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Condensate harvesting from large dedicated outside air-handling units with heat recovery.

INTRODUCTION

With the adoption of building service systems requiring designs to minimize their environmental impact, innovative as well as obvious natural resource conservation measures are being explored.

One element of conservation considered in design is to minimize water usage in buildings. Modern designs often use green design practices suggested by organizations like the US Green Building Council and their LEED new construction guidance, where points toward receiving an overall rating are assigned to reducing annual water consumption by 20% and 30% from a baseline fixture flow rates determined by the Energy Policy Act of 1992 (USGBC, 2007). Hydronic systems makeup for equipment such as cooling towers is not generally included in building baseline water consumption rates for LEED, however should be considered in whole building water consumption estimates. Typical water conservation measures are utilizing low and zero flow fixtures, utilizing grey water for non-potable uses, and minimizing high water demand landscaping. Other, less common, water conservation measures utilizes the harvesting of water producing sources to offset the annual water consumption to achieve a net annual water reduction. These water producing sources include storm water recovery and air conditioning condensate harvesting.

Condensate from air conditioners, dehumidifiers, and refrigeration units can provide facilities with a steady supply of relatively pure water for many processes. Laboratories are excellent sites for this technology because they typically require dehumidification of large amounts of 100% outside air (DOE, 2005).

Another element considered in building design is indoor air quality and building envelope pressurization. Ventilation for Acceptable Indoor Air Quality, ASHRAE Standard 62.1-2004 recommends appropriate ventilation levels for various building and occupancy types. If buildings have excessive exhaust requirements due to fume hoods, user required air change rates, or other process exhaust, the minimum ventilation air requirements are potentially higher. Outdoor ventilation air commonly contains a higher moisture concentration and temperature than what is desired in the space. Conditioning this air by both reducing the temperature as well as reducing the moisture content is required.

Because some designs require increased levels of outside ventilation air for a variety of reasons, the potential to recover the energy being expelled by the exhaust system and transfer to the incoming ventilation air is high. Due to this potential, ASHRAE has incorporated policy regarding energy recovery, outlined in the Energy Standard for Buildings except Low-Rise Residential Buildings, ASHRAE Standard 90.1. To satisfy ASHRAE 90.1, fan systems that have a design air flow rate of 5000 cfm (2358 L / sec) or greater and have a minimum outside ventilation air flow that is equal to 70% or more of the supply air shall have an energy recovery system with at least 50% recovery effectiveness (ASHRAE, 2004). There are exceptions to this requirement, however for the purpose of this paper, no exceptions are taken.

To centralize the process of pre-conditioning ventilation air and the recovery of exhaust air energy, large energy recovery units or DOAHUs with energy recovery means are commonly used. These units are designed to provide preconditioned ventilation air either directly to the occupied space or ducted to an additional AHU. All building exhaust is also ducted to these units, which exchanges sensible and latent energy with the incoming ventilation air.

There are several types and configurations of DOAHU available for use. They have heat recovery means, either by an enthalpy wheel, glycol run around loop, air-to-air heat exchanger, or others. In addition, they often have cooling and heating coils, as well as humidifiers depending on the application. The type of DOAHU studied in this paper is one with an energy recovery device, pre-heat and cooling coil downstream of the recovery device, and a heating coil in the reheat position, shown in Figure 1.

[FIGURE 1 OMITTED]

These units pre-condition large quantities of moisture laden air down to a more moisture neutral state, where the humidity ratio (1) is close to the delivered supply air humidity ratio. This process produces large quantities of condensate, which is commonly discharged to the sanitary sewer systems. Some water treatment facilities operate at near capacity and do not allow or discourage condensate disposal into the sanitary sewer system (ICC, 2006).

The focus of this paper is to determine the feasibility of coupling both the water conservation and indoor air quality and building envelope pressurization elements of design by harvesting air conditioning condensate for non-potable water supplementation in large commercial, institutional, and medical buildings, where large volumes of outside air are required.

