Month: June 2019

You are probably all familiar with radiant barriers. Sometimes it is thin foil draped under the roof deck, sometimes it’s used on the inside of stud walls or over furring strips before drywall goes up and there is even plywood with a radiant barrier attached to one side that is used for roof decking.

The point of this article is to remind you that you eliminate the benefit of a radiant barrier when you sandwich it between materials in other words when there is no “air gap”, but I also want to help you understand why this is.

How Radiant Heat Transfers 

Heat energy is the “force” that makes the atoms move and molecules jiggle and it’s in everything over absolute zero (-460°F). Heat is transferred or moved in one of three  ways but heat itself isn’t these things, these are methods by which heat is moved like walking, flying in a plane or riding a surfboard.

  • Conduction – Heat moving when one molecule bumps into another and imparts some it’s force. It’s like standing in a line and shoving someone, they move because you impart force directly on them.
  • Convection – Heat moving when the molecules in a fluid are free to move around. It’s like flying on a plane, you are moving freely through the air and bringing your energy with you.
  • Radiation – Heat moving through the air or a vacuum via electromagnetic waves. It’s like surfing because your energy is riding a wave DUDE!…. and that stupid metaphor was the whole reason for the other two lame ones…

So from a practical standpoint in a building we control conductive heat transfer with insulation, convective heat transfer by air sealing to the unconditioned spaces and radiation with low emissivity barrier with the shiny side facing an air gap, this is if you need a radiant barrier at all.

Radiant heat can only transfer when you have two surfaces pointed at one another that have a different temperatures. The rate at which heat will transfer between them is a function of the temperature difference, the distance between them and the emissivity of each surface. A suface with an emissivity of 1 is a so called “perfect black body” and is a theoretical perfect emitter and absorber of radiant heat.

A surface with an emissivity of of 0 is perfect reflector of radiant heat energy and neither absorbs or emits radiant heat. In practice we do not see 1 or zero but a fraction of 1 with a black dull surface being close to one and a shiny, reflective radiant barrier generally being around 0.10 meaning only 10% of the radiant energy is absorbed or emitted.

So why can’t we sandwich a radiant barrier? 

Imagine getting a pan on a stove nice and hot and then hovering your hand over it, you would feel the radiant heat emitting from the pan. Now place a sheet of aluminum foil over the pan and hover your hand again, very little radiant will be absorbed and emitted by the foil and your hand will be much cooler.

Now….

Push your hand down on the foil and squeeze it into the pan…

NO DON’T DO IT! ARE YOU CRAZY?

Spoiler alert, it will burn you.

While aluminum foil has a low emissivity it is very thermally conductive and heat travels through it easily via conduction (molecule to molecule). This means that the only way it helps you block heat is when one shiny, low emissivity side faces an air gap (or vacuum or other fluid that allows the electromagnetic waves to pass easily through). This is why you see white radiant roofs on shopping centers that face the sky, or plywood for roof decking with a radiant layer that faces down into the attic.

If you press anything solid up against both sides of a radiant barrier you make it a conductive layer and it does NO GOOD.

Some of you may (incorrectly) assume that a radiant barrier must be pointed at a light source (like the sun) to do any good. Remember, you don’t need visible light to have radiant heat transfer just a temperature difference. So a radiant layer on the underside of roof decking will help block radiant heat from leaving that roof decking and entering the ceiling and trusses and whatever else is in that attic even if it is pitch black up there because the radiant barrier is bad at absorbing AND emitting radiant heat so even though the radiant barrier on the underside of the roof deck would be hot to “touch” (conductive) it does much less emitting then wood so more of the heat stays put.

— Bryan

 


We have been discussing a lot of methods for checking a refrigerant charge without connecting gauges over the last few months. This got me thinking about the “approach” method of charging that many Lennox systems require.

Approach is simply how many degrees warmer the liquid line leaving the condenser is than the air entering the condenser. The approach method does not require gauges connected to the system but it does require a good temperature reading on the liquid line and suction line (Shown using the Testo 115i clamp and 605i thermo-hygrometer smart probes).

