Aaon Rolls Out 410a Digital Scrolls!

Digital Scrolls offer great advantages over standard scroll compressors for rooftop packaged systems. Aaon was the first manufacturer to offer digital scrolls for their R-22 systems--but the phaseout of that refrigerant is looming.

The latest version of the Aaon Ecat32 software includes new offerings of digital scrolls for selected units using R-410a. These advanced compressors are available in the following units for the HFC refrigerant:

• 230/3/60 – RM-006; RM-013; RN-026
• 460/3/60 – RM-006; RM-007; RM-013; RM-015; RM-016; RN-026; RN-031

Stay tuned for more offerings as Copeland rolls out their R-401A digital scrolls!

Greening Small Rooftop Packaged Units: Heat Recovery

This article on 'greening' rooftop packaged units is the third of the ‘Greening Small Packaged Units’ series and addresses the use of exhaust air heat recovery in these types of systems.

Heat recovery is a well-understood and accepted method of energy conservation. However, the energy saved comes at a cost. Generally, an air conditioning system that has heat recovery capabilities operates with higher pressure drops than a system without heat recovery, and there may be other parasitic loads that are required to run the heat recovery equipment.

Energy codes generally require heat recovery on systems that use a significant amount of outdoor air, since it is a reasonable assumption that on such systems, which have very large ventilation loads, the amount of energy saved will greatly outweigh the additional energy required to operate the heat recovery equipment. However, depending on the operating conditions, there usually are energy benefits for systems that operate with even very minimal outdoor air requirements.

For an owner or designer trying to decide whether heat recovery is right for a particular application, it is important to know what these benefits are in terms of energy cost reductions, payback or return on investment, and, more and more frequently, carbon emission reductions.

For rooftop packaged units, the heat recovery product of choice is the heat wheel. The industry has settled on this product for many reasons, including first cost, footprint, efficiency and layout considerations. Aaon uses the Airxchange wheel, which is an ARI 1060 certified heat recovery device.


As with their rooftop economizers, Aaon provides this efficiency option as an integrated, factory installed option. This greatly reduces on site labor, eases commissioning, and ensures the owner of the energy benefits of their investment.


(If field-installed RTU economizers have a high rate of failure, imagine how often field installed heat recovery wheels are a commissioning problem!)

To aid in the heat recovery analysis, Airxchange has provided a free software program (registration required) to calculate the energy and cost benefits of applying their heat wheels on air-handling systems. This makes it very easy for an engineer to do a bin-data analysis of the benefits of this option. Given a particular heat wheel and some basic information about the RTU it is serving, it will calculate the gross heat recovery for cooling and heating hours, as well as calculate the additional fan energy required to operate the wheel. It will also perform a simple economic analysis calculating a net dollar savings when using the heat wheel.

An analysis of a 16 ton Aaon RM unit (pdf) shows the net energy savings available using a wheel on this type of unit. In the above analysis, a 5,200 CFM supply air system is compared looking at conditions of 100% OA and 30% OA. In both cases the analysis (using Seattle bin data, a 5 day week and typical office hours of operation) shows net energy cost savings, about $500/year on the 30% OA case, and about $1,700/year on the 100% OA case. Almost all of those savings come from the heat required to offset the ventilation load during the winter—the cooling savings are small by comparison.

However, the effect of the wheel on cooling is important in one respect--the use of the heat wheel may allow the designer to reduce the cooling (and, of course, heating) capacity of the RTU. In this example, the wheel adds 1.4 tons and 84 MBH to the cooling and heating capacity of the 30% OA system, and 3.7 tons and 230 MBH to the 100% OA system.

These ‘free’ tons of capacity that you gain by using the heat wheel effectively allows your cooling system to operate at a higher actual IPLV than is calculated in the ARI rating of the unit. ARI has acknowledged this in the publication of ARI Guideline V (Calculating the Efficiency of Energy Recovery Ventilation and Its Effect on Efficiency and Sizing of Building HVAC Systems). This guideline basically defines an efficiency rating for the heat recovery system (RER) and a ‘combined efficiency’ rating (CEF) for the entire system, accounting for the EER of the RTU and the RER of the heat wheel. This CEF is calculated in the Airxchange software linked above


If the goal of a design is not just energy savings, but carbon emission reduction, the wheel’s advantage is obvious. Every btuh that is recovered from the exhaust air is less natural gas that would need to be burned in a gas burner (the most common form of heat for these units in this region). But there is one other powerful way in which wheels can leverage energy savings or reduce carbon emissions: they can be used to greatly increase the applicability of a heat pump cycle for heating operation. In an Aaon unit, the entering air into the refrigerant coil needs to be 45º F or higher for the heat pump system to provide any heat. In the example reviewed above (RM16) the mixed air at a design heating day in Seattle is pre-heated to nearly 50 º F for the 100% OA case—well above the minimum needed for HP operation! And although capacity drops off, an air-source Aaon heat pump will still operate at conditions as low as 17 º F ambient. Converting the system to a water-source HP greatly improves the heat capacity at even the coldest days—and by reducing the amount of heat required from the ground, the use of the heat wheel can help keep ground loop costs down, too!

Converting a system from gas heat to heat pump operation has a large energy and carbon reduction benefit. First, it transfers the heating energy source from a high embodied-carbon fuel to electricity, which in the Pacific Northwest is considered a nearly carbon-free energy source. And it provides an advantage over electricity because, even with heating COP’s on the order of 1.5*, it greatly reduces the amount of utility electricity required to do the same amount of heating.

*at extreme conditions—moderate conditions greatly improve this performance

Adding Mechanical Cooling to Indirect/Direct Evaporative Systems

On this website, I have discussed indirect evaporative cooling, direct evaporative cooling and systems that combine the two into indirect/direct evaporative cooling. As we saw from the last article, however, there will likely be many applications where additional cooling beyond what can be attained by evaporative methods is necessary to keep a space comfortable. In these cases, we need to add mechanical cooling into the mix. But it is important to understand how to do this—there are tricks that can preserve most of the energy benefit of the evaporative cooling systems you have designed into your system.