METHODOLOGY

The first step in the analysis was to gather temperature and humidity data for several geographical regions in Texas. The areas chosen for study were north Texas represented by Dallas, central / south Texas represented by San Antonio, and east Texas represented by Houston. The data used was annual daily average dry bulb temperature and relative humidity data obtained using an online weather archiving site. Since condensation production is the focus of this paper, outside air temperatures of 65[degrees]F (18.3[degrees]C) or below are excluded in the analysis. In highly humid climactic zones where deep de-humidification is used to prevent mold and mildew, additional water production may be realized at temperatures as low as 55 [degrees]F. All remaining data was used to calculate the average humidity ratio for each day using the following equations.

Eq. 1 determines the saturation pressure of the water vapor in the air as a function of the dry bulb temperature in the absolute scale of either Rankine or Kelvin.

Ln([p.sub.ws]) = [[C.sub.1]/[T.sub.ab]] + [C.sub.2] + [C.sub.3] * [T.sub.db] + [C.sub.4] * [T.sub.[db].sup.2] + [C.sub.5] * [T.sub.[db].sup.3] + [C.sub.6] * Ln([T.sub.db]) (1)

where

[T.sub.db] = dry bulb temperature, [degrees]R (K)

[p.sub.ws] = saturation pressure of water vapor, psia (Pa)

The saturation pressure is used to calculate humidity ratio using Eq. 2.

[omega] = 0.62198 * ([p.sub.ws]/[[p.sub.atm] - [p.sub.ws]]) * [[phi]/100] (2)

where

[omega] = humidity ratio, [lb.sub.v] /[lb.sub.a] ([g.sub.v] /[kg.sub.a])

[p.sub.atm] = atmospheric pressure, psia (kPa)

[phi] = relative humidity, %

Once the humidity ratios based on the daily average dry-bulb temperature and relative humidity level are calculated, the annual performance of the AHU is evaluated. The first calculation performed was to determine the leaving enthalpy wheel air conditions using Eq. 3 and Eq. 4. The recovery effectiveness was initially modeled at the ASHRAE 90.1 minimum effectiveness of 50%; however total enthalpy wheels can have a sensible and latent effectiveness as high as 75% (VanGeet and Reilly, 2006). The higher the latent recovery effectiveness is, the lower the moisture content in the air entering the cooling coil is, which ultimately reduces the amount of condensate produced. The return air temperature is assumed during cooling days to be 75[degrees]F (23.4[degrees]C) at 55% RH, since this operating point is very common for Texas and falls within the acceptable range of operating temperature and humidity level defined in ASHRAE 55-2004. The assumed return air condition does not account for any sensible gain during transport back to the DOAHU.
Table 1. Saturation Pressure Constants for I-P and SI Units

Constants        I-P Units               SI Units

[C.sub.1]  -1.044 x [10.sup.4]   -5.800 x [10.sup.3]
[C.sub.2]  -1.130 x [10.sup.1]   -5.516 x [10.sup.0]
[C.sub.3]  -2.702 x [10.sup.-2]  -4.864 x [10.sup.-2]
[C.sub.4]  1.289 x [10.sup.-5]   4.176 x [10.sup.-5]
[C.sub.5]  -2.478 x [10.sup.-9]  -1.445 x [10.sup.-8]
[C.sub.6]          6.546                   6.546


[T.sub.ECC] = [[epsilon].sub.s] * [T.sub.RA] + [T.sub.OA] * (1 - [[epsilon].sub.S]) (3)

[[omega].sub.ECC] = [[epsilon].sub.L] * [[omega].sub.RA] + [[omega].sub.OA] * (1 - [[epsilon].sub.L]) (4)

where

[T.sub.ECC] = temperature leaving enthalpy wheel, entering cooling coil, [degrees]F ([degrees]C)

[[omega].sub.ECC] = humidity ratio leaving enthalpy wheel, entering cooling coil, [lb.sub.v]/[lb.sub.a] ([g.sub.v]/[kg.sub.a])

