Month: October 2018

Here is another excellent article from Michael Housh, owner of Housh Home Energy and a regular contributor to HVAC School. Thanks Michael!


I thought I would go through a simple example of sizing a hydronic circulator for an application.  This is a made up scenario, but I sketched out a 20’ x 20’ square home with baseboard radiators that encompass mostly all of the exterior walls.  It should also be mentioned that this article does not go into detail on choosing a radiator/panel, this is a hypothetical situation to show the steps involved in determining the total head loss (resistance) through a piping arrangement.

I equate the following to what Manual-D would be for duct-design (the air-side of HVAC).  Before I move on there are a few points that I need to make clear. Because hydronic systems are a closed loop it’s important to realize that the lift (or height) from the pump to the highest point of the system has no effect on sizing the circulator, the only thing that matters is the drag/resistance of the pipe, valves, and fittings in the system.  So whether a system is like the one in the drawing or you have a loop that travels 1000’ in the air, the process to size the circulator is the same.

 

Another thing I will just gloss over here is that for any of the following formulas it’s assumed that the piping system has turbulent flow (Reynold’s number between 2,300 – 200,000).  Turbulent flow means the water molecules are not traveling in a straight line, they gravitate from center to the wall of the pipe and back again, this adds increased head/resistance but it allows heat to be transferred through the wall of the pipe/radiator more easily.

 

While this is an imaginary project and I’m going with ¾” piping for the entire loop, it is important when designing and selecting the size of the loop that we keep our velocity inside of the pipe between 2-4 ft/sec.  The 2 ft/sec is the minimum velocity required to be able to get air to move downwards throughout the piping system, if we get above 4 ft/sec then we can have noise issues or damage the pipes.

 

The primary difference between this exercise and doing a Manual-D is that we must consider every pipe, valve, and fitting in the entire circuit, as opposed to only using the longest Total Equivalent Length (supply and return) for design.

 

For some it may be easier to think of this like an electrical circuit (Ohm’s Law), we add the resistance through the entire circuit to come up with the total resistance (ft. head) our system will have to pump against.

 

The next step is totaling up all the equivalent lengths of everything using charts available (most all of the charts that I use are in Modern Hydronic Heating by John Siegenthaler and because of copyright I can’t publish them, but engineeringtoolbox.com is a good place to look for these types of resources).  I created a spreadsheet from the data in my drawing and came up with the following data.


You may notice that the air-separator has a Cv (Flow Coefficient), I’m not going to go into details in this article on how I determined it was equivalent to 11’ of ¾” copper, but I will give the definition of Cv. Cv is the flow rate of 60° water to make a 1 psi pressure drop across the device.  If you read my last article on setting up a pump, then you should know that a pressure difference can be converted to ft. of head, so there are some formulas that can be used to determine the equivalent length of a device for a certain pipe-size but is a bit complex for this article. Below I combined the important totals to give us our total equivalent length of 162.8’ of ¾” copper.

While this article is getting long, the above information is gold!  It allows us to create a fingerprint for our system and determine the performance over a range of flow rates.  With this information, we can create the System Head-loss Curve, which is defined by the following equation.

The reason I say this is gold is that we are able to create our system fingerprint over an array of flow rates, which will give us the head-loss we must deal with, as in the following graph.

In our hypothetical system we are going to design around a 180° output temperature from the boiler and a 20° Delta-T, giving us an average temperature of 170°.  Since density and viscosity are dynamic this is the number we will use to determine the Fluid Properties Factor for our application. We can solve for this using the following formula.

Next all we need is to determine our pipe size coefficient, which is just a multiplier based on the internal dimensions and roughness of our tubing.  This is something that is looked up in a chart as well, and for ¾” copper pipe is equal to 0.061957. I’m also going to pick a generic flow rate for the application of 4 GPM, so we can wrap things up.  So now we can substitute all of our values into our head-loss equation, solving for our design flow rate of 4 GPM.

So currently we would need to select a pump that would produce 4 GPM at 5.13 ft. of head.  We would do so by matching up manufacturer’s data/pump curve to match our desired output. This is a deep subject, some of which carries over into the air-side of HVAC, but some of the topics I have just skimmed the surface on.  While it may seem overwhelming, the main take away from this article is going to be to relate the piping system to an electrical diagram.