The first trick is to determine where to put the mechanical cooling coil. The temptation might be to think it would be a mistake to put it in as the last component in the air flow. This is because it seems you would then simply be removing a lot of the latent effect that you are putting into the air in the direct evap portion of the system. However, inspection of psychrometric processes indicates that this is not the case. In fact, anytime the dew point of the air leaving the indirect evap is lower than the desired design air drybulb, you will see an advantage in running the direct evap section if it is located upstream of the cooling coil. A quick example will demonstrate this effect: Let's look at a 12,000 cfm system providing 60º supply air at an extreme sensible ambient weather condition in Seattle, WA. We will first look at this system with a standard mixed air (20% OA) arrangement with a traditional cooling coil. Then we will superimpose a three-stage indirect-mechanical cooling-direct system and compare the energy performance:

(click for larger image)

In the above image, the traditional mixed-air psychrometric process is indicated in red and the evaporative process in purple (with blue indicating the mechanical cooling portion). Both systems start at the ambient OA condition of 95º/68º. The standard cooling system then mixes the OA with return air in a 20%/80% proportion and then cools sensibly to a 60º leaving air temperature. (The cooling process in the traditional system is shown as purely sensible, but in reality, the system would likely use a return air bypass configuration to allow the portion of the supply air to be super-cooled to achieve latent cooling and thus prevent any latent load in the space from steadily building humidity through multiple passes through this pyschrometric process.) Note that the sensible cooling load in this system requires about 20 tons of mechanical cooling.

In the three-stage evaporative system, 100% OA is first indirectly cooled to to the condition at point I/D evap + R2. Then, mechanical cooling takes over to point I/D evap + R3, after which the direct evap section evaporatively cools to a 60º LAT condition at I/D evap + R4. I have then shown a sensible heating process from the LAT to represent the zone load to demonstrate that this will provide a very comfortable resultant air condition in the space at I/D evap + R5. Note that this is true even if there is a significant latent component to this load (The resultant room temperature is approximately centered in the pink zone that represents the ASHRAE summer comfort envelope). For this analysis, the direct evaporative system is operating at full capacity, and the cooling coil is modulating to provide the desired LAT DB. (Since this system is 100% OA, we are not concerned about humidity levels building up in the space as in the recirculating system.)

The first thing that should just jump out of this is that the evaporative system requires less than HALF the mechanical cooling of the traditional system--while providing the increased ventilation benefit of 100% OA! And this is neglecting the additional latent cooling load that would probably be needed to maintain humidity levels in the space with the traditional recirculating system. To add to the IEQ benefit, the direct evap section works as an air washer and effectively increases the filtration of the air to improve IEQ beyond that of a traditional 100% OA system.

Now lets compare the three-stage system we just examined (with a indirect/cooling coil/direct arrangement) to that of a three-stage system with an indirect/direct/cooling coil configuration:

(click for larger image)

Two things should be obvious in this example. First, the mechanical cooling load is even lower than in the previous example: down to under 5 tons! Thats about a quarter of the load for the traditional system, and a little more than half of the load of the evaporative system with the cooling coil before the evap section. Second, and this is the secret behind the reduction in mechanical load, the air leaving this system is significantly closer to saturation than the previous example. In other words, despite the fact the leaving dry bulb temperatures are the same in both cases, in the latter case there is more latent heat in the supply air. The evaporative cooling process before the coil allows the system air to hold more latent energy but yield the same sensible condition for conditioning the space. A quick inspection of the comfort zone indicates this air is perfectly suitable to provide an acceptable comfort condition, even with a reasonable latent load.

Earlier, I said there should be an advantage to running the direct evap upstream of the cooling coil if the dew point of the air entering the direct evap was colder than the design air DB. In this case the dew point of the entering air is 53 degrees--which is quite a bit cooler than the 60 degree design point we are looking for, so thus we gain the advantage seen. What about the case where the entering air is too moist? Let's look at a system where the entering air dew point is well above the supply air temperature:

(click for larger image)

In this case, the OA enters with a 62º dew point. It cools through an indirect section to about 72º/65º, and then directly to a cooling coil to reach the leaving air temperature of 60ºF db. Since the enthalpy and WB lines are nearly parallel, it seems there is very little advantage to using direct evaporation to get the air leaving the indirect evap section to saturation and then cooling it. But, importantly, there is certainly no disadvantage, (other than the electrical draw of the pump). Also notice that again, this system provides a significant load reduction compared to a standard system even while providing 100% OA!

Running through that process actually shows a slight advantage for using the direct evap section:

(click for larger image)

This cooling advantage should be confirmed for your specific system since it is highly dependent on the latent capacity of the cooling coil, and is offset by the pump energy and some small increase in system static pressure when the direct evap media is wet.

However, keeping the direct evap pump running even in these conditions provides several advantages besides energy savings:

• Simplifies the control scheme
• Provides IEQ benefit of air washing
• Increases the life of the direct evap media by reducing cycling of the evap pump

Whether or not it makes sense to use the direct evap portion of your system in times of high ambient moisture is a decision that can change depending on the particulars on any given project. But if there is a net energy penalty for using this system when the OA dew point is high, one can see from this analysis that the penalty is slight and that it would only occur for very few hours a year.

Greening Small Rooftop Packaged Units: Variable Air Volume

Second in the "Greening Small Rooftop Packaged Units" series.

Variable air volume systems are an accepted energy conservation strategy that has gained wide acceptance in the HVAC industry. And HVAC systems provide other benefits, too, including improved occupant comfort and flexibility.

The energy benefit of VAV systems comes primarily from the ability to reduce fan energy use when the full capacity is not needed. Since the fan system is typically sized at peak load, using a constant volume system means that you essentially waste fan energy for 95% of the operating hours of your system. Since fan power decreases with the cube of the speed (theoretically--motor amp draws at low speeds plateau, reducing savings in practice), the fan savings can be significant.


In fact, ASHRAE considers the potential for energy savings with variable volume systems so great, that they are considering revising standard 90.1 to require this feature on single-zone systems, in addition to the current requirement on multiple zone systems.

But there is a catch for designers using rooftop packaged DX units. Very few manufacturers provide VAV enabled units for smaller tonnages. Below about 20-50 tons, there is very little on the market to service this need. Aaon, on the other hand, offers a full line of VAV units down to capacities as low as 2 tons. And they configure their units to use either air-cooled DX refrigeration, water-cooled DX refrigeration, or chilled water cooling!

