Electric Motor Primer

Aaon has published an electric motor primer that discussed the differences between AC, DC and ECM motors. It is a useful resource for engineers who want to optimize their choice of motors in the products they select and specify.


This document can be found here.

Greening RTU's: Aaon High Efficiency Units

Today's codes and customers are demanding higher and higher efficiencies from their units. And Aaon has been continually improving their product offering to stay one step ahead of the curve. Not only do they offer energy-saving features like heat recovery, foam core panel construction, digital scrolls and the best RTU economizers on the market, but they also offer industry-leading cooling efficiencies.



To help customers select the unit that is right for their project, Aaon has created this quick select guide that shows the available efficiency levels of their RTU products (Energy Star, ASHRAE 90.1, ASHRAE 189, CEE Tier 1 or CEE Tier II) for all of their units from 2 tons to 70 tons.

See what else Aaon is doing to revolutionize the efficiency of rooftop units at our Greening RTU's section of this blog!

Aaon Completes Conversion to Foam Core Construction with Announcement of New RQ

A few years back, Aaon made a committment to convert all of their air handling products to high-performance foam core panels. With the introduction of the new RQ rooftop unit (1-6 tons) this conversion is complete!


We've mentioned the many benefits of the foam core technology before, but it is worthwhile to revisit the subject to understand how much better this R-13 double-wall cabinet is than the standard single wall R-1 to R-3 batt insulation cabinet design in the industry. This detailed report on the performance of the foam core panel shows that in Seattle's environment, as much as 20% energy savings can be realized just from the better thermal performance and low leakage of this design.


But the RQ is not just better because of its high-performance cabinet. This product also makes available all of the energy saving advantages of the larger AAON RN product, like:

• Digital scroll compressors


• True Variable-Air-Volume Performance


• Integrated, high-reliability economizers


• Heat Recovery


• Heat Pump Operation


• Ground Loop Operation


• Direct Drive Plenum Fans

And offers a new energy savings feature: ECM Fans!

And while SEER's are of relatively small importance to the overall efficiency of a rooftop system in a heating dominated climate like Seattle's, the RQ boasts SEER's that are in line with CEE's Tier 1 and Tier 2 efficiency levels

With the introduction of the RQ, Aaon has set a completely new standard in energy efficient rooftop air conditioning systems.

Greening Small Rooftop Package Units: Foam Core Panels in Depth

A while back, I summarized the benefits of foam core panels in comparison to the industry standard of fiberglass batt insulation.



Now Aaon has published an in-depth, seventeen page study of the benefits of the foam core panel.


Foam core panels have many advantages over standard insulation:

• Superior R-value
• No thermal breaks
• Greater Rigidity
• Lower leakage
• Stronger damage resistance

This new study calculates the effect of these advantages over the course of a year, in heating and cooling, for buildings in Atlanta, Chicago, Houston, Los Angeles, Miami, Minneapolis, New York, Seattle and Tulsa.

This study quantifies the benefit of this advanced cabinet construction to assist engineers and owners asses the benefit of demanding higher performance out of their roof top systems.

Aaon Fan Engineering White Paper

Aaon has pushed the envelope of packaged rooftop unit design by providing a wide selection of fan types in their systems, including forward curved housed centrifugals, backward inclined un-housed centrifugals (plenum fans) and axial fans. And, for their larger units, they have generally moved towards direct-driven plenum fans for supply air in arrays as best fits the application. This move has not been random, but is based upon solid engineering reasoning based on the inherent stability and efficiency of these types of fans. The newly published Aaon white paper, Value in the Air provides a thorough justification for this emphasis on direct-drive plenum fans. But in doing so, it also provides an excellent primer on general fan engineering topics including overloading vs. non-overloading fan curves, fan stability in single and dual fan applications, fan control for VAV systems, and the overall energy impact of fans on our HVAC systems. It is an excellent resource and one that engineers should review to enhance their understanding of fan systems.

Greening Small Rooftop Units: Digital Scrolls

Digital scroll compressors offer many benefits to compressorized HVAC equipment, as I have discussed in previous articles. Some of these advantages are obvious, and need little computational support--like the modulating capacity control.

Others, however, could use some numerical support to quantify the advantages they confer. The efficiency advantage these compressors confer is one of these sorts of advantages. While it is easy to conceptually understand how this technology can improve efficiency of compressorized units, how much of an advantage this is depends on a myriad of factors, including the capacity of the unit, the operating schedule of the system, the climactic conditions the system experiences and the application of the system.

The energy advantage will change depending on whether the system operates seven days a week or five, whether the system has multiple or single compressors, whether the system is in Atlanta or Seattle, whether the system is VAV or constant volume and whether the system has hot-gas bypass or not.

Aaon realized that the complexity of this calculation made it difficult to quantify the advantage of this advanced feature. To make it easier to see the energy advantage, Aaon added a simple energy calculation tool to the Engineering Toolkit that they provide with their ECAT32 selection software.