In future articles this will allow us to break down complex piping systems into more manageable pieces.  As with most things in our industry, it’s not always about the information you have in your head, but your ability to find the correct information/formulas/resources when they’re needed. Keep on learning!

— Michael

 

This article is written by one of the smartest guys I know online, Neil Comparetto. Thanks Neil!


Recently I posted a question in the HVAC School Group on Facebook, “when designing a residential duct system what friction rate do you use?”. As of writing this, only one answer was correct according to ACCA’s Manual D.


I feel there is some confusion on what friction rate is and what friction rate to use with a duct calculator. Hopefully, after reading this tech tip you will have a better understanding.

So, what is friction rate?

Friction rate (FR) is the pressure drop between two points in a duct system that are separated by a specific distance. Duct calculators use 100′ as a reference distance. So, if you were to set the friction rate at .1″ on your duct calculator for a specific CFM the duct calculator will give you choices on what size of duct to use. Expect a pressure drop of .1″ w.c. over 100′ of straight duct at that CFM and duct size / type.

Determining the Friction Rate

First, you need to know what the external static pressure (ESP) rating for the selected air handling equipment is. ( external static pressure means external to that piece of equipment. For an air handler, everything that came in the box is accounted for, including the coil and typically the throwaway filter. For a furnace the indoor coil is external and counts against the available static pressure)

Next you have to subtract the pressure losses (CPL) of the air-side components (coil, filter, supply and return registers/grilles, balancing dampers, etc.). Now you will have the remaining available static pressure (ASP). ASP = (ESP – CPL)

Now it’s time to calculate the total effective length (TEL) of the duct system. In the Manual D each type of duct fitting has been assigned an equivalent length value in feet. This is done with an equation converting pressure drop across the fitting to length in feet (there is a reference velocity and a reference friction rate in the equation). Add up both the supply and return duct system in feet. It is important to note that this is not a sum of the whole distribution system. The most restrictive run, from the air handling apparatus to the boot is used. Supply TEL + Return TEL = TEL

The formula for calculating the friction rate is FR= (ASP x 100) / TEL
This formula will give you the friction rate to size the ducts for this specific duct system. If you test static pressure undersized duct systems are very common, almost expected. This is because a “rule of thumb” was used when designing the ducts.

This is just an introduction to the duct design process. I encourage you to familiarize yourself with ACCA’s Manual D and go build a great system!

— Neil Comparetto

This article is written by Michael Housh, owner of Housh Home Energy and a regular contributor to HVAC School. Thanks Michael!


This is another article in trying to relate the air-side of HVAC with the water-side.  In this article I’m going to talk about how to determine pump flow (GPM) based on the pressure difference across the pump.

 

I relate checking pump flow to checking air flow with static pressure on the air-side.  Without knowing how many gallons per minute we’re moving the less we can truly tell about what our system is doing.  Most of the equations that revolve around hydronic systems (or air-conditioning systems) revolve around how many pounds of a substance we’re moving through the system.

 

The trouble with many of the systems out there is that most installers/designers don’t add anyway for us to check this in the field.  I’ve primarily only seen any sort of pressure port or gauge on larger commercial/industrial systems, or sometimes older pump flanges that have a pressure port built-in.  Most people today are installing isolation flanges on their pumps, so I strongly encourage everyone to spend the few extra dollars on isolation flanges with a drain port built-in, this can allow a technician or installer access to checking the pressure difference across the pump.

 

Measuring pump flow (just like air flow) requires looking at manufacturer’s charts on the pump you are working on.  But instead of checking static pressure across an appliance we use the pressure difference across the pump. In the U.S. the amount of energy produced by a pump is measured in ft. of head, however all of the gauges that are used measure in PSI.  Luckily there is a formula to convert PSI -> Ft. of Head.