When taking an approach reading make sure to take the air temperature in the shade entering the coil and ensure you have good contact between your other sensor and the liquid line.

The difference in temperature between the liquid line and the outdoor temperature can help illustrate the amount of refrigerant in a system as well as the efficiency of the condenser coil. A coil that rejects more heat will have a leaving temperature that is lower and therefore closer to the outdoor temperature. The liquid line exiting condenser should never be colder than the outdoor air, nor can it be without a refrigerant restriction before the measurement point.

Here is an approach method chart for an older 11 SEER Lennox system showing the designed approach levels.

While most manufacturers don’t publish an approach value, you can estimate the approach by finding the CTOA (Condensing Temperature Over Ambient) for the system you are servicing and subtracting the design subcooling.

6 – 10 SEER Equipment (Older than 1991) = 30°F CTOA

10 -12 SEER Equipment (1992 – 2005) = 25°F CTOA

13 – 15 SEER Equipment (2006 – Present) = 20°F CTOA

16 SEER+ Equipment (2006 – Present) = 15°F CTOA

I did this test on a Carrier 14 SEER system at my office so the CTOA would be approximately 20°

Then Find the design subcooling. in this case, it is 13°F

Subtract 13°F from 20°F and my estimated approach is 7°F +/- 3°F. I used the Testo 115i to take the liquid line temperature and the 605i to take the outdoor temperature using the Testo Smart Probes app and I got an approach of 4.1°F as shown below.

More than anything else, the approach method can be used in conjunction with other readings to show the effectiveness of the condenser at rejecting heat.

If the system superheat and subcooling are in range but the approach is high (liquid line temperature high in relation to the outdoor air), it is an indication that the condenser should be looked at for condition, cleanliness, condenser fan size and operation and fan blade positioning. If the approach is low it can be an indication of refrigerant restriction when combined with low suction, high superheat and normal to high subcooling.

If the approach value is low with normal to low superheat and normal to high suction pressure and high subcooling it is an indication of overcharge.

The approach method is only highly useful by itself (without gauges) on a system that has been previously benchmarked or commissioned and the CTOA and subcooling or the approach previously marked, or on systems (like Lennox) that provide a target approach specific to the model.

— Bryan

In order to wrap my head around diagnostic issues it really helps me to engage in thought experiments where I think of more extreme examples of an issue or situation or consider the ideal in order to find the “edges” of a concept. Once I find the edges of the extreme then I can begin to sort down to a more exact conclusion.

So let’s consider compressors and mass flow

First, don’t get overwhelmed by the phrase “mass flow” I’m not going to start in with confusing words and fancy math. As techs we rarely need to do advanced calculations anyway, it’s more about understanding relationships between factors or IFTTT (If this then that). If this thing occurs or I change this what happens to that.

Mass flow just means how much fluid is moving over a a given amount of time. The “stuff” in this case is refrigerant and the mass measurement is generally lbs in the USA and the rate could be minutes, seconds or hours.

Our goal with a compressor is to move as many lbs of refrigerant, as quickly as possibly with the minimal amount of watts in energy used in order to move the greatest # of btus/hr we can … pretty straight forward so far.

The typical single speed, single stage compressor with no unloading capability runs ESSENTIALLY the same speed and with the same volume in the compression chamber (cylinder, scroll etc…) this means that a traditional compressor has a fairly constant volume in the compression chamber and rate of compression. I say “fairly” constant because as the compressor moves greater mass or works against greater pressure the motor will tend to slip more resulting in a slower rotational speed of the rotor.

So let’s imagine an old single speed, single stage reciprocating compressor with no unloading. It’s compressing refrigerant with a constant volume in the cylinder that goes from its largest cylinder capacity at the bottom of the down stroke (suction stroke) to its minimum capacity at the top of the up stroke (discharge stroke). This variation in volume in the cylinder as the pistons actively move and down is what creates the pressure differential between the high side and the low side and this pressure difference is what allows the refrigerant to move through the circuit.

