals, tiny particles and live organisms caused by moisture issues. Knowing this will make you better tech and more capable of advising your customers well.

— Bryan

We have discussed DTD (Design Temperature Difference) quite a bit for air conditioning applications, but what about refrigeration? Let’s start by defining our terms again

Suction Saturation Temperature

Saturation temperature is the temperature the refrigerant will be at a given pressure if it is in the process of changing state. This change of state would be from liquid to vapor (boiling) in the case of the low side (evaporator / suction line). When we look at saturation temperatures instead of pressures we can use similar rules and we will see similar saturation temperatures across all refrigerants when the application is the same. Experienced HVAC and refrigeration techs pay far closer attention to the saturation temperatures than they do pressures.

Evaporator TD and DTD

Evaporator TD (temperature difference) is the measured difference between the suction saturation temperature (evaporator boiling temperature) and the box temperature. DTD (design temperature difference) is the designed or expected TD.

Delta T

Many A/C techs will confuse TD with Delta T. Delta T is the difference between the evaporator AIR temperature entering the coil to the air temperature leaving the coil. The Delta T will vary based on the humidity in the box where TD will not.

Target Box Temperature 

The temperature the refrigeration box should maintain when the system is operating properly

Superheat

The increase in temperature between the suction saturation temperature and the suction line temperature leaving the evaporator. Superheat is the temperature (sensible heat) gained between the point that all of the liquid boiled off in the evaporator coil and the suction line at the outlet of the coil. in refrigeration, like HVAC 10°F(5.5°K) of superheat  is average with a range from 3°F to 12°F(1.65°K – 6.6°K) depending on the equipment type (10°F(5.5°K) for med temp, 5°F(2.75°K) for low temp, 3°F(1.65°K) for ice machines ).

Hot Pull Down

Refrigeration equipment is unlike HVAC equipment in that the evaporator will spend most of its life running in a very stable environment with minimal fluctuation in the box temperature.

On occasion a refrigeration system will see a huge change in load in cases where it was off and needs to “pull down” the temperature, or when doors are left open or when a large quantity of warm product is placed in the box. When a piece of refrigeration equipment is in hot pull down it cannot be expected to abide by the typical DTD or superheat rules and must be allowed to get near the design box temperature before fine adjustments are made to the charge, TXV superheat settings or to the EPR (Evaporator Pressure Regulator) if there is one.

Design Temperature Difference (DTD)

In air conditioning applications a 35°F DTD is a good guideline for systems that run 400 CFM(679.6 m3/h) of air per ton of cooling (12,000 btu/hr). In refrigeration the DTD is much lower than in air conditioning.

There are several reasons for this but one big reason is the desire to maintain relatively high relative humidity levels in refrigeration to keep from drying out and damaging product. Keep in mind that NOTHING is a substitute from manufacturer’s data but here are some good DTD guidelines for traditional / older refrigeration equipment. Keep in min dthat the trend is toward lower evaporator TD on newer equipment.

Walk-ins  10°F(5.5°K) DTD +/- 3°F(1.65°K)
Reach-ins  20°F(11°K) DTD +/- 5°F(2.75°K)
A/C 35°F(13.75°K)) DTD +/- 5°F(2.75°K)

You then subtract the DTD from your box temperature / return temperature to calculate your target suction saturation. You can then use this target saturation / DTD and compare it to your actual measured saturation and DT once the box is within 5°F – 10°F(2.75°K – 5.5°K) of it’s target temperature to help you set your charge, TXV and EPR as well as diagnose potential airflow issues when compared with suction superheat and subcooling / clear site glass.

For Example –

If you have a medium temp walk-in cooler with a 35°F(1.66°C) box temperature you would expect to see a suction saturation of  25°F +/- 3°F

When doing a quick inspection of a piece of refrigeration equipment without gauges you can use this data to do the following calculation –

35°F – 10°F DT + 10°F superheat = 35°F suction line temperature +/- 3°F 

In this particular case logic tells us that the suction line could be no WARMER than 35°F(1.66°C) because that is the temperature of the air the refrigerant is transferring its heat to. However by the time you factor in the the accuracy of your box thermometer and line thermometer and the assumed saturation temperature you would still expect a 35°F(1.66°C) suction line temperature +/- 3°F(1.65°K)

For a -10°(-23.33°C) box, low temp reach-in you would calculate it this way

-10°F- 20°F DT + 5°F superheat = -25°F suction line temperature +/- 5°F 

Clearly, this is NOT the way to commision a new piece of equipment or to benchmark a system you haven’t worked on before, but it can give you a quick glimpse at the operation of a piece of refrigeration equipment without attaching gauges, especially on critically charged or sealed systems.

