Labs 21 Highlights Benefits of Low Pressure Lab Ventilation Design

Labs21 is the leading organization focusing on low-energy, sustainable laboratory design. Cosponsored by the EPA and the DOE, Labs21 has their sights set on these highly energy-intensive buildings and strives to find technically sound solutions to minimize the energy impact of these facilities without sacrificing their mission.

And Labs21 has set their sights on pressure drop.



Their document, Low Pressure Drop HVAC Design for Laboratories, identifies huge opportunities for savings by paying attention to this design criteria. They feel that 30-65% of the laboratory HVAC energy use could be saved by designing low pressure drop into these systems.

Tek-Air's inherently low-pressure drop air-valves can be an integral part of a low-pressure lab HVAC design, and one of the few identified that is essentially cost-neutral.

'Green' is where the lab market is going. Labs21 can help point your way there.

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.

'Greening' Lab Design

Laboratory fume hoods are energy intensive. In order to provide safety for their operator, they need to ensure a constant face velocity of air at the sash--air that must first be conditioned to keep the space temperature acceptable for comfort, moved via mechanical means to the lab, and then exhausted out of the building.

A common comparison used to highlight the energy costs of these systems is to compare the energy impact of a single fume hood with that of a typical US household. On average a single lab fume hood uses as much energy as three typical US houses. And when you consider that a given facility may have many lab hoods in a single laboratory space, you can see how these energy impacts quickly add up.


In order to minimize the wasted energy associated with these laboratories, high-precision VAV lab controls have been developed to ensure operator safety, and to only provide the minimum amount of air necessary--And great savings have been realized by this sort of measure. But the energy efficiency of these systems can be improved even more.

Once the airflow has been taken down to a minimum, the energy associated with conditioning that air has been greatly reduced. But the energy associated with moving that air still can be reduced further. ASHRAE 90.1 states:

ASHRAE Standard 90.1 - 6.5.3.2.3:
“For systems with direct digital control of individual zone boxes reporting to the central control panel, static pressure setpoint shall be reset based on the zone requiring the most pressure; i.e., the setpoint is reset lower until one zone damper is nearly wide open.”

This calls for static pressure reset for VAV systems to minimize fan energy--ensuring that only the minimum amount of static pressure is provided to move the air. And this strategy is perfectly applicable to laboratory VAV systems as well as commercial air conditioning--as long as the system components are selected appropriately.

Tek-Air has published a white paper entitled Demand Based Static Pressure Reset Control for Laboratories That explores the energy benefits of this type of control scheme.


This paper analyzes system component selection, including control valves and sensors and illustrates the impact of these decisions on the overall energy use of the VAV system. In an analysis of a 50,000 cfm exhaust system, the reduced static from a pressure reset strategy can result in nearly $9,000 per year savings in fan energy (based on 0.75" savings, and $0.06/kwh electric costs).

These sorts of static pressure savings are easily attainable with a wise selection of air valve components. The commonly specified venturi-type valve has a minimum operating pressure that prevents these savings from being realized, and this added pressure drop often creates objectionable noise, which requires even more pressure drop for the system in the form of sound attenuators. This pressure reset strategy requires valves that can operate accurately and safely at low pressures.

The Tek-Air PRD valve provides unmatched pressure performance, and a quick examination of a cross section of the valve shows why:


Each blade of the damper is a smooth airfoil, greatly reducing turbulence and keeping the pressure and acoustic profile of the valve to a minimum.

If pneumatic air is not available, Tek-Air's new Accuvalve provides very similar performance with the convenience of electronic actuation. (And it won an innovation award at the 2008 AHR expo!)



A peek at the cross section of this valve shows how it attains these low pressure drops:


The airfoil shape of the valve assembly assures minimal pressure drop and sound generation for great efficiency in the fan system.

Energy savings cannot come at the cost of safety, and it is imperative that systems utilizing this method of pressure reset have sensors that can operate accurately and effectively in a wide range of pressure regimes. Tek-Air uses vortex shedding flow sensor technology to ensure the most accurate and linear control on the market.

Energy conservation is only going to become a bigger and bigger issue for designers of all building systems, and fume hood systems are a large opportunity for savings. It is important that designers and owners consider all the impacts of their design decisions and their system selections.

(Don't forget about checking the fan for stability: See this article for a review on this issue.)

