Chris Muller of Purafil Honored by ASHRAE

Purafil's Chris Muller was honored with an ASHRAE distinguished service award. Muller, a 12 year ASHRAE member, serves as a Distinguished Lecturer (DL) for ASHRAE and as chair of the ASHRAE Standard Committee 145P, which was charged with the task of engineering the ASHRAE Standard 62.1. In addition to these ongoing responsibilities, Muller remains thoroughly involved within the organization as a voting member of several committees and as a co-author of the ASHRAE Standard 62.1-2004 User’s Manual.

Muller’s continued service and commitment to ASHRAE contributed to his DSA nomination. ASHRAE members’ involvement in key roles such as society president, chapter officers and committee chair are assessed on a tally system. Each role accounts for a specified number of points and DSA nominees must score at least 15 to be considered for the honor.

Muller is a seasoned 20-year Purafil veteran responsible for technical support services as well as specific research and development functions. He has also written and edited over 100 articles and technical papers, which have been published by leading industry print media on environmental air quality and gas-phase air filtration applications.

Read more here (pdf).

Foam-Core Air Handler Panels: A Coming Standard

If you've been specifying or installing commercial modular or semi-custom air handlers recently, you've probably noticed that quite a few manufacturers have switched away from the traditional single (or double) wall fiberglass batt insulation design, to a newer double-wall injected foam core panel. What is driving this change in the industry?

In a word: Performance.

To examine this further let's look at a typical design:



What you have is a sandwich construction with thin gage sheet metal enclosing the rigid foam core. The foam actually acts as a structural component, adhering to the exterior sheet metal and causing the entire composite assembly to function as a single structural unit. This means that you can achieve much greater rigidity with 20 or 22 gage steel than can be attained in a traditional batt insulation design with 16 gage steel or thicker! In fact, panel deflections for this type of panel are generally around L/240 or less (1/240th of the longest panel dimension) when subjected to an 8" static pressure load. That's actually better than the deflection spec for most custom equipment.

So why do we care about deflection? Well, a minor advantage is that this type of construction resists dings and dents much better. A slightly more important criteria is that the panels, especially floor panels, are much less likely to 'oil-can' when under pressure or under the weight of foot traffic. Even more important is that these panels make the unit itself more rigid and less susceptible to deflection or deformation in shipping or rigging. But the real advantage of this construction comes in the realm of energy savings.

These panels in general just blow away the thermal performance of batt-type panels. The following table compares the R-value improvement of rigid foam insulation over that of batt insulation:

(click on image for larger view)

Generally speaking, you get twice the insulation from foam in the same depth. Note that this only takes into account the performance of the insulation itself--further advantage is gained by the thermal break that prevents heat conduction from occurring at the panel seams--which is almost impossible to prevent in a traditional batt-insulation panel.

What does this amount to? Well, for a rooftop unit operating in a heating climate, this difference in R-value could amount to as much as 2% of the total unit energy over the course of a year.

But let's talk about panel rigidity again--it is in thermal performance where this really becomes important. Because the overall thermal performance of an air handler casing is really a function of two things--Overall U-value (defined by the insulation and thermal break) and the leakage rate.

Think about it: Every cubic foot of air leaked out of a cabinet is a cubic foot that had system energy applied to it to condition it, but now will not reach the occupied space. Conversely, every cubic foot of unconditioned air that leaks into a cabinet is a cubic foot that needs to be compensated for by more work by the air conditioning system. And this is a criteria that is critically affected by better panel design.

Traditional batt-type commercial air handlers catalog leakage rates of about 3-5% at 4" of static pressure. But because foam-core panels flex much less, and therefore don't open up leaks at panel seams as much, they typically exhibit much smaller leakage rates. Aaon catalogs leakages of less than 1% of the design airflow at 8" of static pressure. (To make a true comparison with the batt panels described above, you have to remember that static pressure increases with the square of leakage).