Part of the problem with using VAV at smaller tonnages is that for DX systems, the size of the smallest compressor in the system is a considerable portion of the entire cooling load--as much as 100% for single-compressor systems. This means that as you vary the leaving air volume, the capacity of the cooling system stays the same, greatly decreasing the leaving air temperature. In most cases, this will cause the DX coil to frost, which leads to all sorts of problems for the system. This drawback is generally dealt with by installing a hot gas bypass on the first cooling circuit. However, this strategy works against the energy conservation intent of using a VAV system in the first place, since the HGBP imposes a false load on the compressor system, and the compressor draws full amps even at partial load.

In the example below (click here for full pdf of selection), the compressor on a 5-ton VAV unit draws more energy than the supply and exhaust fans together--nearly twice as much!

(click image for larger view)

You can easily see that in some systems a VAV unit operating with a hot gas bypass could actually use more energy than a constant-volume system with simple on-off compressor control. Of course the latter system may cause some comfort problems that the VAV system would avoid, but it would cost you energy to gain the added comfort.

Aaon has elegantly addressed this drawback by their use of digital scroll compressors allowing you to vary compressor capacity linearly to match system load and avoid freezing your coils--and to do so in an extremely energy-efficient manner.

In 2004, the ASHRAE Journal published a study (pdf) that examined possible advances in energy efficiency in rooftop packaged DX units. In it the researchers created a high-efficiency 10 ton unit configuration:

Based on the initial energy and cost analyses, we developed a design configuration incorporating the best design options:
• Increased heat exchanger size to achieve an EER of at least 10.3, consistent with the ASHRAE 90.1-1999 requirement for 10-ton electric-heat rooftop units;
• Variable air volume using an induction motor and inverter;
• Energy recovery wheel (ERW); and
• Economizer.

This unit was also tested in a configuration that included a variable speed compressor similar to the Aaon digital scroll. The researchers concluded that the base unit, without the variable speed scroll, reduced energy costs by 25% compared to a constant-volume unit. The variable speed compressor was shown to further improve the part load performance.

The proposed unit configuration, significantly, is extremely similar to the example unit above. In other words, the 'future energy-efficient unit' of 2004 is available as an Aaon catalog unit today!

High Efficiency VAV Unit of the Future

Aaon High Efficiency VAV unit of Today

Customers LOVE their Aaon

We know that people love their Aaon units, but this seems to be taking things a little too far!



Seriously, though, congratulations to Digital Forest and many happy years of cooling!

Read more about the installation here, including this time-lapse video of the installation and crane pick:

Greening Small Rooftop Packaged Units: Economizers

Introduction: Small rooftop packaged air conditioning units are sold in staggering numbers in the United States. As such, they represent a very large portion of the installed and future energy use in the built environment. This article on 'greening' rooftop packaged units is the first of a series that will address opportunities to increase the efficiencies of these units, and highlight JB products that can address these opportunities. Each article will discuss a different facet of efficient rooftop packaged unit design. This first installment will discuss the impact of effective economizers for rooftop packaged units

It is well established that air-side economizers save energy in the Pacific Northwest. And this stands to reason when you look at a graph of where the bulk of Seattle weather bin data lies:

(click for larger image)

The majority of the bin hours per year lie to the left of the 55º line, indicating that an economizer system would eliminate the need for mechanical cooling altogether during these hours. And nearly all of the hours are located to the left of the 75º line, where ambient temperatures would be lower than return air temperatures in a cooling system--allowing the system to offset some mechanical cooling load by using outside air.

When you consider that the use of outside air also brings IAQ benefits, it is clear why air-side economizers are such a compelling strategy for Northwest mechanical systems.

But there is a problem with air economizers in small packaged units: Too many of them don't work properly in the field. The reason for this is that for most small rooftop packaged cooling units do not have factory installed economizers. The standard of the industry is a bolt-on option that is shipped as a separate assembly to the jobsite for installation by the installing contractor. In some cases, they may not even be available at all for some duct configurations.

Typical small packaged unit economizer instalation

In practice, these economizers have a high rate of failure. The issue of non-functional economizers for small rooftop packaged units is significant enough that Puget Sound Energy includes re-commissioning of these devices in their Commercial HVAC Rooftop Unit Premium Service Rebate (program developed with the assistance of NEEC). And the Califorina Public Interest Energy Research program (PIER) goes further, recommending to owners and designers:

Specify reliable, factory-installed and -tested economizers with direct-drive actuators and low-leakage dampers.

That's exactly what Aaon provides on all of their units down to 1 ton.

(click for larger image)

Aaon's rooftop unit design provides inherent energy advantages over the competition. And factory-installed economizers are just one of many.

Extra: PIER software to estimate economizer savings.
Want free psychrometric software? See our offering here.

The Importance of Cooling Tower Maintenance

Cooling tower maintenance is not just critical for extending the life of your equipment, but it also can significantly improve the energy performance of your mechanical system.

BetterBricks, a non-profit venture of the Northwest Energy Efficiency Alliance has summarized the energy benefits of cooling tower maintenance in their article, Optimizing Cooling Tower Performance.

This article highlights the negative effects of:

• Scale Deposits
• Clogged Nozzles
• Poor Airflow
• Poor Pump Performance
  

If you own or maintain cooling towers, this article is well-worth reviewing. And remember that Johnson-Barrow's FluidTek tower service group is a certified Evapco Mr. Goodtower service center.

Introduction to Modular Chillers

Modular chillers are a product innovation that has recently gained wide acceptance in the HVAC industry. But since they cost more than standard chillers on a per-ton basis, it might seem a unlikely that this equipment would be a very popular cooling solution. However, modular chillers offer advantages that are not available with standard chiller equipment.

These advantages can be summarized in a few points:

• Ease of Installation
• Compact footprint
• Redundancy
  

Ease of Installation

Modular chillers were originally developed as replacement chillers for existing building chiller plants. Many chillers are located in the bowels of the buildings they serve. It is often far easier to remove the existing equipment in pieces than to find a rigging path suitable to take it out of the building in one piece. Of course, this only helps if it is also possible to move the new chilling capacity into the chiller room in pieces!

Modular chillers were designed to fit through standard doors and to have a small turning radius to negotiate internal corridors without requiring demolition of existing walls.





Additionally, since these modules are light enough to ride in a freight elevator, it is usually possible to avoid crane costs for the installation project. Further cost savings are realized once the modules are in the room. Since each module has very low refrigerant volumes, the retrofit usually does not trigger codes requiring refrigerant monitoring or emergency ventilation.