Rethinking Air-Cooled Chillers

Air Cooled Advantages

Air cooled chillers offer many advantages to owners and designers. The first, and perhaps most compelling for many jobs is lower installed cost. Lower installed costs (compared to water cooled chillers) are driven by the following advantages:

• No Cooling Tower, Tower pumps, Tower and Pump Starters
• No equipment room required for the chillers
• Mounted starters

They also are easier to maintain, since the systems are significantly simpler than water-cooled systems:

• No on site Systems Engineer required
• No water treatment or make up water required
• No leaks on the roof
• No cooling tower, condenser pumps, associated starters

Generally, however, these advantages have come with significant trade-offs: Efficiency and Sound performance.

However, the introduction of Variable Speed oil-free air-cooled chillers by Smardt changes the balance.


First off, the Smardt Chiller is efficient. With IPLV's as low as 0.65 kw/ton, these chillers rival water-cooled system when the parasitic loads of the condenser pumps and cooling tower are considered. These chillers gain their efficiencies both from the inherent efficiency of the Turbocor compressor and the elimination of oil return issues that prevent other air-cooled chillers from capitalizing on the reduced head pressures available at low ambients.

This means these chillers use about 60-65% energy of other air-cooled chillers for the same load, and can nearly eliminate the energy benefit typically provided by moving to water-cooled systems. When you consider the cost of water (nearly $15/1000 gallons in Seattle, including sewer charges) this means the yearly cost of operation of these units is unrivaled. And energy conservation rebates are extremely attractive for these chillers.

The other major traditional trade off with using air-cooled equipment is sound. Screw chillers especially are known for their unfavorable sound characteristics. In most municipalities, sound ordinances are driven by occupancy and time of day. The most stringent criteria must be met during evening hours, typically when the units are not at their peak load. However, with constant-speed systems, the compressor is either on or off. This means it is either putting out its full sound or none at all. At full speed, such compressors can often exceed the evening sound criteria--even if they are on only momentarily. And the staging between on and off can be objectionable in its own right, regardless of sound level.

The Smardt chiller minimizes the problems with compressor sound in two ways. First, the variable speed drive allows the compressor to ramp slowly up and down to match the required output, eliminating the objectionable switching between compressors that constant-speed chillers exhibit. And secondly, they are just extremely quiet to begin with. Since no moving mechanical part is in contact with the chiller casing, very little mechanical noise is transmitted. Ninety-ton Turbocor compressors have been tested at 72 dBa at one meter, compared to screw compressors that can be as high as 80 dBa or higher in the same test. Five of these compressors operating together yield a sound level of 75 dBa at 10’.

More Benefits

But efficiency and sound are not the only benefits from using the Turbocor technology on air-cooled chillers. Other, less obvious ones exist.

Turbocor compressors have only one moving part, yielding un-matched reliability.

Reliability is enhanced by the elimination of oil in the refrigerant system. And the frictionless bearing requires almost no maintenance.

Since Turbocor compressors are variable speed driven, they provide an inherent soft-start on the compressor. Instead of kicking the motor up to full speed when power is applied to the system, the VSD slowly ramps the compressor up to the required speed for the load sensed by the system. This reduces stress on the already greatly simplified system to reduce wear and tear on the components.

But this soft start has another, very important advantage over standard air-cooled chiler systems--the use of the VSD eliminates inrush amperage. When an electrical motor is at rest, there is very little inductive resistance to current flow through the windings. As the motor starts to turn, this inductive resistance increases with the increase in RPM. What this means is when power is applied across the line (or even with a reduced voltage starter) to a stopped motor, there is a spike of electrical current far greater in amplitude than the design amp draw of the motor:

(example graph of inrush on a well pump motor)

This temporary increased amp draw heats the motor beyond where it is designed to operate for extended periods. This forces the chiller designer to provided anti-recycle timers to prevent rapid re-starts that could fatally overheat the motor. In practice, this usually means constant speed compressors cannot be started more often than every half-hour or so.

Additionally, this increased amp draw has effects that need to be addressed electrically. This becomes even more significant if the chillers are being served by emergency power. The emergency generators that serve the chiller must be sized to handle the inrush amperage. This can be a very costly addition, especially since the added amperage is only required for the first 30 second of operation or so.

Generators = $$$

Turbocor compressors on the Smardt air-cooled chillers eliminate inrush and provides a soft-start. This both heightens reliability and reduces electrical costs. For jobs where reliability is a primary concern, like data centers, this technology makes a lot of sense. First, it eliminates the need for increased generator sizing, it is an inherently more reliable compressor, and it frees the cooling system from reliance on a water utility service that could be disrupted.