After using the simple drop-down windows to select geographic location (by city), Aaon model number, 5 or 7 day schedule, 12 or 24 hour operation, Constant or VAV fan control you then can select the variable and constant capacity units to compare against each other. Options available include assigning hot gas bypass or not to the lead and/or lag compressors and variable speed or cycling fan control to the condenser fans. Once these options have been selected for both units, you simply hit the "calculate" button and the energy performance summary is generated.



This gives you the energy improvement conferred by the digital scroll as a percentage, and also in a comparative EER the constant capacity compressor system would have to be rated at in order for the energy performance of the two systems to be equal.

Further graphs allow you to examine the bin hours at a given OADB for the geographic location and operating schedule you specified,



and the relative energy performances of both systems at a given OADB.



A little playing around with the system allows a user to quickly find where the energy benefits are greatest. In general, the digital scroll confers the most benefit to single-compressor systems, and systems that require HGBP for constant-capacity compressors, such as VAV systems. The system also highlights the point that HGBP is a very expensive way to to capacity control, since the compressor draws full amps whenever it is running. It is the unique modulating capacity of digital scrolls that really makes a difference in these sorts of applications.

Energy Code and Relief air in Rooftop Units

In a warm building served by rooftop units in a cold ambient condition, air wants to do an unfortunate thing. Warm, less dense air tends to rise through the colder, denser air. If unchecked, this causes an air migration up through the ductwork, to the RTU and out into the environment, while cold air is infiltrated into the building to make up the vacated volume. This causes a loss of heating energy detrimental to the energy performance of the building.


If the rooftop unit is operating, this stack effect is usually more than overcome by the pressures developed by the fans. But when the unit is off, often the only thing preventing warm air from working its way up through the ductwork is the action of the unit dampers. Most units installed under the Seattle Energy code have an automatic OA damper associated with the economizer that can be driven closed. But the relief air path is a different story. Often, this air path is controlled with a simple gravity damper configured to relieve air when the building is under positive pressure, but to prevent air from entering while under negative pressure.

The problem with this arrangement is that it does nothing to prevent a stack effect from occurring when the unit is not operating. The damper will act to let warm air out, which is exactly what is what we would hope to avoid. To address thisl, the 2006 Seattle Energy code has a section that requires a positive-closing damper on all air openings on building air systems:

1412.4.1 Dampers: Outside air intakes, exhaust outlets and relief outlets serving conditioned spaces shall be equipped with motorized dampers which close automatically when the system is off or upon power failure. Stair shaft and elevator shaft smoke relief openings shall be equipped with normally open (fails open upon loss of power) dampers. These dampers shall remain closed until activated by the fire alarm system or other approved smoke detection system.

EXCEPTIONS:

1. Systems serving areas which require continuous operation.
2. Combustion air intakes.
3. Gravity (non-motorized) dampers are acceptable in systems with a design outdoor air intake or exhaust capacity of 300 cfm or less buildings less than 3 stories in height.
4. Gravity (non-motorized dampers are acceptable in exhaust and relief outlets in the first story and levels below the first story of buildings three or more stories in height. Reserved
5. Type 1 grease hoods exhaust.

Dampers installed to comply with this section, including dampers integral to HVAC equipment, shall have a maximum leakage rate when tested in accordance with AMCA Standard 500 of:

1. Motorized dampers: 10 cfm/ft² of damper area at 1.0 in. w.g.
2. Non-motorized dampers: 20 cfm/ft² of damper area at 1.0 in. w.g., except that for non-motorized dampers smaller than 24 inches in either dimension: 40 cfm/ft² of damper area at 1.0 in. w.g.

Dampers used as a component of packaged HVAC equipment shall comply with the damper leakage requirements, unless it is the lowest leakage available as a factory option. Drawings shall indicate compliance with this section.

This has caused some disruption in the rootop packaged unit market, because the option of providing automatically closing motorized dampers on all air openings on these sorts of units is not one that is easily available from most manufacturers.

Johnson-Barrow has worked with Aaon to provide an engineered option on most configurations of the Aaon RM and RN rooftop packaged line to meet this requirement.

So if you have a project where a packaged RTU is the right solution, there is a code-compliant option likely available from Aaon.

Aaon Helps Project Achieve LEED Platinum

Signature Center, in Golden Colorado, recently received a LEED™ Platinum rating. This ambitious goal was realized with inspired design, and wise choice of mechanical systems.

The design featured:

• Underfloor Air Distribution
• Chilled Beams
• Evaporatively Cooled Chillers with Variable Speed Pumping
• Evaporatively Cooled Rooftop DX Air Handlers
• Non-CFC R-410a Refrigerant
  

Aaon made a natural choice because of their commitment to efficient packaged equipment. This project utilized a packaged LL Chiller Plant as well as a penthouse-type RL air handler. Both units utilized Aaon's unique, water-saving evaporative condensing system. Major project savings were achieved not just in energy, but in this other increasingly scare resource.

Not only did this project reach LEED™ Platinum--It also received the 2007 top award in the institutional building category from the Colorado Renewable Energy Society (CRES).

You can read more about this notable project in this Aaon case study.

Aaon Goes 'Outside the Box' in HPAC article

Aaon makes some news in the HPAC Magazine article 'Outside-the-Box' Thinking Produces Outside-the-Building Mechanical Room published in the December 2008 issue.