If you assume that 4" is a typical pressure rating for a commercial air handler, you can see that you would typically waste about 3-5% more energy in leakage with a traditional design than you would with a newer foam-core panel design. And this is additive to the direct thermal losses due to conduction through the cabinet. (Check the cataloged leakage rating even of foam-core designs, details matter and the seal and fastening design can also affect these leakage numbers. Not all manufacturers meet the standards described here).

But what about custom equipment? While generally the design details of that class of equipment can reduce the conduction, deflections and leakages even with standard batt insulation, there are still some advantages to foam core design. Energy Labs has introduced a foam-core panel for jobs that demand this premium construction, if needed. The traditional batt panel design still affords more flexibility in layout and does not come at the premium price necessary for a fully custom foam core unit.

So, Why Use Direct-Drive, Anyway?

Why indeed?

If you have read my posts on the Hollisterian and Florentine effects on direct-drive fans, you may be wondering if it is really worth the complication to chose this type of fan-drive system for your project.

But when you think about it, the lesson from those two effects is to strive to use fan selections that are at a synchronous speed when at design. And since you know you can vary the width of the wheel to get the CFM you need, this should be a simple trick that you (or any decent AHU provider) can do when laying out your equipment. And if this isn't possible, you know how to compensate for an asynchronous fan selection in your motor size and system design. The only question left is what do you gain for this effort?

Plenty.

Efficiency for one. Belts are a source of inefficiency. Friction between the belt and the pulleys causes heat and erosion of the belt which is exhibited in the system by a reduction in efficiency of the system. How much friction are we talking about? A sample chart might help quantify this:



That's right--About 4% of your motor energy is lost on systems with brake HP's around 50, and more than seven percent on systems less than two BHP. That's a lot of energy to just throw away.

Another reason is maintainability.

Guess how much belt maintenance needs to be done on a direct-drive system? Guess how many belts need to be stocked to replace broken belts? Guess how much time needs to be spent adjusting belt tension to spec? Guess how many times someone needs to be called in to correct a squealing belt?

A rule of maintenance is that if something is hard to do, it won't be done. A good corollary to that might be that if something doesn't need to be done at all, there's a good chance it won't cause a problem downstream because someone didn't do it.

A related advantage is the longevity of the system. For starters, A belt drive system requires at least four bearings, two at the motor and two at the fan. Since a direct drive fan only has the motor bearings, simple mathematics would indicate that you would have at least half the bearing failures. But, as is often the case, the real situation is a bit more complicated than simple math. In each system there is a force on the motor bearing perpendicular to the shaft. In the belt drive system, this force is from the belt tension, in the direct drive system it is from the fan wheel weight.

This is where real life makes things more complicated--because the belt tension is usually several times the weight of the fan wheel. This means that the bearings in the direct-drive fan motor see much lower stress than in the belted case. This translates to many times the expected life for these critical components.

Another advantage? No belt dust. This means that projects with critical air quality concerns may be able to avoid final filters

And let's not forget belt noise--Not just the squeaking that is caused by the slipping of an incorrectly tensioned or worn belt, but the inherent noise that is added to the system from the normal operation of the belt drive itself. (Which is not, by the way, accounted for in the fan sound data you get from the manufacturer. Those were measured on direct-drive fan wheels....)

So let's review:


(Some people might argue that changing pulleys to adjust speed is less of an advantage than a disadvantage)

Now, of course, there will always be applications for belt-drive fans. Sometimes they just make a better fit for the project than direct-drive. Since the motor on a belt-drive fan is supposed to be operating at its synchronous speed (by design) you don't need to oversize motors to reach operating points where a full wheel width is truly the most efficient or quietest solution possible. But direct-drive sure makes sense when it makes sense.

Fun With Mechanical Engineering

At least I think that's what this is.

The Florentine Effect, Explained

This post is a follow-on to my earlier posting on the "Hollisterian" effect and is a further explanation of the tricks involved in properly selecting direct-drive fans in air handlers.