The Dark Side of Chiller oil

Most compressorized systems operate with oil mixed in with the refrigerant. This oil is required for lubrication of the shaft bearing and, in positive displacement compressors like scrolls or screws, it actually provides the seal that is necessary to effect the compression of the refrigerant. The oil is miscible with the refrigerant, and travels with it throughout the system to provide lubrication and compression sealing. That means the oil is everywhere the refrigerant is—Which can lead to several operation problems that need to be addressed in the design and operation of the system.

The first issue is oil transport. The oil travels with the refrigerant whenever the velocity of the refrigerant is high enough to carry the oil with it. This means that the system must be designed to ensure adequate velocities are met at all times. The biggest obstacle to refrigerant flow occurs, however, at the expansion device between the condenser and the evaporator. This is by design a big restriction on refrigerant flow—the greater the restriction, the greater the pressure drop across it and therefore the larger the temperature difference between condenser and evaporator. Its function is similar to that of a flow restrictor in a shower head or an orifice plate in a commercial piping system. Therefore, in order to ensure an adequate flow rate for oil transport between the condenser and evaporator, the pressure between these components must be kept above a minimum. If the pressure difference falls below this minimum, oil starts stacking up in the condenser and the machine will trip out on low oil pressure.

While there are methods that can be used to minimize this problem, such as oil pumps and eductors that siphon the accumulated oil out of the condenser, these are usually only provided on larger tonnage centrifugal machines, and even then there is a practical limit to the operating environment the compressor is designed for. In practice, head pressure control is used where low condensing temperatures are expected, such as fan cycling on air-cooled condensers or bypass lines on water cooled compressors. And even then, oil transport problems are a major cause of chiller downtime, especially in cool climates.

These head-pressure control mechanisms ensure a minimum condensing pressure to assure oil transport—which is good, because this ensures trouble-free operation. But it does this at the cost of energy efficiency.

See, a compressor is very similar to a pump. The total mass flow of refrigerant can be thought of as analogous to the flow rate of the pump and the pressure difference between the evaporator and condenser is analogous to the static head of the pump. In order to reduce energy use on a pump, you can either reduce the flow rate through the pump, or you can reduce the static head. In a chiller system you have the same options, but remember that the refrigerant mass flow rate is what determines the cooling capacity of the chiller. If you want to reduce energy use, but still provide the same cooling capacity, your only option is to reduce condenser pressure. One of the great advantages of using the ambient air as a heat sink is that for over 99% of the year, you have access to air temperatures (wet bulb or dry bulb, depending on the heat rejection of the system) below your design condition. This is why part load ratings on chillers are so much better than peak load ratings.


Head pressure controls, however, artificially limit how low this head pressure can go. Just when you are really getting efficient, the system kicks in a mechanism to keep your system from getting any more efficient! Without the need to maintain oil flow, a compressor can take full advantage of the available head pressure relief and operate extremely efficiently.

The second major issue with oil in your system is that it actually inhibits heat transfer, reducing the overall efficiency of the refrigeration system. In fact, a nominal 3.5% charge of oil in a chiller system equates to about an 8% loss of efficiency from this insulating effect (from ASHRAE 601).


And this only gets worse with time. An ARI study found that this insulating effect increases for the first 5-6 years of operation reducing chiller efficiency by about 20% due to oil fouling of the heat transfer surfaces in the chiller.

These numbers assume a constant oil charge—which is a big assumption. In practice, chiller oil charges often exceed the recommended oil charge by very significant amounts—since a common method of correcting a ‘low oil pressure’ alarm is to add more oil—despite the fact that the usual cause of this alarm is the stacking of refrigerant in the condenser, not any loss of oil in the machine. The same ASHRAE study above sampled many machines in the field and found that the average charge of oil was nearly 13% which equates to a total efficiency loss of 21%!

The last major issue with oil in the compressor is the chance of motor burnout. If a refrigerant system experiences a small leak in a low pressure location, moisture and air can enter the refrigerant system. Enough moist air can react with the refrigerant and oil to form hydrochloric and hydrofluoric acids which then travel through the system to the motor windings and eat away the insulation, eventually causing massive arcing and a catastrophic failure of the compressor motor. This acidic residue will be deposited throughout they system, requiring extensive flushing of the chiller shells before the system can safely be brought back on line. This is not only very expensive, but amounts to a very long downtime in what may be a critical building component.