The best practice is to know the equipment you are working on, read up on it and properly log benchmark data the first time you work on a piece of equipment or during commissioning.

It should also be noted as Jeremy Smith pointed out, in recent years TD’s have been decreasing as manufacturers seek higher efficiency through higher suction and lower compression ratios.

This means that TD’s as low as 5 can be designed into some units but keep in mind… the suction line can still be no warmer than the box so as DTD drops so does superheat and the critical nature of expansion valve operation.

— Bryan

 

When a system has abnormally high head pressure (high condensing temperature over ambient) and compression ratio, one of the easiest things to look for is a dirty condenser coil and more often than not, that will be the cause.

However…

There is another category of issues that can cause high condensing temperature (high head pressure) that result from improper practices rather than dirt and grime.

When a tech comes across a failed condensing fan motor or a damaged blade they will often go to their van and see what they have as a “universal” replacement part. I don’t have an issue with using aftermarket repair parts in some cases but you need to make sure that the part you are using will operate properly without sacrificing capacity, efficiency and longevity.

Often when using aftermarket parts a tech may be sacrificing one or more of these things and that can lead to issues.

When replacing a fan blade you need to ensure –

  1. The pitch is a match
  2. The number of blades is a match
  3. The Diameter is a match

If you change the pitch you will also need to change the # of blade and vice versa to end up with the same CFM airflow output which can be very tough to determine in the field.

The diameter really cannot change or you won’t have the proper gap between the blade edge and the shroud (Usually 1/2″ – 1″) which can greatly impact air movement.

When replacing a motor you need to ensure –

  1. The RPM (# of poles)  matches
  2. Voltage and phrasing matches
  3. The HP is the same or greater
  4. The physical size will allow proper installation

In some cases, the technical specs may work but the motor body may be deeper. When this happens you need to make sure that the fan blade can still sit high enough in the fan shroud for proper movement of air. In many cases the blade/shroud are designed so that the middle/center of the blade matches up with the bottom of the fan shroud (cowling) and if it isn’t it can decrease airflow.

This issue comes into play often in cases where a factory motor fails on smaller tonnage residential units with a less than 1/4 HP motor. In these cases when you replace the factory motor with a universal motor the larger physical depth of the motor and sometimes the width can result in less than designed airflow. Make sure when replacing the motor that you are still able to place the motor blade in the same position in relation to the blade to ensure proper air flow/condensing temperature.

Sometimes you will come across systems that are running higher head pressure than they should be. In these cases you will want to check and make sure the motor HP and RPM are correct and that the blade is properly sized and positioned in the shroud.

As always, being attentive is key to finding issues, even issues caused by others

— Bryan

 

This tech tip is written by one of the best all-around HVAC minds out there. Neil Comparetto.

I think that we all can agree that duct leakage is not ideal. Our job is to condition the space. If we can’t control the air, that becomes difficult. On top of that anytime you are losing already paid for conditioned air. But really, how bad could it be?

I’m in Richmond Virginia, so we’ll use that as our example location. According to ACCA Manual J summer design conditions our outdoor design temperature is 92° Fahrenheit, with a moisture content of 106 grains per pound. (grains is a measurement of absolute moisture). Let’s use the indoor conditions 75° F and 50% relative humidity, which converts to 65 grains of moisture.

Our example system will be a 3-ton air conditioner moving 1200 CFM with ducts in a vented attic. For this exercise, we won’t get into duct sensible heat gain that even a 100% tight duct system will have to overcome.

This system will have a modest 10% supply duct leakage into the attic (Energy Star estimates that the typical duct system has 20-30% duct leakage). Assume 0% return leakage (which is unlikely). So we already know that 10% of our capacity is gone, never to return again into the attic.