By now, you should be familiar with the Hollisterian reduction in available horsepower that occurs when you select a design point at an asynchronous motor speed. And you know that if you ever do need to design such a case, you will also need to oversize your motor by a factor equal to the motor design RPM divided by the actual operating RPM. Thus, if you are laying out an 1800 RPM motor to operate at design at 1500 RPM (let's ignore the fact an 1800 RPM motor is actually 1775 RPM, for now), you will need to select a motor nameplate HP that exceeds the brake HP of the design point by a factor of 1.2 (1800/1500). And because you are such a careful engineer, you've thrown in a bypass around the VFD in order to allow fan operation even if the VFD burns out.

So you are ready to go, right? What else could possibly trip you up now?

This is where the Florentine effect can get you.

The Florentine Effect

Let's go back to that 13.5 HP selection at 1500 RPM we discussed in the last post. Turns out, that's pretty close to a 77% width 30" fan operating at 14,000 cfm and 4" of static. When I select an EPQN fan using the Twin City Fan software, I get a brake horsepower of 13.22 HP. You know that in order to account for the derate due to the RPM that you need to pick a motor that can provide a nominal HP at least 1.2X this brake, or 15.9 BHP. You select the next larger size, or the 20 HP motor. That's a full 50% bigger than the design brake, and at least 25% bigger than you need, when you account for Hollisterian effects.

So you install the fan, and everything works just fine. In fact, you might go several years before any trouble raises its head. But then, suddenly, one day it does.

Let's say the VFD serving the fan burns out, or is taken out of service for a short while. The owner, wanting to preserve function of his fan system, even if he has to operate it at constant speed, does the obvious thing and flips the fan to bypass...

Suddenly the fan kicks on, starts roaring, and then the motor burns out. A follow up inspection might even find that some of the duct fittings have blown apart. What happened?

Well, let's think about what happened. The fan normally operated at 1500 CFM at peak design. The bypass, which is essentially a standard motor starter wired in parallel to the VFD, kicked the motor on at the line power frequency of 60 HZ, or 1800 rpm. This means that you weren't supplying air at 14,000 CFM at 4", but something greater. Following the fan laws, you would ride the system curve up to about 17,000 CFM at 5.5"! And that is assuming that you are operating a constant volume system or a VAV system at peak cooling load. If you were in heating on a VAV system with the boxes choked back to their minimum flows, you might develop even higher pressures!

Let's look at what happened:

(click to see larger image)

The above is a fan curve plot of the two operating conditions of your fan: 1500 RPM and 1800 RPM. The HP curves are plotted for both cases also. They share a common system curve (this assumes an unchanging duct system--a bad assumption for a VAV system). I've highlighted the resulting flow rates and brake horsepowers at the two resultant operating points where the system curve intersects the fan curve.

The first thing that should jump out is that the brake horsepower for this fan operating at 1800 RPM jumps up to 22.9 HP*--even greater than the 20 HP that you picked to protect from the Hollisterian effect! The other thing you should notice is that if this is, in fact, a variable volume system, the system curve shown at design is not the system curve that would be seen by the fan unless the system was at full cooling. If the boxes neck back at part load or in a heating condition, the system curve shifts to the left, pushing the intersection between the fan curve and the system curve closer to the fan curve peak. This greatly increases the amount of static pressure the fan can develop at this higher speed--up to about 9" in the case shown here.

*this can be calculated by the fan law formula:

bhp₂=bhp₁(rpm₂/rpm₁)³

Thus in a VAV system, this can be a double whammy, kicking out your motor and damaging your duct system.

How to avoid this problem? Well, you could size the motor for the even larger size demanded by the Florentine effect--but that would still leave you with the possible problem of overpressurization of the ductwork in bypass. Probably the first thing to consider is whether or not you really need a bypass, anyway. With today's more reliable VFD's, putting in a bypass is far less of a necessity. Many drives can function for the life of the equipment with no failures at all. If redundancy is absolutely necessary, consider providing a second, parallel VFD instead of a standard starter. VFD prices have come down considerably since the parallel-starter bypass concept was developed. This is no longer the cost-prohibitive strategy that it was at one time.