So you may think to yourself (as I have in the past)

“If pressure differential is what causes the refrigerant to move then don’t we want a big pressure difference between the compressor discharge and suction so that more refrigerant will move?”

The answer is absolutely NO

We actually want the minimum pressure differential we can get away with while still accomplishing the task of maintaining an evaporator (or evaporators in the case of multi-circuit systems) at the desired temperature and (nearly) full of boiling refrigerant.

The reason we want lower pressure differential has to do with mass flow rate, if the compressor has a fixed volume in the cylinders and the pistons are pumping away at the same speed then that part of the equation is fairly fixed. The only way to increase the amount of refrigerant being moved by the compressor is to

#1 – Increase the density of the refrigerant

#2 – Reduce the amount or re-expansion waste in the cylinder

#3 – Reduce the pressure to overcome in the discharge

The first part of that equation is simple, when suction gas is higher pressure it is also higher density, when the suction pressure entering the compressor drops the density also drops. When then density of the refrigerant drops entering the compressor the compressor moves less refrigerant because there is just less there for it move.

Think of this like an old PSC blower motor on undersized duct work. When the static pressure on the return increases the amount of air being moved decreases because the density of the air is decreased. The blower is still spinning the same speed (on a PSC), heck, it may even be spinning faster due to the motor experiencing less resistance, but the airflow decreases. This happens not becasue the motor is doing anything different, it moves less air mass becasue the air is less dense entering the blower and therefore you are moving less air.

When you drop the suction pressure entering a typical compressor you drop the mass flow rate becasue the mass entering the compressor is reduced, lower mass flow rate means moving fewer lbs of refrigerant which (by itself) means lower capacity.

Now let’s move to the second part which is re-expansion and this one applies more to reciprocating compressors where there is a clear compression and expansion stroke vs. a scroll , rotary or screw where the compression is essentially a continuous cycle.

Imagine a compressor sitting in a room with no tubing connected just pumping air. The compressor would be pulling from 14.7 PSIA and discharging into 14.7 PSIA (atmospheric air pressure at sea level). When the piston draws down it would pull in air and fill up and then as the piston pushes up it would start to discharge air out of the cylinder really quickly in the up stroke because the only thing pushing against the discharge is 14.7 PSIA and therefore the highest pressure that would build up inside the compressor is slightly more than 14.7 for it to overcome the pressure of the discharge valve and push out into the air.

If that same compressor were pumped into a chamber where the pressure built up to 200 PSIA what would change?

The compressor would move less air even if  the suction was still left open to atmosphere (and therefore the same air density) because now the discharge valves wouldn’t open until the pressure in the cylinder went above 200 psi meaning that the effective stroke would be reduced due to the pressure being pushed against (#3 on the list above). It would also need to pull down further to re-expand the gas left over in the cylinder to below 14.7 PSIA for more air to enter the cylinder again.

In a scroll, rotary or screw there isn’t valves and cylinders in the same way but the amount of refrigerant being moved is still impacted by changes in suction density (suction pressure) and the pressure exiting the compressor… in other words the COMPRESSION RATIO.

Have you ever noticed that the BTU ratings on compressors have dropped over the last 10 years as units become more efficient? Where a 3-ton unit may have previously had a compressor with a 36 in nomenclature for a nominal three tons you may now find it has closer to 30 or even less.

You may also notice that high efficiency systems often have larger condensing coils and larger evaporators which bring the head pressure and therefore the condensing temperature closer to the outdoor temp and the evaporators are also running a higher temperature bringing up the suction pressure. Manufacturers are increasing how much refrigerant the compressor can move (mass flow rate) by bringing the design head pressure down and the design suction suction pressure up. They can then afford to downsize the compressor achieving the same capacity with less input watts also known as greater energy efficiency.