Compact Footprint

In order to fit through doors and down corridors, modular chillers are designed to be extremely compact. They use highly efficient brazed plate heat exchangers to minimize their size as much as possible:



The result is a chiller plant with the smallest footprint per ton of any current option available--even if you are installing the modules in a new project instead of a retrofit.

330 ton chiller plant comparison (click for larger image)

Redundancy

With multiple, independent modules, modular chillers provide unmatched redundancy. If a single circuit is down there are always multiple other circuits operating. And providing N+1 redundancy to a modular chiller plant is far less expensive in first costs and mechanical room space than for any other chiller type. This inherent modularity allows fantastic turndown capabilities, and the part-load efficiency of a modular chiller plant is comparable with that of a large constant speed centrifugal or screw chiller.

ClimaCool Advantages

ClimaCool modular chillers were designed to take full advantage of the modular chiller design. For example, some manufacturers design their chillers for modular installation, but not modular operation. These chillers are designed with an electrical bus bar system to power all of the modules from a single power source. This may mean a slight savings at installation, but significantly degrades the redundancy advantage of this type of chiller. With a bus bar system if one chiller needs to be worked on, the entire array needs to be powered off.

ClimaCool avoids this disadvantage by powering each module independently of all the others:



Similarly, ClimaCool offers full redundancy on the water side, too, by providing isolation valves for the heat exchangers as a standard feature. Some manufacturers offer these valves as a first-cost add and they may significantly affect the chiller's footprint dimensions if added. Providing these valves as standard provides for yet another ClimaCool advantage: Easy conversion to a variable primary flow system! Modular chillers have a tight flow envelope on the brazed plate heat exchangers--each heat exchanger should essentially be considered a constant-flow device. By providing electric actuators controlled by the chiller controller on these isolation valves, the modular chiller plant can easily adjust for variable primary flow.

Another way in which ClimaCool offers advantages over other modular designs is in heat exchanger protection. Brazed-plate heat exchangers are highly efficient and very compact, but they demand very clean water to prevent clogging. All manufacturers of modular chiller equipment require straining of the system water before it enters the exchanger. Some manufacturers provide large-mesh strainers that are mounted in the headers serving the heat exchangers at each module. This approach requires an annual back-flush of the heat exchangers to clean out the debris that inevitably passes through the mesh. It also discourages proper maintenance, since the strainers are hard to get to and are therefore often ignored until clogging causes flow problems. ClimaCool takes a different approach, using small-mesh basket-type strainers outside of the headers to prevent heat exchanger fouling. This eliminates the need for annual back-flushing, and greatly eases maintenance. They also offer a deluxe 80-mesh high-capacity strainer option for especially dirty water or for systems where maintenance man-hours are limited:


Additionally, ClimaCool provides, as standard, convenient back-flush hose-bibs to allow this sort of maintenance as needed without requiring disassembly of the chiller header or taking the other modules off-line.



And, of course, ClimaCool offers chillers that comfortably exceed minimum energy code requirments:



Efficiency, redundancy, compact size and ease of installation: All reasons to consider ClimaCool modular chillers for your next chiller project.

Your Next Energy Conservation Measure May be a Quiet Fan

It might sound strange, but a super low sound axial cooling tower fan is an energy-saving device--But not because it uses less energy than the fan it replaces, because it doesn't. The reason is a little more complicated than that.

But first it makes sense to review a few basics about cooling towers.

The Basics

There are two major types of cooling towers and fluid coolers: Induced Draft and Forced Draft.


Forced-Draft towers utilize centrifugal fans to blow air through the tower. The air is forced into a pressurized plenum inside the tower and then through the fill. This means that access into these towers is limited, since doors must be able to resist pressure without leakage and tend to be small and difficult to use. This also makes it difficult to observe the basin of these towers while operating in order to troubleshoot problems if necessary.



Induced draft towers use an axial fan to pull air through the tower, creating a negative pressure within the tower. This allows the unit to be built in an open configuration, making access and observation far easier. In general, induced draft towers cost less, are easier to maintain and, importantly, require about half the fan horsepower to do the same cooling as a forced draft unit.

In fact, there are only a few reasons why you wouldn't use an induced draft tower in preference to a forced draft tower:
1. Height restrictions
2. Static pressure capacity for ducted installations
3. Noise Control

If you project requires an extremely short cooling tower or needs a tower to be installed indoors with ducted inlets and/or outlets, there is a good chance you will need to use the less efficient forced-draft tower. And, until recently, it used to be that the same was true of sound-critical installations. But not any more.

The acoustical benefit of forced draft units are twofold: First, they are quieter than induced draft units right out of the box. (Low-profile forced-draft units are especially quiet.) And, secondly, they can easily accept sound attenuators to make their already quiet performance even quieter. The price you pay, of course, is fan energy and dollars. Attenuators require that you expend even more money and fan energy than the already more expensive and less efficient bare forced draft unit.



Th super low sound fan (SLSF) changes the playing field. The addition of the SLSF on an Evapco induced draft cooling tower does not affect the efficiency at all--the performance is the same with and without the quieter fan. And since the fan knocks 9-15 dBa off of the sound power of the tower, suddenly induced draft fans are competitive in sound level with a forced-draft unit. Generally speaking (and each application is different) a SLSF induced draft unit is just about as quiet (if not quieter) than a forced-draft unit of the same capacity--and very competitive in first cost. And further sound abatement is available to shave a few more dB off of the sound level.

This development makes it very possible to meet demanding noise criteria and still retain the sizable energy benefits of the axial fan. And with innovative products like the Evapco ESWA, the lowest-sound option can even be the energy leader!

Hearing is believing, so Evapco has provided a few video clips to help you get an idea of how significant this sound improvement is [videos may require Internet Explorer to work properly]:

Video 1
Video 2

More information on low-sound options is also available here (pdf).

After 10 years, Johnson-Barrow and JCI/York Announce Split

JCI Worldwide services has recently announced a new HVAC products distribution strategy and support for its Building Efficiency business division which includes York international and JCI controls. In an effort to integrate its marketing efforts towards owners, contractors and engineers, JCI will be reorganizing its go-to-market strategy in 15 plus major markets around the nation, including Washington State. The plan is for York to be more directly marketed through the JCI offices with the assistance of an independent rep organization as a support service to the controls division. This will also include the integration of the unitary (Consumer Products) division into their new business strategy.