 

H = PSI / (0.433 * SG)

 

Where:

 

H = feet of head

PSI = pounds per square inch

0.433 = constant

SG = specific gravity (for rule of thumb we can use 1.02 for water)

 

The constant (divisor) can be simplified by multiplying 0.433 * 1.02 = 0.44:

 

H = PSI / 0.44

 

At this point I must give a WARNING about some of the pictures I’m going to show, as they will surely be controversial, but until my mind is changed on this subject Testo will have to remove H2O from their measurable media type.

Before I move on, let me lay some groundwork.  My latest system had a design load of @ 80K BTU.  This system utilized two zones each having a pump.  By using the “rule of thumb” sensible heat rate equation for hydronics, I can layout what my total system GPM should be to deliver the load.  Most hydronic systems are designed around a 20° deltaT on the distribution side, which is what I’ll use in the below equation (for more on the sensible heat rate equations you can read this tech-tip I wrote).

 

80,000 / (500 * 20) = 8 GPM

 

So this means my total system needs to move 8 gallons per minute in order to meet demand.  In my scenario I actually have 2 zone pumps (one does a bonus room which needs about 2-3 GPM, and the other the main house needing about 5 GPM).  I utilized the same 3-speed pump on both zones, and I will cover how I set the speed based on the above GPM that I need to deliver. Below is the pump curve (similar to a fan chart) that we need to compare to, each line representing one of the speeds of the pump.

I take the pressure at the inlet and the outlet of the pump to get my deltaP, using the drain port on the isolation flanges.    

DeltaP = 3.5 PSI (as you can see this is very hard to take from an analog style gauge).  I can then convert my deltaP to ft. of head using the formula above.

 

3.5 / 0.44 = 7.95 ft. of Head

 

We can then compare to the above pump curve (for low speed, lowest curve), which shows a flow rate of about 2.8-3 GPM, which is where I want to be for that zone.  I can then do the same for the other zone pump (set on medium speed).

 

This gives me a deltaP of 4.8 PSI and converted ft. of head of 10.9.  When comparing my deltaP to the medium speed pump curve my flow rate is about 5 GPM, giving me a total system flow rate of my desired 8 GPM.

 

To take it a step further, I then compare my system loop deltaT, which in this case actually worked out perfectly to 20°, so I won’t go through the formula again, but this proves that we’re delivering the desired 80K BTU to the space.

— Michael

 

In Brazing and soldering tubing, we have a few things we need to accomplish to make a proper connection

  1. We can’t overheat the joint to the point that it overheats the base metal or the flux where applicable
  2. We must bring the entire joint above the melting temperature of the brazing alloy
  3. Draw the alloy deep into the joint

New techs often underheat, overheat or take too long to complete a joint because they don’t use visual cues to apply the alloy at the proper time.

Every metal and alloy responds a little differently but we always use indicators of some sort to know when to start and stop.

Some things you need to know before making a connection are

  • What is the melt temperature of the base metal?
  • What is the working temperature of the alloy?
  • Is flux required and is it external or integral to the alloy you are using?
  • What is the thermal conductivity of the metals?

Most metals we work with will respond to heat in the same way with a color change shown in the chart below. Notably, aluminum will show no change in color before it hits its melt point.

Copper to Copper

Copper to copper connections require no flux when rods that contain phosphorus are used. This is why rods using a small amount of silver with the remainder of the rod being made of copper and phosphorus are common.

Most of these phos/copper alloys have a working temperature around 1200°F and copper has a melting temperature of around 1950°F. A quick look at the chart above will show you that a good brazing indicator would be copper in the “cherry” range of 1175 – 1275°F when applying the rod. We use the Solderweld 15% round alloy at Kalos.

Copper to Brass 

Brass is a metal that is made up of a mixture of copper and zinc. Brass has a lower melting point than copper but is great for casting so many valves and other refrigeration components will be made of brass. It is preferable to use a high silver content alloy with either an external flux or a flux coated rod like the Solderweld 56% rod.

You will then heat up both sides of the joint until you see the proper color on the copper and to a lesser extent in the brass as well. The flux will also act as an indicator because it will go completely clear and flat giving both base metals a “wetted” look at about 1100°F (for most appropriate fluxes). Both the color change and the clear flux can act as indicators that the temperature is correct on copper to brass.