[[epsilon].sub.s] = enthalpy wheel sensible recovery effectiveness

[[epsilon].sub.L] = enthalpy wheel latent recovery effectiveness

[T.sub.RA] = return air temperature from building, [degrees]F ([degrees]C)

[[omega].sub.RA] = return air humidity ratio, [lb.sub.v]/[lb.sub.a] ([g.sub.v]/[kg.sub.a])

[T.sub.OA] = outdoor air temperature, [degrees]F ([degrees]C)

[[omega].sub.OA] = outdoor air humidity ratio, [lb.sub.v]/[lb.sub.a] ([g.sub.v]/[kg.sub.a])

The next step is to determine the humidity ratio for the cooling coil leaving dry bulb temperature and relative humidity. It is assumed that the cooling coil leaving conditions is a 55[degrees]F (12.8[degrees]C) dry bulb temperature with a 90% relative humidity. This assumption is made since a cooling coil leave temperature of 55[degrees]F (12.8[degrees]C) and essentially saturated at 90% RH would provide adequate dehumidification and supply dry enough air to account for any latent gain encountered in the space. This assumption also implies that the entering air humidity ratio is not less than the assumed cooling coil leaving air humidity ratio, resulting in a dry-coil.

The final step is to calculate the condensate production potential from the dehumidification process. Condensate production potential represents the potential for condensate production per cubic feet of treated air flow. Condensate production potential is calculated using Eqs.6 and 7 to be applied to any dedicated outdoor air handling unit with similar configuration and performance characteristics. To determine the condensate production rate for a specific air flow quantity, simply multiply the air flow rate by the condensate production density using like units.

[DELTA][[omega].sub.cc] = [[omega].sub.ECC] - [[omega].sub.LCC] (5)

[[for all].sub.c] = [DELTA][[omega].sub.cc] * [[[rho].sub.a]/8.33](I - P) (6)

[[for all].sub.c] = [DELTA][[omega].sub.cc] * [[[rho].sub.a]/[1.0 x [10.sup.6]]](SI) (7)

where

[[DELTA][omega].sub.cc]= change in humidity ratio across the cooling coil, [lb.sub.v]/[lb.sub.a] ([g.sub.v]/[kg.sub.a])

[[omega].sub.LCC] = cooling coil leaving air humidity ratio, [lb.sub.v]/[lb.sub.a] ([g.sub.v]/[kg.sub.a])

[[rho].sub.a]= density of dry air, [lb.sub.a]/[ft.sup.3] (kg/[m.sup.3])

[[for all].sub.c] = condensate production potential, gal/[ft.sup.3] [.sub.a]([L.sub.v]/[L.sub.a])

Figures 2 and 3 show the condensate production potentials for San Antonio, TX, Houston, TX, and Dallas / Fort Worth, TX in 2007. Figure 2 shows the average daily potential for each location in 2007. Figure 3 show monthly averages for 2007, which produces a more easily interpreted plot. In Houston and San Antonio, TX the maximum average monthly condensate production potentials were approximately 6.0 x [10.sup.-5] gal / [ft.sup.3] [.sub.a](8.0 x [10.sup.-6] [L.sub.v] / [L.sub.a]) and 5.6 x [10.sup.-5] gal / [ft.sup.3] [.sub.a](7.5 x [10.sup.-6] [L.sub.v] / [L.sub.a]) respectively, occurring in mid August, 2007. The average condensate production potential for Dallas / Fort Worth was 5.0 x [10.sup.-5] gal / [ft.sup.3] [.sub.a](6.7 x [10.sup.-6] [L.sub.v] / [L.sub.a]) occurring in early to mid August, 2007.