The Hollisterian Effect, Explained

One of the best things about working with seasoned experts is that you get to benefit from their previous, um, experiences. You don't always have to learn the hard way yourself.

Sometimes, these not-quite-the-way-I-planned it episodes are actually elegant illustrations of physical principles, and deserve something more fitting than being remembered as that one time someone screwed something up. Two particular examples certainly fit this bill, and, as it turns out, they both have to do with applying direct drive in custom air handlers. The principles that they illustrate have been christened the "Hollisterian" and "Florentine" effects by our own Jake Marley, in honor of certain colleagues who shall remain unidentified for the purposes of this post. I will discuss the Hollisterian effect here, and the Florentine effect in a future post.

And, instead of dredging up the actual events that gave rise to the discovery of these principles, the gist of which I am sure most readers could figure out, I will instead focus on the principles themselves.

The Hollisterian Effect

Direct Drive fans offer some great advantages to a system designer. There are no belts to maintain, no belt dust to foul the discharge air, no inefficiencies from the belt drive and far less vibration than a belted system. But they also do carry some design limitations that must be dealt with appropriately.

The first limitation? Direct drive fans are direct drive. In other words, they are directly coupled to the motor shaft, and therefore turn at the speed of the motor. Which is great, if you have a design condition where the fan needs to turn at 1800 or 1200 or 900 rpm. If you have a design condition that requires a fan selection at, say, 1500 RPM, then you need to do pick a motor/fan system at an 'asynchronous' design condition.

No big deal, right? We've got VFD's today, so this is a piece of cake.

This is exactly where the Hollisterian effect can get you. See, VFD's are not constant horsepower devices. They are, up to 60 Hz, constant torque devices.

Let's look at the equation for motor power:

hp = (Torque x Speed)/5250

If you have a constant-torque motor, this equation simplifies to"

hp=C x Speed

Where C is a constant equal to the torque constant divided by 5250.

So, what you have got is something like this:

(click on image for larger view)

Where the HP available (the blue line) increases linearly up to 60 Hz, at which point the HP then remains constant and the available torque drops away.

So what does this mean to a designer?

Well, let's say you selected a direct-drive fan to meet your design criteria at a 1500 rpm design condition. Let's say the brake horsepower of that fan selection is 13.5 HP. You select a 15 HP, 1800 RPM motor driven by a VFD. You're good, right?

Well, let's look back at our HP equation--Applying the math, you now have only 1500/1800 (or 5/6th) of the motor hp available at 1500 RPM, or, in this case, 11.7 HP. You really needed a 20 HP motor!

If you are working with low speed fans, and you are selecting in the 400-500 RPM range, you can see that you are going to be robbing about half of the nameplate HP from the selected motor, assuming you are going to select a reasonably available standard motor speed. In the above example, that would turn the 20 HP motor into a 30 HP motor!

So what do you do? Well, one way to attack this problem is to select the bigger motor (and VFD) and call it good. Other than some additional first costs, this might be the right solution. A more elegant solution might be to see if you can't select a fan wheel with slightly shorter blades to bring your design condition in closer to a synchronous speed. Energy Labs provides direct drive systems regularly, and thus will allow you to select plug fans from 50% to 105% of the standard AMCA wheel width to address these sorts of issues. Aaon's fan selection routine in their Ecat32 software allows for variable width wheels and actually takes into account any Hollisterian or Florentine effects (discussed later) in the sizing of their motors!

Variable-width wheel selections allow you to shift the whole fan curve leftwards on the page without losing height--reducing CFM to match your needs, but preserving peak static pressure. This means you could select a fan wheel at 1800 RPM, but reduce the total air delivered by providing a 80% wheel width so that you don't exceed your design flow at the faster speed.

Or, lastly, you could instead chose a 1200 RPM motor, and just select it for 72 Hz service. As long as the motor and drive manufacturer are happy with this selection, there is nothing preventing you from over-speeding your motor.

Too Much Seriousness

Laughing babies needed.