Wayne Garret, Western Regional Sales Manager for JCI, says the move is designed to better integrate the three aspects of the company. “We needed to get a single face to the customer regarding who JCI is. Unitary, Controls, and Engineered products needed to be more closely tied than was presently the case. This will help our customers get a better understanding of the JCI depth.”

Mark Johnson NW sales manager was asked what things will happen as a result of these changes. “In Markets such as Seattle, major changes are already underway to integrate controls, engineered products, and unitary equipment. Air Cold, a division of Ferguson has discontinued its representation of the York Unitary products. Johnson Barrow will be terminating it’s relationship with the Engineered products division, and finally JCI in Bothell, WA will direct marketing strategies for the Washington and other Pacific NW markets.”

Patrick Hollister of Johnson-Barrow commented that this announcement did not surprise their organization. “For years we have been figuring that York would reposition itself to better integrate the unitary and controls division into a more uniform marketing organization. Thus, over the last couple of years we have been positioning ourselves to be more diversified in our product and service offering. Evidence of this can be seen with the addition of AAON, Smardt Chiller, and Climate Cool. We want to maintain our independence as a solutions oriented company focused on unique products that provide value for our customers”.

When asked to comment on the JCI announcement, Gary Bodenstab of Johnson-Barrow echoed Hollister’s observations. “Look at the magnitude of change around the country. US Air has replaced Air cold in the major SW markets, Ferguson has dropped York. The controls division is in flux trying to regain market share in the NW markets. We figured some major change was underway in the NW—it was only a matter of time. We wish the best for JCI and its new strategy. It’s now time for Johnson-Barrow to focus on our roots of independent companies dedicated to market innovation, energy conservation, and customer value.”

FREE Psychrometric Software

If you like the charts that I created to show psychrometric processes (like here), you're in luck.

Johnson-Barrow has made a deal to make this software available to our customers. We've also made some major aesthetic changes that we feel are a huge improvement: We've changed the chart colors and added a dynamic new logo!:




Pretty spiffy, huh?

The free version of the software allows you to do some simple analysis and process charting--allowing you to create high quality psychrometric charts for presentations or personal use. Additionally, a copy of the free version gives you a sizable discount off of the full version that is available here (chose HDpsychchart Pro Edition OEM upgrade SKU# HD1001, select Johnson-Barrow as OEM company).

The full version allows you to do the following:

• Model mixing, direct evaporative and humidification processes
• Create charts at any elevation
• Add climactic bin data
• Create flow charts
  
• Create detailed psychrometic process data points tables
• Vary the limits and extents of the chart axes
• Show ASHRAE winter and Summer comfort zones
• Project constant condition lines for ease of analysis

And many more tasks that make psychrometric chart analysis easy. The software also comes with additional tools like fan law calculators and even a loan payment calculator!

The free software is available for direct download on the toolbar to the right, on our www.jbarrow.com main page or right here.

UPDATE:

Some users have reported a problem with the software that prevents proper registration of the file. A new file that does not have this problem will shortly be uploaded.

A Giant Golf Ball in Your Outside Air Opening

"What the heck is that thing"

That's often what we hear when we introduce people to the Tek-Air IAQ-Tek outside air monitor.

Sometimes elegant solutions to difficult problems are a little surprising.

The Problem: Outside Air Measurement

Getting an accurate measurement of outdoor air flow is a vexing problem for HVAC professionals. It's about the most difficult airflow measurement situation around.

The air is often moist, dirty, and at extreme temperatures. Most air inlets, especially on packaged air handling units, are poorly designed for accurate flow measurement. The airflows in the inlets are usually highly turbulent, non-uniform and at very low velocities. Wind impinging on the inlets can cause large flow fluctuations. This is tough duty for any flow measurement system.

To make matters worse, the outdoor air flow is one of the more important measurements in an building HVAC system with big implications to the indoor environmental quality and energy use of the building. Understandably, LEED® guidelines encourage the use of outdoor air monitoring.

Generally, the air flow velocities available at an OA probe need to be slow enough to prevent moisture carryover--this makes traditional pressure measurements with pitot-type sensors very unreliable, because the signal from these probes varies with the square of the velocity. At low velocities, the noise from turbulence, wind and other sources simply drowns out the signal with a very low signal-to-noise ratio.

This has led to the use of hot-wire anemometers (thermistors) in this application. These products provide excellent low velocity air flow measurement, but this application provides challenges unique to this technology. In particular, dirt and moisture build-up on the sensors will cause the calibration to stray and upstream filters are usually recommended. Additionally, since the sensors measure the velocity at a discrete point in the air opening, a large number of sensors are required to adequately provide a representative flow measurement for large openings. And even with a large number of sensors, the turbulence and non-uniformity of the airflow in an outdoor air hood or behind a louver makes it very difficult to get a useful reading, no matter how accurate each sample measurement is.

A Different Way

Tek-Air saw the above difficulties and looked for a new solution. And that's why they developed this unique airflow sensing device.

Most flow sensors are designed to minimize the disturbance they create in the airflow. Tek-Air realized they needed to take a different approach for this difficult challenge:



The IAQ-Tek probe is large--really large. In fact each sensor body is about 8" in diameter and has over a dozen pressure ports in it. It dampens out the effect of localized turbulence on the airflow measurement by forcing a large-scale diversion of the airflow in the inlet. The measured variable is the average pressure difference between the ports on the front of the sensor body and the ports on the back of the sensor body. The 'golf-ball' dimples in the face of the sensor ensure stagnation of the airflow to significantly decrease the effect of localized turbulence and ensure a steady, accurate reading.

The unique design of this probe allows accurate readings at 6-8" behind an oudoor air louver, and directly in front of dampers. No prefilters, air straighteners or sections of straight duct are required.

So what does this give you?

• Accurate and stable low velocity readings from 75 to 750 fpm
• Immunity to signal noise
• Great flexibility in application

These probes can get accurate measurements in places you wouldn't even consider other OA probes:






The units come with a temperature and density compensating transducer (-40º to 120º F), for accurate measurement in all conditions. And each system comes with a Nema 4x monitor with LCD readout for local observation. They are rugged devices that need no significant maintenance requirements and can even be hosed down, if needed, for cleaning.

Can they really be accurate in such tight conditions? A test with the unit installed 4" behind a louver outlet, with 18" between the louver and an OA damper yielded the following results:



That's from -6% to +4% (of full range) error at velocities of 100 to 700 fpm with damper positions from full open to 45º. That's fantastic accuracy in an extremely difficult measurement condition.