Let’s give some real world examples of altering mass flow rate by impacting these factors in the field –

  1. Dirty Condenser Coil – Decreases mass flow rate and system capacity because the head pressure and compression ratio go up
  2. Low Indoor Airflow – Decreases mass flow rate because refrigerant density goes down entering the compressor and compression ratio goes up (to a degree). Keep in mind that when there is low air flow or low load head pressure will also tend to drop as the mass flow rate drops. It is held up by the outdoor temperature as a limitation on how low the condensing temperature will drop however.
  3. Overcharge – The impact of overcharge on mass flow rate will vary depending on the metering device and how overcharged the system is. On a TXV / EEV system it will always result in lower mass flow becasue the head pressure will increase. On a fixed orifice it may result in a slight increase in mass  flow initially as suction pressure increases.
  4. High Indoor (Evaporator) Load – Increases mass flow unless there is some control preventing it from doing so like a CPR (compressor pressure regulator). Increased heat entering the evaporator will increase the pressure and density of the refrigerant returning to the compressor, this will increase the mass flow rate, system capacity and head pressure if all else remains the same.

What happens if we change compressor capacity on the fly?

For years in residential and light commercial we’ve been used to fairly fixed compressor volume flow rates but nowadays we see many different types of multi-stage and variable capacity technologies from a simple dual capacity unloading scroll to a digital scroll all the way to variable frequency, variable speed scroll compressors. These compressors have their “rated” capacity which is the state at which they are tested for bench-marking against other units. They can then reduce their capacity below their rating and some can every produce a higher capacity then their rating.

In all of these cases the compressor is altering the amount of refrigerant it is moving by making a change within the compressor itself resulting in lower mass flow when the compressor stages or ramps down and higher mass flow when it ramps up.

Lets imagine a theoretical 4-ton rated unit with a compressor that can ramp down to 2-tons or it ramp up to 5-tons.

What that means in practice is that the compressor is capable of moving an amount refrigerant consistent with two tons of capacity up to a mass flow that can produce 5-tons of capacity at the same rated conditions.

So here is what you would see change when that compressor changes mass flow in comparison to rated capacity if everything else remained the same –

Low Stage (2-ton)

High Suction Pressure

Low Head Pressure

Low Subcool

High Superheat (potentially)

Low Evaporator Delta T

Poor Dehumidification due to high coil temperature

Low compressor amps

Low Compression Ratio

Low Discharge Temperature

Low Approach (liquid line temperature above outdoor temperature)

High Efficiency (EER / SEER)

 

High Stage (5-ton)

Low Suction Pressure

High Head Pressure

High Subcool

Low Superheat (potentially)

High Evaporator Delta T

Strong Dehumidification due to lower coil temperature

High compressor amps

High Compression Ratio

High Discharge Temperature

High Approach (liquid line temperature above outdoor temperature)

Low Efficiency (EER/SEER)

 

Now think about how a system responds when the compressor isn’t pumping properly. It is almost exactly the same as the low stage / low mass flow example listed above with the exception of the efficiency.  When we have lower mass flow than rated these are symptoms we will see whether it is by design or due to a failure.

In practice these variable capacity systems will often be matched with a variable speed blower and a wide range TXV or EEV so that the coil temperature and feeding can adjust with the change in mass flow to help mitigate some of the negative effects of staging down.

There are come interesting things that can be done with modern controls and variable mass flow compressors. One example is Bosch branded condensing units that vary the compressor mass flow to set a fixed evaporator temperature, effectively adjusting the capacity to match the load on the evaporator coil. Another is Carrier Greenspeed heat pumps that ramp the compressors up during heat mode to drive up the pressure on both coils to increase the heat produced inside and reduce defrost requirements.

— Bryan

 

There has been much written and many jokes made about the misdiagnosis of TXV (Thermostatic expansion valves) and rightly so. This article will cut straight to the point to help those of you who may still need a bit of clarification and hopefully, we will save the lives of a few TXVs and the pocketbooks of some customers.

Q: What is a TXV?

A: A TXV (TEV) is a type of metering device. The metering device’s job is to create a pressure drop from the liquid line into the evaporator which will result in refrigerant boiling (changing from liquid to vapor) through the majority of the evaporator coil. This low temperature “boiling” absorbs heat from the space or product being cooled.


Q: How does a TXV Function?