Why You Can't Buy an NC 35 Air Handler

No, it's not because we can't make quiet air handlers.

It's because NC isn't the right criteria to use to specify an air handler's sound level.

Why not? Well, to understand that, we have to discuss what NC is, exactly.

NC levels are defined by a series of curves that define the maximum sound level at a given frequency that an ambient sound can exhibit. Stated like this, it seems simple in the extreme to apply this rating to the sound level created by an air handler--but there is one important point missing: NC is a property of spaces, not equipment. Typically, allowable NC values are determined from charts like these:

It doesn't matter if the air handler whispers or is a screamer--if the sound levels in the space are below an acceptable NC curve, the sound level is acceptable. But this resultant sound level depends on a myriad of factors--the discharge sound level from the air handler, the duct layout, the selection of diffusers, attenuation devices, and, importantly, the room itself.

Hard surfaces and small volumes will tend to result in louder overall conditions, while large volumes and soft surfaces quiet a space. And, of course, sound generated in the space or from outdoor sources (traffic, etc.) will affect the overall NC level of a space.

How can you account for these effects? Well IAC has created a simple worksheet to determine the required insertion loss criteria needed in an attenuator array to meet a given target NC level. The SNAP sheet (Systemic Noise Analysis Procedure) is a simple method of calculating the resultant sound levels due to the HVAC system in a space. We've created a simple spreadsheet that helps keep the calculations straight here. Just simply copy the NC level you wish to meet from the green-tinted table at the bottom and paste those cells into the green bar at the top of the sheet. Then enter the discharge sound level from the air handler in the blue cells. If you follow the step-by step instructions on the SNAP form, you should be able to fill out the yellow cells with the sound attenuating characteristics of the system you are designing.

What you will have after putting in all this data is a required insertion loss criteria that should help you select a sound attenuator. Once you have picked one, just input its insertion loss performance into the first red bar, and the self-noise criteria in the next two bars. A successful selection will give you a sound attenuator that brings the sound level below the NC curve you are trying to hit, and does not generate enough self-noise to bring the sound levels back up above it!

This procedure only accounts for sound generated and transmitted by the HVAC system serving the space--You will need to account for other noise sources separately. But at least it takes care of the noise source you as an HVAC designer/contractor control!

How should you specify your air handler sound performance? By specifying the outlet, inlet and radiated sound power levels. As long as you have verified that the appropriate NC level in the space will not be exceeded with the specified values, you can be assured that any air handler meeting or beating your specification will be an appropriate fit.

Zero Pressure Drop Sound Traps?!

There's an old adage in life: There ain't no such thing as a free lunch.

In the world of sound traps, the "lunch" is insertion loss (the amount of sound attenuation provided) and the "bill" is pressure drop. Generally speaking, the more insertion loss you get at a given air velocity, the more static pressure drop you pay. And, when you consider that these pressure drops are greatly increased if the sound attenuator is located close to duct fittings (see page 8-9 of this document for examples), this bill can be very high indeed for any project where duct space is at a premium.

You know, all of them.

Enter the ZAPD™ series silencer from Industrial Acoustics (Z12A series cutsheet linked).


These silencers eliminate added static pressure drop by eliminating the air constriction that typical silencers impose upon the airstream:

Traditional Silencer

This is done by keeping the fill out of the airstream, so that the airflow sees no disturbance, and thus experiences no pressure drop:

New ZAPD™ Silencers

This design can provide significant insertion loss performance, with up to 12-14 dB in the first band for some 10' models, and 35-50 dB in the center bands for some 10' models--all with negligible self-noise and no additional pressure drop!

The exterior-baffle design does create functional limits to the size of these silencers, but they are great for use in systems where energy efficiency is a high priority, or where there is very little static pressure available, such as in VAV terminal boxes or water-source heat pumps.

Your free lunch just arrived. And it will pay you in energy savings for the rest of its life.

Evapco ESWA: The Most Efficient Fluid Cooler on the Market

Recently, Evapco introduced a new fluid cooler design that blows away other traditional units in efficiency and sound performance.