Copper or Brass to Steel 

Working with copper or brass to steel will definitely require a brazing alloy with no phosphorus and flux. Steel changes color in much the same way as copper but it has less thermal conductivity which means that the heat you apply tends to concentrate in one spot rather that travel the way it does with copper. Steel doesn’t melt until 2500°F but the working range for the flux is generally 1100°F – 1600°F (depending on brand/type) so you can easily overheat the flux when working with steel as well as bump into the copper melting temperature of 1950°F if you aren’t careful.

When working with copper or brass to steel use the metal color in that “cherry” zone as well as the quiet, clear flux as an indicator of proper brazing temperature.

Aluminum

Aluminum gives you no indication of when it’s going to melt which makes it more tricky to work with. It also melts at 1220°F which means that if you are working with aluminum to other metals you are in the danger zone as soon as there is any redness in the other metals. In brazing aluminum to aluminum, patching aluminum or working with aluminum to other metals you need to rely heavily on the aluminum flux to tell you when it is time to apply the alloy.

For aluminum to aluminum work we use Alloy-sol from Solderweld because you can apply as much flux as you need and it gives you a great indication of when to start applying the rod when the flux goes clear at around 600°F.

For aluminum to copper we use Al-Cop, it has a flux built into a channel on the rod that can be helf to the joint and will melt and run out when the proper temperature is reached.

In all cases we are looking for visual cues rather that overheating and damaging the base material or burning flux or underheating and globbing up rod onto the joint.

— Bryan

I’m not sure who first started calling evaporator coil odors “dirty sock syndrome” but I really wish they hadn’t. Nowadays every tech out there can’t help but categorize every odor from the system as dirty sock syndrome” and customers just LOVE to hear that they have a “syndrome” named after a dirty locker room smell.

What is Dirty Sock Syndrome? 

Loosely defined, dirty sock syndrome is a smell that comes from an evaporator coil, especially when it first comes on or shifts from heat to cool. The smell comes from biological material that dries out on the coil when the system is off or in heat mode.

Customers will notice this smell when it first comes on more than once it has been running a while and often describe it as smelling musty, moldy or like….. well…. dirty socks.

True dirty sock syndrome is almost exclusively found in humid climates and is most common with heat pump systems. Gas furnace air temperatures are generally high enough to kill the biological material (at least that’s the working theory).

Many have noticed that the occurrences appear to be increasing which some have attributed to porous recycled aluminum used in coils though this hasn’t been proven.

What Can be Done to Stop It?

There are many different approaches used by manufacturers and contractors but the first thing to do is to give all of the obvious parts a good inspection and cleaning.

Inspect and Clean-

  • Both sides of the evaporator
  • The return box
  • Drain Pan
  • Blower Wheel

Also check for

  • Standing water around the unit
  • Dirty filters (duh)
  • Gaps in return ducts
  • Signs of rodent or other pest intrusions

Many manufacturers have turned to using special coatings to help prevent the biological material (or biofilm) from attaching to the metal. These coatings can be factory or aftermarket applied and can provide pretty reliable results.

In the field, it is often more practical to use a combination of filtration, cleaning and UV-C bulb installations to help keep stuff from growing on and bonding to the coil. Many techs fail to realize that a HUGE part of the effectiveness of UV-C is how it is installed and keeping the bulbs regularly changed.

The effectiveness of UV-C depends completely on it’s proximity to the coil. It will do a great job of irradiating portions of the coil it is close to but the effectiveness will fall off VERY quickly the further the bulb is from the surface.

When installing UV make sure to –

  • Protect or shield any wires or plastic parts that may be damaged
  • NEVER look at the light when it’s on
  • Install according to manufacturers specs
  • Inform the customer of the ongoing maintenance and costs of bulb replacement

Specialty Cleaning 

As long as “dirty sock” syndrome has been around, folks have been spraying stuff on coils trying to stop it. I read one place that bleach works well… I bet it does until your coil corrodes out and your customers canary dies!

There are many products that techs swear by but largely the industry is going away from spraying harsh chemicals or chemicals known to cause human reactions because whatever you spray on that coil is going to make it into the customers home or business.

