Cost and Energy Consumption Comparison
To compare these units and decide which is the most advantageous to use, it is necessary to consider the cost of the equipment and its impact on the associated chiller plant cost, and the annual energy consumption cost. For this purpose, a 10,000 cfm (4720 L/s) model was chosen for comparison, and it was assumed that the fan brings this amount of outdoor air for 24 hours/day, everyday of the year. Different selections and quotations were obtained from various major manufacturers.

Based on these selections, and by computing the air conditions at different sections of each unit, the annual cooling energy and electrical consumption were calculated using the Abu Dhabi bin hour data that were computed using weather data covering a span of 10 years, which were provided by6 the Abu Dhabi Ministry of Communications – Meteorological Department for the period 1985 – 1994.

Tables 1a and 1b showed an example of a calculation for the cooling energy demand of the double-wheel energy recovery unit and total energy wheel with horseshoe arrangement. Design conditions are shown in the first row according to ASHRAE climatic design conditions.

Air conditions are obtained at each section of the unit, and the total annual cooling energy consumption was determined in that manner. It is assumed that the unit is cycled off when the ambient temperature drops below that required for the indoor comfort conditions in accordance with ANSI/ASHRAE Standard 55-1992, Thermal Environmental Conditions for Human Occupancy, at 76°F (24.4°C) and 50% RH.

Electrical energy was calculated taking into account the power consumption of the supply fan, the exhaust fan, the runaround pump, and the total energy and sensible wheel’s energy.  Total energy was calculated based on the operating hours for each unit.

Table 2 shows a summary for the electrical energy consumption calculation.  Two cases were considered for the dual energy recovery units with runaround coils and heat pipe options:

1. Face velocity through the coils 315 fpm (1.6 m/s); and
2. Face velocity through the coils 510 fpm (2.6 m/s).

Because of the large size of the energy wheel, enough cross-sectional area is available for the coils, which helps reduce the face velocity and consequently the energy consumption of the fan, but for the cost of a bigger a coil. Thus, these two options were considered to study feasibility of increasing the coil’s size for reducing energy consumed.
Table 3 shows a summary of results for the cooling coil design load, annual cooling energy and electrical energy required for each unit.

Table 4 shows a comparison between the units considered. For comparison, the cost of the air-cooled chiller plant being considered was estimated at $1,200/ton ($341/kW), including civil, mechanical, electrical, and utility connections costs. The equipment costs were obtained from manufacturers’ quotations. Electrical consumption was assumed to be 1.7 kW/ton (6 kW/kW) for the total chilled water plant (including chillers, pumps and auxiliaries). The electric power cost was determined based on flat rate of $0.0543/kWh. The total capital and operating costs were calculated, for a 20-year period. The net present value was determined using an 11.75% discount rate.

As shown on Table 4, the fresh air-handling unit resulted in the highest life-cycle cost. This value was reduced by 58% when the total energy wheel with the horseshoe heat pipe arrangement was used, which resulted in the lowest life-cycle cost. However, the drawback when using this unit is the difficulty of controlling the temperature and moisture content of the supply air, which will vary depending on the ambient conditions.

The higher coil face velocity resulted in an increase of the life-cycle cost by about 8%. This lead into concluding that exceeding a coil face velocity of 400 fpm (2 m/s) is not recommended.

We noticed that Unit A (conventional fresh air-handling unit) and Unit E (fresh air-handling unit with total-energy wheel) both dehumidify the outdoor air to the same dew point as the other systems, but they deliver it cold-56.3°F (13.5°C) dry bulb – rather than reheating it to a neutral 70°F (21.1°C). In these two systems, this cold air is able to offset a portion of the space sensible cooling loads.

However, many designers don’t take into consideration the cooling effect for sizing the secondary (local) HVAC systems nor it is used to reduce the chilling plant size. In this case, it has no impact on the capital cost. In off peak hours, when the room temperature condition is satisfied and the secondary system is turned off, this cool, conditioned outdoor air may overcool the space. At such time, the dehumidified outdoor air should be reheated or the local HVAC equipment needs to add heat to avoid overcooling the space. For this analysis, we decided to list the cooling energy in the tables but to ignore its impact. The reader may choose otherwise.

Recommendations
Based on the results and assumptions of this specific analysis:

• When no need exists for a constant supply temperature and RH, the total energy wheel with a horseshoe heat pipe arrangement with the lowest energy consumption costs and capital costs can be used (supply temperature was found to vary between 66°F and 70°F [18.9°C to 21.1°C] and RH between 59% and 67%). The recommended spacing between the heat pipe coils is 4.6 ft (1.4 m) for easier cooling coil’s maintenance, although this leads to having a slightly longer unit.
  