So maybe you do need a giant golf ball, after all.

Rethinking Air Handler Pressure Testing Specifications

Pressure testing is one of the most important ways to ensure the quality of the cabinet of an air handler provided on your job. Leakage costs money and energy. Every CFM that leaks out of an air handler is air that energy has been expended on that is now lost. Likewise, every CFM of air that leaks into an air handler displaces air that has been conditioned, requiring more air volume to do the same duty. And air leakage can have other negative effects, like causing condensation on surfaces or allowing unfiltered air to enter the system.

So to prevent leaks as much as possible, an experienced engineer will specify leak testing on the air handlers provided on their jobs. This is accomplished by blocking off all openings into the unit and pressurizing it (positively or negatively) with a pressure blower and then measuring the airflow into (or out of) the unit to maintain a test pressure:


Several decisions have to be made when deciding how to test the unit:

1. When to test (At factory or at jobsite)
   
2. How many sections to test
3. Positive or negative test pressure
4. What pressure to test to
5. How to set the failure criteria

When to test the unit

There is a very strong argument to be made that the only pressure test that matters is the pressure test performed on the site. After all, it really matters little to the final built out project if the unit performed flawlessly on the factory floor. The actual performance in the field is all that anyone really cares about. If only a single pressure test can be fit into the budget, it stands to reason that the field test is the one most critical to the overall quality of the delivered project.

However, there is a case to be made for a factory test, too. The field pressure performance of an air handler is not only a function of the manufacturing process, but is also strongly dependent on site conditions, including the flatness of the support the air handler sits on and the rigging and mating procedures used by the contractor. Once a unit is on site, it is sometimes very difficult to determine where the failure lies if it doesn't meet the specified leakage rates. In a worst case, you might have the manufacturer, the shipper, the installing contractor and the general contractor all pointing fingers at each other. If the fault actually lies in a factory defect, the problem could be found and corrected in the most controlled environment possible with a factory test.

So the real answer? Both, if you can afford it. And if you catch a problem before it gets to the field, you might feel you couldn't afford not to do both.

How many sections to test?

Many pressure tests specifications treat the entire air handler as a single section, and require a single test for the whole unit. Some break the unit up into positive and negative pressure areas (upstream and downstream of fans, respectively) and call for them to be tested separately. Each method has its advantages and its disadvantages.
The first consideration is cost. Each test costs time and energy that will be reflected in the overall price for the job. Requiring multiple tests on a single unit will raise the cost of the air handlers to the owner. This will also require more time at the factory and on site, and could affect overall completion dates in some instances.
A second consideration is accuracy. Since multiple tests allow the unit to be tested to the actual pressure condition the sections will see, presumably this will give you a better idea of the leakage than a single test. However, there is an appreciable amount of leakage within the air handler at the internal wall that will be factored into this measurement (and double counted!) that will unrealistically penalize the performance of the unit. This is especially significant, since the internal openings within an air handler (at fan walls, usually) that need to be blocked off to perform the test are rarely built in a fashion that allows for an effective air seal to be created for these temporary tests. Even a small amount of leakage at these internal walls can mean the difference between passing or failing a tight leakage criteria.
Consider a single test for units unless special requirements drive a need for multiple positive and negative tests.

Positive or negative test pressure?

If you are making multiple tests on multiple sections of an air handler, the answer to this question is simple: Test to the conditions each section will see in operation. If you are performing a single test on the entire unit, then you may want to carefully consider how you wish to test the unit. In general, there are some leaks that will open up under one pressure condition and will close under the other. Generally it is accepted that leaks at panel seams tend to close under negative pressure and tend to open under positive. Doors that swing out tend to behave similarly, while doors that swing in behave in the opposite fashion. Seams at test closures can do either based upon the method of construction of the closure. So there is no easy rule of thumb that says one method is preferable to the other. In many cases it makes sense just to find the point of extreme pressure in the unit under normal operation, determine if this is positive of negative, and test to that condition. This condition is easy to find by simply calculating the pressure condition at each section in the unit by starting at the external static pressure and working towards the inlet, adding back pressure losses at each internal component, and subtracting the fan static pressure increase at the fan wall. Each open section of air tunnel will have a pressure associated with it, with the extremes usually falling at the inlet or discharge plenum of the supply fan. These are the maximum pressures the air handler will see in operation--and usually one will be significantly further from ambient pressure than the other.

What Pressure to test to and
How to set failure Criteria

The above two considerations go hand in hand, so I will deal with both of them together. Traditionally, specifications are written so that a certain percentage of the total air flow is allowed in leakage (usually around 1-2%) at a specific test pressure. How the leakage percentage and the test pressure are determined varies from engineer to engineer and job to job. Sometimes the test pressure is based on the total fan static, sometimes it is based on the actual cabinet pressure, and sometimes it is based on a nominal test pressure (like, say 10"). In either of the first two cases, there is usually a sizable safety factor applied.
The allowed leakage percentage varies, but it is usually in the low single digits.
While this has been the standard in the industry for many years, there are some significant weaknesses in this approach to testing. First, the actual leakage rate measured in the field is determined essentially by the total face area of all the leaks in the system--this is a function of cabinet size and construction quality, not of supply air flow. By tying success or failure of the test to the fan capacity of the system, you are favoring simpler, smaller air handlers over larger, more complicated air handlers.
Imagine a simple 10,000 CFM air handler operating at 8" of total static with just a fan, a heating coil, prefilters and a mixing box. Then imagine that same air handler, but with a cooling coil, high efficiency final filters, return fan and air blender:

(click for larger image)

In the example above, you have two 10,000 CFM air handlers, one with 297 square feet of cabinet area, the other with 630 square feet of cabinet area. The large air handler has more than double the cabinet area, and more leak points such as doors, dampers, coil penetrations and shipping splits--yet both would be required to meet the same leakage rate in a test--in this case, say, 200 cfm at a 2% leakage criteria.

A further complication would arise with the pressure selection. A common pressure test criteria is 1.5x the maximum fan static pressure--or in this case, 12". It is often difficult to find a pressure blower with a static capability in this range--It might be impossible to effectively provide this test in a timely manner on a job site. Additionally, many components (especially doors, when tested in a pressure condition opposite that they would see in operation) leak uncharacteristically at higher pressures. And the unit would never see pressures anywhere near 12" in real operation, anyway, since the fan will typically create an area of negative pressure in the inlet plenum, and positive at the discharge. The maximum amplitude of either pressure is, by necessity, less than the total static pressure capability of the fan. Thus the leakage rate measured in the test will be a very significant overestimate of the actual leakage that will be experienced in operation.