[FIGURE 2 OMITTED]

[FIGURE 3 OMITTED]

CASE STUDY

Building Description

The case study building was a large medical research laboratory located in San Antonio, TX. Building parameters were obtained using the contract documents. The cooling is provided using a 1400 ton capacity chilled water system and the heating is provided using a combination of low pressure steam and hot water. The hot water is produced using a steam to water heat exchanger. Both hydronic systems employ the use of variable frequency drives on all distribution pumps. The air delivery is provided using six energy recovery units (ERU) and two traditional variable volume air handling units, located in the building penthouse on the fifth floor. The two variable volume air handling units are ignored in this paper since they are not dedicated outdoor air handling units; however the harvesting of condensate from these types of units is a potential future paper topic. Three of the six ERUs are enthalpy wheel types and three are glycol run around loop style heat recovery. One of each type of ERU is a stand-by unit. For the simplicity of this paper, the ERUs are modeled as one unit, all with the same recovery effectiveness. The maximum total outside air processed through the ERUs is 110,000 cfm (51700 L / sec). The total recovery effectiveness was modeled at 59%, with a sensible and latent effectiveness on 73% and 33% respectively calculated using procedures defined in ARI Standard 1060-2005. Based on these system characteristics and using the San Antonio, TX condensate production density, adjusted for the new sensible and latent effectiveness, the monthly condensate productions are shown in Figure 4, with a total annual estimated production of 1,887,031 gallons (7.15 x [10.sup.6] L).

[FIGURE 4 OMITTED]

What to Do with the Condensation Collected?

Two possible uses for the condensate water collected are explored. The first is water closet and urinal supplementation and the second is cooling tower makeup water usage.

Water Closet and Urinal Usage. The case study building has a designed full time occupancy level of 250 people. Based on the occupancy level, the annual domestic water consumption for the water closets and urinals was calculated and was compared to the estimated annual condensate production quantity. It was assumed that 50% are male and 50% are female. To determine the estimated daily water consumed by water closets and urinals, Eq. 8 was used:

[Q.sub.DC] = [n.sub.m] * [[([f.sub.u] * [q.sub.u] + [f.sub.wc] * [q.sub.wc])].sub.m] + [n.sub.f] * [[([f.sub.wc] * [q.sub.wc])].sub.f] (8)

[Q.sub.DC] = daily water consumption, gal (L)

[n.sub.m] = male occupancy

[n.sub.f] = female occupancy

[f.sub.u] = urinal frequency of use

[f.sub.wc] = water closet frequency of use

[q.sub.wc] = water closet volume per use, gal/flush (L/flush)

[q.sub.u] = urinal volume per use, gal/flush (L/flush)

The fixture volume per flush was in accordance with the Energy Policy Act of 1992. The frequency of use for male and female occupants was estimated according to Table 3.
Table 3. Daily Frequency of Use Estimation

              Male  Female

Urinal          2       0
Water closet    1       3

Source: Vickers (2001); USGBC (2007)


The total daily estimated water use for water closets and urinals was 1050 gallons. To calculate the estimated annual consumption, the daily usage was multiplied by the number of occupied days per year. The case study building operates on a five day work week, which equates to 260 days per year. The total estimated annual consumption for water closets and urinals is 273,000 gallons (1.03 x [10.sup.6] L).

Based on an estimated annual condensate production of 1,887,031 gallons (7.15 x [10.sup.6] L) and an estimated annual water closet and urinal consumption of 273,000 gallons (1.03 x [10.sup.6] L) it was concluded that the potential to completely supplement the closet and urinal consumption existed, with an excess of 1,614,031 gallons (6.12 x [10.sup.6] L) potentially used for other non-potable systems such as landscaping irrigation systems.