• Double energy recovery systems resulted in better humidity control with a constant fresh air supply temperature and RH year-round regardless of fresh ambient conditions (70 °F [21°C] and 61% RH). Although costs are higher, they are recommended for use when constant supply temperature and humidity are necessary. Of course, the results of the analysis may differ with climate operating hours, utility costs, and installed costs.
  

The following are summarized design parameters recommended based on the preceding analysis and our experience installing and maintaining these systems.

• Total energy or sensible wheel to have a maximum air face velocity of 600 fpm (3 m/s). This limits the pressure drop, blower power and cross leakage to modest levels.
• Heat pipe and runaround coil to have a maximum air face velocity of 400 fpm (2 m/s):

1. As the area defined by the wheel allows larger rectangular coil area;
2. Heat pipe runaround coils typically have lower effectiveness than sensible wheel at the same face velocity, for runaround coil or heat pipe (typically eight rows or less): and
3. Life-cycle analysis justifies use of lower face velocity.
   

• Wheels and heat pipes should be tested and rated according to the following:

1. ANSI/ARI Standard 1060-2001, Performance Rating of Air-to-Air Heat Exchangers for Energy Recovery Ventilation Equipment; and
2. ANSI/ASHRAE 84-1991, Method of  Testing Air-to-Air Heat Exchangers

• Recommended maximum air duct velocity to be 1,200 fpm (6 m/s) and total external air static pressure drop should not exceed 1.5 in. w.g. (380 Pa) for each of the fresh air and exhaust air ductwork. It is recommended to use the static regain method for duct sizing.
• Recommended maintenance spacing between wheel, coils, heat pipe and fans is 1.7 ft to 2 ft (500 to 600 mm).
• The energy wheel edges should be protected with an epoxy coating (or equivalent) to climate edge corrosion.
• Proper filtration should be provided at the fresh air intake of the wheel as well as the exhaust air intake for proper wheel operation and for reducing the need for frequent cleaning and maintenance.
• The wheel purge system should be field adjusted to get the design purge air based on actual field differential air pressure between fresh air and exhaust airstream. The consultant or designer should specify air pressure taps extended to the unit’s outer panel to allow measurement of differential air pressure between upstream fresh air and upstream exhaust air of the wheel.
• An optional speed detector with alarm function and interface to the building management system is recommended to guard against motor or belt failure.

Controls for Fresh Air-Handling Unit
Necessary Controls

• It is necessary to control the leaving air temperature from the cooling coil to a dry-bulb temperature of 56.3°F (13.5°C), which corresponds to the absolute humidity level of the comfort indoor condition of 76°F (24.4°C) and 50% RH.
• Fresh air fan motors start/stop, exhaust air fan start/stop, total energy wheel motor start/stop, sensible wheel motor start/stop or runaround pump start/stop with necessary electric protection and allowance for local as well as remote controls and communication.
• Ambient dew point (or grains transmitter) sensor to shut off the cooling motorized valve and enthalpy wheel motor when the ambient dew point is below 56°F (13.3°C).
  
  

Optional Control

• Should the occasional need arise to control the final leaving supply fresh air temperature (which, if the component is properly selected, should be achieved automatically), then the consultant or designer can opt to control the capacity via a variable speed drive of the sensible wheel motor, a solenoid valve for the heat pipe and a three-way bypass valve at the runaround coil.
• Variable speed drives for both fresh air and exhaust air can be adopted for variable occupancy applications such as office buildings, theaters, conference rooms, sport arenas, restaurants and others that are possible to control via CO2 sensors located in the exhaust air ducts.

Installation and Maintenance

• Adjust the purge to actual site conditions.
• Multipass labyrinth seals or adjustable brush seals are important elements for effective and efficient wheel operation to minimize leakage rate to a range of 0.05% to 0.2%. Seals require field adjustment.
• Wheel media cleaning can be done annually using vacuum or pressurized air (hot water is accepted by some manufacturers). The wheel is designed for laminar flow and resists plugging and accumulation of dust particles because of the back-flushing done by the incoming and outgoing of airstreams, which help minimize the need for frequent wheel cleaning.
• Other wheel components requiring routine maintenance involve bearing lubrication, verifying bearing bolt and sheave tightness, belt condition, rotor runout and flatness, media tightness, etc.
• Coils and heat pipe require minimum maintenance such as scheduled cleaning and runaround pump maintenance.

References

1. Berbari, G.J. 1998.  “Fresh air treatment in hot and humid climates.”  ASHRAE Journal 40(10):64-70

Bibliography

        2001 ASHRAE Handbook-Fundamentals
        Selection and Pricing Software from SEMCO Inc. Heat Pipe and SPC Inc. and Bry-Air.