The article discusses the conversion of an old Apollo Mission facility into a modern printing facility. Great economies were realized by using Aaon LL chillers and air handlers. The LL chillers can be provided with pumps, boilers and accessories to create a full 'mechanical room in a box' that can be shipped to the site, pre-designed and pre-piped. As the article describes:

Recognizing that a different type of solution was called for, Ecogenia, a Montreal-based distributor of HVAC products and controls, specified two 335-ton LL Series chillers from AAON Inc. The LL Series integrates mechanical-room components into a single packaged outdoor unit that includes a heat exchanger, a pumping package, boilers, expansion tanks, controls, and an air-cooled or a high-efficiency evaporatively cooled condenser section. A cooling tower is not required.

Aaon LL chillers can be laid out using the Ecat32 software in just a few minutes, and the installation of an entire mechanical room is a easy as a crane pick and utility connection. This may be just the solution for your next project!

Read more about LL chillers here (pdf).

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

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.

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.

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.

Saving Water With Evaporative Condensing

Using evaporative cooling makes energy sense, not just in directly cooling the air, but also in cooling the heat rejection portion of a compressorized cooling system. This is because refrigerant compressors can be thought of, simplistically, as pumps. In a pump, two things govern energy use--the flow rate of fluid through the pump, and the head that pump needs to overcome to move the water. The same is essentially true for compressors.

In a compressor, the mass flow rate of the refrigerant essentially determines the cooling capacity it is providing. So for a given cooling load, we can't reduce the flow rate to increase our efficiency. All that is left is the compressor head. And that is something that we can effect.

In a refrigerant system, the condensing pressure of the refrigerant vapor is determined by the temperature of that refrigerant.

Let's look at this relationship for R-410a (from DuPont .pdf here):




We can see that the condensing pressure for 120º R-410a (which is around where an air-cooled condenser would operate) is about 450 psi. Compare that to a condensing pressure of about 280 psi or so at a temperature of 90º-which is an easily attainable condensing temperature in an evaporative condenser in the Pacific Northwest. When you consider that a reasonable suction temperature might be about 42º (or 150 psi) reducing your condensing temperature from 120º to 90º represents a head reduction of nearly 60%! In practice, moving to an evaporatively condensed piece of equipment from a comparable air-cooled piece of equipment can improve NPLV's by about 20% or so.

So, obviously, this makes sense from an energy conservation point of view. But what about water use? The obvious trade-off is that you are now using water where before, in the air-cooled case, you weren't. So how can we reduce the use of water while still benefitting from the reduced head pressure on the compressors?

To understand how how to improve water utilization in these systems, it is important to know what the state of the industry is for evaporative condensing. A typical system is illustrated below:


What you typically have is an induced draft evaporative unit that sprays water over a refrigerant coil. The water evaporates, and that evaporation cools the refrigerant in the coil. The remaining water falls into a basin where a pump then sprays the water back up over the coil again. All the while, a fan operates to draw air past the water and enhance the evaporative process. Importantly, every 1000 btu rejected by this system reflects about pound of water evaporated.

So what can we do to reduce water use? Well, the obvious thing is to reject less heat via evaporation. However, this might seem problematic because we want to maintain the low condensing temperatures that we can reach using evaporation. This is where it is important to understand a little about how refrigerant systems really work.

One of the main concerns with refrigerant compressors is they are designed to move gas--not liquid. A very effective way to break a compressor is to introduce liquid into it. So to be sure that no liquid enters the compressor, refrigeration systems are designe to operate with a few degrees of superheat. This takes the refrigerant safely away from the saturation line, and assures that the compressor will not see any liquid. However, this adds some extra heat into the system that then need to be rejected. Then, through the operation of the compressor, even more heat is added into the system, taking the system even further away from saturation.

However, the additive effect of the intentional pre-compressor superheat and the heat added by the compressor itself means that there is a significant amount of heat in the refrigerant that needs to be rejected before refrigerant condensing can even start. This de-superheat process is illustrated below:

(click for larger image)

If this heat can be removed without requiring evaporation, a significant amount of the water use can be eliminated. And this is exactly the approach that Aaon has taked with their evaporative condensing design. Their solution is illustrated below:



The main difference between the first system and this one is the addition of a finned desuperheater coil located above the spray system, in the cool, saturated air stream above the wetted portion of the condenser. This coil allow the system to reject the superheat without using any water--saving, on average, about 20% of the water use at peak load.

However, the benefits extend beyond there, since at about 70º ambient, this coil can reject about 50% of the total heat in the system, and it can reject 100% at about 30º ambient. So the true water savings range from about 20% to 100% depending on the operating profile and ambient conditions of the unit.

There are still other benefits: If chemical water treatment is being used on this system, the lower water use will translate to lower chemical use. And, since the tube surfaces in the wetted portions are at lower temperature in the Aaon system, there is a corresponding lower chance of creating scale--which is formed primarily from calcium carbonate which exhibits inverse solubility, depositing much more readily at higher temperatures. This lower fouling, in turn means the system will operated more efficiently for years to come, since less scale means better heat transfer at the tube surfaces which means lower head pressure on the compressor!

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.

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.