A challenge for this application is accounting for periods of time where condensate is not being produced with sufficient volume to supplement water closet and urinal usage. Condensate storage systems with capacity to provide uninterrupted service during these times should be considered during design. Based on the cost of municipal water, engineering fees, materials and labor, and maintenance, it would be difficult to implement a system which is both functional as well as economically practical. Other factors, such as LEED rating, or other customer or local design requirements might lend toward incorporating this type of system.
Table 2. Baseline Water Flow Requirements

Fixture                            Energy Policy Act of 1992 Flow
                                             Requirement

Water closets, gpf (lpf)                     1.6 (6.1)

Urinals, gpf (lpf)                           1.0 (3.8)

Showerheads, gpm * (lps)                     2.5 (0.16)

Faucets, gpm * (lps)                         2.5 (0.16)

Replacement aerators, gpm * (lps)            2.5 (0.16)

Metering faucets, gal/cy (l/cy)              0.25 (0.95)

* At flowing water pressure of 80 psig (552 kPa)


Cooling Tower Makeup Water Usage. To estimate the annual makeup water consumption used for cooling towers, the condenser system load was first determined. Since the case study building contained other chilled water using devices in addition to the DOAHUs, simply using the coil loads for this equipment as the cooling tower load was incorrect. A whole building cooling tower estimated load was necessary to determine the makeup water requirement. Based on 2007 weather data, the sensible and latent coil loads for the DOAHUs were first determined using the operating parameters described previously. These daily coil loads were converted to percent load based on the maximum design coil load for the DOAHUs. The design maximum chiller capacity is 1400 tons (4928 kW). The percent load profile was used to determine the daily chiller load by multiplying the maximum chiller capacity by the percent load building profile determined previously. Once the daily chiller load was determined, the cooling tower makeup water demand was calculated using the following equations.

[q.sub.evap] = [m.sub.a] * [h.sub.fg@85 [degrees]F] * ([DELTA][omega]) (9)

where

[q.sub.evap] = heat transfer due to evaporation, Btu/h

[h.sub.fg@85[degrees]F] = 1045 Btu/[lb.sub.v] (65.7 kJ/[kg.sub.v]) enthalpy of evaporation at 85[degrees]F (29[degrees]C)

[m.sub.a] = mass flow rate of dry air, [lb.sub.a]/h ([kg.sub.a]/h)

[DELTA][omega] = condenser water evaporation rate, [lb.sub.v]/[lb.sub.a] ([kg.sub.v]/[kg.sub.a])

Equation 9 represents heat transfer due to evaporation. It is assumed that all heat transfer dissipated by the cooling tower is 90% due to evaporation; however the actual ratio of sensible and latent components varies with outdoor air temperature and humidity (Marley, 1983).

[q.sub.tower] = 1.2 * ([q.sub.bldg]) (Accounting for compressor heat) (10)

[q.sub.evap] = ([q.sub.tower]) * 90% (11)

To account for heat generated by the chiller, the total heat rejection load on the cooling tower is 1.2 times the building chilled water load. Using equation 9, the following relationship can be determined, where the evaporation rate is the total cooling tower heat rejection load divided by the enthalpy of evaporation, shown in equation 13. For consistency, system was analyzed only during days where the outside air temperature was above 65 [degrees]F (18.3 [degrees]C).

[[for all].sub.evap] = [m.sub.a] * ([DELTA][omega]) (12)

[[for all].sub.evap] = [[q.sub.evap]/[h.sub.fg@85[degrees] F]] (13)

where

[[for all].sub.evap] = cooling tower makeup due to evaporation, [lb.sub.v]/h ([kg.sub.v]/h)

As the water of evaporation exits the cooling tower in a pure vapor state, the dissolved solids left behind increases in concentration in the re-circulating water. Given no control, the total dissolved solids (TDS) level in the re-circulating water can potentially damage the tower, condenser, and related equipment. To prevent this, a portion of the re-circulated water is continuously drained from the system and replenished with clean water. This process is called "blowdown". To calculate the quantity of blowdown, the following equations were used:

B = [[[[for all].sub.evap] - ((C - 1) * D)]/(C - 1)] (14)

where

C = cycles of concentration (estimated at 4, local municipal requirement)

D = drift (estimated at 0.0002 times condenser water flow) (Marley, 1983)

Figure 5 shows the daily estimated cooling tower makeup water demand in 2007, as well as the potential condensate volume. The annual makeup water required for the cooling towers due to evaporation and blowdown was estimated at 11,584,374 gallons (4.39 x [10.sup.7] L). When compared to the annual condensate produced by the DOAHU, supplementing the cooling tower makeup water with condensate has the potential of an estimated 16% water savings.