The secret? They re-thought how to design a fluid cooler.

In their testing, they found that the most efficient heat transfer occurred in a fluid cooler coil when the coil was completely flooded with water. However, in a traditional fluid cooler design, this condition could not be attained because air flow was needed over the coil in order to evaporate a portion of the spray water pouring over it.

Traditional Fluid Cooler Design

A little out-of-the box thinking led their engineers to realize that there were two heat transfer processes that really mattered in a fluid cooler:

1. The spray water cooling the fluid in the coil by conduction
2. The air cooling the spray water water by evaporation

Both of these processes were optimized in different conditions. So they decided to separate the two processes from each other:

New ESWA Design

The new ESWA cools the spray water with conventional cooling tower fill and then, only after the water is cool, floods the coil for optimal heat transfer.

The result? A fluid cooler that uses 30%-50% less energy than a traditional induced draft cooler, and up to 80% less than a forced draft tower!

And there are other benefits, too. Since the water basin is completely enclosed, the splash noise from the basin is attenuated, making the ESWA one of the quietest fluid coolers on the market. The basin is also accessible, making the ESWA coil extremely easy to inspect and clean. And since the air inlets are above the coil, very little 'stack' effect is created, making the heat loss from a standard ESWA in heating season less than that of a traditional fluid cooler equipped with positive closure dampers!

Evapco Hits the Big Time

Funny things happen when you type "Evapco" in to the search box at Youtube:



No, I don't understand it either

HVAC Engineering Calculations Software

Engineering Power Tools is a third party shareware program that some of us at Johnson-Barrow have found useful. It is an engineering program with an HVAC module that includes the following HVAC modules:

☆ Air Flow Thru Perforated Plate
☆ Blower Wheels
☆ Clean Room Standards**
☆ Control Valve Sizing**
☆ Convection Coefficients
☆ Duct Sizing**
☆ Fan Law Calculations
☆ Fluid Flow in Pipes
☆ Heat Index**
☆ Heat Loss From Insulated Pipes**
☆ Inert Gas Purge Rate**
☆ Orifice Flow
☆ Pipe Sizing Tables

• Air
• Natural Gas
• Water

☆ Pressure & B.P. vs. Altitude**
☆ Properties of Air
☆ Psychrometrics
☆ Psychrometrics II**
☆ Refrigerant Vapor Pressure**
☆ Safety Ventilation
☆ Saturation Tables
☆ Solar Radiation
☆ Standard Atmosphere Data

• Calculations**
• Table (SI Units)**
• Table (US Units)**

☆ Steam Pipe Sizing
☆ Temperature Conversions
☆ Water Hammer**
☆ Wind Chill Factor**

**Available in "Plus" version

If it looks useful to you, you can try before you buy.

Best Practices for Data Center Design

A 2006 study published by Lawrence Berkley National Laboratory, Environmental Energies Technologies Division examines best practices for energy-efficient data center design. Some of the best practices highlighted include well-understood and accepted practices like ‘right-sizing’ central plants and using hot and cold aisles. However, the article makes a strong case for the use of air-side economizers for the minimization of energy-using refrigeration and direct evaporative cooling for humidification. Read more here:

Best Practices for Data Centers: Lessons Learned from Benchmarking 22 Data Centers

VAV Static Pressure Control

Most VAV systems require that a minimum duct static pressure point be maintained to ensure operation of the air terminals. This means that even at very low flow rates, an elevated static pressure needs to be provided, above that expected by following the traditional system curve. As the fan speed drops to match the required flow, the intersection of the actual system curve and the fan curve moves the fan closer and closer to stall.



What is the effect of this interaction? And how can you select a fan to best deal with this problem? Energy Labs has done a valuable study of this effect and it can be found here.

Cool Ways to Conserve Water

A few years back, I had an article published in the April 2005 issue of Plumbing Systems and Design Magazine that highlighted the many ways to optimize the water saving performance Cooling towers.



You can read that article right here.