On a 3 ton air conditioner that will be roughly 3,600 btuh. We are now delivering 1080 CFM of supply air to the living space, and returning 1200 CFM. Where does the additional 120 CFM of return air come from? You guessed it, outside. The supply duct leakage into the attic, outside of our thermal and pressure boundary, has now brought the living space into a negative pressure. No big deal, it’s only 120 CFM… but have you ever done the math!?

Stick with me, it’s not as bad as it looks. Here are the formulas for the sensible and latent heat required to bring the infiltration air back to indoor conditions (75°/ 50%RH).

Sensible BTUH = 1.08 x CFM x (Outdoor temp – indoor temp) Latent BTUH = 0.68 x CFM x (Outdoor grains – Indoor grains)

Let’s use 92° F as our outdoor air temperature number. In all likelihood, considering that the attic floor/ceiling plane is one of the leakiest parts of the house, and the attic is typically > 120° F, that in real life it will be higher than whatever outdoor temperature is.

Our example will look like this:

1.08 x 120 CFM x (92°-75°) = 2,203 btuh of sensible heat

.68 x 120 CFM x (106 grains – 65 grains) = 3,346 btuh of latent heat

2,203 + 3,346= 5,549 btuh of total heat.

That is an additional 5,549 btuh of total heat. The 3,346 btuh of latent heat is the more difficult number to deal with. Next time you are bored flip through your favorite air conditioner’s product data and see what it can produce, you may be surprised. Don’t forget about the 3,600 btuh that’s up in the attic somewhere. And just think, this is from only 10% supply duct leakage, considerably more is very possible.

As you can imagine in the heating season this problem doesn’t go away. Typically outside air is much drier than indoor air, and duct leakage will dry out the indoor space. If the heating system is a heat pump the capacity loss is corrected by electric strip heat, which is bad. That means when you seal the ducts auxiliary heat is reduced, which is good.

Leaky ducts can contribute to many more issues than just energy loss and comfort. Did you know that a one square inch hole in the duct system is equal to thirty-inch hole in the building envelope? The potential to create pressure imbalances in the building is tremendous. Pressure imbalances can cause many issues, like flues backdrafting, excess dust and allergens, uneven temperatures, and moisture issues to name a few.

Something as simple as sealing ducts can solve many issues, hopefully, you include it in your scope of work.

— Neil

We had a really great conversation on the HVAC School Facebook Group about some belt tension best practices and it turns out that even a lot  of really smart and experienced techs are not aware of all the factors related to belt tensioning.

Myth #1 is that amperage is used to set belt tension. Now don’t get me wrong, checking amperage before and after changing belt tension is an excellent practice to ensure you are not binding the bearings from over tension, it does not tell you whether or not the belt is at optimum tension.

I think Browning summarizes it best in this statement from their Browning tool box technician app

Ideal tension is the lowest tension at which the belt will not slip under peak load conditions

Getting a belt too tight shortens the life of the belt and bearings and can cause high amperage. Leaving a belt too loose will shorten the belt life and result in loss of airflow and noise.

Many techs confuse the sheave adjustment, designed to alter the pulley ratio and the airflow with the belt tension adjustment. These are not the same thing and serve separate purposes.

The adjustable sheave allows the pulley faces to adjust closer or further from one another, resulting in a belt that rides closer to the hub when looser (halves further apart) or closer to the edge when tighter  (halves further separated) THIS ADJUSTMENT IS FOR FAN SPEED ONLY NOT TENSIONING

With a properly tensioned belt the belt should not slip significantly when starting, it should not be noisy and it should not bounce around. If you tighten the belt check the amps before and after and the motor should not overamp.

The correct tension method is to get the belt close to the correct tension by feel with a deflection of 1/64 of an inch for every 1″ of distance between the two pulley centers. You can then use an app or a chart like THIS ONE to find the proper force to generate this deflection.

You would then use a belt deflection tool like the one shown above to test the deflection force required and adjust accordingly. The video below demonstrates this.

I like what Jeremy smSmithtated in the group “Belt tension has less impact on motor amperage than pitch diameter of the sheave and how that affects total airflow.Use the Emerson tool and the app (or paper chart if you’re all stone age) Record tension and other data (sheave diameter, center to center length, rpm and proper tension) on blower housing.”

Check those belts.