CONCLUSION

The analysis performed showed that the condensate production from a large dedicated outdoor air handling unit or series of units can completely supplement water closet and urinal water demand annually with 1,614,031 gallons (6.12 x [10.sup.6] L) of annual excess, which could potentially supplement the landscape irrigation system, or reduce the cooling tower makeup demand by 16% annually. It is shown that water conservation as well as indoor air quality and building pressurization design elements can be coupled to produce a better overall building service design that would routinely not be considered. With 1,887,031 gallons (7.15 x [10.sup.6] L) of relatively clean water is to be sent to the sanitary sewer system, condensate harvesting is an environmentally conscious option and should be considered in addition to low and zero flow plumbing fixtures when making water conservation design decisions.

This paper focuses on dedicated outdoor air handling unit condensate harvesting, however units with return air and outside air are widely used in commercial buildings. To apply the proposed methodology using a whole building approach to condensate harvesting, systems of these types should be considered. In addition, retrofits of existing building with large outside air requirements, such as hospitals, laboratories, and dormitories should be considered.

REFERENCES

ARI, 2005, "Standard for Performing Rating of Air-to-Air Heat Exchangers for Energy Recovery Ventilation Equipment", Arlington: Air-Conditioning & Refrigeration Institute.

ASHRAE, 2001, "Standard 62.1-Ventilation for Acceptable Indoor Air Quality", Atlanta: American Society of Heating, Refrigeration, and Air-Conditioning Engineers.

ASHRAE, 2007, "Standard 90.1-Energy Standard for Buildings except Low-Rise Residential Buildings", Atlanta: American Society of Heating, Refrigeration, and Air-Conditioning Engineers, pg 41.

AHSRAE, 2004, "Standard 55-Thermal Environmental Conditions for Human Occupancy", Atlanta: American Society of Heating, Refrigeration, and Air-Conditioning Engineers, pg 5.

ASHRAE, 1993, "ASHRAE Handbook: Fundamentals", Atlanta: American Society of Heating, Refrigeration, and Air-Conditioning Engineers, pg 6.11.

DOE, 2005, "Laboratories for the 21st Century: Best Practices, Water Efficiency Guide for Laboratories", U.S. Department of Energy and U.S. Environmental Protection Agency, DOE/GO-102005-2008.

ICC, 2003 "International Plumbing Code", Chicago, International Code Council.

ICC, 2006 "International Mechanical Code Commentary" pg 3-27, Chicago, International Code Council.

Marley, 1983, "Cooling Tower Fundamentals", Marley Cooling Tower Co.

USGBC, 2007, "LEED-New Construction & Major Renovation Reference Guide", Washington D.C., Third Edition, Version 2.2.

Vickers, Amy, 2002, "Water Use and Conservation." Amherst, MA Waterplow Press.

VanGeet, Otto, Sue Reilly, 2006, "Ventilation Heat Recovery for Laboratories", ASHRAE Journal March 2006 pgs 44-53.

(1.) Humidity ratio is defined as for given moisture sample it is the ratio of the mass of water vapor to the mass of dry air in the sample (ASHRAE, 1993).

Frank L. Painter, PE

Associate Member ASHRAE

Frank L. Painter is a mechanical engineer with the U.S. Army Corps of Engineers, Fort Worth District, San Antonio Construction Management Office, San Antonio, TX.
COPYRIGHT 2009 American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Inc.
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Author:Painter, Frank L.
Publication:ASHRAE Transactions
Article Type:Report
Geographic Code:1USA
Date:Jul 1, 2009
Words:3891
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