This is why we use the following approach to prevent and remedy dirty sock coil odors –

  • Keep good filters in the system (MERV 8 or better, 4″ whenever possible)
  • Keep drain pans and drain lines very clean
  • Use Evap+ from Refrigeration Technologies on the Evaporator Coils
  • Install UV-C at the coil if the issue persists

In 13 years in business this strategy has worked well for us although I’m sure there are more extreme cases where coatings or coil replace necessary.

— Bryan

 

 

 

In my recent classes with my employees at Kalos, we are going over finding target pressures and temperatures for an air conditioning system with the goal being to get techs to have “target” readings in mind before they start connecting tools. This step is an important part of being able to “check a system without gauges” like we have talked about so often. Much of this list makes more sense if you are already familiar with our 5 pillars of diagnosis.

It is important that we start using these terms when speaking to one another, writing notes and diagnosing because these will translate better between systems and between A/C and Refrigeration. Some readings we take in HVAC like static pressure and delta t do not apply to refrigeration, while others like target condensing and evaporator temperature are key in both disciplines.

Keep in mind that when I give a “rule of thumb” you should always consider manufacturer specs, charts, panels, install and service manuals as superior to a rule of thumb. You will be amazed what you might learn reading the service and installation manuals of the systems you install and work on.

 

Target Evaporator Temperature or DTD (Design Temperature Difference) = The temperature the evaporator coil should be based on the return temp (A/C 35°F below ambient) or below box temp in the case of refrigeration 10° below box on a walk in 20°below box on a reach in). on a typical A/C system with a 75°F return temperature, the DTD would be 35°F which means the target evaporator temperature would be 40°F. The DTD will vary based on airflow and evaporator coil size.

 

Measured Evaporator Temperature or TD (Temperature Difference) = The suction saturation temperature (not pressure) as measured on the suction gauge for that particular refrigerant it can then be compared to box or return temperature to calculate your measured or actual TD

 

Target Condensing Temperature Over Ambient (CTOA) = This is the target temperature the liquid saturation (condensing temperature) measured on the gauge SHOULD be above the outdoor air temperature measured in the shade entering the condenser coil. This will be 30° over ambient on VERY old units, all the down to as low as 15° on new very high-efficiency units. This is LIQUID LINE ONLY the discharge line will be higher pressure.

 

Target Condensing Temperature – The outdoor temperature + the CTOA = Target condensing temperature Example: Outdoor Temp 95° + 15° for a 16 SEER system = 110° target condensing temp.

 

Target Superheat – This is the superheat you SHOULD have and it varies based on if the system has a TXV or Piston metering device. If a piston you MUST use a use a superheat chart as well as a thermo-hygrometer / psychrometer to measure the indoor wet bulb and dry bulb because those charts require those readings. If the system is a TXV then set your target at 5-15° if measured inside and 10°-20° if measured outside

 

Measured Superheat – The increase of the suction line temperature when compared to the suction saturation.

 

Target Subcool – The subcool you wish to achieve. Many units will have this marked on the data tag, if not then use 10° subcool on TXV systems and 5° – 15° on piston systems recognizing that on a piston system this rule will not always apply.

 

Measured Subcool – The measured difference between the liquid line temperature and the condensing temperature (liquid saturation temp) off of the high side gauge. This is liquid line only, not the discharge line.

 

Outdoor Ambient – the outdoor dry bulb temperature, in the shade entering near the center of the condenser coil

 

Return DB – Return dry bulb. The temperature of the return air without taking evaporation or humidity into account. Best taken in the return right before the unit and not in the space.

 

Return WB – Return wet bulb. The temperature of the return air + the evaporative effect. Lower WB, when compared to DB, means lower relative humidity. Wet bulb and dry bulb will be the same at 100% RH. Best taken in the return right before the unit and not in the space.

 

Return RH% –  The relative humidity of the air in the return. Relative humidity is the percentage of moisture in the air compared to how much moisture is in the air. Hotter air can hold more moisture than the same air at a lower temperature.

 

Target Supply Air Temperature – Target supply air temperature is calculated using a delta t chart and comparing the return DB and the return WB temperatures. The target supply air temperature is dry bulb and can be compared to the return DB to calculate the target delta t.