— Bryan

Download the podcast Directly HERE

As always if you have an iPhone subscribe HERE and if you have an Android phone subscribe HERE

Condenser Flooding / Motormaster Podcast Companion

This article and podcast is courtesy of Jeremy Smith, one of the most knowledgeable and helpful refrigeration techs I know.

It’s my feeling that, no matter how well explained, this topic really requires a treatment that is more in depth and one that can be absorbed slowly with the ability to continually return and re-read certain sections to allow for best understanding of the subject matter.

As discussed in the podcast, as the outdoor temperature drops, the capacity of the condenser increases dramatically causing it to be, essentially, oversized for normal operation.   To counteract that, we use a valve (headmaster) or valves (ORI/ORD) to fill the condenser with liquid to effectively reduce the amount of coil that is actively rejecting heat and condensing refrigerant.   This also maintains a high enough liquid pressure feeding our TEV.   This prevents wild swings in TEV control because it is a pressure operated mechanical device.

First things first, let’s open up Sporlan’s 90-30-1 … seriously go ahead and click it , it will open in another tab so you can go back and forth.

This is a document I reference all the time when dealing with condenser flooding problems.  If you’re tech savvy, save it on your mobile device.  If you’re more of a low-tech guy, listening to a podcast and reading an internet publication on your flip phone or whatever, go ahead and print this out, laminate it and keep it in your clipboard.   Heck, even if you are a high tech guy, sometimes nothing beats a hard copy of this the first few times you work through it.

If ,after the podcast, you haven’t read through this to familiarize yourself with it, take the time to do so.   It seems like a really complicated procedure to work through, and the first few times that you do it on your own, it can be.  With practice, however, you’ll get used to it.

We’ll work through a condenser flooding calculation here in slow time, outlining all the different calculations taken into account.

First lets lay out the basic info we need.  The measurements and counts will vary, of course, depending on the equipment that you have.

If we have an R22 unit, 44 condenser passes ⅜” in diameter each are 38 ¾” long with 42 return bends.   Our evaporator temperature is 20°F, current temp is 35°F and the lowest expected ambient is -20°F.

Now, that seems like a lot of information, but we’ll break it all down.

First, we need to figure the total length of the condenser tubing in feet.   So, we take 44 x 38 ¾ and get 1705” of tubing.   1705 ÷ 12” per foot gives us 142.083 feet of tubing.   Now, that’s just the straight tubing.   We’ve got return bends to account for.

Refer to our Sporlan document.   In TABLE 1, you’ll find an equivalent foot length per return bend.   In the case of a ⅜” return bend, it’s. 2 feet per bend, so 42 x .2 gives us 8.4 feet more.

Add those together for total length of 150.483.  Back to TABLE 1 look in the R22 section under ⅜” tubing and follow the line for -20°F across.   You’ll find a density factor of 0.055.   This number is how many pounds of liquid refrigerant is needed to fill one foot of tubing at that temperature.   So, 150.483 x 0.055.  This gives us 8.28 pounds.  This is the amount required to fill the entire condenser with liquid, but we don’t really need to fill the WHOLE coil….

Back to the document..TABLE 2 this time.

Across the top, find 20° evaporating temp, now follow that down to the -20°F row.   This gives us a percentage.   82%  so, this unit at -20% will have 82% of its condenser filled with liquid.   So let’s take 8.28 x 0.82 to get our flooding charge.

6.78 pounds.

Now, what does this number really mean.   This is the amount of refrigerant we need to add to a system that we’ve JUST cleared the sightglass on when the ambient temperature is 70°F or higher.  If our ambient temperature were 70 degrees or warmer, we could add just that amount past a clear sight glass and walk away, satisfied in knowing that the unit will run properly no matter what the weather throws at it.

Remember, though, that our current ambient is 35°F.   So, now what?

Time to stop.  Get your Sharpie out and WRITE THIS NUMBER DOWN!   Record it on the unit somewhere.  Somewhere easy to see but somewhere that the sun doesn’t degrade the ink over time.   That way, you only have to go through this one time.  If you’re doing a new installation and startup, do the next guy a favor and write both this AND the total system charge down somewhere so that I don’t have to guesstimate the charge when it all leaks out.

Now, let’s go back to TABLE 2 and look at the 35°F row.   We find that at 35°, we need to have 63% flooded.   Well, we’ve got a clear sight glass and it’s 35° ambient so, we’re already 63% flooded.