Developed in Australia in the mid-90’s, the Turbocor™ compressor was born out of a desire to avoid the energy penalty and maintenance headaches associated with oil in the refrigerant circuit of traditional compressorized cooling systems.


These centrifugal compressors utilize a unique magnetic bearing to completely avoid the need for oil in the refrigerant system. These systems require very little maintenance and provide excellent efficiency both initially and many years down the road.

Smardt manufacturers water-cooled and air-cooled chillers exclusively utilizing these innovative compressors. These chillers avoid all of the drawbacks of oil, and eliminate much of the cost of ownership that is commonly associated with chiller systems. Since the bearings do not wear, traditional scheduled stop-major chiller teardowns are unnecessary, and all of the downtime and cost associated with oil (around 50% of the maintenance cost on these systems) are avoided. Smardt chillers provide an owner with excellent efficiency, reliability and economy.

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).

Saving Water in Evaporatively Cooled Systems

Water is a limited resource, just like energy. Engineers are very aware of the need to save energy in their designs, and one of the best ways to do this is to take advantage of evaporative heat rejection for their cooling systems. The traditional cooling tower is an extremely effective way to reduce energy use at the compressors in a traditional cooling system. But introducing a cooling tower introduces a need for water to the system. It would be advantageous if this water use could be kept to an absolute minimum.

Especially since, in Seattle, water is expensive. As of this posting, the water utility rate per thousand gallons is $4.48 (summer) and the Sewer costs tack on an additional $9.96. When you consider that a cooling tower consumes a minimum of 1.8gph/ton (evaporation required to reject that heat), you can see that over a 1900 hour cooling season, these costs can really add up for a reasonably-sized cooling tower.



Earlier, I posted an article that highlighted ways to reduce water use in traditional cooling tower systems. For the most part, these recommendations address keeping the actual water use as close to the theoretical 1.8gph/ton evaporation figure as possible. Reducing the water use any further requires reducing the load on the tower, since evaporation is the only way a cooling tower can reject heat.

There are two ways to reduce load on a cooling tower--Reducing the total building load, or rejecting heat through some other method other than the cooling tower. Assuming the first option has already been exhausted through good engineering practices, the only other option is the second.

This is the approach taken by Aaon in their evaporative condenser systems. They essentially use a dry finned coil as the first stage of cooling before the refrigerant is cooled by evaporative methods. This essentially allows the system to reject as much heat as possible through a non-evaporative method before water is used. Every btuh that is rejected in this manner means less water used in the system.


This idea could be borrowed and applied to an open cooling tower by the use of a dry-cooler as a pre-cooler before a cooling tower. This way, the system rejects as much heat as possible in a dry fashion, and only uses water for what the dry-cooler can't do. This system gets to take advantage of the strengths of both methods of heat rejection--the water conserving function of a dry-cooler, and the lower water temperatures and more efficient heat rejection provided by a cooling tower.

Evapco has capitalized on this approach by creating a new, water-saving fluid cooler called the WDW:


This unit is a hybrid between a dry-cooler and an evaporative fluid cooler. It is provided with a control panel that controls both wet and dry sides of the unit, varying fan speeds with a VFD and determining when to run the evaporative pumps to optimize both water efficiency and fan energy.

Cutaway of an Evapco WDW unit

In practice, the evaporative system is only used for a small portion of the year, only when the design condenser water temperatures cannot be met by the dry-cooler side alone. What you see is a major reduction in water use compared to the same system served by a fully evaporative system:


Other advantages of this approach besides reduced water use are reduced chance of tower plume (since there are far fewer hours in which water is being evaporated, and when this does occur, it occurs in warmer temperatures) and the ability to provide some cooling even if city water is lost due to a service disruption.

But since a dry-cooler uses more fan energy per ton of cooling than a cooling tower, this system will inevitably use more energy to save water. Does this approach pay off?

An example from a real project might help demonstrate the economies involved. Below are the utility cost calculations from a project utilizing a 240 ton WDW installed in Seattle on a heat pump system with a portion of the load serving a 24/7 cooling application:



Note that even with the reduced water cost (to approximate the effective cost of using a deduct meter to avoid being charged wastewater charges for evaporated water) the hybrid system saves about 18% of the annual operational utility costs compared to a fully evaporative system. This affords a relatively quick payback for the added equipment costs associated with the hybrid system.

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.