 

Target Delta T (Air Temp Split) – Don’t confuse TD or Evaporator split above with delta t or air temp split. Keep in mind that the 18°- 22° rule that many use only applies to homes with 45% to 55% relative humidity. As RH% goes up the target split will go down and as RH% goes down the split will go up. Delta T will also vary based on airflow with higher airflow resulting in a lower Delta T.

 

Measured Delta T – The measured difference between the supply and return air DB. Keep in mind this should be taken a few feet before and after the unit to allow for air mixing and reduce radiant gains/losses.

 

Delta H – Delta H is an advanced measurement that calculates the change in enthalpy (heat content) of the air between the return and the supply. You can do this with two digital thermos-hygrometers like the 605i and it takes into account the temperature and humidity of the air entering and leaving the evaporator.

 

Delivered Capacity – Delivered capacity is the calculation of BTUs of heat being removed from an air stream which combines the Delta H with the CFM of air to give you the total “work” being done across the evaporator coil.

 

Discharge Temperature – The measured temperature (with a line clamp) of the discharge line leaving the compressor, not the liquid line

 

Target Liquid Line Temperature  –  When “checking a system without gauges” the target liquid line temperature is the target condensing temperature minus the target subcool. This is usually measured at the condensing unit.

 

Target Suction Line Temperature –  When “checking a system without gauges” the target suction temp is the target evaporator temperature plus the target superheat. This is most accurate when measured inside but is also valuable when measured outside.

 

Approach – Approach is just another name for target liquid line temperature and it a reading that Lennox publishes a target for on many of their units in the installation manual and on the back of the panel of the condensing unit. Systems with larger or more efficient condenser coils tend to have a lower approach (cooler liquid line) while those with smaller, less efficient coils ten to have a higher approach (warmer liquid line)

 

Suction Line TD (Temperature Rise) – The difference between the suction line temperature inside after the evaporator coil and outside by the condensing unit. When a suction TD is more than 10°F compressor overheating and oil carbonization can occur under some load conditions.

 

Liquid Line TD (Temperature Drop) – The difference between the liquid line temperature outside by the service valve and inside before the metering device. ideally, the liquid would have VERY LITTLE temperature drop and any drop of more than a few degrees should be looked into. Long line length, vertical risers, running the liquid line through a low ambient space, contact between the liquid line and the suction line or restrictions can lead to higher than normal LL TD.

 

Static Pressure – The positive or negative pressure exerted on all surfaces equally within a duct system. Static pressure does not measure flow, it is like the pressure inside a balloon that inflates or deflates. Static pressure is generally measured in inches of water column in the USA (“wc)

 

Design TESP – This is the total external static pressure, both positive (supply) and negative (return) that a particular furnace or air handler are designed to work under external to the appliance. Most typical residential units are designed for 0.5”wc.

 

Measured TESP – This is the total external static measured using a manometer or magnahelic gauge and static pressure tips. On a furnace this would be measured inside the furnace before and after the blower but before the coil. In an air handler or fan coil it will generally be measured before and after the unit in the ducts. This is the total difference between the negative and positive reading so if return static was -0.2” and supply static was +0.3” the total would be 0.5”wc TESP.

 

Static Pressure Drop – This is the measured pressure change across a portion of the air system. For example across a coil, filter, duct etc… This is helpful in diagnosing airflow issues and changes over time.

 

While this may seem like a long list, most of it is pretty common sense. One thing to mention is the fact that if you do not have a thermos-hygrometer like the 605i and accurate temp clamps you cannot properly check a Delta T on any system or set the superheat on a fixed orifice system. In order to properly set a charge or diagnose a system, you need a way to accurately test line temperatures and measure return / indoor Wet Bulb, Dry Bulb and Relative Humidity.

 

Step one on diagnosing a refrigerant issue, checking or setting a charge should be to get an accurate return (or box) DB, WB and RH as well as the outdoor ambient temperature and then working from there taking appropriate readings. When calling a senior tech or your manager please be prepared will all relevant readings to make a quick and correct diagnosis.

— Bryan

 

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