Since the most we need is 82% flooded, 82%-63% gives 19% so, we take our total, 8.28 x 0.19 to get 1.57 pounds.  At our current conditions, that’s all the flooding charge that we need to add because we’ve already got some flooding going on to have a clear sightglass because we’re under the 70 degree mark and the low ambient controls are in play and doing their job.

Some techs claim that just spraying water on the coil will flood the condenser enough to allow the use of that as a charging technique.    Let’s think about it for a minute.   What variables come into play with a method like that?  Variables that we can’t control…  for starters, what is the wet bulb temperature of the air entering the condenser?  How well is the condenser wetted? With the stakes being what they are, I’m not excited about the prospect of using this because I’m probably going to be the guy who winds up on the roof when it’s -20 and the wind is howling and this unit is low on gas because someone tried to use this method to figure a flooding charge, didn’t get enough gas in the unit and now it’s short.    I’ve still got to my due diligence as a service tech, do a full leak check, not find anything, and walk away wondering if I missed a leak somewhere all because someone else didn’t take a couple minutes to do a little work to do the job properly.  This is a totally preventable service call.

What about TABLE 3, you ask?  Very astute and that tells me that you’re reading ahead. Excellent.  I have never had to use it.

It gives a different flooding percentage for units with an unloader and low ambient controls where they’ll be running in low ambient conditions.  With the unloader, remember that we’re really moving less heat, changing the condenser dynamic and making it even MORE oversized than it would be if there weren’t an unloader, so more refrigerant needs to be added to properly flood the condenser.

— Jeremy Smith

P.S. – You can checkout the Testo 770-3 multimeter we mentioned in the middle by going here

Part 1:

We talk with Trevor Matthews with Emerson about causes of air conditioning and refrigeration compressor failure and the causes Verifying System Operation Sheet from Emerson https://hvacrschool.com/EmersonVerify Diagnosing Compressor Failures from Emerson https://hvacrschool.com/CompFailures

 

Part 2:

Part 2 of the discussion with Trevor Matthews with Emerson about causes of air conditioning and refrigeration compressor failure and the causes Verifying System Operation Sheet from Emerson https://hvacrschool.com/EmersonVerify Diagnosing Compressor Failures from Emerson https://hvacrschool.com/CompFailuresWe talk with Trevor Matthews with Emerson about causes of air conditioning and refrigeration compressor failure and the causes Verifying System Operation Sheet from Emerson https://hvacrschool.com/EmersonVerify Diagnosing Compressor Failures from Emerson https://hvacrschool.com/CompFailures

If you have an iPhone subscribe to the podcast HERE and if you have an Android phone subscribe HERE.

When I first started in the trade we used to run into shielded control wires on the Carrier Comfort Zone 1 zoning systems and also on a Carrier VVT system I used to maintain at a bank. I knew it has something to do with electrical “noise” and that communicating systems often called for it but I never looked any further into it.

Over the last decade there has been a lot of different residential communicating systems that have come out. Some require shielded cable, some recommend it and others don’t mention it all.

The fact is that whenever controls work on a low voltage “signal” rather than a simple “on/off” control they are more susceptible to induced charges from other nearby conductors, electronics and even transients from electrical storms.

A shielded cable has a  metallic jacket that surrounds the individual conductors and routes the induced charges to ground, keeping it away from the conductors inside.

As an example of this, I installed a Carrier Infinity system at my own house WITHOUT using shielded cable and almost every time there are lightning strikes nearby the unit will throw a communications fault, since I’m in florida that happens quite often.

If you do have the wisdom to run shielded cable you need to remember to bond (ground) one side of the shield securely to a good equipment ground on one end and ONE END ONLY. If you ground both ends you risk the sheild becoming a path in the case of a ground fault which could cause some bigger issues. If you ground both ends you can also create a “ground loop” that can cause the very noise you set out to eliminate.

In some cases, you can perform a similar function by grounding leftover/unused conductors on one end if you failed to run a shielded cable. There is no guarantee it will solve the issue depending on the severity because the other conductors don’t fully surround the conductors being utilized.

The lesson being, when working with communicating “signal” controls run shielded cable whenever possible. I was looking around and found this spec sheet from Southwire on their shielded 8 wire.

— Bryan

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