A: A TXV “measures” the temperature and (usually) the pressure at the end of the evaporator coil with a bulb and a tube called an external equalizer. The bulb measures temperature and provides an opening force, the equalizer measures pressure and provides a closing force. There is also a spring that may have an adjustable tension that provides additional closing force. When working properly these forces achieve a balance and maintain the evaporator superheat to the designed of set levels at the end of the evaporator.  The TXV’s job is to maintain superheat within certain operational ranges and conditions. 


Q: How do they fail?

A: A TXV may fail either too far open or too far closed. Too far open is also called “overfeeding” and it means that boiling refrigerant is being fed too far through the evaporator coil, this would show up in low superheat. If the TXV fails closed it can be said to be “underfeeding” which means not enough boiling refrigerant is fed through the evaporator coil and superheat will be too high at the evaporator outlet. 

These failures can and do occur, but they are usually caused by contaminants or moisture in the system that have worked their way to the valve and caused it to stick or become restricted. Another cause of valve failure is a rub out on bulb tube and an external equalizer without a core depressor installed on a port that has a Schrader core in place.  

When a valve is overfeeding the first thing to check is bulb insulation, placement and strapping. If the numbing isn’t properly sensing the suction line it can lead to the valve remaining too far open.


Q: Why are they misdiagnosed so often? 

A: TXV’s are often incorrectly condemned in cases of low evaporator airflow or load. This happens because techs will find a system with low suction pressure and assume that means it is low on refrigerant. They will then start to add refrigerant and the TXV will respond by closing further the more refrigerant is added. The tech will see that the suction isn’t increasing and they will conclude that the TXV is failed. 

This occurs because the tech is paying too much attention to suction pressure without considering the other readings.


Q: What is the correct way to diagnose a TXV? 

A: First take all of your refrigerant readings as well as your liquid line and suction temperature at both ends (on a split system). This means superheat, subcooling, suction saturation (evaporator coil temp) and liquid saturation (condensing temp). For a TXV to do what it is supposed to you need a full line of liquid before the TXV, this means you need at least 1° of subcooling in theory but in reality, you will want to make sure that you have the factory specified subcooling which is usually around 10°. In refrigeration, we do this same thing by looking for a clear sight glass. On a split system checking the subcool outside and then confirming there is no big temperature difference inside to out is a great way to ensure that kinked lines or plugged line driers aren’t an issue. 

The next thing that a TXV needs is enough liquid pressure to have the required pressure differential. This amount of required pressure differential will vary a bit based on the valve but usually, we want to see a 100 PSI minimum difference between the liquid line pressure and the desired evaporator pressure. If the head pressure drops too low due to low ambient conditions this can come into play and impact the ability of the valve to do its job. 

Once this is all confirmed then it is simply a matter of checking the superheat at the end of the evaporator. Most A/C systems will be maintaining 6-14° of superheat at the evaporator outlet. If it is in that range then the valve isn’t bad, it’s doing its job. 

If it is lower than 6° of superheat at the evap outlet then it could be overfeeding (double check your thermometer and gauges) and if the superheat is well above 14° at the evaporator outlet, with the proper subcool and liquid pressure entering… then you have a failed closed (underfeeding valve). Keep in mind that some valves will have a screen right before the valve and this can be the cause of the restriction rather than the valve. You can intentionally freeze the coil and try to see the freezing point or use thermal imaging to help spot if it’s the valve or the screen. When you find the point of temperature you find the point of pressure drop, just remember that the TXV is DESIGNED to provide pressure to maintain a fairly fixed superheat. 


Q: Do TXVs Ever Fail

A: They can fail internally but most often they fail because of a blocked inlet screen (if they have one), contaminants entering the valve, loss of charge from the power head, bulb location and positioning issues and overheating of the valve. In commercial and refrigeration applications you can often replace or clean the screen and replace the power head rather than replacing the entire valve. 


As I have said many times before diagnosis make sure your tools are well calibrated and working and that you are ACTUALLY reading the pressure correctly. I’ve seen many misdiagnoses just because a Schrader wasn’t pushing in or a multi-position valve cracked properly.

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