A different way of specifying pressure performance can address both of these complications--and that is to tie the performance to the cabinet itself, as opposed to the air flow. There is already a criteria to do exactly this. ClimateCraft recognized the difficulty with specifying pressure test criteria to arbitrary pressures and airflow percentages. They realized that SMACNA already had a pressure test criteria, the SMACNA leak class rating. The leak class of a pressure plenum (or air handler, in this case) is calculated using the following formula (from ANSI/ASHRAE 111-1988):

Leak Class	=	(leak CFM) x 100
		(Area sq.ft.) x (Test Pressure) ^ 0.65
OR
Leak CFM	=	(Leak Class) x (Area sq.ft.) x (Test Pressure) ^ 0.65
		100

ClimateCraft has adopted this method of rating pressure performance and builds their units to meet or exceed a leak class of 6. For the units above, that equates to about 146 CFM at 8" for the larger unit, and 69 cfm for the smaller (or 1.5% and 0.7% of the supply air volume, respectively).

In practice, we have found that a leak class of 6 represents excellent performance for high-quality custom air handling units of any of the manufacturers we represent and quite often can be met by the high-quality foam panel semi-custom air handlers by Aaon, too. This is, however, a very high bar for traditional commercial-grade batt-insulated air handlers.

Perhaps the biggest advantage of the leak class specification is that it encompasses both the allowable leakage critiera and the test pressure in a single number. A leak class 6 air handler will perform to the same leakage class whether it is tested at 4" or 10"--the test pressure can be chosen to meet realistic pressure conditions and to facilitate testing of the unit. It is purely determined by the design of the air handler cabinet and the execution of assembly.

Reducing Ground-Loop First Costs

Ground-Loop heat pump systems perhaps have the greatest potential for reducing energy use in the built environment than any other space-conditioning technology now in use. This potential has been long recognized by the EPA and the DOE, and represents a great opportunity for owners and designers attempting to create systems that out-perform those that are commonly built in this region.

They also have a reputation for being expensive--very expensive.

And with drilling costs in this region historically being quoted as high as $15/lineal foot, this reputation is well deserved. These prices usually put this technology out of the range of economic justification for typical projects.

So what can a designer do to minimize costs, yet still provide the energy benefits of this technology?

Add a cooling tower.

Hybrid Systems

To understand how adding a cooling tower to a ground loop saves costs, first you have to understand a simple concept about closed ground-loop systems. While the ground loop is often referred to as a "heat exchanger", the ground-loop (and the ground it occupies) acts more as a leaky heat storage battery. Unless there is sufficient ground-water movement through the well-field, most of the heat that is rejected into the ground remains there throughout the year unless it is later removed by the ground loop itself.

That means that over time, if the heat added is not balanced by heat removed, the ground temperature will continually increase over the seasons, increasing loop temperatures and decreasing system efficiency.

(graph showing increase in temperature over time for imbalanced loop of differing bore hole numbers. From here)

The best situation for a designer is when the heat added to the ground over the course of the year (by the process of cooling the building) is balanced by the amount of heat removed from the loop (by the process of heating the building). But a heat-pump does not just move heat from one source to another. Because a compressor is needed to perform this work, a heat pump always adds the heat of compression to the equation. This is a benefit in heating, since the heat of compression is added to the amount of heat moved from the loop to the building. This is a hindrance in cooling, since this compressor heat is added to the heat moved from the building to the loop. In practice, about 1.2 to 1.8 tons of heating are needed to balance out 1 ton of cooling. This means that many ground loops will see an imbalance where more heat is rejected to the loop than is removed from the loop over the course of a year. This effect can be significantly compounded (or mitigated) by the configuration and use of the building served--buildings with significant yearly cooling loads will be more affected than by buildings dominated by heating loads.

A ground-loop designer typically combats this effect by increasing the volume of the well field by increasing the number wells to a point where the relatively small amount of heat-leakage out of the well-field and added volume is enough to account for the imbalance of the system and minimize the heat gain. Thus ground loop well-fields are often sized due to the minimum requirements of either heating or cooling demand for the building. Cooling-dominated well-fields are more common throughout the US, especially in the southern portion of the country.

If the designer could correct for this imbalance and build the loop to the smaller size required by the heating load of the building, then fewer wells would be needed, and thus the overall cost of the loop would come down. One of the most cost-effective ways to provide extra cooling to balance out the loop on such a system is by way of a cooling tower or fluid cooler. When a cooling tower is used in conjunction with a ground loop, you have what is called a hybrid system.

Hybrid systems can be extremely effective at bringing down first costs of ground loop systems. A study by Kevin Rafferty of the Oregon Institute of Technology found that hybrid systems can reduce the cost of a ground loop by as much as half for some systems:


But can we expect similar reductions in first cost for the Puget Sound region, where we have a generally cool climate and a long heating season? For some systems, it appears the answer is yes. A presentation by Scott Hackel of the University of Wisconsin at the ASHRAE 2007 summer meeting investigated the cost savings possible using hybrid systems throughout the country. His study showed very significant reductions in ground heat exchanger (GHX) loop lengths for school, retail and office applications in the Seattle region:

(Click for larger image)

Hybrid loops may just make the next ground loop you consider pencil out.

Digital Scroll Compressors: Just Plain Cool

Copeland Compressors (now a Division of Emerson) has recently introduced their digital scroll compressor technology. This is one of the most interesting products to come out in a long time for the DX cooling market. But to understand why it is so cool, you first need to understand a little bit about scroll compressors in general.

Scroll compressors have essentially displaced the older reciprocating compressor designs for small-tonnage air conditioning systems. Which is much to the operator's benefit, because scrolls are inherently more reliable and require none of the maintenance that the piston-type reciprocating compressors required. But this advantage comes with a price: comprehension. Reciprocating compressors were so much easier to understand--since the compression stroke in a piston is easy to grasp and most people are familiar with this process from the similar function of pistons in gas engines.

Some smart guy had to come along and invent a highly efficient and low-maintenance compressor technology that no one can describe easily--even using curious arm gestures and words like "orbit"!

The secret to a scroll compressor is two high-precision spiral "scrolls" that are designed to mesh with each other to extremely close tolerances:


The upper scroll is stationary and the lower scroll 'orbits' in a rotary fashion:

Comparison of scroll to piston compressors showing relationship of upper and lower scroll (click for larger image)

The upper and lower scrolls continually 'pinch' off volumes of low pressure gas and move them towards the center of the scrolls, compressing the volume further and further as they work. This compression requires extremely close tolerances between the sides and ends of the scroll surfaces, since the only seal is the lubricating oil in the refrigerant circuit. If tolerances are to great, no seal is effected and the compression is lost.

Still hard to picture? This animation should make things a bit clearer:

So, great: We have a highly efficient compressor with two capacity settings: 'On' and 'Off'. If you are trying to meet a close control spec, this may not be close enough control. You would typically overshoot and then undershoot the required cooling capacity as the compressor kicks on and off. And since anti-recycle timers are required to prevent overheating the compressor motor, there is a limited number of times the cooling capacity can be switched on and off in an hour.

A new innovation allows two-step unloading to 66% capacity--but this can still be a pretty big step of control on a small refrigeration system--especially ones with only a single compressor. Wouldn't it be nice to get a fully modulating compressor with all of the advantages of the scroll compressor?

That's where the digital scroll comes in. Copeland's engineers were clever enough to realize that they could achieve this performance out of a scroll, not by modulating its capacity directly, but instead by modulating the time during which this capacity is provided. They found that if they quickly turn on and off the compression cycle, without having to turn off and on the compressor motor, they could modulate the output very closely to meet the needed capacity. The trick was finding a mechanism by which this rapid switching between active and inactive compression could be accomplished.

The solution they arrived at was elegant. They found that by merely moving the scrolls apart axially, they could defeat the oil seal between the scrolls, and turn off the compression. Then they simply needed to move the scrolls back together and compression would immediately restart.


The above graphic shows the scrolls separated to cancel out the compression cycle, and a visual representation of how the scroll would operate to provide 50% capacity--Operating 10 seconds on and then 10 seconds off in a repeating cycle.

What's the result? Very efficient operation down to 10% of full capacity:

(click for larger image)

See the savings noted above? That's compared to the commonly used DX modulation method of hot gas bypass (HGBP). It's important to note that the HGBP works by applying a false load to the compressor--it does not reduce compressor energy at all! As far as the compressor motor is concerned, it is doing just as much work as when the compressor is providing full output. In fact, the HGBP system is even more of an energy hog than is suggested by the graph above--since compressors with this device will operate at full load for extended periods of time, drawing full amps all the time, as opposed to a standard system where the compressor would turn on and off to match the load.

The digital scroll gives a DX system all the fine control capability of a chilled water system, without sacrificing energy performance like HGBP systems do. It allows effective operation of VAV airflow systems without frosting coils. It provides efficiencies unmatched in the DX market. For these reasons Aaon was quick to incorporate these compressors into their RM and RN rooftop packaged AHU lines. Unfortunately, digital scrolls are not available in all standard scroll compressor sizes and voltage ratings. And they are currently only available in R-22 compressors. This handy chart indicates where the digital scrolls are available in each RM/RN model size and for which voltages. This file is valid as of 9/22/07, and is definitely subject to change in if/when new digital scrolls are released. Additionally, R-410a compressors are expected out in the near future for 6, 7, 13, 15, 16 and 25 ton sizes in 460/3ph electrical services only. Stay tuned for the availability of those units!

Update: Aaon has rolled out units using R-410a digital scrolls, as well as new software to calculate the efficiency benefits of these compressors.

Aaaah...They're Just Jealous

Admit it: Did your mother react this way, too?

HVAC System Efficiency Tool

Designers and owners are continually bombarded by claims of equipment efficiencies. Industry groups such as ARI, AMCA or CTI have been set up to validate these equipment efficiencies to give these claims credence.

However, it isn't the equipment efficiency that drives the energy use of the building, but the overall efficiency of the system that the equipment is part of.

Steve Kavanaugh, University of Alabama Professor, ASHRAE fellow and author of the ASHRAE design guide for ground-loop heat pump systems (with Rafferty) stresses the importance of the system efficiency. He has also provided (free of charge) a handy tool to calculate the system efficiency of typical HVAC systems, given the efficiencies of their component equipment.



This tool (HVACSysEff06)is available at Steve's Geokiss website software download page.

Take a look around--there are some other interesting tools there, too.

Heating Your Showers with Your Cooling Tower

Most large buildings are throwing heat away for many hours of the year. In a large facility, this is most often accomplished by way of a cooling tower. Commonly, the cooling tower cools water from about 95º to around 85º. Many hundreds of thousands of btuh's from lighting, solar loads, equipment and any of the myriad heat load sources in these facilities are rejected to the atmosphere in this cooling process. Wouldn't it be nice if you could reclaim some of that heat and use it for a something that always requires heat input, like domestic water heating?

Sure, you could take the cooling tower water and run it through a heat exchanger to preheat the makeup water from the city utility before it enters your hot water heater, but that would only offset part of the heating load. The highest temperature you could reach would be on the order of 93º--any higher would require artificially allowing the condenser water to heat up, penalizing the efficiency of the chiller it serves.

It would be a lot more convenient if there were some way to use the heat in the condenser water loop to create higher temperature water--water that could be directly used to heat domestic water. And that is exactly why Colmac developed their HPW series of water-to-water heat pumps, specifically designed for domestic service.

These heat pumps include a circulating hot water pump and a double-wall heat exchanger as required for domestic service. They can directly heat the domestic water to temperatures of 140º or higher, using water as cold as 55º. This means that they can actually be used to pre-cool chilled water to reduce load on a chiller, as well as take waste heat out of a condenser line.

Florida Heat Pump also has a full line of water-to-water heat pumps for similar heat recovery jobs. These are a competitive alternative when domestic water service is not required, or where an external heat exchanger can be provided to meet domestic service requirements. These are also very flexible alternatives to traditional central plant chillers, with the ability to reverse cycle and provide hot water or cold, and come in convenient modular sizes for ease of installation and efficient capacity staging.

And there is no reason to stop at considering condenser water systems for sources of heat. Using water-to-water heat pumps, any source of flow that carries waste heat can be utilized to provide usable energy for your system. Why not pump heat out of your sewer lines? Luckily for the creative energy engineer, smells aren't transfered by the refrigeration cycle!